Learn System Curve on Page 4

CENTRIFUGAL PUMPS HANDBOOK Table of Contents Nomenclature and Definitions....................................................................................1 Centrifugal and Positive Displacement Pumps in the Operating System.................4 Cavitation and NPSH in Centrifugal Pumps..............................................................6 Pump Suction Conditions .........................................................................................10 Elements of Minimum Flow .....................................................................................13 Effects of Oversizing ..................................................................................................17 Fluid Viscosity Effects on Centrifugal Pumps .........................................................20 Pump Balancing Criteria...........................................................................................24 Bearing Basics ............................................................................................................26 Centrifugal Pump Efficiency.....................................................................................30 Motor Size Selection for Centrifugal Pumps ............................................................32 Setting the Minimum Flows for Centrifugal Pumps ...............................................34 Estimating Maximum Head in Single-and Multi-Stage Pump Systems ..................37 Tips on Pump Efficiency ...........................................................................................39 Examining Pump Capacity Problems.......................................................................42 Venting Pump Systems ..............................................................................................44 Installation and Start-Up Troubleshooting ...............................................................46 Upgrading Utility and Process Pumps......................................................................50 Variable Speed Pumping ...........................................................................................55 Self-Priming Centrifugal Pumps ...............................................................................60 Centrifugal Pump Testing..........................................................................................64 The Canned Motor vs. Mag Drive Debate ................................................................68 National Electric Code Impact on Sealless Centrifugal Pumps ..............................70 Pumping Hydrofluoric Acid ......................................................................................75 User Perspective: When to Apply Mag Drive Pumps...............................................79 Interpreting Sealless Pump Failures .........................................................................81 Magnetic Couplings for Sealless Pumps ...................................................................83 Suction Side Problems – Gas Entrainment ..............................................................87

CENTRIFUGAL PUMPS HANDBOOK Table of Contents Nozzle Loading – Who Sets the Standards?..............................................................91 Low Flow Options......................................................................................................94 Pump Design Changes Improve Lubrication............................................................98 CPI Pumping............................................................................................................101 Pump Buying Strategies ..........................................................................................107 A Common Sense Approach to Combatting Corrosion and Abrasion ..................111 Recommendations For Vertical Pump Intakes.......................................................115 Hydraulic Instabilities and Cavitation ....................................................................118 High Speed/Low Flow Pumps: Top 10 Issues.........................................................125 Pump Ratings Vital When Pressure's On...............................................................131 Communicating Your Pump Needs ........................................................................136 Impellers and Volutes: Power with Control............................................................139 Fully Lined Slurry Pump for FCCU Bottoms Use .................................................142 Options for Sealless Centrifugals ............................................................................146 Tips for Selecting ANSI Process Pumps..................................................................148 ANSI Upgrades Require More Than Technology....................................................153 Selecting Mag-Drive Pumps.....................................................................................160 Operation Protection for Mag-Drive Pumps...........................................................166 Sealless Options Optimize Solutions.......................................................................170 Vertical Turbine Pumps Power Petrochemical .......................................................175 Chopper Pumps Digest the Solids ..........................................................................180 So You Need Pumps for a Revamp .........................................................................184 Pumping Hot Stuff–Another Perspective .....................................................190 The Power of Speed and Staging ..................................................................193 A Guide to ANSI Centrifugal Pump Design and Material Choices ..............198 Self-Priming Pumps: It’s in the System!.......................................................205 Know the Inside Story of Your Mag Drive Pumps .......................................209 Pump Rebuilding at Avon ............................................................................213 Evaluating Sealless Centrifugal Pump Design & Performance .....................222

CENTRIFUGAL PUMPS HANDBOOK Table of Contents Vertical Motor-Under Pumps Expand Their Range......................................227 Centrifugal Pump Suction ............................................................................231 Pumping Options for Low Flow/High Head Applications ............................234 Anti-Friction Bearings in Centrifugal Pumps ..............................................237 Centrifugal Pumps Operating in Parallel ....................................................248 Fire Pump Systems–Design and Specifications ............................................251 Well Pump Applications for Mine Dewatering ............................................256 Design Practices For Safe Handling of Hazardous Fluids in API Applications..260 Bearing Reliability in Centrifugal Pumps ............................................................265 Magnetic Drive Pumps Without the Hype ..........................................................271 Reliability Tips for Operating Magnetic Drive Pumps ........................................277 Canned Motor Pumps: Back to Basics ................................................................278 Selecting the Right Thermoplastic Sump Pump for Corrosive, Abrasive and Ultrapure Services ............................................................................................282 Successful Submersible Operation Part 1: Pump Installation and Start-Up ......288 Successful Submersible Operation Part 2: Inspection and Maintenance ..........295 Precision Solutions for Low-Flow Handling ........................................................299 Key Centrifugal Pump Parameters and How They Impact Your Applications– Part 1 ................................................................................................................304 Key Centrifugal Pump Parameters and How They Impact Your Applications– Part 2 ................................................................................................................310 Continuous Monitoring of Sealless Pumps ..........................................................315 Selecting Sealless Pumps and Circulation Systems for Difficult Pumping Applications ......................................................................................................324 Installation: The Foundation of Equipment Reliability......................................336

All materials © 2002 Pumps & Systems, LLC. No part of this publication may be reproduced without the written consent of the publisher. The publisher does not warrant, either expressly or by implication, the factual accuracy of the articles or descriptions herein, nor does the publisher warrant the accuracy of any views or opinions offered by the authors of said articles or descriptions.

CENTRIFUGAL PUMPS HANDBOOK

Nomenclature and Definitions BY PAT FLACH

V

apor pressure, cavitation, and NPSH are subjects widely discussed by engineers, pumps users, and pumping equipment suppliers, but understood by too few. To grasp these subjects, a basic explanation is required.

their parent liquid and boil at a lower temperature. While vapor pressure curves are readily available for liquids, they are not for solutions. Obtaining the correct vapor pressure for a solution often requires actual laboratory testing.

VAPOR PRESSURE

Cavitation can create havoc with pumps and pumping systems in the form of vibration and noise. Bearing failure, shaft breakage, pitting on the impeller, and mechanical seal leakage are some of the problems caused by cavitation. When a liquid boils in the suction line or suction nozzle of a pump, it is said to be “flashing” or “cavitating” (forming cavities of gas in the liquid). This occurs when the pressure acting on the liquid is below the vapor pressure of the liquid. The damage occurs when these cavities or bubbles pass to a higher pressure region of the pump, usually just past the vane tips at the impeller “eye,” and then collapse or “implode.”

CAVITATION

Knowledge of vapor pressure is extremely important when selecting pumps and their mechanical seals. Vapor pressure is the pressure absolute at which a liquid, at a given temperature, starts to boil or flash to a gas. Absolute pressure (psia) equals the gauge pressure (psig) plus atmospheric pressure. Let’s compare boiling water at sea level in Rhode Island to boiling water at an elevation of 14,110 feet on top of Pikes Peak in Colorado. Water boils at a lower temperature at altitude because the atmospheric pressure is lower. Water and water containing dissolved air will boil at different temperatures. This is because one is a liquid and the other is a solution. A solution is a liquid with dissolved air or other gases. Solutions have a higher vapor pressure than

NPSH

FIGURE 1

Water Sp. Gr. = 1.0

100 FEET STATIC HEAD

100 FEET STATIC HEAD

43 psi

32.5 psi

Gasoline Sp. Gr. = .75

Net Positive Suction Head is the difference between suction pressure and vapor pressure. In pump design and application jargon, NPSHA is the net positive suction head available to the pump, and NPSHR is the net positive suction head required by the pump. The NPSH A must be equal to or greater than the 100 FEET NPSHR for a pump STATIC to run properly. One HEAD way to determine the NPSH A is to measure the suction pressure at the suction 52 psi nozzle, then apply the following formula: SaltWater

Sp. Gr. = 1.2

Static head using various liquids.

NPSHA = PB – VP ± Gr + hv The Pump Handbook Series

where PB = barometric pressure in feet absolute, VP = vapor pressure of the liquid at maximum pumping temperature in feet absolute, Gr = gauge reading at the pump suction, in feet absolute (plus if the reading is above barometric pressure, minus if the reading is below the barometric pressure), and hv = velocity head in the suction pipe in feet absolute. NPSH R can only be determined during pump testing. To determine it, the test engineer must reduce the NPSHA to the pump at a given capacity until the pump cavitates. At this point the vibration levels on the pump and system rise, and it sounds like gravel is being pumped. More than one engineer has run for the emergency shut-down switch the first time he heard cavitation on the test floor. It’s during these tests that one gains a real appreciation for the damage that will occur if a pump is allowed to cavitate for a prolonged period.

CENTRIFUGAL PUMPING Centrifugal pumping terminology can be confusing. The following section addresses these terms and how they are used: Head is a term used to express pressure in both pump design and system design when analyzing static or dynamic conditions. This relationship is expressed as: head in feet =

(pressure in psi x 2.31) specific gravity

Pressure in static systems is referred to as static head and in a dynamic system as dynamic head. To explain static head, let’s consider three columns of any diameter, one filled with water, one with gasoline, and one with salt water (Figure 1). If the columns are 100 ft tall and you

1

measure the pressure at the bottom of each column, the pressures would be 43, 32.5, and 52 psi, respectively. This is because of the different specific gravities, or weights, of the three liquids. Remember, we are measuring pounds per square inch at the bottom of the column, not the total weight of the liquid in the column. The following four terms are used in defining pumping systems and are illustrated in Figure 2.

FIGURE 2

Total Static Head Static Discharge Head

Static Suction Head

Static Discharge Head

Total Static Head

Total static head is the vertical distance between the surface of the suction source liquid and the surface level of the discharge liquid. Static discharge head is the vertical distance from the centerline of the suction nozzle up to the surface level of the discharge liquid. Static suction head applies when the supply is above the pump. It is the vertical distance from the centerline of the suction nozzle up to the liquid surface of the suction supply. Static suction lift applies when the supply is located below the pump. It is the vertical distance from the centerline of the suction nozzle down to the surface of the suction supply liquid. Velocity, friction, and pressure head are used in conjunction with static heads to define dynamic heads. Velocity head is the energy in a liquid as a result of it traveling at some velocity V. It can be thought of as the vertical distance a liquid would need to fall to gain the same velocity as a liquid traveling in a pipe. This relationship is expressed as: hv = V2/2g where V = velocity of the liquid in feet per second and g = 32.2 ft/sec2. Friction head is the head needed to overcome resistance to liquid flowing in a system. This

2

Static Suction Lift

Total static head, static discharge head, static suction head, and static suction lift. resistance can come from pipe friction, valves, and fittings. Values in feet of liquid can be found in the Hydraulic Institute Pipe Friction Manual.

at a pump suction flange, converting it to head and correcting to the pump centerline, then adding the velocity head at the point of the gauge.

Pressure head is the pressure in feet of liquid in a tank or vessel on the suction or discharge side of a pump. It is important to convert this pressure into feet of liquid when analyzing systems so that all units are the same. If a vacuum exists and the value is known in inches of mercury, the equivalent feet of liquid can be calculated using the following formula:

Total dynamic discharge head is the static discharge head plus the velocity head at the pump discharge flange plus the total friction head in the discharge system. This can be determined in the field by taking the discharge pressure reading, converting it to head, and correcting it to the pump centerline, then adding the velocity head.

vacuum in feet =

in. of Hg x 1.13 specific gravity

When discussing how a pump performs in service, we use terms describing dynamic head. In other words, when a pump is running it is dynamic. Pumping systems are also dynamic when liquid is flowing through them, and they must be analyzed as such. To do this, the following four dynamic terms are used. Total dynamic suction head is the static suction head plus the velocity head at the suction flange minus the total friction head in the suction line. Total dynamic suction head is calculated by taking suction pressure The Pump Handbook Series

Total dynamic suction lift is the static suction lift minus the velocity head at the suction flange plus the total friction head in the suction line. To calculate total dynamic suction lift, take suction pressure at the pump suction flange, convert it to head and correct it to the pump centerline, then subtract the velocity head at the point of the gauge. Total dynamic head in a system is the total dynamic discharge head minus the total dynamic suction head when the suction supply is above the pump. When the suction supply is below the pump, the total dynamic head

is the total dynamic discharge head plus the total dynamic suction lift. Centrifugal pumps are dynamic machines that impart energy to liquids. This energy is imparted by changing the velocity of the liquid as it passes through the impeller. Most of this velocity energy is then converted into pressure energy (total dynamic head) as the liquid passes through the casing or diffuser. To predict the approximate total dynamic head of any centrifugal pump, we must go through two steps. First, the velocity at the outside diameter (o.d.) of the impeller is calculated using the following formula: v = (rpm x D)/229 where v = velocity at the periphery of the impeller in ft per second, D = o.d. of the impeller in inches, rpm = revolutions per minute of the impeller, and 229 = a constant. Second, because the velocity energy at the o.d. or periphery of the impeller is approximately equal to the total dynamic head developed by the pump, we continue by substituting v from above into the following equation: H = v2/2g where H = total dynamic head developed in ft, v = velocity at the o.d. of the impeller in ft/sec, and g = 32.2 ft/sec2. A centrifugal pump operating at a given speed and impeller diameter will raise liquid of any specific gravity or weight to a given height. Therefore, we always think in terms of feet of liquid rather than pressure when analyzing centrifugal pumps and their systems. ■ Patrick M. Flach is the western hemisphere Technical Services Manager for the Industrial Division of EG&G Sealol.

Pumping Terms Have you had a momentary (or continuing) problem with converting gallons per minute to cubic meters per second or liters per second? Join the crowd. Though the metric or SI system is probably used as the accepted system, more than English units, it still presents a problem to a lot of engineers. Authors are encouraged to use the English system. Following is a list of the common conversions from English to metric units. This is far from a complete list. It has been limited to conversions frequently found in solving hydraulic engineering problems as they relate to pumping systems.

PUMPING UNITS FLOW RATE (U.S.) gallons/min (gpm) x 3.785 = liters/min (L/min) (U.S.) gpm x 0.003785 = cubic meters/min (m3/min) cubic feet/sec (cfs) x 0.028 = cubic meters/sec (m3/s)

HEAD feet (ft) x 0.3048 = meters (m) pounds/square inch (psi) x 6,895 = Pascals (Pa)

POWER horsepower (Hp) x 0.746 = kilowatts (kW)

GRAVITATIONAL CONSTANT (g) 32.2 ft./s2 x 0.3048 = 9.81 meters/second2 (m/s2)

SPECIFIC WEIGHT lb/ft3 x 16.02 = kilogram/cubic meter (kg/m3)

VELOCITY (V) ft/s x 0.3048 = meters/second (m/s)

VELOCITY HEAD V2/2g (ft) x 0.3048 = meters (m)

SPECIFIC SPEED (Ns) (gpm–ft) x 0.15 = Ns(m3/min–m) Ns = N(rpm)[(gpm)0.5/(ft)0.75] J. Robert Krebs is President of Krebs Consulting Service. He serves on the Pumps and Systems Editorial Advisory Board.

TABLE 1. ENGLISH TO METRIC CONVERSION Basic Units Length Mass Force Pressure Time Gallon (US) Gallon (US)

Multiply English Feet Pound Pound Pound/Square In. (psi) Seconds Gallon Gallon

The Pump Handbook Series

x Factor x 0.3048 x 0.454 x 4.448 x 6,895 x1 x 0.003785 x 3.785

= Metric = Meter (m) = Kilogram (Kg) = Newton (N) = Pascal (Pa) = Seconds (s) = Meter Cubed (m3) = Liter (L)

3

CENTRIFUGAL PUMPS HANDBOOK

Centrifugal and Positive Displacement Pumps in the Operating System BY ROSS C. MACKAY

I

n the many differences that exist between centrifugal and positive displacement pumps, one which has caused some confusion is the manner in which they each operate within the system. Positive displacement pumps have a series of working cycles, each of which encloses a certain volume of fluid and moves it mechanically through the pump into the system, regardless of the back pressure on the pump. While the maximum pressure developed is limited only by the mechanical strength of the pump and system and by the driving power available, the effect of that pressure can be controlled by a pressure relief or safety valve. A major advantage of the positive displacement pump is that it can deliver consistent capacities because the output is solely dependent on the basic design of the pump and the speed of its driving mechanism. This means that, if the required flow rate is not moving through the system, the situation can always be corrected by changing one or both of these factors. This is not the case with the centrifugal pump, which can only react to the pressure demand of the system. If the back pressure on a centrifugal pump changes, so will its capacity. This can be disruptive for any process dependent on a specific flow rate, and it can diminish the operational stability, hydraulic efficiency and mechanical reliability of the pump.

CENTRIFUGAL PUMP PERFORMANCE CURVE The interdependency of the system and the centrifugal pump can be easily explained with the use of the pump performance curve and the system curve. A centrifugal pump performance curve is a well known shape which shows that the pressure the pump

4

can develop is reduced as the capacity increases. Conversely, as the capacity drops, the pressure it can achieve is gradually increased until it reaches a maximum where no liquid can pass through the pump. Since this is usually a relatively low pressure, it is rarely necessary to install a pressure relief or safety valve. When discussing the pressures developed by a centrifugal pump, we use the equivalent linear measurement referred to as “head,” which allows the pump curve to apply equally to liquids of different densities. [Head (in feet)=Pressure (in p.s.i.) x 2.31+ Specific Gravity of the liquid]

When the pump curve is superimposed on the system curve, the point of intersection represents the conditions (H,Q) at which the pump will operate. Pump Curve H

System Curve Q O

Pumping conditions change ONLY through an alteration in either the pump curve or the sysSYSTEM CURVE tem curve. When considering possible The system curve represents the movements in these curves, it pressures needed at different flow rates should be noted that there are only to move the product through the sysa few conditions which will cause tem. To simplify a comparison with the pump curve to change its posithe centrifugal pump curve, we again tion and shape: use the ‘head’ measurement. The sys• wear of the impeller tem head consists of three factors: • change in rotational speed • static head, or the vertical eleva• change in impeller diameter tion through which the liquid • change in liquid viscosity must be lifted Since these conditions don’t nor• friction head, or the head required mally develop quickly, any sudden to overcome the friction losses in change in pumping conditions is the pipe, the valves and all the fitlikely to be a result of a movement tings and equipment in the system curve, which means • velocity head, which is the head something in the system has required to accelerate the flow of changed. liquid through the pump (Velocity Since there are only three ingrehead is generally quite small and dients in a system curve, one of often ignored.) which is minimal, it follows that As the static head does not vary either the static head or the friction simply because of a change in flow head must have changed for any rate, the graph would show a straight movement to take place in the sysline. However, both the friction tem curve. head and the A change in the static velocity head head is normally a result of will always a change in tank level. If vary direct- Head the pump is emptying a System ly with the Friction & Curve tank and discharging at a capacity. The Velocity Head fixed elevation, the static combination head against which the Static Head of all three pump must operate will be creates the Capacity gradually increasing as the system curve. The Pump Handbook Series

suction tank empties. This will cause the system curve to move upwards as shown.

An increase in friction head can be caused by a wide variety of conditions such as the change in a valve setting or build-up of solids in a strainer. This will give the system curve a new slope.

When the operating conditions of a system fitted with a centrifugal pump change, it is helpful to consider these curves, focus on how the system is controlling the operation of the pump, and then control the system in the appropriate way. ■ Ross C. Mackay is an independent consultant located in Tottenham, Ontario, Canada. He is the author of several papers on the practical aspects of pump maintenance, and a specialist in helping companies reduce their pump maintenance costs.

Both sets of events produce the same result: a reduction of flow through the system. If the flow is redirected to a different location (such as in a tank farm), it means that the pump is now operating on an entirely new system which will have a completely different curve.

Thus, it is clear that regardless of the rated capacity of the centrifugal pump, it will only provide what the system requires. It is important to understand the conditions under which system changes occur, the acceptability of the new operating point on the pump curve, and the manner in which it can be moved.

The Pump Handbook Series

5

CENTRIFUGAL PUMPS HANDBOOK

Cavitation and NPSH in Centrifugal Pumps BY PAUL T. LAHR

C

FIGURE 1 985

800

800

O

N

O RB

300

UR

L HY

NE PA O A PR NI O M M A

40

ET

RM

L

Y TH

E

E

R

O

L CH

LO

NE

ET

M

H

E ID

1

NE

LE

U TH

E

R.

AT

70

F=

15" 20"

G

P.

LO

CH

I TR

R TE

(S

22.5"

A

W VY

25"

A

HE

E

26" 27"

BO

N

HY

TE

LE N

(T

0 .1

AR

ET

0 5 10"

6

RO

C

( TR CIS ) AC H LO R

O R LO H IC RO

28"

ER

O

RM

O OF

R

O HL

YL

H

ET

M

O

AC

E YL

HL

C RI

NE

ET

CH

10 5 2

H

E

E

HY

E AN

O OR

D RI

ET

20 14

)

Y TH

E ID

EN E

E

L

30

ER

H ET

AT

L

FO

DI

M

AN

T BU

D

28.5"

W

.60 .50

50

M

CH L

1.0 .80

60

E AT

3 2

80

O DI

C

ET

6 5 4

100

SU

O HL

IS

10 8

140

LF

E

D RI

N

20

200

XI

DI

1. Mechanical damage occurs as the imploding bubbles remove segments of impeller material. 2. Noise and vibration result from the implosion. Noise that sounds like gravel being pumped is often the user’s first warning of cavitation.

30

300

E

FID

DE

TA

60 50 40

NE

A TH

M

BU

Flow is reduced as the liquid is displaced by vapor, and mechanical imbalance occurs as the impeller passages are partially filled with lighter vapors. This results in vibration and shaft deflection, eventually resulting in bearing failures, packing or seal leakage, and shaft breakage. In multi-stage pumps this can cause loss of thrust balance and thrust bearing failures.

400

O

TR

NI

L NE ME SU HA RO EN ET LUO G O IF DR TR E NE RI HY RO EN LO LO YL H CH C OP R O P ON

O

80

ABSOLUTE PRESSURE–LBS. PER SQ. IN.

BUBBLE FORMATION PHASE

DI

O US

CA

200

100

600 500

E

XID

E XID

29" 29.1" 29.2" 29.3"

.40 .30

29.4"

.20

29.5" 29.6"

NET POSITIVE SUCTION HEAD

29.7" 29.72"

.10

When designing a pumping 60 30 0 30 60 90 120 150 180 system and selecting a pump, one TEMPERATURE–F must thoroughly evaluate net positive suction head (NPSH) to pre- Vapor pressures of various liquids related to temperature. vent cavitation. A proper analysis

6

The Pump Handbook Series

210

240

VACUUM–INCHES OF MERCURY

600 500 400

GAUGE PRESSURE–LBS. PER SQ. IN.

-60° to 240°F 1000

CAVITATION EFFECTS

BUBBLE COLLAPSE PHASE

friction in the suction pipe is a common negative component of NPSHA, the value of NPSHA will always decrease with flow. NPSHA must be calculated to a stated reference elevation, such as the foundation on which the pump is to be mounted. NPSHR is always referenced to the pump impeller center line.

involves both the net positive suction heads available in the system (NPSHA) and the net positive suction head required by the pump (NPSHR). NPSHA is the measurement or calculation of the absolute pressure above the vapor pressure at the pump suction flange. Figure 2 illustrates methods of calculating NPSHA for various suction systems. Since

C

avitation is the formation and collapse of vapor bubbles in a liquid. Bubble formation occurs at a point where the pressure is less than the vapor pressure, and bubble collapse or implosion occurs at a point where the pressure is increased to the vapor pressure. Figure 1 shows vapor pressure temperature characteristics. This phenomenon can also occur with ship propellers and in other hydraulic systems such as bypass orifices and throttle valves—situations where an increase in velocity with resulting decrease in pressure can reduce pressure below the liquid vapor pressure.

FIGURE 2 4b SUCTION SUPPLY OPEN TO ATMOSPHERE -with Suction Head

4a SUCTION SUPPLY OPEN TO ATMOSPHERE -with Suction Lift

CL

PB

NPSHA=PB + LH – (VP + ht) NPSHA=PB – (VP + Ls + ht)

PB

CL

4d CLOSED SUCTION SUPPLY -with Suction Lift

4c CLOSED SUCTION SUPPLY -with Suction Lift

p

CL

NPSHA=p + LH – (VP + ht) NPSHA=p – (Ls + VP + ht)

CL

p

Calculation of system net positive suction head available (NPSHA) for typical suction conditions. PB = barometric pressure in feet absolute, VP = vapor pressure of the liquid at maximum pumping temperature in feet absolute, p = pressure on surface of liquid in closed suction tank in feet absolute, Ls = maximum suction lift in feet, LH = minimum static suction head in feet, hf = friction loss in feet in suction pipe at required capacity.

FIGURE 3 ENTRANCE LOSS

FRICTION

TURBULANCE FRICTION INCREASING ENTRANCE PRESSURE LOSS AT DUE TO VANE TIPS IMPELLER

E

B C

POINT OF LOWEST PRESSURE WHERE VAPORIZATION STARTS

D

A

INCREASE PRESSURE

It is a measure of the pressure drop as the liquid travels from the pump suction flange along the inlet to the pump impeller. This loss is due primarily to friction and turbulence. Turbulence loss is extremely high at low flow and then decreases with flow to the best efficiency point. Friction loss increases with increased flow. As a result, the internal pump losses will be high at low flow, dropping at generally 20–30% of the best efficiency flow, then increasing with flow. The complex subject of turbulence and NPSHR at low flow is best left to another discussion. Figure 3 shows the pressure profile across a typical pump at a fixed flow condition. The pressure decrease from point B to point D is the NPSHR for the pump at the stated flow. The pump manufacturer determines the actual NPSHR for each pump over its complete operating range by a series of tests. The detailed test procedure is described in the Hydraulic Institute Test Standard 1988 Centrifugal Pumps 1.6. Industry has agreed on a 3% head reduction at constant flow as the standard value to establish NPSHR. Figure 4 shows typical results of a series of NPSHR tests. The pump system designer must understand that the published NPSHR data established above are based on a 3% head reduction. Under these conditions the pump is cavitating. At the normal operating point the NPSHA must exceed the NPSHR by a sufficient margin to eliminate the 3% head drop and the resulting cavitation. The NPSHA margin required will vary with pump design and other factors, and the exact margin cannot be precisely predicted. For most applications the NPSHA will exceed the NPSHR by a significant amount, and the NPSH margin is not a consideration. For those applications where the NPSH A is close to the NPSH R

A

B

C

D

E

POINTS ALONG LIQUID PATH RELATIVE PRESSURES IN THE ENTRANCE SECTION OF A PUMP

The pressure profile across a typical pump at a fixed flow condition.

The Pump Handbook Series

7

specific speed by substiFIGURE 4 tuting design flow rate and the system designer’s NPSHA. The pump speed N is generally determined Q1 by the head or pressure Q2 required in the system. For a low-maintenance 100% CAP Q3 pump system, designers and most user specificaQ4 3% tions require, or prefer, Ss values below 10,000 to 12,000. However, as indicated above, the pump Ss is dictated to a great extent by the system conNPSH ditions, design flow, head, and the NPSHA. Figures 5 and 6 are Typical results of a four-point net posiplots of Ss versus flow in tive suction head required (NPSHR) test gpm for various NPSHA based on a 3% head drop. or NPSH R at 3,500 and 1,750 rpm. Similar plots gpm if the maximum Ss is to be can be made for other common maintained at 12,000. Various pump speeds. options are available, such as Using curves from Figure 5 and reducing the head to allow 1,750 Figure 6 allows the system designer rpm (Figure 7). This would allow to design the system Ss, i.e., for a sysflows to 4,000 gpm with 20 feet of tem requiring a 3,500 rpm pump NPSHA. with 20 feet of NPSHA, the maximum flow must be limited to 1,000

SUCTION SPECIFIC SPEED The concept of suction specific speed (Ss) must be considered by the pump designer, pump application engineer, and the system designer to ensure a cavitation-free pump with high reliability and the ability to operate over a wide flow range. N x Q0.5 Ss = —————— (NPSHR)0.75 where

N = pump rpm Q = flow rate in gpm at the best efficiency point NPSHR = NPSHR at Q with the maximum impeller diameter

The system designer should also calculate the system suction

NPSHR

TOTAL HEAD

(2–3 feet), users should consult the pump manufacturer and the two should agree on a suitable NPSH margin. In these deliberations, factors such as liquid characteristic, minimum and normal NPSH A, and normal operating flow must be considered.

3 FIGURE 5

2

2

V=

S, Suction specific speed

HS

1 9 8

3

4

7

5

6

6 7

5

8

4

9

24

V=

HS

10

12 V= HS 14

16

20

18

28 32 36

50 55

45

V=

HS

40

60

65

3

Solution for S=N

2

Q Hsv0.75

for N=3,500 rpm 1 1

2

3

4

5

6

7 8 9 1

2

3

4

5

6

7

8 9 1

2

3

4

5

6

7

8

9 1

Q, Capacity, gpm

A plot of suction specific speed (Ss) versus flow in gallons per minute (gpm) for various NPSHA or NPSHR at 3,500 rpm. (Single suction pumps. For double suction use 1/2 capacity). Hsv=NPSHR at BEP with maximum impeller diameter.

8

The Pump Handbook Series

5

FIGURE 6 4 3

S, Suction specific speed

2

1 9 8 7

V=1

HS

6

2

5 3

4

4

3

14

7 8 9

16

50

HS

HS

6

2

5

V=4

2

V=1

5

18

28

4

V=2

HS

20

32

40

36

Solution for

10

S=N

Q Hsv0.75

for N=1,750 rpm 1 1

2

3

4

5

6

7

8

9 1

2

3

4

5

6

7

8 9 1

2

3

4

5

6

7

8

9 1

Q, Capacity, gpm

A plot of suction specific speed (Ss) versus flow in gallons per minute (gpm) for various NPSHA or NPSHR at 1,750 rpm. (Single suction pumps. For double suction use 1/2 capacity.) HSV=NPSHR at BEP with the maximum impeller diameter.

FIGURE 7

2 3

NPSH - FEET

HEAD

1

4

C

GPM

A

B

A typical plot of the suction and discharge systems. Curve 1 = pump head capacity performance, curve 2 = total system curve, curve 3 = suction system curve NPSHA, and curve 4 = pump NPSHR.

It is important for the pump user to understand how critical the system design requirements are to the selection of a reliable, trouble-free pump. Matching the system and pump characteristics is a must. Frequently, more attention is paid to the discharge side. Yet it is well known that most pump performance problems are caused by problems on the suction side. Figure 7 is a typical plot of the suction and discharge systems. It is important that points A, B, and C be well established and understood. A is the normal operating point. B is the maximum flow for cavitation-free operation. C is the minimum stable flow, which is dictated by the suction specific speed.

The Pump Handbook Series

As a general rule, the higher the suction specific speed, the higher the minimum stable flow capacity will be. If a pump is always operated at its best efficiency point, a high value of Ss will not create problems. However, if the pump is to be operated at reduced flow, then the Ss value must be given careful consideration. ■

REFERENCES 1. Goulds Pump Manual. 2. Durco Pump Engineering Manual. 3. Hydraulic Institute Test Standards—1988 Centrifugal Pumps 1.6. Paul T. Lahr is the owner of Pump Technology, a consulting firm. He serves on the Pumps and Systems Editorial Advisory Board.

9

CENTRIFUGAL PUMPS HANDBOOK

Pump Suction Conditions BY ROSS C. MACKAY f a wide receiver has the right a function of the system design on speed and good hands, all that’s the suction side of the pump. needed from the quarterback is Consequently, it is in the control of to throw the ball accurately, the system designer. and the team will probably gain To avoid cavitation, the NPSH good yardage, maybe even a available from the system must be touchdown. greater than the NPSH required by Believe it or not, much the the pump, and the biggest mistake same is true of a pump and its sucthat can be made by a system designtion conditions. If it has the right er is to succumb to the temptation to speed and is the right size, all provide only the minimum required that’s required from the quarterat the rated design point. This leaves back is to deliver the liquid at the no margin for error on the part of the right pressure and with an even designer, or the pump, or the system. laminar flow into the eye of the Giving in to this temptation has impeller. proved to be a costly mistake on If the quarterback’s pass is off many occasions. target, badly timed, or the ball’s In the simple system as shown turning end over end in the air, in Figure 1, the NPSH Available can the receiver may not be able to be calculated as follows: hang on to it, and there’s no gain on the play. When that happens, the quarterback FIGURE 1 knows he didn’t throw it Ha properly and doesn’t blame the receiver. Unfortunately, that’s where the comparison ends. The engineering Hvp ”quarterbacks” tend to Hs blame the pump even when its their delivery that’s bad! Hf Just as there are techniques a quarterback must learn in order to throw accurately, there are rules which ensure that a liquid arrives at the impeller eye with NPSHA = Ha + Hs - Hvp - Hf the pressure and flow characteristics needed for reliable operation. where RULE #1. Ha= the head on the surface of the PROVIDE SUFFICIENT NPSH liquid in the tank. In an open Without getting too complicatsystem like this, it will be ed on a subject about which comatmospheric pressure. plete books have been written, Hs= the vertical distance of the let’s just accept the premise that free surface of the liquid every impeller requires a miniabove the center line of the mum amount of pressure energy pump impeller. If the liquid is in the liquid being supplied in below the pump, this order to perform without cavitabecomes a negative value. tion difficulties. This pressure Hvp= the vapor pressure of the liqenergy is referred to as Net uid at the pumping temperaPositive Suction Head Required. ture, expressed in feet of The NPSH Available is suphead. plied from the system. It is solely

I

10

The Pump Handbook Series

Hf=

the friction losses in the suction piping. The NPSH Available may also be determined with this equation:

NPSHA= Ha + Hg + V2/2g - Hvp where Ha= atmospheric pressure in feet of head. Hg= the gauge pressure at the suction flange in feet of head. V2= The velocity head at the point of measurement of 2g Hg. (Gauge readings do not include velocity head.)

RULE #2. REDUCE THE FRICTION LOSSES When a pump is taking its suction from a tank, it should be located as close to the tank as possible in order to reduce the effect of friction losses on the NPSH Available. Yet the pump must be far enough away from the tank to ensure that correct piping practice can be followed. Pipe friction can usually be reduced by using a larger diameter line to limit the linear velocity to a level appropriate to the particular liquid being pumped. Many industries work with a maximum velocity of about 5ft./sec., but this is not always acceptable.

RULE #3. NO ELBOWS ON THE SUCTION FLANGE Much discussion has taken place over the acceptable configuration of an elbow on the suction flange of a pump. Let’s simplify it. There isn’t one! There is always an uneven flow in an elbow, and when one is installed on the suction of any pump, it introduces that uneven flow into the eye of the impeller. This can create turbulence and air

entrainment, which may result in impeller damage and vibration. When the elbow is installed in a horizontal plane on the inlet of a double suction pump, uneven flows are introduced into the opposing eyes of the impeller, upsetting the hydraulic balance of the rotating element. Under these conditions the overloaded bearing will fail prematurely and regularly if the pump is packed. If the pump is fitted with mechanical seals, the seal will usually fail instead of the bearing-but just as regularly and often more frequently. The only thing worse than one elbow on the suction of a pump is two elbows on the suction of a pump— particularly if they are positioned in planes at right angles to each other. This creates a spinning effect in the liquid which is carried into the impeller and causes turbulence, inefficiency and vibration. A well established and effective method of ensuring a laminar flow to the eye of the impeller is to provide the suction of the pump with a straight run

FIGURE 2

of pipe in a length equivalent to 5-10 times the diameter of that pipe. The smaller multiplier would be used on the larger pipe diameters and vice versa.

FIGURE 3 Air Pocket

RULE #4. STOP AIR OR VAPOR ENTERING THE SUCTION LINE Any high spot in the suction line can become filled tices are more difficult to trouwith air or vapor which, if transbleshoot in a closed tank simply ported into the impeller, will create because they can’t be seen as an effect similar to cavitation and easily. with the same results. Services Great care should be taken which are particularly susceptible in designing a sump to ensure to this situation are those where the that any liquid emptying into it pumpage contains a significant does so in such a way that air amount of entrained air or vapor, entrained in the inflow does not as well as those operating on a sucpass into the suction opening. tion lift, where it can also cause the Any problem of this nature may pump to lose its prime. (Figure 3) A similar effect can be caused by a concentric FIGURE 4 reducer. The suction of a pump should be fitted with an eccentric reducer positioned with the flat side uppermost. (Figure 4). If a pump is taking its suction from a sump or tank, the forrequire a change in the relative mation of vortices can positions of the inflow and outlet draw air into the sucif the sump is large enough, or tion line. This can usuthe use of baffles. (Figure 5) ally be prevented by providing sufficient RULE #5. submergence of liquid CORRECT PIPING ALIGNMENT over the suction openPiping flanges must be accuing. A bell-mouth design rately aligned before the bolts on the opening will are tightened and all piping, reduce the amount of valves and associated fittings submergence required. Suction should be independently supThis submergence is ported, so as to place no strain completely independent on the pump. Stress imposed on of the NPSH required by the pump casing by the piping the pump. reduces the probability of satisIt is worthwhile factory performance. noting that these vor-

The Pump Handbook Series

11

RULE #6. WHEN RULES 1 TO 5 HAVE BEEN IGNORED, FOLLOW RULES 1 TO 5.

FIGURE 5 Inflow

To Pump Suction

Inflow

To Pump Suction

Under certain conditions the pump manufacturer may identify some maximum levels of forces and moments which may be acceptable on the pump flanges. In high temperature applications, some piping misalignment is inevitable owing to thermal growth during the operating cycle. Under these conditions, thermal expansion joints are often introduced to avoid transmitting piping strains to the pump. However, if the end of the expansion joint closest to the pump is not anchored securely, the object of the exercise is defeated as the piping strains are simply passed through to the pump.

12

Piping design is one area where the basic principles in-volved are regularly ignored, resulting in hydraulic instabilities in the impeller which translate into additional shaft loading, higher vibration levels and premature failure of the seal or bearings. Because there are many other reasons why pumps could vibrate, and why seals and bearings fail, the trouble is rarely traced to incorrect piping. It has been argued that because many pumps are piped incorrectly and most of them are operating quite satisfactorily, piping procedure is not important. Unfortunately, satisfactory operation is a relative term, and what may be acceptable in one plant may be inappropriate in another. Even when ”satisfactory” pump operation is obtained, that

Baffle

The Pump Handbook Series

doesn’t automatically make a questionable piping practice correct. It merely makes it lucky. The suction side of a pump is much more important than the piping on the discharge. If any mistakes are made on the discharge side, they can usually be compensated for by increasing the performance capability from the pump. Problems on the suction side, however, can be the source of ongoing and expensive difficulties which may never be traced back to that area. In other words, if your receivers aren’t performing well, is it their fault? Or does the quarterback need more training? ■

Ross C. Mackay is an independent consultant who specializes in advanced technology training for pump maintenance cost reduction. He also serves on the editorial advisory board for Pumps and Systems.

CENTRIFUGAL PUMPS HANDBOOK

Elements of Minimum Flow BY TERRY M. WOLD

M

inimum flow can be determined by examining each of the factors that affect it. There are five elements that can be quantified and evaluated: 1.

Temperature rise (minimum thermal flow)

2.

Minimum stable flow

3.

Thrust capacity

4.

NPSH requirements

5.

Recirculation

mechanical handbooks. What is the maximum allowable temperature rise? Pump manufacturers usually limit it to 15°F. However, this can be disastrous in certain situations. A comparison of the vapor pressure to the lowest expected suction pressure plus NPSH required (NPSHR) by the pump must be made. The temperature where the vapor pressure equals the suction pressure plus the NPSHR is the maximum allowable

temperature. The difference between the allowable temperature and the temperature at the pump inlet is the maximum allowable temperature rise. Knowing ∆T and C p , the minimum flow can be determined by finding the corresponding head and efficiency. When calculating the maximum allowable temperature rise, look at the pump geometry. For instance, examine the vertical can

FIGURE 1

The highest flow calculated using these parameters is considered the minimum flow.

TEMPERATURE RISE Temperature rise comes from energy imparted to the liquid through hydraulic and mechanical losses within the pump. These losses are converted to heat, which can be assumed to be entirely absorbed by the liquid pumped. Based on this assumption, temperature rise ∆T in °F is expressed as:

H

1

778 x Cp

η–1

SUCTION Low Pressure Lower Temperature

DISCHARGE High Pressure Higher Temperature

∆T = ————— x ——————

where H = total head in feet Cp = specific heat of the liquid, Btu/lb x °F η = pump efficiency in decimal form 778 ft–lbs = energy to raise the temperature of one pound of water 1°F To calculate this, the specific heat and allowable temperature rise must be known. The specific heat for water is 1.0, and other specific heats can be as low as 0.5. The specific heats for a number of liquids can be found in many chemical and

A high-pressure vertical pump. Asterisks (*) denote where lowtemperature fluid is exposed to higher temperatures. Flashing and vaporization can occur here. Temperature increases as fluid travels from A towards B. The Pump Handbook Series

13

pump in Figure 1. Although pressure increases as the fluid is pumped upward through the stages, consider the pump inlet. The fluid at the inlet (low pressure, low temperature) is exposed to the temperature of the fluid in the discharge riser in the head (higher pressure, higher temperature). This means that the vapor pressure of the fluid at the pump inlet must be high enough to accommodate the total temperature rise through all the stages. If this condition is discovered during the pump design phase, a thermal barrier can be designed to reduce the temperature that the inlet fluid is exposed to. Some books, such as the Pump Handbook (Ref. 5), contain a typical chart based on water (Cp = 1.0) that can be used with the manufacturer’s performance curve to determine temperature rise. If the maximum allowable temperature rise exceeds the previously determined allowable temperature rise, a heat shield can be designed and fitted to the pump during the design stage. This requirement must be recognized during the design stage, because once the pump is built, options for retrofitting the pump with a heat shield are greatly reduced.

MINIMUM STABLE FLOW Minimum stable flow can be defined as the flow corresponding to the head that equals shutoff head. In other words, outside the ”droop“ in the head capacity curve. In general, pumps with a specific speed less than 1,000 that are designed for optimum efficiency have a drooping curve. Getting rid of this ”hump“ requires an impeller redesign; however, note that there will be a loss of efficiency and an increase in NPSHR. What’s wrong with a drooping head/capacity curve? A drooping curve has corresponding heads for two different flows. The pump reacts to the system requirements, and there are two flows where the pump can meet the system requirements. As a result, it ”hunts“ or ”shuttles“ between these two flows. This can damage the pump and other equipment, but it will happen only under certain circumstances:

14

1.

The liquid pumped must be uninhibited at both the suction and discharge vessels.

2.

One element in the system must be able to store and return energy, i.e., a water column or trapped gas.

3.

Something must upset the system to make it start hunting, i.e., starting another pump in parallel or throttling a valve.

FIGURE 2

Note: All of these must be present at the same time to cause the pump to hunt. Minimum flow based on the shape of the performance curve is not so much a function of the pump as it is a function of the system where the pump is placed. A pump in a system where the above criteria are present Recirculation zones are always on the presshould not have a droop- sure side of the vane. A shows discharge ing curve in the zone of recirculation (the front shroud has been left out for clarity). B shows inlet recirculation. operation. Because pumps with a drooping head/capacity curve have higher effitistage) with integral bearings. These ciency and a lower operating cost, it bearings can be sized to handle the would seem prudent to investigate the thrust. Thrust can be balanced by the installation of a minimum flow bypass. use of balanced and unbalanced stages or adding a balance drum, if THRUST LOADING necessary. These techniques for Axial thrust in a vertical turbine thrust balancing are used when high pump increases rapidly as flows are thrust motors are not available. It is reduced and head increased. Based on worth noting that balanced stages the limitations of the driver bearings, incorporate wear rings and balance flow must be maintained at a value holes to achieve lower thrust; therewhere thrust developed by the pump fore, a slight reduction in pump effidoes not impair bearing life. Find out ciency can be expected, and energy what your bearing life is and ask the costs become a factor. pump manufacturer to give specific thrust values based on actual tests. NPSH REQUIREMENTS If a problem exists that cannot be How many pumps have been handled by the driver bearings, conoversized because of NPSH available tact the pump manufacturer. There (NPSHA)? It seems the easiest soluare many designs available today for tion to an NPSH problem is to go to vertical pumps (both single and multhe next size pump with a larger sucThe Pump Handbook Series

impeller design. The problem is the result of a mismatched case and impeller, too little vane overlap in the impeller design, or trimming the impeller below the minimum diameter for which it was designed. Recirculation is one of the most difficult problems to understand and document. Many studies on the topic have been done over the years. Mr. Fraser’s paper (Ref. 1) is one of the most useful tools for determining where recirculation begins. In it he describes how to calculate the inception of recirculation based on specific design characteristics of the impeller and he includes charts that can be used with a minimum amount of information. An example of Fraser calculations, which show the requirements to calculate the inception of suction and discharge recirculation, is shown in Figure 3.

FIGURE 3 B2

R2

D2 B1

R1

D1 h1

.14 .12 .10

Cm2 U2 .08

Ve U1

.06 .04 .02 10

15

20

25

30

Discharge Angle β2

.32 .30 .28 .26 .24 .22 .20 .18 .16 .14 .12 .10 .08

RECIRCULATION CALCULATIONS

10

15

20

25

30

Inlet Angle β1

35

40

Incipient recirculation. Minimum flow is approximately 50% of incipient flow, while minimum intermittent flow is approximately 25% of incipient flow. See text under “Recirculation Calculations” for details

tion, thereby reducing the inlet losses. A couple of factors become entangled when this is done. A larger pump means operating back on the pump curve. Minimum flow must be considered. Is the curve stable? What about temperature rise? If there is already an NPSH problem, an extra few degrees of temperature rise will not help the situation. The thrust and eye diameter will increase, possibly causing damaging recirculation. When trying to solve an NPSH problem, don’t take the easiest way out. Look at other options that may solve a long-term problem and reduce operating costs.

RECIRCULATION Every pump has a point where recirculation begins. But if this is the case, why don’t more pumps have problems?

Recirculation is caused by oversized flow channels that allow liquid to turn around or reverse flow while pumping is going on (Figure 2 shows recirculation zones). This reversal causes a vortex that attaches itself to the pressure side of the vane. If there is enough energy available and the velocities are high enough, damage will occur. Suction recirculation is reduced by lowering the peripheral velocity, which in turn increases NPSH. To avoid this it is better to recognize the problem in the design stage and opt for a lower-speed pump, two smaller pumps, or an increase in NPSHA. Discharge recirculation is caused by flow reversal and high velocities producing damaging vortices on the pressure side of the vane at the outlet (Figure 2). The solution to this problem lies in the The Pump Handbook Series

Figure 3 indicates the userdefined variables and charts required to make the Fraser calculations for minimum flow. Information to do the detailed calculations include: Q = capacity at the best efficiency point H = head at the best efficiency point NPSHR = net positive suction head required at the pump suction N = pump speed NS = pump specific speed NSS = suction specific speed Z = number of impeller vanes h1 = hub diameter (h1 = 0 for single suction pumps) D1 = impeller eye diameter D2 = impeller outside diameter B1 = impeller inlet width B2 = impeller outlet width R1 = impeller inlet radius R2 = impeller outlet radius F1 = impeller inlet area F2 = impeller outlet area β1 = impeller inlet angle β2 = impeller outlet angle The above information is obtained from the pump manufacturer curves or impeller design files. The impeller design values are usually considered proprietary information. KVe and KCm2 can be determined from the charts in Figure 3.

15

With all of the above information at hand, suction recirculation and the two modes of discharge recirculation can be determined. As previously mentioned, Fraser has some empirical charts at the end of his paper that can be used to estimate the minimum flow for recirculation. A few of the design factors of the impeller are still required. It is best to discuss recirculation with your pump manufacturer before purchasing a pump, in order to reduce the possibility of problems with your pump and system after installation and start-up.

is economical, efficient, and insures a trouble-free pump life. It takes a coordinated effort by the user and the manufacturer to come up with an optimum system for pump selection, design, and installation.

REFERENCES 1.

F.H. Fraser. Recirculation in centrifugal pumps. Presented at the ASME Winter Annual Meeting (1981).

2.

A.R. Budris. Sorting out flow recirculation problems. Machine Design (1989).

3.

J.J. Paugh. Head-vs-capacity characteristics of centrifugal pumps. Chemical Engineering (1984).

4.

I. Taylor. NPSH still pump application problem. The Oil and Gas Journal (1978).

SUMMARY Minimum flow can be accurately determined if the elements described above are reviewed by the user and the manufacturer. Neither has all the information to determine a minimum flow that

16

The Pump Handbook Series

5.

I.J. Karassik. Pump Handbook. McGraw-Hill (1986). ■

Terry Wold has been the engineering manager for Afton Pumps for the last four years. He has been involved in pump design for 25 years. Mr. Wold graduated from Lamar Tech in 1968 with a bachelor’s degree in mechanical engineering and is currently a registered engineer in the State of Texas. Thanks to P.J. Patel for his comments and assistance in preparing the graphics.

CENTRIFUGAL PUMPS HANDBOOK

Effects of Oversizing BY: IGOR J. KARASSIK

POWER CONSUMPTION After all, we are not primarily interested in efficiency; we are more interested in power consumption. Pumps are designed to convert mechanical energy from a driver into energy within a liquid. This energy within the liquid is needed to overcome friction losses, static pressure differences and elevation differences at the desired flow rate. Efficiency is nothing but the ratio between the hydraulic energy utilized by the process and the energy input to the pump driver. And without changing the ratio itself, if we find that we are assigning more energy to the process than is really necessary, we can reduce this to correspond to the true requirement and therefore reduce the power consumption of the pump. It is true that some capacity margin should always be included, mainly to reduce the wear of internal clearances which will, with time, reduce the effective pump capacity. How much margin to provide is a fairly complex question because the wear that will take place varies with the type of pump in question, the liquid handled, the severity of the service and a number of other variables. A centrifugal pump operating in a given system will deliver a capacity corresponding to the

FIGURE 1 H – Q Curve

Head

SystemHead Curve

Capacity

Pump H-Q curve superimposed on system-head curve

FIGURE 2 Head-Capacity

Head

O

intersection of its head-capacity curve with the systemhead curve, as long as the available NPSH is equal to or exceeds the required NPSH (Figure 1). To change this operating point in an existing installation requires changing either the headcapacity curve or the system-head curve, or both. The first can be accomplished by varying the speed of the pump (Figure 2), or its impeller diameter while the second requires altering the friction losses by throttling a valve in the pump discharge (Figure 3). In the majority of pump installations, the driver is a constant speed motor, and changing the system-head curve is used to change the pump capacity. Thus, if we have provided too much excess margin in the selection of the pump head-capacity curve, the pump will have to operate with considerable throttling to limit its delivery to the desired value. If, on the other hand, we permit the pump to operate unthrottled, which is more likely, the flow into the system will increase until that capacity is reached where

at Full Speed

(N1)

SystemHead Curve

Head-Capacity at Full Speed (N2) Head-Capac ity at Full Sp eed (N ) 3 H 3

H1 H2

} Friction Losses Static Pressur e or Head

Capacity

Q3

Q2

Q1max

Varying pump capacity by varying speed

FIGURE 3 Head-Capacity

at Constant Sp

eed

SystemHead Curve

H3 H2 H1

SystemHead Curve by Throttling Valve

} Friction Losses

Head

ne of the greatest sources of power waste is the practice of oversizing a pump by selecting design conditions with excessive margins in both capacity and total head. It is strange on occasion to encounter a great deal of attention being paid to a one-point difference in efficiency between two pumps while at the same time potential power savings are ignored through an overly conservative attitude in selecting the required conditions of service.

Static Pressur e or Head

Capacity

Q3

Q2

Q1max

Varying pump capacity by throttling

The Pump Handbook Series

17

FIGURE 4 240

B

H-Q 1800 R.P.M.

220

3

14 /4"Impeller

C

200

A

14"Impeller

D

180

% Efficiency

Feet Total Head

H-Q 1800 R.P.M.

160 140

90 Static Head

80

180

70 lle

r

lle

"Im

pe

60

Im 3 / 4" 14

4

4 3/ Q1

η−

140

η−

B.H.P.

Q1

4"I

160

pe

mp

ell

r

er

200

Q

η−

4

Q1

η−

120

50

er

ell

p "Im

40

100

30

80

20

60

10

0

1000

2000 Capacity in G.P.M.

3000

4000

Effect of oversizing a pump the system-head and head-capacity curves intersect.

EXAMPLE Let’s use a concrete example, for which the maximum required capacity is 2700 gpm, the static head is 115 ft and the total friction losses, assuming 15-year old pipe, are 60 ft. The total head required at 2700 gpm is therefore 175 ft. We can now construct a systemhead curve, which is shown on curve A, Figure 4. If we add a margin of about 10% to the required capacity and then, as is frequently done, we add some

18

margin to the total head above the system-head curve at this rated flow, we end up by selecting a pump for 3000 gpm and 200 ft. total head. The performance of such a pump, with a 14-3/4 in. impeller, is superimposed on the system-head curve A in Figure 4. The pump develops excess head at the maximum required capacity of 2700 gpm, and if we wish to operate at that capacity, this excess head will have to be throttled. Curve “B” is the system-head curve that will have to be created by throttling. If we operate at 3000 gpm, the pump will take 175 bhp, and we will have to drive it with a 200 hp motor. The Pump Handbook Series

If we operate it throttled at the required capacity of 2700 gpm, operating at the intersection of its head-capacity curve and curve B, the pump will require 165 bhp. The pump has been selected with too much margin. We can safely select a pump with a smaller impeller diameter, say 14 in., with a head-capacity curve as shown on Figure 4. It will intersect curve A at 2820 gpm, giving us about 4% margin in capacity, which is sufficient. To limit the flow to 2700 gpm, we will still have to throttle the pump slightly and our system head curve will become curve C. The power consumption at 2700 gpm will now be only 145 bhp instead of the 165 bhp required with our first overly conservative selection. This is a very respectable 12% saving in power consumption. Furthermore, we no longer need a 200 hp motor. A 150 hp motor will do quite well. The saving in capital expenditure is another bonus resulting from correct sizing. Our savings may actually be even greater. In many cases, the pump may be operated unthrottled, the capacity being permitted to run out to the intersection of the head-capacity curve and curve A. If this were the case, a pump with a 14-3/4 in. impeller would operate at approximately 3150 gpm and take 177 bhp. If a 14 in. impeller were to be used, the pump would operate at 2820 gpm and take 148 bhp. We could be saving more than 15% in power consumption. Tables 1 and 2 tabulate these savings. And our real margin of safety is actually even greater than I have indicated. Remember that the friction losses we used to construct the system-head curve A were based on losses through 15-year old piping. The losses through new piping are only 0.613 times the losses we have assumed. The system-head curve for new piping is that indicated as curve D in Figure 4. If the pump we had originally selected (with a 14-3/4 in. impeller) were to operate unthrottled, it would run at 3600 gpm and take

TABLE 1. COMPARISON OF PUMPS WITH 143/4 IN. AND 14IN. IMPELLERS, WITH THE SYSTEM THROTTLED Impeller Curve BHP Savings

Throttled to 2700 GPM 143/4" “B” 165

14" “C” 145 20 hp or 12.1%

TABLE 2. COMPARISON OF PUMPS WITH THE SYSTEM UNTHROTTLED Impeller GPM BHP Savings

Unthrottled, on Curve “A” 143/4" 3150 177

14" 2820 148 29 hp or 16.4%

C l e a r l y , important energy savings can be achieved if, at the time of the selection of the conditions of service, r e a s o n a b l e restraints are exercised to avoid incorporating excessive safety margins into the rated conditions of service.

EXISTING INSTALLATIONS

But what of existing installations in which the pump or pumps have excessive margins? Is it too late to 133/4" achieve these savings? Far from it! As 3100 a matter of fact, it is 147 possible to establish 40.5 hp more accurately the 21.6% true system-head curve by running a performance test once the pump has been installed and operated. A reasonable margin can then be selected and several choices become available to the user:

TABLE 3. EFFECT OF DIFFERENT SIZE IMPELLERS IN SYSTEM WITH NEW PIPE AND RESULTING SAVINGS NEW PIPE (UNTHROTTLED OPERATION, CURVE “D”) Impeller GPM BHP Savings

143/4" 3600 187.5

14" 3230 156.5 31 hp 16.5%

187.5 bhp. A pump with only a 14 in. impeller would intersect the system-head curve D at 3230 gpm and take 156.6 bhp, with a saving of 16.5%. As a matter of fact, we could even use a 13-3/4 in. impeller. The head-capacity curve would intersect curve D at 3100 gpm, and the pump would take 147 bhp. Now, the savings over using a 14-3/4 in. impeller becomes 21.6% (See Table 3).

1.

The existing impeller can be cut down to meet the more realistic conditions of service.

2.

A replacement impeller with the necessary reduced diameter can be ordered from the pump man-

The Pump Handbook Series

ufacturer. The original impeller is then stored for future use if friction losses are ultimately increased with time or if greater capacities are ever required. 3.

In certain cases, there may be two separate impeller designs available for the same pump, one of which is of narrower width than the one originally furnished. A replacement narrow impeller can then be ordered from the manufacturer. Such a narrower impeller will have its best efficiency at a lower capacity than the normal width impeller. It may or may not need to be of a smaller diameter than the original impeller, depending on the degree to which excessive margin had originally been provided. Again, the original impeller is put away for possible future use. ■

Igor J. Karassik is Senior Consulting Engineer for IngersollDresser Pump Company. He has been involved with the pump industry for more than 60 years. Mr. Karassik is Contributing Editor Centrifugal Pumps for Pumps and Systems Magazine.

19

CENTRIFUGAL PUMPS HANDBOOK

Fluid Viscosity Effects on Centrifugal Pumps BY: GUNNAR HOLE hen sizing a pump for a new application or evaluating the performance of an existing pump, it is often necessary to account for the effect of the pumped fluid’s viscosity. We are all aware that the head-capacity curves presented in pump vendor catalogs are prepared using water as the pumped fluid. These curves are adequate for use when the actual fluid that we are interested in pumping has a viscosity that is less than or equal to that of water. However, in some cases—certain crude oils, for example—this is not the case. Heavy crude oils can have viscosities high enough to increase the friction drag on a pump’s impellers significantly. The additional horsepower required to overcome this drag reduces the pump’s efficiency. There are several analytical and empirical approaches available to estimate the magnitude of this effect. Some of these are discussed below. Before beginning the discussion, however, it is vital to emphasize the importance of having an accurate viscosity number on which to base our estimates. The viscosity of most liquids is strongly influenced by temperature. This relationship is most often shown by plotting two points on a semilogarithmic grid and connecting them with a straight line. The relationship is of the form:

W

FIGURE 1

µ = AeB/T where µ = the absolute viscosity of the fluid A and B = constants T = the absolute temperature of the fluid Plotting this relationship requires knowledge of two data points, and using them effectively requires some judgement as to

20

Reproduced from the Hydraulic Institute Standards (Figure 71) the normal operating temperature as well as the minimum temperature that might be expected during other off-design conditions such as start-up. The Pump Handbook Series

The effect of pressure on the viscosity of most fluids is small. For mineral oils, for example, an increase of pressure of 33 bars (≈480 psi ) is equivalent to a tem-

NON-NEWTONIAN

FIGURE 2

These are fluids where the shear rate-shear stress relationship is nonlinear. They can be divided into four categories: • Bingham-plastic fluids are those in which there is no flow until a threshold shear stress is reached. Beyond this point, viscosity decreases with increasing shear rate. Most slurries have this property, as does America’s favorite vegetable, catsup. •

Dilatant fluids are those of which viscosity increases with increasing shear rate. Examples are candy mixtures, clay slurries, and quicksand.

•

Pseudo-plastic fluids are similar to Bingham-plastic fluids, except there is no definite yield stress. Many emulsions fall into this category.

•

Thixotropic fluids are those of which viscosity decreases to a minimum level as their shear rate increases. Their viscosity at any particular shear rate may vary, depending on the previous condition of the fluid. Examples are asphalt, paint, molasses, and drilling mud.

There are two other terms with which you should be familiar:

Reproduced from the Hydraulic Institute Standards (Figure 72 ) perature drop of 1°C. The following definitions are used when discussing fluids and viscosity. There are five basic types of liquid that can be differentiated on the basis of their viscous behavior; they are:

NEWTONIAN These are fluids where viscosity is constant and independent of shear rate, and where the shear rate is linearly proportional to the shear stress. Examples are water and oil. The Pump Handbook Series

•

Dynamic or absolute viscosity is usually measured in terms of centipoise and has the units of force time/length2.

•

Kinematic viscosity is usually measured in terms of centistokes or ssu (Saybolt Seconds Universal). It is related to absolute viscosity as follows:

kinematic viscosity = absolute viscosity/mass density The normal practice is for this term to have the units of length2/ time. Note:

1 cSt = cP x sp gr 1 cSt = 0.22 x ssu – (180/ssu) 1 cP = 1.45E-7 lbf – s/in2 1 Reyn = 1 lbf – s/in2 21

The explanation further deWater scribes the motion of fluid in the Curve-Based immediate neighPerformance % of BEP Capacity borhood of the 60% 80% 100% 120% spinning impeller. Capacity, gpm 450 600 750 900 There Stepanoff Differential Head, ft. 120 115 100 100 mentions the experEfficiency 0.70 0.75 0.81 0.75 imental results of Horsepower 18 21 21 27 others demonstratViscous (1,000 ssu) ing that, by reducPerformance ing the clearance Capacity, gpm 423 564 705 846 between the staDifferential Head, ft. 115 108 92 89 tionary casing and Efficiency 0.45 0.48 0.52 0.48 the impeller, the reHorsepower 25 29 28 36 quired power can Note: Pumped fluid specific gravity = 0.9 be reduced. He also writes about the details of some investigations The process of determining that demonstrate the beneficial the effect of a fluid’s viscosity on effect of good surface finishes on an operating pump has been studboth the stationary and rotating suried for a number of years. In the faces. Included is a chart prepared book Centrifugal and Axial Flow by Pfleiderer, based on work by Pumps, A.J. Stepanoff lists the Zumbusch and Schultz-Grunow, losses that affect the performance that gives friction coefficients for of pumps as being of the followcalculating disk friction losses. The ing types: chart is used in conjunction with • mechanical losses the following equation: • impeller losses 2 3

TABLE 1. WATER-BASED AND VISCOUS PERFORMANCE

(hp)d = KD γ u

•

leakage losses

•

disk friction losses

where

Of all external mechanical losses, disk friction is by far the most important, according to Stepanoff. This is particularly true for pumps designed with low specific speeds. Stepanoff gives a brief discussion of the physics of a rotating impeller and emerges with a simple equation that summarizes the drag force acting upon it:

(hp)d = Kn3D5 where K = a real constant n = the pump operating speed D = the impeller diameter

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who need a quick answer to a particular problem may need to look elsewhere for help. In the book, Centrifugal Pumps, V. Lobanoff and R. Ross discuss the effect of viscous fluids on the performance of centrifugal pumps. They make the point that because the internal flow passages in small pumps are proportionally larger than those in larger pumps, the smaller pumps will always be more sensitive to the effects of viscous fluids. They also introduce a diagram from the paper “Engineering and System Design Considerations for Pump Systems and Viscous Service,” by C.E. Petersen, presented at Pacific Energy Association, October 15, 1982. In this diagram, it is recommended that the maximum fluid viscosity a pump should be allowed to handle be limited by the pump’s discharge nozzle size. The relationship is approximately:

viscositymax = 300(Doutlet nozzle –1) where viscosity is given in terms of ssu D is measured in inches

K = a constant based on the Reynolds number D = impeller diameter γ = fluid density u = impeller tip speed Like most of Stepanoff’s writing, this presentation contains great depth with considerable rigor. It makes interesting reading if you are willing to put forth the time. Those of us

With respect to the prediction of the effects of viscous liquids on the performance of centrifugal pumps, Lobanoff and Ross direct the reader to the clearly defined methodology of the Hydraulic Institute Standards. This technique is based on the use of two nomograms on pages 112 and 113 of the 14th edition (Figures 71 and 72). They are reproduced here as Figures 1 and 2. They are intended

TABLE 2. POLYNOMIAL COEFFICIENTS Correction Factor Cη CQ CH0.6 CH0.8 CH1.0 CH1.2

Dx1 1.0522 0.9873 1.0103 1.0167 1.0045 1.0175

Dx2 -3.5120E-02 9.0190E-03 -4.6061E-03 -8.3641E-03 -2.6640E-03 -7.8654E-03

Dx3 -9.0394E-04 -1.6233E-03 2.4091E-04 5.1288E-04 -6.8292E-04 -5.6018E-04

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Dx4 2.2218E-04 7.7233E-05 -1.6912E-05 -2.9941E-05 4.9706E-05 5.4967E-05

Dx5 -1.1986E-05 -2.0528E-06 3.2459E-07 6.1644E-07 -1.6522E-06 -1.9035E-06

Dx6 1.9895E-07 2.1009E-08 -1.6611E-09 -4.0487E-09 1.9172E-08 2.1615E-08

for use on pumps with BEPs below and above 100 gpm, respectively, which permits the user to estimate the reduction of head, capacity, and efficiency that a viscous fluid will produce on a pump curve originally generated with water. A variation on this technique is described below. The following example is taken from pages 114-116 of the Hydraulic Institute Standards section on centrifugal pump applications. There, the use of Figure 72, “Performance Correction Chart For Viscous Liquids,” is discussed. Table 1 was calculated using polynomial equations developed to replace the nomogram presented in Figure 72. The results of the calculation are within rounding error of those presented in the standard. And the approach has the additional benefit of being more convenient to use, once it has been set up as a spreadsheet. In the course of curve-fitting Figure 72, it was convenient to define a term known as pseudocapacity:

pseudocapacity = 1.95(V)0.5[0.04739(H)0.25746(Q)0.5]-0.5 where V = fluid viscosity in centistokes H = head rise per stage at BEP, measured in feet Q = capacity at BEP in gpm

TABLE 3. CORRECTION FACTOR COMPARISON Per Table 7 of HI Standards Per Polynomial Expressions

Cη 0.635 0.639

Pseudocapacity is used with the following polynomial coefficients to determine viscosity correction terms that are very close to those given by Figure 72 in the Hydraulic Institute Standards. These polynomials have been checked throughout the entire range of Figure 72, and appear to give answers within 1.0% of those found using the figure. The polynomial used is of the form:

C x = D x1 + D x2P + D x3P 2 + D x4P 3 + Dx5P4 + Dx6P5 where Cx is the correction factor that must be applied to the term in question Dxn are the polynomial coefficients listed in Table 2 P is the pseudocapacity term defined above For comparison, the correction factors for the example above (tabulated in Table 7 of the Hydraulic Institute Standards) and those calculated using the polynomial expressions above are listed in Table 3. The problem of selecting a pump for use in a viscous service is relatively simple once the correction coefficients have been calculated. If, for example, we had been looking for a pump that could deliver 100 feet of

The Pump Handbook Series

CQ 0.95 0.939

CH0.6 0.96 0.958

CH0.8 0.94 0.939

CH1.0 0.92 0.916

CH1.2 0.89 0.887

head at a capacity of 750 gpm, we would proceed as follows:

Hwater = Hviscous service/CH1.0 Qwater = Qviscous service/CQ The next step would be to find a pump having the required performance on water. After determining the efficiency of the pump on water, we would correct it for the viscous case as shown above:

ηviscous service = ηwater x Cη The horsepower required by the pump at this point would be calculated as follows:

hpviscous service = (Qviscous service x Hviscous service x sp gr) (3,960 x ηviscous service) As with water service, the horsepower requirements at offdesign conditions should always be checked. ■ Gunnar Hole is a principal in Trident Engineering, Inc. in Houston, TX. He has been involved in the selection, installation, and troubleshooting of rotating equipment for the past 15 years. Mr. Hole is a graduate of the University of Wisconsin at Madison and is a Registered Professional Engineer in Texas.

23

CENTRIFUGAL PUMPS HANDBOOK

Pump Balancing Criteria BY GUNNAR HOLE

T

he subject of balancing rotors is one of the fundamentals of rotating equipment engineering. A number of balancing standards have been developed over the years to meet the requirements of pump manufacturers and users, and the idea of balancing is simple. Unfortunately, the definitions and mathematics used in describing balancing problems can be confusing. This article compares these criteria so the end user can use consistent reasoning when making balancing decisions.

TABLE 1. BALANCING CRITERIA Unbalanced Force Method

Specified Eccentricity Method

As per API 610 As per 6th Edition AGMA 510.02

Residual Unbalance (RUB), in.–oz where: Wj = rotor weight per balance plane, Ibf N = rpm ε = eccentricity, in. Eccentricity (ε) or Specific Unbalance in.–oz/lbm

in.–lbm/lbm where RUB = εWj Unbalance Force (UBF), lbf where: UBF = εMω2 and M = Wj/386 lbf–s2/in. ω = 2 π N/60 rad/s Circular Velocity (CV), in./s

mm/s where CV = εω ISO Standard 1940 Balance Grade

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56347 Wj N2

56347 N2 3522 N2

16 ε Wj

16 ε

ε see Table 2

0.10 Wj

εWjN2 35200

368 N 9347 N G – 9347 N

εN 9.54 2.66 εN

G – 2.66 εN

Presented below is a description of the problem, definitions of some of the more important terms used, and references that can be consulted for a more thorough review. A table also compares three of the most common balancing criteria used in the pump industry. Perhaps the least controversial comment that can be made to an experienced equipment specialist is that “accurate rotor balancing is critical to reliable operation.” I could add some spice to the conversation by giving my opinion on how good is good enough, but I would rather address the standards used in the pump industry and show how they take different approaches to resolve the Specified problem of balancing rotors. Circular I use the term rotor repeatedVelocity ly in this discussion. For the purMethod pose of this article, I include As per API 610 partially and fully assembled 7th Edition pump shaft/sleeve/impeller assemblies as well as individual 4 Wj pump components installed on N balancing machine arbors in this definition. The three major criteria used will be referred to as the Unbalanced Force Method (UFM), the Specified Eccentricity Method (SEM), and the Specified Circular 4 Velocity Method (SCVM). N In the UFM the allowable unbalance permitted in a rotor is 0.25 the amount that will result in a dynamic force on the rotor system N equal to some percentage of the rotor’s static weight. This allowable unbalance is therefore related to the operating speed of the rotor. WjN An example of this method can be 140800 found in API Standard 610 6th Edition, where the unbalance force contributed to a rotor system 0.26 by a rotating unbalance is limited to 10% of the rotor’s static weight. The SEM attempts to specify 0.665 balance quality by limiting the distance by which the center of mass of the rotor can be offset G – 0.665 from the center of rotation of the rotor. This method is used in AGMA Standard 515.02, which is The Pump Handbook Series

commonly referenced by flexible coupling vendors. It has the advantage of being conceptually simple. For the gear manufacturers who developed this standard, it allowed the use of manufacturing process tolerances as balancing tolerances. In Paragraph 3.2.7, API 610 7th Edition suggests that couplings meeting AGMA 515.02 Class 8 should be used unless otherwise specified. The SCVM is based on considerations of mechanical similarity. For geometrically similar rigid rotors running at equal peripheral speeds, the stresses in the rotor and bearings are the same. This method is described in ISO Standard 1940—Balance Quality of Rigid Rotors. It also forms the basis of API Standard 610 7th Edition’s very stringent 4W/N balancing requirement. Standards based on this methodology are becoming more common. In Table 1 the three balancing criteria discussed above are compared with respect to their effect on the various parameters involved in balancing. The terms used in the table are defined as follows:

RESIDUAL UNBALANCE This is the amount of unbalance present or allowed in the rotor. It has the units of mass and length. It is computed by taking the product of the rotor mass (per balance plane) times the distance from the rotor’s center of mass to its center of rotation. Note that 1 in.–oz is equivalent to 72.1 cm–g.

ECCENTRICITY This is the distance that the center of mass of the rotor is displaced from the rotor’s center of rotation. It has the unit of length. It can also be considered as a measure of specific residual unbalance, having the units of length–mass/mass. This term is the basic criterion of SEM balancing rules (see Table 2). Note that 1 in. is equivalent to 25.4 mm.

TABLE 2. BALANCE QUALITY CLASSES

FLEXIBLE ROTOR

The elastic deflection of flexible Note: AGMA 515.02 refers to several Balance rotors sets up additional centrifugal Quality Classes. They are summarized as follows: forces that add to the original unbalance forces. Such rotors can be balEquivalent ISO anced in two planes for a single speed AGMA Balance Quality Grade only. At any other speed they will Class ε, µ-in. 1,800 rpm 3,600 rpm become unbalanced. Balancing the 8 4,000 19.2 38.3 rotor to allow it to run over a range of 9 2,000 9.6 19.2 speeds involves corrections in three 10 1,000 4.8 9.6 or more planes. This process is called 11 500 2.4 4.8 multi-plane balancing. 12 250 1.2 2.4 One important point is that the pump/coupling/driver system must UNBALANCE FORCE be considered as a whole when evalThis is the force that is exertuating balance quality. A simple ed on a rotor system as a result of pump rotor can be balanced to meet the non-symmetrical distribution API 610 7th Edition’s 4W/N criteria of mass about the rotor’s center of in a modern balancing machine withrotation. The units of this term are out too much trouble. An electric force. This term is the basic criterimotor rotor may be even easier to on of UFM balancing rules. Note balance due to its simple constructhat 1 lbf is equivalent to 4.45 tion. But the coupling connecting Newton. them can be a completely different CIRCULAR VELOCITY matter. This is the velocity at which The coupling will likely have the center of mass of the rotor more residual unbalance than either rotates around the center of rotathe pump or the motor. And every tion. You can think of it as a tantime you take the coupling apart and gential velocity term. It has the put it back together you take the units of length per unit time. It chance of changing its balance condiforms the basis for balancing rules tion. As written, API 610 7th Edition based on the ISO Standard 1940 allows a coupling to have a specific series. In fact, the Balancing residual unbalance nearly 60 times Grades outlined in ISO 1940 are higher than for a 3,600 rpm pump. referenced by their allowable circuThis can be a significant problem if lar velocity in millimeters per secyou use a relatively heavy coupling. ond. The balance quality called for These balancing methods are priin API 610 7th Edition is better marily intended for use on rigid than the quality that ISO 1940 recrotors—those operating at speeds ommends for tape recorder drives under their first critical speed. and grinding machines. ISO 1940 Flexible rotors, which operate above recommends G–6.3 and G–2.5 for their first critical speed, are considermost pump components, where ably more complicated to balance. API 610 calls for the equivalent of The process of balancing flexible G–0.67. Note that 1 in./s is equivarotors is discussed in ISO Standard lent to 25.4 mm/s. 5406–The Mechanical Balancing of Flexible Rotors and ISO Standard RIGID ROTOR 5343–Criteria for Evaluating Flexible A rotor is considered rigid Rotor Unbalance. when it can be balanced by makThe basic concepts of rigid and ing mass corrections in any two flexible rotor balancing are the same. arbitrarily selected balancing The main difference is that with rigid planes. After these corrections are rotor balancing we are only conmade, the balance will not significerned with the rigid body modes of cantly change at any speed up to vibration. With a flexible rotor, we the maximum operating speed. have to consider some of the higher With the possible exception of modes of vibration as well. In these home ceiling fans, I believe that cases the deflection of the rotor affects two-plane balancing is the minithe mass distribution along its length. mum required for rotating equipIn general, each of the modes has to ment components. be balanced independently.

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Appendix I of API 610 7th Edition briefly discusses some of the implications of operating a rotor near a critical speed. The guidelines given there recommend separation margins that specify how far away from a critical speed you can operate a rotor. These margins depend on the system amplification factors (also known as magnification factors), which are directly related to the damping available for the mode or resonance in question. The net result of these recommendations is to limit the maximum operating amplification factor to a maximum of about 3.75. The amplification factor can be thought of as a multiplier applied to the mass eccentricity, ε, to account for the effect of system dynamics. Algebraically, the physics of the situation can be represented as follows: x = X sin (ωt – Φ)

(ω/ωn)2 X = ε ———————————— ([1 – (ω/ωn)2]2 + (2ζω/ωn)2)0.5 2ζω/ωn Φ = tan–1 ———————— 1 – (ω/ωn)2

where x is the displacement of a point on the rotor X is the magnitude of the vibration at that point ε is the mass eccentricity ω is the operating speed or frequency of the rotor Φ is the phase angle by which the response lags the force ζ is the damping factor for the mode of vibration under consideration X/ε is the amplification factor ω/ωn is the ratio of operating speed to the critical speed under consideration A more detailed discussion on the topic of damped unbalance response (or whirling of shafts) can be found in any introductory vibration textbook. ■ Gunnar Hole is a principal in Trident Engineering, Inc. in Houston, TX.

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CENTRIFUGAL PUMPS HANDBOOK

Bearing Basics BY RAY RHOE

A

ntifriction bearings, which can utilize either balls or rollers, are used to transfer radial and axial loads between the rotating and stationary pump and motor assemblies during operation. Even under the best of installation, maintenance, and operating conditions, bearing failures can and will occur. The purpose of this article is to provide a working-level discussion of bearings, the types of failures, and how bearings should be installed and maintained for optimum life expectancy. Due to space limitations, we cannot address all the different sizes and types of bearings available, or all the constraints currently utilized in design. However, because electric motors are used more often to drive centrifugal pumps, our discussion will be based on bearings typically used in quality motors. These bearings usually include a single radial bearing and a matched set of duplex angular contact bearings (DACBs). Together, these bearings must: • allow the unit to operate satisfactorily over long periods of time with minimum friction and maintenance • maintain critical tolerances between rotating and stationary assemblies to prevent contact and wear • transmit all variable radial and axial loads in all operating conditions, which include reverse rotation, startup, shutdown, maximum flow, and maximum discharge pressure Each bearing has a specific purpose. The radial bearing, which is located at one end of the motor, only transfers radial loads such as minor unbalanced rotor loads—and the weight of the rotor itself in the case of horizontally oriented components. The DACBs

26

Photo 1. Typical radial bearings

must transfer radial loads at the other end of the motor, and they must transfer all axial loads. Photo 1 shows several typical radial bearings, and Photo 2 shows DACBs.

DIFFERENT BEARING CONFIGURATIONS Radial bearings may be provided with either 0, 1, or 2 seals or shields that are effectively used to prevent entry of foreign material into the bearings. If the bearing is equipped with one seal or shield, the installer should determine which end of the motor the seal or shield should face. Failure to install radial bearings properly in the correct orientation may result in the blockage of grease or lubricant to the bearings during routine maintenance. The orientation of DACBs is more complex, DACBs must be installed in one of four configurations, as determined by design: 1. face-to-face 2. back-to-back 3. tandem: faces toward the pump 4. tandem: faces away from the pump The “face” of the DACB is that side that has the narrow lip on the

The Pump Handbook Series

outer race. The “back” of the bearing has the wider lip on the outer race and usually has various symbols and designators on it. Photo 2 shows two pairs of DACBs. The pair on the left is positioned faceto-face while the pair on the right is back-to-back. Note that the lip on the outer races of the first pair is narrower than on the second pair. This distinguishing characteristic provides an easy identification of which side is the face or back. In tandem, the narrow lip of one bearing is placed next to the wide lip on the other. In other words, all bearing faces point either toward the pump or away from it. To facilitate the installation of DACBs, the bearing faces should be marked with a black indelible marker showing where the burnished alignment marks (BAMs) are on the back. This is because the four BAMs, two on each bearing, must be aligned with their counterparts, and not all BAMs are visible during installation. For example, when the first bearing is installed in a face-to-face configuration, the BAMs are on the back

ing cannot be hammered into position or removed and reused because it will be destroyed internally by these actions.

2. DACBs

Photo 2. Two pairs of DACBs, with the pair on the left positioned face-to-face, the pair on the right back-to-back

side, hidden from the installer. Marking the face of each bearing allows the installer to see where the BAMs are, so that all four BAMs may be aligned in the same relative position, such as 12 o’clock.

BEARING PRELOAD Under certain operating conditions (hydraulic forces, gravity, and movement of the pump and motor foundation such as on a seagoing vessel), the rotor may be loaded in either direction. If this occurs, the balls in a DACB with no preload could become unloaded. When this happens, the balls tend to slide against the races (ball skid) rather than roll. This sliding could result in permanent damage to the bearings after about five minutes. To prevent ball skid, bearing manufacturers provide bearings that have a predetermined clearance between either the inner or outer races. Face-to-face bearings have this clearance between the outer races. When the bearings are clamped together at installation (the outer races are clamped together), the balls are pressed between the inner and outer races, causing the preload. Back-to-back bearings have the clearance on the inner races, which are usually clamped together with a bearing locknut. By increasing the clearance between races, the preload can be increased from zero to a heavy load. This way, when conditions

cause the rotor loads to change direction or be eliminated, the bearing balls will still be loaded and ball skid should not occur. One disadvantage of using preloaded bearings is that bearing life will be reduced due to the increased loading. Preloaded bearings should not be used unless design conditions require them. If uncertain about the need for preload, users should contact manufacturers.

BEARING INSTALLATION Once the proper bearings have been obtained and the correct orientation determined, installation is relatively simple. The shaft and especially the shaft shoulder should be cleaned and any welding or grinding operations secured. The bearings must be installed in a clean environment, and the shaft must be free of nicks and burrs that may interfere with installation.

1. RADIAL BEARINGS To install radial bearings, they should be heated in a portable oven to 180–200°F. Then, using clean gloves and remembering the correct orientation, quickly slide each bearing over the shaft and firmly onto the shaft shoulder. Do not drop or slap them into position. Experience indicates that you have about 10 seconds after removing the bearing from the oven before it cools and seizes the shaft. If it seizes the shaft out of location, remove it and scrap it. The bear-

The Pump Handbook Series

Installation of DACBs follows the some procedure, except that additional care must be taken to position the bearings properly, line up the burnished alignment marks, and not erase the indelible marks added on each bearing face. After the first bearing has been installed, rotate the rotor (if necessary) so the alignment mark on the inner race is at 12 o’clock, then rotate the outer race so it too is at 12 o’clock. Before proceeding with the second bearing, mentally walk through the procedure. Remember which direction the face goes and that the burnished alignment marks must be in the same position as the first bearing marks. Also remember you have about 10 seconds before the bearing seizes the shaft. The purpose of aligning the four burnished alignment marks is to minimize off-loading (fight) and radial runout loads that will occur if the true centers of the bearings are not lined up. Minor imperfections will always occur, and they must be minimized. Failure to align the marks will result in the bearings loading each other. DACBs come only in matched pairs—they must be used together. To verify that a pair is matched, check the serial number on the bearing halves—they should be the same, or properly designated, such as using bearing “A” and bearing “B.”

NEW BEARING RUN-IN After new bearings have been installed, they should be run in while monitoring their temperature, noise, and vibration. Run-in is often called the “heat run” or “bearing stabilization test.” To perform this test, first rotate the pump and driver by hand to check for rubbing or binding. If none occurs, operate the

27

installed and the balls ride on the ball ridge located on the outer race. Evidence of reverse loading appears as “equator” bands around the balls. Contamination (failure to follow Rule 3): Contamination of bearings almost always occurs during installation, but can also occur when liquids or other constituents from the pump leak or are present in the surrounding environment. If contamination is found in a new bearing before installation, the bearing should be carefully cleaned and repacked. Evidence of solid contamination in used bearings usually appears as very small, flat dents in the races and balls. Photo 3. Radial bearing disassembly

unit at the design rating point and record bearing temperatures every 15 min. Bearing temperatures should increase sharply and then slowly decline to their normal operating temperature, usually 20–60°F above ambient. During the heat run care should be taken to ensure that the temperature does not exceed the value specified by the manufacturer. If it does, the unit should be secured and allowed to cool to within 20°F of ambient, or for 2 hours. The unit may then be restarted and the test repeated as necessary until the bearing temperature peaks and begins to decline. If, after repeated attempts, bearing temperatures do not show signs of stabilization, too much grease may be present. The bearing should be inspected and corrective actions taken as necessary. Now let’s cover some basic rules to follow when working with bearings: 1. Never reuse a bearing that has been removed using a gear puller, even if it is new. The bearing has been internally destroyed in the removal process. See “True Brinelling” under “Bearing Failures.”

28

2. DACBs must be installed in the correct orientation. If not, they may experience reverse loading and fail. See “Reverse Loading” under “Bearing Failures.” 3. Bearings must be installed in a clean environment. Contamination is a leading cause of premature failures. 4. Do not pack the bearing and bearing cups full of grease. Excessive grease will cause overheating and ball skid. See “Excessive Lubrication” under “Bearing Failure.”

BEARING FAILURES Failure to follow these four basic rules will result in premature bearing failures. These and other failures will occur for the following reasons: True Brinelling (failure to follow Rule 1): This type of bearing failure occurs when removing bearings with a gear puller. The force required to remove a bearing from a shrink-fit application is great enough that when it is transferred through the balls to the inner races, the balls are pressed into the inner and outer races forming permanent indentations. Reverse Loading (failure to follow Rule 2): Reverse loading occurs when DACBs are improperly The Pump Handbook Series

Excessive Lubrication (failure to follow Rule 4): Too much grease in a bearing may cause the balls to “plow” their way through the grease, resulting in increased friction and heat. If the bearings and bearing caps are packed full of grease, ball skid could occur. When it does, the balls do not roll, but actually slide against the races. Experience shows that the bearings may be permanently damaged after more than five minutes of ball skid. Finding packed bearings and bearing caps is a good indication that too much grease caused the bearing to fail. Bearing manufacturers usually recommend that bearings have 25-50% of their free volume filled with grease. Excessive Heat: Failure to provide adequate heat transfer paths, or operating the component at excessive loads or speeds may result in high operating temperatures. Evidence of excessive temperature usually appears as silver/gold/brown/blue discoloration of the metal parts. False Brinelling: False brinelling occurs when excessive vibrations cause wear and breakdown of the grease film between the balls and the races. This condition may be accompanied by signs of corrosion. A good example of how false brinelling could occur would be when a horizontally positioned component is shipped

Zero-leakage magnetic liquid seal developed to retrofit process pumps across the country and not cushioned from a rough road surface. The load of the rotor is passed through the bearing balls, which wear away or indent the races. Evidence of false brinelling looks similar to true brinelling, but may be accompanied by signs of corrosion where the grease film has not been maintained. Correction simply involves protecting the unit from excessive vibration and using specially formulated greases where past experience demonstrates the need. Fatigue Failure: Even when all operating, installation, and maintenance conditions are perfect, bearings will still fail. In this case, the bearings have simply reached the end of their useful life, and any additional use results in metal being removed from the individual components. Evidence of fatigue failure appears as pits.

BEARING DISASSEMBLY FOR INSPECTION Now that we know what to look for in failed bearings, let’s see

how we disassemble a bearing for inspection. Before disassembling any bearing, however, turn it by hand and check it for rough performance. Note its general condition, the grease (and quantity thereof), and whether there is any contamination. If solid contamination is present, the particles should be collected using a clean filter bag as follows: 1. Partially fill a clean bucket or container with clean diesel fuel or kerosene. 2. Insert a clean filter bag into the kerosene container. This ensures that no contamination from the container or the kerosene gets into the filter bag. 3. Using a clean brush, wash the grease and contamination out of the bearing. The grease will dissolve and any contamination will be collected in the filter bag for future evaluation.

DACB DISASSEMBLY To disassemble DACBs, support the face of the outer race and press down against the inner race. The back of the bearing must be on top.

HANDLING, TRANSPORTATION, AND STORAGE Common sense applies in handling, storing, or transporting precision bearings. They should not be dropped or banged. They should be transported by hand in cushioned containers, or on the seat of vehicle—not in a bike rack. They should be stored in a cool, clean, dry environment. Because nothing lasts forever, including bearing grease, bearings should not be stored for more than a few years. After this, the grease degrades and the bearings may become corroded. At best, an old bearing may have to be cleaned and repacked, using the correct type and amount of grease.

RADIAL BEARING DISASSEMBLY

MAINTENANCE

After removing the grease and any contamination, you should disassemble radial bearings by removing any seals or shields, which are often held in place by snap rings. Then, to remove a metal retainer, drill through the rivets and remove both retainer halves. Then the bearing should again be flushed (in a different location) to remove any metal shavings that may have fallen between the balls and races when drilling out the rivets. If the bearing does not freely turn by hand, some metal particles are still trapped between the balls and races. Next, place the bearing on the floor as shown in Photo 3 with the balls packed tightly together on the top. Insert a rod or bar through the inner race and press down, hard if necessary. Note: If the balls are not packed tightly together, disassembly will not occur.

Routine maintenance of bearings usually involves periodic regreasing (followed by a heat run) and monitoring bearing vibrations, which will gradually increase over long periods of time . To maintain pumps and drivers that are secured for long periods of time, simply turn the rotor 10–15 revolutions every three months by hand. This will ensure than an adequate grease film exists to prevent corrosion of the bearing. If this action is not taken, the bearings may begin to corrode due to a breakdown in the grease film. ■

The Pump Handbook Series

Ray W. Rhoe, PE, has a BSCE from The Citadel and 15 years’ experience with pumps, testing, and hydraulic design.

29

CENTRIFUGAL PUMPS HANDBOOK

Centrifugal Pump Efficiency BY DAVID CUMMINGS

T

he efficiency of centrifugal pumps of all sizes is becoming more important as the cost and demand for electricity increases. Many utilities are emphasizing conservation to reduce the number of new generating facilities that need to be built. Utilities have increased incentives to conserve power with programs that emphasize demand side conservation. These programs often help fund capital equipment replacements that reduce electrical consumption. Demand side management programs make replacing old pumping equipment more feasible than ever before.

DETERMINING PUMP EFFICIENCY

The efficiency of a pump ηp is ratio of water horsepower (whp) to brake horsepower (bhp). The highest efficiency of a pump occurs at the flow where the incidence angle of the fluid entering the hydraulic passages best matches with the vane angle. The operating condition where a pump design has its highest efficiency is referred to as the best efficiency point (BEP). ηp = whp/bhp The water horsepower (whp) may be determined from the equation: whp = QHs/3,960 where Q = capacity in gallons per minute, H = developed head in feet, and s = specific gravity of pumped fluid. Preferably, the brake horsepower supplied by a driver can be determined using a transmission dynameter or with a specially calibrated motor. Brake horsepower determined in the field by measuring kilowatt input and multiplying by the motor catalogue efficiency can be inaccurate. If motor power is determined in the field, data should be taken at the motor junction box, not at the motor control center. Overall pump motor efficiency ηo may be determined from the equation: ηo = whp/ehp

30

where ehp = electrical power in horsepower.

EFFICIENCY LOSSES Pump efficiency is influenced by hydraulic effects, mechanical losses, and internal leakage. Each of these factors can be controlled to improve pump efficiency. Any given design arrangement balances the cost of manufacturing, reliability, and power consumption to meet users needs. Hydraulic losses may be caused by boundary layer effects, disruptions of the velocity profile, and flow separation. Boundary layer losses can be minimized by making pumps with clean, smooth, and uniform hydraulic passages. Mechanical grinding and polishing of hydraulic surfaces, or modern casting techniques, can be used to improve the surface finish, decrease vane thickness, and improve efficiency. Shell molds, ceramic cores, and special sands produce castings with smoother and more uniform hydraulic passages. Separation of flow occurs when a pump is operated well away from the best efficiency point (BEP). The flow separation occurs because the incidence angle of the fluid entering the hydraulic passage is significally different from the angle of the blade. Voided areas increase the amount of energy required to force the fluid through the passage. Mechanical losses in a pump are caused by viscous disc friction, bearing losses, seal or packing losses, and recirculation devices. If the clearance between the impeller and casing sidewall is too large, disc friction can increase, reducing efficiency. Bearings, thrust balancing devices, seals, and packing all contribute to frictional losses. Most modern bearing and seal designs generate full fluid film lubrication to minimize frictional losses and wear. Frequently, recirculation devices such as auxiliary impellers or pumping rings are used to provide cooling and lubrication to bearings and seals. Like the main impeller, these devices pump fluid The Pump Handbook Series

and can have significant power requirements. Internal leakage occurs as the result of flow between the rotating and stationary parts of the pump, from the discharge of the impeller back to the suction. The rate of leakage is a function of the clearances in the pump. Reducing the clearances will decrease the leakage but can result in reliability problems if mating materials are not properly selected. Some designs bleed off flows from the discharge to balance thrust, provide bearing lubrication, or to cool the seal.

EXPECTED EFFICIENCIES The expected hydraulic efficiency of a pump design is a function of the pump size and type. Generally, the larger the pump, the higher the efficiency. Pumps that are geometrically similar should have similar efficiencies. Expected BEPs have been plotted as a function of specific speed and pump size. A set of curves that may be used to estimate efficiency is provided in Figure 1. The specific speed (Ns) of a pump may be determined from the equation: Ns = NQ0.5/H0.75 where N = speed in rpm, Q = capacity in gpm, and H = developed head in feet. Using a pump performance curve, the highest efficiency can be determined and the specific speed calculated using the head and capacity at that point. Using the specific speed and the pump capacity, the expected efficiency can be estimated. If the pump has bearings or seals that require more power, such as tilting pad thrust bearings or multiple seals, this should be considered when calculating the expected efficiency.

MOTOR EFFICIENCIES Efficiencies for new “premium efficiency motors” are provided in Figure 2. Using these values, with anticipated pump efficiencies in Figure 1, the expected power consumption for a well designed pump

and motor can be determined. The calculated power consumption can be compared with an existing installation to determine the value of improving pump performance or replacing the unit.

EXAMPLE CALCULATION OF PUMP EFFICIENCY A single-stage end-suction process pump will be used as an example for an efficiency calculation. The pump uses a mechanical seal and an angular contact ball bearing pair for thrust. The pumped fluid is water with specific gravity of 1.0. The pump operates at its BEP of 2,250 gpm, developing 135 feet of head. The pump speed is 1,750 rpm (note: with the new motor the speed may change, but to simplify the example it will be assumed the new and old motor both operate at 1,750 rpm). The expected power consumption for a new unit can be calculated. First the pump specific speed will be calculated: [1,750 rpm x (2,250 gpm)0.5] Ns = (135 ft)0.75 Ns = 2,096 Figure 1 can be used at Ns = 2,100 and interpolated for 2,250 gpm. The expected efficiency is 86%. The water horsepower is: (2,250 gpm x 135 ft x 1.0) whp = 3,960

TABLE 1. EXPECTED EFFICIENCY FOR “PREMIUM EFFICIENCY MOTORS” Motor Horsepower 5 10 15 20 25 30 40 50 75 100 125 150 200 Over 200

1200 rpm 88.0 90.2 91.0 91.7 92.4 93.0 93.6 93.6 94.5 94.5 94.5 95.0 95.0 95.0

Minimum Acceptable Efficiency 1800 rpm 3600 rpm 88.0 87.0 91.0 90.2 92.0 91.0 93.0 91.7 93.5 92.0 93.6 92.4 94.1 93.0 94.1 93.0 95.0 94.1 95.0 94.5 95.4 94.5 95.4 94.5 95.0 95.0 95.4 95.0

The antifriction bearings and typical mechanical seal do not require a power adjustment. However, if a tilting pad thrust bearing or other device, such as a special seal, that used more power was part of the design, the correction would be made here by adding the additional horsepower to the calculated value. Using Figure 2, the minimum efficiency for a 100 Hp motor is 95%. The efficiency value may change slightly for the operating condition and should be verified with a motor manufacturer. The 95% efficiency will be used in this example. The whp = 76.7 Hp expected electrical horsepower is: The expected brake horsepower is: ehp = 89.2 Hp/0.95 bhp = 76.7 Hp/0.86 ehp = 93.9 Hp bhp = 89.2 Hp or ehp = 93.9 Hp x FIGURE 1 .7457 kw/Hp ehp = 70.0 kw The last time this pump was rebuilt and put in service, power was measured at 79.5 kw. The difference in power consumption between the existing unit and a new unit can be calculated: ehp = 79.5 kw 70.0 kw ehp = 9.5 k. Efficiency of various pumps sizes and specific speeds The Pump Handbook Series

If power costs 8 cents a kilowatt hour and the pump operates continuously, the savings of replacing this unit on an annual basis can be calculated: cost = 9.5 kw x $0.08 per kw hr x 8,760 hr/yr cost = $6,658 year This figure can be used to determine if the additional power consumption justifies replacing the pump. If a replacement pump and motor of this size can be purchased and installed for $40,000, and the electric utility offers a 50% rebate program, the net cost of $20,000 for the user is certainly worth considering.

SUMMARY Remember, for any centrifugal pump to operate efficiently it needs to be properly applied. When processes require changing flow rates frequently, variable speed drives can be a solution. A pump operating far from its BEP will be neither efficient nor reliable. Many times changing the pump size to better match the system will reduce power costs dramatically. ■ David L. Cummings is an independent consultant who provides engineering services and equipment for specialized applications.

31

CENTRIFUGAL PUMPS HANDBOOK

Motor Size Selection for Centrifugal Pumps BY ROBERT J. HART

Q:

How do I select the proper motor size for my centrifugal pump applications? In some applications we have experienced driver overload, while other applications appear satisfactory using the same selection method. What seems to be a straightforward requirement of selecting a pump motor too frequently results in a major problem when commissioning the pumping system and the installed motor overloads and is tripped off the line. Correcting the problem can be as simple as adjusting a manual valve. And it can be as complex and time consuming as replacing the motor, motor starters, and service wire with a larger size, and altering the flow control system. A simple rule of thumb of supplying a motor size that exceeds the pump manufacturers rated point brake horsepower by some fixed margin, or of supplying a motor size equal to the brake horsepower at the end of the selected impeller diameter curve, may not work in all cases. The person selecting the motor must have a thorough knowledge of the pumping system and its characteristics, pump industry practices, and limitations of the generalized data provided by the pump manufacturer with the quotation.

A:

GENERALIZED PERFORMANCE CURVE LIMITATIONS The brake horsepower values published by the pump manufacturer on the generalized hydraulic performance curves (TDH, efficiency, BHP vs. capacity) are the basis for the rated point BHP returned to a potential customer when responding to a request for quotation for a specific application. The data is as accurate as

32

practical for the designated equipment design, but does have a tolerance range which may be experienced for any specific pump for such characteristics as the Total Dynamic Head (TDH) at a specific flow rate. Brake horsepower is related to the TDH by the following formula: BHP =

(TDH) x Flow x Specific Gravity 3960 x Efficiency x Viscosity Correction Factors*

*See Hydraulic Institute Standards for Values Most pump users will not accept a lower than specified rating point TDH, and the manufacturer is frequently required to test the equipment prior to shipment to confirm that the pump meets the specified requirements. If the installed impeller should produce less TDH than specified, the manufacturer must replace the impeller with a larger diameter. If high alloy materials are involved, there may be considerable expense and delay involved. Hence the practice in the pump industry is to publish a performance curve (TDH vs. Capacity) for a given impeller diameter that is somewhat less than can actually be achieved by the specified diameters. Should a test then be specified and the impeller TDH test higher than allowed tolerances at a given flow rate, the impeller can be reduced to a smaller diameter to provide the required values without replacing it. One of the current pump industry acceptance test criteria, the Hydraulic Institute Standards, permits the TDH to exceed the design point requirements by as much as 8%. Sometimes pump impellers will exceed the published data by as much as 20% when first tested. This is not usually the case, but there is a range of performance, especially on

The Pump Handbook Series

relatively new or infrequently tested pump designs. The user should be informed of this potential variation if the impeller requires replacement due to normal operating wear of the pump, especially if it is to be purchased from a source other than the original equipment manufacturer, which may not have historical test records of the original hydraulic design. Industry practice is to guarantee only the TDH (with a tolerance range) at a specified flow rate and the pump efficiency. The resulting brake horsepower is guaranteed only by the same tolerance, and then only if the pump is tested.

MECHANICAL SEAL HORSEPOWER LOSSES The performance curve published by the pump manufacturer and described above does not provide allowances for the power required to turn a mechanical seal that is loaded to typical process conditions. For a high suction pressure, double mechanical shaft seal pump installation, this can be a measurable amount and must be added to the horsepower required to move the liquid. An allowance of one to two horsepower, for example, may be required for some ANSI style pump designs to compensate for seal losses. Hence, if the generalized performance curve rating point results in a BHP of 7.5 Hp, a motor of 10 Hp may be considered for the application if no allowance is given for factors like seal drag. As part of the equipment quotation, an estimate for the seal horsepower drag should be requested for all pumps requiring mechanical seals. If a double mechanical seal has been specified with a buffer fluid pressurized flush, the buffer fluid pressure must be specified by the seal and pump manufacturer and observed

by the user to assure the estimated seal drag horsepower is not exceeded. Over pressurizing the double mechanical seal buffer system at the site can result in a motor overload condition not anticipated during the motor selection phase. Seal horsepower losses typically have a greater impact on the installations at or below 25 Hp, but they should be considered for all installations.

FLOW CONTROL Since the BHP of most pump designs increases with increasing flow through the pump, it is the user’s responsibility to assure that the actual system flow does not exceed the rated flow originally specified when the pump was purchased. Pumping systems that limit flow only by the resistance of installed piping have a tendency to be sized with safety factors to “assure” the pump selected will provide adequate flow (see the above comments regarding generalized performance curves). A motor may overload when the pump operates at a higher flow rate than anticipated and requires a greater horsepower. Should the actual system curve extend beyond the end of the published pump curve and not intersect the pump curve, the actual horsepower will be greater than the “end of curve” horsepower frequently used as basis for motor selection. With adequate NPSH available to the pump, the performance curve and corresponding horsepower may extend to greater than published values (Figure 1).

FLUID CHARACTERISTICS Both specific gravity and viscosity can affect the required pump brake horsepower (see equation above). Motors are normally selected on the basis of rated conditions of head, flow, specific gravity, temperature, and viscosity. The off-design conditions of these characteristics should be

examined and those FIGURE 1 fluid characteristics which affect brake horsepower evaluated before selecting a driver. As examples: TDH Is there an alternate start-up or shutA down flush liquid required which has a higher specific gravity B liquid than the rated flow material? What is the actual liquid viscosity at a lower temperature than rated conditions, and will it increase the BHP of the BHP pump? Even though the pump and piping Flow is well insulated, without heat tracing the system will be at Pump performance curve. A=calculated system ambient temperature curve with safety factors, B=actual system curve. during a start-up. This will cool the incoming liquid below the control of the pump manufacturer. continuous on-line conditions that There is no simple rule of thumb. would exist once the piping system Oversizing motors to compenis in operation and at equilibrium. sate for all of the conditions that PUMP WEAR may or may not exist on every installation can be a major addiA certain amount of internal tional expense when considering recirculation takes place inside a the total electrical system. This centrifugal pump casing at all times. article has illustrated the variables As internal clearances change due to that must be taken into account. ■ wear, the rate of this circulation increases. If the system demands Robert J. Hart is a Senior down stream of the pump remain Consultant at the DuPont Company. constant and the system is designed He also serves on the Pumps and to maintain process flow, the pump Systems Editorial Advisory Board. must flow at a higher rate to compensate for this recirculation. Because of this, it may require a corresponding higher horsepower. See I.J. Karassik’s recent article “When to Maintain Centrifugal Pumps” (Hydrocarbon Processing, April 1993) for additional information on this topic.

SUMMARY Motor selection for centrifugal pumps involves many considerations, some of which are beyond the

The Pump Handbook Series

33

CENTRIFUGAL PUMPS HANDBOOK

Setting the Minimum Flows for Centrifugal Pumps BY: IGOR J. KARASSIK

J. P. Messina, Professional Engineer, Pump and Hydraulics Consultant, Springfield, NJ.

Until about 25 years ago, there were only four factors to consider when setting an acceptable minimum flow for centrifugal pumps:

A: •

34

higher radial thrust developed by single volute pumps at reduced flows

desire to avoid overload of drivers of high specificspeed axial-flow pumps

•

for pumps handling liquids with significant amounts of dissolved or entrain-ed air or gas, the need to maintain sufficiently high fluid velocities to wash out this air or gas along with the liquid

QRSA

QSRB

Unstable region pump B

Safe Zones of Operation Min. Flow Pump B

•

FIGURE 1

Pump A

temperature rise in the liquid pumped

Min. Flow

Q:

•

Head, H

I am presently involved in replacing a newly purchased pump. It was accepted by the purchaser, but the shop test was noisy. The manufacturer said this was due to the poor suction piping. The field test was unacceptable and noisy, and there was disagreement about whether the noise was due to improperly placed elbows in the suction piping or if the pump was inappropriately selected. The pump, probably designed for flows much greater than system requirements, was recirculating. The noise was a very low frequency, random banging. The single-stage, double-suction, twin-volute design had four times more NPSH than required. How would your experts have diagnosed this costly problem? What witness shop test should be conducted so that the pump purchaser can be assured of a safe continuous flow as quoted in the proposal? What measurements, observations (both audible and visual), and instrumentation should be used to detect the onset of suction and/or discharge recirculation? If the pump does not perform as quoted, are minor shop alterations conceivable? I would like to suggest to the Hydraulic Institute that a minimum nonrecirculating flow test be added to the standards. Your thoughts?

HRSA HRSB

Since then, a new 25% 100% phenomenon has been Capacity of Q in% discovered that affects the setting of minimum Comparison of safe zones of operation for flows. At certain normal and for high S value impellers reduced flows, all centrifugal pumps are subject to internal strong controversy. Accept my comrecirculation, both in the suction and ments as a personal opinion. discharge areas of the impeller. This One theoretical method exists to produces pulsations at both the sucpredict the onset of recirculation (Ref. tion and discharge, and the vibration 1 and 2). The results of this method can damage impeller material in a have been verified by many tests, way similar to classic cavitation, with actual pumps and plastic transalthough taking place in a different parent models where the onset could area of the impeller. be observed with a strobe light. The Each of these effects may dictate results corresponded within no more a different minimum operating than 5% deviation from the prediccapacity. Clearly, the final decision tions. must be based on the greatest of the Assuming that the pump is propindividual minimums. The internal erly furnished with the necessary recirculation usually sets the recominstrumentation, such as flowmeter, mended minimum, which appears to pressure gauges with sufficient sensibe what happened in your case. tivity to show pulsations, and vibraYou’ve actually raised two tion and noise monitoring devices, an questions: experienced test engineer should be 1. How can one determine the able to pinpoint the onset of internal onset of recirculation? recirculation. 2. After determining the onset, But it is the setting of the miniwhat should be the recommendmum flow which—for the time ed minimum flow in relation to being—remains controversial. the recirculation flow? Obviously, any material damage cannot serve as a standard, because by Unfortunately, the answers to the time the correctness of the deciboth questions remain in the realm of The Pump Handbook Series

a consensus about the acceptable limits of vibration and noise will be difficult. The choice of a minimum flow is much more subjective if it is based on problems arising from internal recirculation than when the temperature rises, and radial thrust and overload of drivers of high specific-speed pumps are concerned. In these situations, the effect of operating at any given low capacity can be quantified. Even the effect of handling liquids laden with air or gas is easy to determine since at some given flow the noncondensible content of the liquid will not be washed away, and will accumulate within the pump, “Bulk-head ring” construction used to elimiwhich will stop pumpnate unfavorable effects of excessively large ing. impeller eye diameter Another major obstacle to overcome in sion has been verified, it is too late. achieving a consensus is to define Therefore, the magnitudes of preswhat is continuous service and what sure pulsation, noise and vibration is intermittent. When Warren Fraser are the only criteria for establishing (who did all the seminal work on the minimum flow. internal recirculation) and I tried to Regarding vibration, the produce a quantitative value that Hydraulic Institute Standards would distinguish between these two, includes a chart, plotting maximum we first tried to define 25% of the permissible peak-to-peak amplitudes time as the breaking point between against frequency, and it is applicathem. At first this seemed reasonable, ble “when the pump is operating at but we soon realized that we had rated speed within plus or minus another problem to face. There was a 10% of rated capacity.” This could difference between running a pump create a serious problem whenever a for six hours per day at or below an pump meets these limits but is subarbitrary flow and running it for three ject to considerably higher vibrations months out of a year for some strictly when operated below the recirculaclimatic conditions. So, we decided to tion flows. The API-610 Standard is avoid making any formal distinction more specific, defining the minimum between continuous and intermittent. continuous stable flow at which the I admit that I do not have a defipump can operate without exceeding nite and final answer to offer on the the noise and vibration limits subject of selecting a minimum flow imposed by the Standard. These limstandard. I continue to use a guideits are expressed in inches per secline that I established some years ago ond rather than mils of (Ref. 3). Because the choice of displacement. required NPSH affects the onset of The Hydraulic In-stitute should internal recirculation, for high suction probably set up rules for establishing specific speeds the minimum flow a minimum flow test. But obtaining should correspond to the onset of the

FIGURE 2

The Pump Handbook Series

recirculation. For cold water, this refers to values over 10,500. And for more conservative S values, such as 8500 to 9500, set the minimum flow at 25% of the best efficiency capacity (Figure 1). These comments represent my personal opinion. I am aware that some users may be more conservative and insist that the minimum flow should never be less than the recirculation capacity. In that case, users should specify this restriction in their

FIGURE 3 Projections from casing wall to reduce axial unbalance

Projections from casing wall provided to reduce problems created by discharge side recirculation

FIGURE 4 Annular ring

Annular ring

Addition of two annular rings to impeller shrouds to reduce axial movement of rotor caused by internal recirculation at discharge

35

requests for bids. But this could only be acceptable if guidelines are published on how to conduct a test for recirculation or a formula becomes widely accepted on how to calculate the onset of internal recirculation. Regarding your question about what minor alterations may be made if the pump does not perform satisfactorily in this connection, there are two possible solutions: 1. For suction recirculation, you can reduce the minimum acceptable flow by incorporating a “bulk-head ring” with an apron overhanging the eye of the impeller (Figure 2). Of course, this does increase the required NPSH and can only be done if there is the necessary margin between available and required NPSH. 2.

36

If the problem is caused by discharge side recirculation,

you can achieve some relief by providing projections from the casing wall (Figure 3). Alternately, annular rings can be fitted to the outer shrouds of the impeller (Figure 4). I hope these comments will serve to open a dialogue between pump users and manufacturers. Such a discussion should lead to the undertaking of a series of tests that will shed additional light on the problem of acceptable minimum flows. These tests, in turn, could permit the Hydraulic Institute to include guidelines in its standards. ■

REFERENCES 1.

W.H. Fraser. Flow recirculation in centrifugal pumps. Proceedings of the Texas A&M Tenth Turbomachinery Symposium (1981).

2.

W.H. Fraser. Recirculation in centrifugal pumps. Presented at

The Pump Handbook Series

the ASME Winter Annual Meeting (1981). 3.

I.J. Karassik. Centrifugal pump operation at off-design conditions. Chemical Processing (1987).

Igor J. Karassik is Senior Consulting Engineer for IngersollDresser Pump Company. He has been involved with the pump industry for over 60 years. Mr. Karassik is a member of the Pumps and Systems Editorial Advisory Board.

CENTRIFUGAL PUMPS HANDBOOK

Estimating Maximum Head in Single – and Multi-Stage Pump Systems BY JAMES NETZEL

The maximum head or discharge pressure of a centrifugal pump can be easily estimated if the impeller diameter, number of impellers used, and rpm of the driver are known.

Q: A:

How can you estimate the maximum (shutoff) head that a centrifugal pump can deliver?

The maximum pressure a centrifugal pump delivers should be known in order to ensure that a piping system is adequately designed. Any pump that operates at a high flow rate could deliver significantly more pressure at zero (0) gpm flow, such as when the discharge valve is closed, than it delivers at operating flow. The maximum head or discharge pressure of a centrifugal pump, which usually occurs at shutoff con-

FIGURE 1 17 16 15 14 Head in Feet x 1000

13 12 11 10 9 8 7 6 5 4 3 2 1 1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 RPM x 1000

Rotations per minute (rpm) vs. head in feet to estimate maximum head The Pump Handbook Series

ditions (0 gpm), can be easily estimated if the impeller diameter, number of impellers used, and rpm of the driver (electric motor, gas engine, turbine, etc.) are known. Let’s say we have a singlestage pump with a 10-in. diameter impeller and an 1,800 rpm driver. To determine the head in feet, simply take the impeller diameter in inches and square it. Our 10-in. impeller at 1,800 rpm would yield 102, or 100 ft of head. An 8-in. impeller would yield 82, or 64 ft of head, while a 12-in. impeller would yield 122, or 144 ft of head. Now let’s assume that our 10-in. diameter impeller is driven by a 3,600 rpm motor. We first determine the head at 1,800 rpm, but then multiply this value by a factor of four. The basic rule is that every time the rpm changes by a factor of two, the head changes by a factor of four. The head at 3,600 rpm for our 10-in. impeller is therefore 102 x 4, or 400 ft of head. Our 8-in. impeller at 3600 rpm would give us 82 x 4, or 256 ft of head, and our 12-in. impeller would give us 122 x 4, or 576 ft of head. For multiple stages (more than one impeller), simply multiply the final head for one impeller by the total number of impellers in the pump. For a pump with three 10-in. impellers and a speed of 3,600 rpm, we get (102 x 4) x 3 = 400 x 3 = 1,200 ft. of head. Now what happens if we reduce the speed below 1,800 rpm? The same rule still applies: a change in speed by a factor of two changes the head by a factor of four. Therefore, a 10-in. diameter impeller spinning at 900 rpm delivers only one fourth the head it would at 1,800 rpm: 102/4 = 25 ft. Plotting several head-versusrpm points on a curve will allow the user to estimate the maximum

37

head at any given speed. Let’s say we have a turbine-driven pump that injects water into the ground to raise the subterranean oil reserves to the surface for processing. The vendor tells you that the maximum head is classified, but you have been requested to resolve system problems that you believe are pressure related. The vendor tells you that the pump has four 8-in. diameter impellers and is driven by the turbine at 13,000 rpm. You would estimate the maximum head as follows: Step 1 Determine the head at 1,800 rpm: 82 x 4 stages = 256 ft Step 2 Multiply the head at 1,800 rpm by four to get the head at 3,600 rpm: 256 x 4 = 1,024 ft Step 3 Multiply the head at 3,600 rpm by 4 to get the head at 7,200 rpm: 1,024 x 4 = 4,096 ft Step 4 Multiply the head at 7,200 rpm by 4 to get the head at 14,400 rpm: 4,096 x 4 = 16,384 ft Step 5 Plot the rpm-versus-head points to obtain the curve shown in Figure 1. As you can see, the estimated head at 13,000 rpm is 12,500 ft. To convert head in feet to psi, simply divide the head by 2.31 to get 5,411 psi. Ray W. Rhoe, PE, has a BSCE from The Citadel and 15 years’ experience with pumps, testing, and hydraulic design.

38

Q: A:

What different types of seal lubrication exist?

A mechanical seal is designed to operate in many types of fluids. The product sealed becomes the lubricant for the seal faces. Many times the fluid being sealed is a poor lubricant or contains abrasives that must be taken into account in the seal design. The design of the seal faces, materials of construction, and seal lubrication play an important role in successful operation. Achieving a high level of reliability and service life is a classic problem in the field of tribology, the study of friction, wear, and lubrication. The lubrication system for two sliding seal faces can be classified as follows: 1) hydrodynamic, 2) elastohydrodynamic, 3) boundary, and 4) mixed film. Hydrodynamic conditions exist when the fluid film completely separates the seal faces. Direct surface contact between seal faces does not take place, so there is no wear, and heat generation from friction is zero. The only heat generation occurs from shearing of the fluid film, which is extremely small. A hydrodynamic seal may rely on design features such as balance factors, surface waviness, or spiral grooves to separate the seal faces. The Society of Tribologists and Lubrication Engineers (STLE) guideline in “Meeting Emissions Regulations with Mechanical Seals” lists hydrodynamic seals as a technology to control emissions. Elastohydrodynamic lubrication (EHD) is also found in sliding surfaces, but more often this involves rolling surfaces separated by an oil film. Here the moving surfaces form

The Pump Handbook Series

an interface region that deforms elastically under contact pressure. This deformation creates larger film areas and very thin films. Such lubrication systems are normally used to control wear in rolling element bearings. In seals where the viscosity of the fluid sealed increases with increasing pressure,elastohydrodynamic lubrication occurs. Boundary lubrication is important for seal faces that are moving very slowly under heavy load. Here, hydrodynamic and elastohydrodynamic lubricant pressures are insufficient to separate the seal faces. The sliding surfaces are protected by the tribological properties of the materials of construction. An example of a seal operating within this lubrication system is a dry-running agitator seal. Mixed-film lubrication, a combination of all the previous systems discussed, occurs in all contact seals. Here the fluid film becomes very thin and is a combination of both the liquid and the gas phases of the fluid sealed. Asperities from one surface may penetrate the lubricating film and contact the opposite surface. The seal face load is then supported partially by the fluid film and partially by solid contact. If the generated head at the seal faces is not removed, surface wear and damage can occur. For applications where the seal face load is too high or the fluid viscosity is too low, designs of seal faces can be changed through balance and face geometry to improve seal performance. ■ James Netzel is Chief Engineer at John Crane Inc. He serves on the Editorial Advisory Board for Pumps and Systems.

CENTRIFUGAL PUMPS HANDBOOK

Tips on Pump Efficiency BY WILLIAM E. (ED) NELSON

GAP “A”

D

D’

GAP “B”

(a)

cised in altering the diameter of a mixed flow impeller.

Q: A:

Q: A:

It depends on the specific speed of the impeller. The specific speed index classifies the hydraulic features of pump impellers according to their type and proportions. Most refinery pumps fall between about 900 and 2,500 on this index. Some vertical multistage pumps are in the 4,000 to 6,000 range. For radial designs, impeller diameter should not be reduced more than 70 percent of the maximum diameter design. Reductions in pump impeller diameters also alter outlet channel width, blade exit angle and blade length and may significantly reduce the efficiency. The greater the impeller diameter reduction from maximum diameter and the higher the specific speed (not suction specific speed), the more the pump will

What is the effect of impeller trimming on NPSH required?

Small reductions in impeller diameter will increase the required NPSH only slightly. Diameter reductions greater than about five to 10 percent will increase NPSH required, which occurs because specific vane loading is raised by the reduced vane length, affecting velocity distribution at the impeller inlet. Not all pump companies consistently show their pump curves the increase NPSHR with reduced impeller diameters. Attention must be paid to this factor when the margin between NPSHR and NPSHA is very narrow or the NPSHR for a pump is extremely low.

What is the effect of trimming an impeller on pump efficiency?

Q: A:

What effect does trimming an impeller have on axial vibration?

Excessive impeller shroudto-casing clearances (Gap “A”) and suction recirculation cause eddy flows around the impeller, which in turn cause low frequency axial vibrations. Flow disturbances related to suction recirculation and cavitation are always present in both diffuser and volute type pumps. As the

FIGURE 2. OBLIQUE CUTS OF VANE Gap “A”

Gap “A”

D’

Gap “B”

The Pump Handbook Series

D

Gap “B”

D

With a constant rotational speed, as is the case with most pumps, the “Affinity Laws” commonly used for calculating the trim do not accurately reflect the relationship between the change in impeller diameter and the hydraulic performance achieved by the pump. The calculations generally dictate more of a cut than required to affect the desired head and flow reduction. The “Affinity Law” errors can be on the order of 20 percent of the calculated reduction. If the calculated reduction trimming calls for a 10 percent reduction in diameter, only seven or eight percent reduction should be made. The lower the specific speed of the impeller cut, the larger the discrepancy. This subject is covered in only a few pump handbooks. The subject is well covered on pages 18 and 19 of Centrifugal Pumps - Design and Application, First Edition, by Val Lobanoff and Robert R. Ross. There are several reasons for the actual head and flow being lower than that calculated: 1. The “Affinity Laws” assume that the impeller shrouds are parallel. In actuality, the shrouds are parallel only in lower specific speed pumps. 2. The liquid exit angle is altered as the impeller is trimmed, so the head curve steepens slightly. 3. There is increased turbulent flow at the vane-tips as the impeller is trimmed, if the shroud-to-casing clearance (Gap “A”) is not maintained. All of these effects contribute to a reduced head development and flow. Pumps of mixed flow design are more affected than the true radial flow impellers found in higher head pumps. More caution has to be exer-

decrease with the trimming of the impeller.

FIGURE 1. SHROUD & VANE REDUCTION

D’

Q: A:

When trimming a pump impeller to change the flow and head, I sometimes get too much of a reduction. What is the problem?

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FIGURE 3. TERMINATING VANES ONLY

impeller diameters are reduced, the flow distribution pattern across the exit width of the impeller becomes more unstable. The tendency for the high-pressure liquid to return to the low pressure side and create tip recirculation is greatly increased. Again, the higher energy level pumps are of major concern (above 200 HP and 650 feet of head per stage).

Q: A:

What are the effects of trimming an impeller on radial vibration?

Careful machining of the volute or diffuser tips to increase Gap “B” while maintaining Gap “A” has ben used for a number of years to greatly reduce the vane-passing frequency vibration. The pulsating hydraulic forces acting on the impeller can be reduced by 80 to 85 percent by increasing the radial Gap “B” from 1 percent to 6 percent. There is no loss of overall pump efficiency when the diffuser or volute inlet tips are recessed, contrary to the expectations of many pump designers. The slight efficiency improvement results from the reduction of various energy-consuming phenomena: the high noise level, shock, and vibration caused by vane-passing frequency, and the stall generated at the diffuser inlet. Table 1 gives recommended dimensions from Dr. Elemer Makay for radial gaps of the pump impeller to casing. Note that if the number of impeller vanes and the number of diffuser/volute vanes are both even, the radial gap must be larger by about 4 percent.

When trimming an impeller from its maximum diameter to adjust the head and flow developed by a centrifugal pump, what is the best way to cut the impeller? Is it best to trim the impeller vanes and the shrouds or just the vanes?

Q:

FIGURE 4. IMPELLER VANE OVERFILING

Length of blend for over filing Impeller diameter, in “A” distance of blend, in 10 & below 1 1/2 10 1/16 through 15 2 1/2 15 1/16 through 20 3 1/2 20 1/16 through 30 5 30 & larger 6

No hard and fast guidelines for the mechanical aspects of impeller trimming exist, but there are several pump construction and hydraulic design factors to consider while making the decision of what to trim. How the impeller is trimmed will greatly influence the hydraulic performance of the pump as well as the vibration levels experienced. You must evaluate the hydraulic characteristics before you decide how to trim the impeller. For volute type pumps, the entire impeller, vanes and shrouds may be cut as shown in Figure 1. However, in some pumps, this method will alter Gap “A” (shroud-tocase clearance), leading to uneven flow distribution at the impeller exit area, which can cause axial vibration and other problems. The double suction impeller type pump is especially sensitive to problems caused by increased Gap “A”, so trimming the entire impeller is not a good choice. It

A:

is best to cut the impeller vanes obliquely (Figure 2), which leaves the shrouds unchanged or to cut the vanes only (Figure 3). Trimming the vanes only tends to even out the exit flow pattern and reduce recirculation tendencies at the exit area. Gap “A” should be about 0.050 inch (radial) for minimum vibration due to vanepassing frequency. In most diffuser type pumps, it is best to trim only the vanes (Figure 3) to control tip recirculation and the ill effects of an increased Gap “A”. This cut yields a more stable head curve. The uniform flow reduces the tendency for tip recirculation, and the possibility of suction recirculation is greatly reduced at the exit area. Structural strength of the shrouds is a factor in the decision in how to trim the impeller. There may be too much unsupported shroud left after a major reduction in diameter. The

TABLE 1. RECOMMENDED RADIAL GAPS FOR PUMPS Type of Pump Design

Gap “A”

Diffusers Volute

50 mils 50 mils

Gap “B” +/- percentage of impeller radius Minimum 4% 6%

*B = 100 (R3-R2) R2 where R3 = Radius of diffuser of volute inlet and R2 = Radius of impeller

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The Pump Handbook Series

Preferred 6% 10%

Maximum 12% 12%

FIGURE 5. SHARPENING OF IMPELLER VANES Normal sharpening

Original thickness

Original outlet width New outlet width

I frequently encounter “vane-passing” frequencies during vibration analysis of a pump. What are some of the methods that can be used to reduce this problem?

Q:

Mill or grind away Max. sharpening Leave at least 1/8 ”

oblique cut leaves the shrouds unchanged and solves the structural strength problem as well as improving the exit flow pattern.

The most effective method of reducing vane-passing frequencies is to carefully maintain proper Gap “A” and Gap “B” clearances to reduce impeller-casing interaction. Sometimes, impellers manufactured with blunt vane tips cause disturbances in the impeller exit area and in the volute area by generating hydraulic “hammer” even when the impeller O.D. is the correct distance from the cut water (Gap “B”). Corrections can be achieved by two methods: 1. Overfiling: This disturbance may be partly or entirely eliminated by tapering the vanes by “overfiling” or removal of metal on the leading

A:

The Pump Handbook Series

face of the vanes (Figure 4). This technique has the additional advantage of restoring the vane exit angle to near that of the maximum impeller design (i.e., before the diameter was reduced). 2. Underfiling: Sharpening the underside of the trailing edge of the vane (Figure 5) can enlarge the outlet area of the liquid channel. This will generally result in about five percent more head near the best efficiency point, depending on the outlet vane angle. At least 1/8 inch of vane tip thickness must be left. Sharpening the vanes also improves the efficiency slightly. Where there are high stage pressures, you must sharpen the vanes carefully because the vanes are under high static and dynamic stresses. ■ Ed Nelson is a consultant to the turbomachinery and rotating equipment industries. He serves on the Pumps and Systems Editorial Advisory Board.

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CENTRIFUGAL PUMPS HANDBOOK

Examining Pump Capacity Problems BY WAYNE C. MICHELETTI We have a 1,200 gpm centrifugal pump that transfers water from a public reservoir up to our makeup reservoir inside the plant. The change in elevation is about 30 ft over a distance of roughly a mile. The pump does not operate continuously, rather it is turned on and off by plant staff who check the makeup reservoir level once per shift. According to operators, the pump seems to deliver ”full flow“ when first started, but is operating at a much lower capacity when checked later. What could cause this consistent decline in pump capacity?

Q:

Your problem could have a couple of causes. One cause might be air that has entered the system and accumulated in the pump. While it is possible for the reservoir water to be saturated with air that will come out of solution in the pump, most centrifugal pumps can handle a small amount of air (2–3% by volume), which will pass through as bubbles with the liquid. Instead, the introduction of air is more likely equipment related. Depending on the pump and piping system, air can get into the water in several ways. For the pump, check the shaft sleeves to ensure that the seal between the sleeves and the impeller hub is adequate. Then examine the stuffing boxes. For pumps operating with a suction lift, lantern rings should be installed and have seal water under positive pressure. Piping can be more difficult to examine because most of it will likely be underground. However, any surface piping should be inspected to assure that it is air-

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tight. And the as-built drawings should also be studied to determine if there might be any irregularities (such as improper pitch or high spots) along the pipeline in which air pockets could form. If air is the cause of reduced pump capacity, this can be confirmed by stopping the pump, opening and closing the vent valve on top of the volute, and immediately restarting the pump (which should run at full capacity). Do not open the vent valve while the pump is operating. Even if air is present in the pump during operation, it will be trapped near the center of the impeller while the heavier water will be forced to the outer edge (and out the vent valve if it is open) (see Ask the Experts, November 1993). A second possible source of your difficulty is the intake at the public reservoir. From the information presented, the system probably has a submerged offshore intake with some form of screening to prevent the entrainment of unwanted materials. Underwater plants, particularly filamentous grasses, can be drawn into and entangled in the intake screening, blocking flow. When the pump is not operating, the natural underwater currents can clear some or all of the blockage so that full flow is temporarily restored at pump startup. The intake can also contribute air to the system. If the level of the public reservoir has dropped, the distance between the surface of the water and the submerged intake might not be adequate to prevent the formation of vortices whenever the pump is operating. Such vortices can bring significant amounts of air into the system.

Q:

We recently decided to move a relatively old but infrequently used standby service water system pump to an auxiliary

The Pump Handbook Series

cooling water application. The centrifugal pump was rated at 1,000 rpm for a suction lift of 15 ft against an 80 ft total head when running at 750 rpm. Using the rules governing the relation of capacity, head, and speed, we thought it should be possible to obtain 1,400 gpm against a total head of 150 ft by replacing the original motor with a larger, 1,050 rpm motor. However, in its new service, the pump has not provided anywhere near the anticipated capacity. What could be wrong?

Switching a pump from one service to another frequently appears to be an easy and costeffective way of avoiding the purchase of a new pump designed for the desired use. Unfortunately, such switches can be tricky business (as discussed in this column in March, 1993). Yet since almost everyone will be tempted to engineer such a switch at least once during a career, it might be helpful to review key calculations that are needed in an effort to determine what went wrong in this case. Summarizing the information you provided, we have one pump intended to handle streams of comparable quality (basically cold water) that has been operated with two different motors. Knowing the original design capacity and total head, we can quickly determine the same information for the new application by the following equations:

A:

Q2 = Q1 x (N2/N1) and H2 = H1 x (N2/N1)2

where Q = pump capacity (gpm) N = motor speed (rpm) H = total head (ft) As you expected, the corresponding capacity and total head for the new motor (at 1,050 rpm) should be: Q2 = 1,000 x (1,050/750) = 1400 gpm 2 H2 = 80 x (1050/750) = 157 ft So far so good. According to these calculations, the pump should be able to provide the desired flow against the estimated head. But before a pump can transfer any fluid, the liquid must have enough outside energy to enter the pumping element at the velocity corresponding to the required pump flow rate. For a centrifugal pump, this energy must be great enough to make the fluid flow into the impeller eye with sufficient force to prevent the fluid pressure from dropping below its vapor pressure when passing the inlet vane edge. This outside energy requirement is known as the Net Positive Suction Head Required or NPSHR. Assuming that your system is at sea level, this value (for the original pump design) can be determined as follows: barometric pressure (abs.) = 33.9 ft – vapor pressure of water = 1.1 ft – suction lift = 15.0 ft NPSHR = 17.8 ft For centrifugal pumps, the NPSHR can be correlated to pump capacity and motor speed by a value known as the suction specific speed (S) according to the following formula: S = (N x Q0.5)/(NPSHR)0.75

Using this equation and the original pump design data (Q = 1,000 gpm; N = 750 rpm), the value of S is 2,737. Since the suction specific speed is constant for a given pump, this equation can be rearranged to calculate the NPSHR for the pump’s new application (Q = 1,400 gpm; N = 1,050 rpm). The increase in pump capacity and speed mean an increase in the NPSHR from 17.8 ft to 34.9 ft. As a result, the conditions of the new application correspond to a suction head as opposed to a suction lift: barometric pressure (abs.) = 33.9 ft – vapor pressure of water = 1.1 ft + suction lift = 2.1 ft NPSHR = 34.9 ft If the NPSH available (the difference between the absolute suction pressure and the liquid vapor pressure) is reduced below the NPSH required, then the pump capacity is reduced, and the pump is likely to cavitate. Unless you can change the suction conditions for the auxiliary cooling water application, it would be better to buy a new pump than attempt this switch. As the result of a recently implemented water management program, several of our older, constant-speed centrifugal pumps now provide significantly more water than is required. What is the best approach for operating these pumps at reduced capacity?

Q:

Congratulations. Many would envy the problem produced by your success in water conservation. Fortunately, there are three well-proven solutions to reducing existing pump capacity. If the system flow is expected to change frequently or irregularly, new adjustable-speed drives might be in order. For variable-torque applications (such as centrifugal pumps), solid-state AC or DC drives are usually best. They will allow the pump

A:

The Pump Handbook Series

to respond quickly and efficiently to lower (and higher) flow demands, enabling you to conserve energy as well as water. However, electrical adjustablespeed drives can be expensive and should be thoroughly evaluated from an economic perspective for ”older“ pumps. A second, more common approach is simply to throttle the discharge. Doing so will introduce a new artificial friction loss component to the head. This will shift the present system–head curve upward to intersect the pump head–capacity curve at a new operating point (corresponding to reduced capacity). It should also reduce the energy requirement slightly. It is never advisable to throttle the pump suction. This approach (occasionally referred to as operating in the ”break“) changes the pump head–capacity curve through cavitation. The resulting operation is not only inefficient but potentially damaging to internal pump components. The third option is more energy-efficient than throttling, but only suitable if a permanent reduction in pump capacity is acceptable. The pump impeller can be cut down, essentially lowering the pump head–capacity curve. However, before trimming an impeller, a number of other factors and resulting implications should be carefully considered. (These were discussed in this column in the January, 1993 issue.) ■ Wayne Micheletti is a water and wastewater consultant and a member of the Pumps and Systems Editorial Advisory Board.

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CENTRIFUGAL PUMPS HANDBOOK

Venting Pump Systems BY MICHAEL D. SMITH umps sometimes suffer damage unnecessarily because they are not 100% full of liquid before they are started. The systems in which they function either are not or cannot be completely vented. A common misconception is that a pump that produces discharge pressure immediately after start-up was sufficiently full of liquid. For some users, this is the working definition of the word “primed.” Igor J. Karassik, an internationally recognized authority on pump systems, has written for years about the need to remove all of the gas or vapor from pumps before starting them. Widespread understanding of the problems trapped gasses can cause developed during 1991 from an effort to understand erosion problems with enlarged, taperedbore seal chambers used on ANSI B73.1M chemical pumps. Testing, independently performed by and

P

for a number of pump and seal companies, showed that the liquid in the seal chamber circulates around the chamber at a large fraction of shaft rotation speed. A secondary flow was observed heading away from the impeller, along the outside diameter of the seal chamber, and toward the impeller along the shaft. Together, these two flow patterns explained how erosion damage was occurring in a few cases where abrasive solids were present (Figure 1). An unexpected byproduct of this testing was the realization that gas or vapor that is present in the seal chamber at the time the pump is started can be trapped there for several minutes by these same flow patterns. Worse, trapped vapors or gases tend to accumulate close to the shaft, near the rear of the seal chamber. For most single mechanical seal installations, that is where the seal faces are located. Here are some common questions on venting pump systems:

FIGURE 1

Q: A:

Why would gas end up close to the shaft?

Imagine the seal chamber is more than half full of liquid as the pump is started. As the pump shaft (and mechanical seal) picks up speed, viscous drag causes the liquid to begin to circulate around the chamber. Soon, centrifugal force overcomes gravity, and the liquid is thrown to the outside of the seal chamber. Any gas is forced inward by the denser liquid.

Q: A:

What problems are caused by the gas?

The most common problem is mechanical seal damage. If the gas bubble is big enough to surround the seal faces, it can prevent the liquid in the seal chamber from cooling and lubricating the faces. Large pockets of gas can damage wear rings and bushings, but gas would tend to be swept out of these areas quickly.

Q: A:

My pumps have flooded suctions. Won’t they fill completely when I open the suction valve?

Primary and secondary flow patterns can result in erosion damage.

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The Pump Handbook Series

Probably not! While it is true most most modern pumps are designed to be completely self-venting, there is an assumption that there is someplace for the gas to go as the liquid enters. Unless the discharge valve is opened slightly and there is no discharge pressure, the gas has nowhere to go. When a horizontal end-suction pump is installed (or re-installed after a repair), and the

suction valve is opened, it will often fill to the top of the suction pipe. When the gas (air, in this case) can no longer escape out the suction pipe, it will compress a small amount in response to the suction pressure. A very large gas pocket remains in the pump at this point, although the pump is probably “primed.”

Q: A:

Why has this cause of seal damage remained hidden?

A big reason why pockets of gas have not been a concern is that they don’t always cause an immediate failure. Seal face damage progresses each time the pump is started while it is not full. Venting is not an issue in many pump starts because the pump was not drained since its last use. If a pump seal fails about once a year, we assume it has a one-year wear life. We don’t even consider that it might be failing every third time the pump is started without being 100% full of liquid.

Q: A:

be capable of being completely vented. When the liquid can be released to the atmosphere, a vent valve is all that may be required. See the sidebar at the end of this article for a procedural solution to a common situation. While discussing system design, it should be noted that the suction line should not have any high points. The suction line should rise continuously either toward the pump or back to the source. If a local high spot is necessary, it will also have to be vented. I have seen many long suction lines that were designed to be level that still had local high spots several pipe diameters above the ends. This can be due to problems with the original installation or the shifting of pipe supports at a later time.

CONCLUSION Whoever has responsibility for the design of the “system” will need information on the pump, the piping, and the operating conditions to assure that it can be vented. ■ Michael D. Smith is a Senior Consulting Engineer at the DuPont Company in Wilmington, DE.

How can the problem be avoided?

Venting Your Pumps It is common to have venting problems when a pump is connected to a system that is pressurized even when the pump is not running. These systems often employ a check valve in addition to a discharge valve. Some users drill a small hole in the check valve flapper to help vent the pump, but this technique is not effective for the most common operating strategies. The following is a simple procedure that can be used to get more complete venting of these hard-to-vent systems. It assumes the pump is empty of liquid and both suction and discharge valves are closed. • Open suction valve (pump fills part way). • Close suction valve. • Open discharge valve part way (once pressure equalizes, air will rise into discharge piping). • Open suction valve. • Start pump.

CAUTION: The pump seal will be exposed to full discharge pressure using this procedure. Never start a pump with the suction valve closed.

The operator must understand why it is important to fill the pump completely. The pump system must

The Pump Handbook Series

45

CENTRIFUGAL PUMPS HANDBOOK

Installation and Start-Up Troubleshooting BY JOHN W. DUFOUR AND LYNN C. FULTON lot of time and money are spent manufacturing and testing centrifugal pumps and developing purchasing specifications for bidding and selecting them. However, events after leaving the manufacturer may result in a pump that won’t perform reliably or deliver the desired hydraulics.

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SHIPPING AND HANDLING Once the pump/driver/baseplate assembly leaves the factory, anything can happen if specific instructions on how it should be shipped, received, stored, and installed are not followed. A document that records what was and must be done, what must be approved and by whom, and when these events should happen is crucial. Without this, work will be missed or duplicated. Manufacturers prepare products for shipping differently. Some mount pumps in custom-made crates, while others hang the shipping tag on auxiliary piping and bolt two-by-fours to the base. The purchaser should define special requirements. Will the pump be shipped overseas? Is long-term storage required? Is there lifting equipment at the site? These questions must be answered ahead of time. In all cases, Material Safety Data Sheets should be included during shipping and installation. Everyone who comes in contact with the pump needs to know what’s in it. There are other questions, too. What form of transportation will be used? A dedicated truck or a common carrier? Who will receive the equipment? When? A dedicated truck usually has two drivers driving around the clock, directly from manufacturer to delivery site. This is costly but quick. A common carrier is less expensive but can take longer. For example, pumps from an East Coast manufacturer, destined for Texas, were loaded on a truck Friday afternoon. The pumps arrived 15 days later. With no one to receive them, the driver left them at a warehouse. It took two

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days to locate them, and they were delivered a week after that. Pumps are easily damaged during transportation, storage, or installation. Most baseplates are designed to be lifted with an overhead device or moved by fork lift. Care must be taken to prevent damage to auxiliary piping from lifting slings or hooks. Storage facilities often don’t have an overhead crane, so a forklift moves the assembly off the truck and around the storage area. Again care must be taken to balance the load before lifting and to avoid bumping or dropping the assembly (falling just an inch can crack the mechanical seal face ring). Never lift the pump by its shaft or auxiliary piping.

STORAGE Sometimes the pump goes directly from truck to foundation, but the assembly is often stored for a time. Storage may be a graveled yard or a warehouse with overhead lifting equipment and a controlled environment. In any case, following three rules will help avoid problems: 1.

Keep oil/grease in the bearings.

2.

Keep water/moisture out of the case (seal, windings, etc.).

Steam turbines often have carbon rings and seals. Remove them to prevent corrosion under the rings, or continuously purge the case with dry nitrogen.

INSTALLATION As mentioned, prepare a document to ensure proper installation. Outline specific requirements, in sequence, for each pump. Define tasks and inspections, who is responsible, and special procedures—grouting plans, cold alignment targets, pre-start-up checks, hot alignment checks, etc. Vendors often give details on installation, and writings on the subject are available. Here is a list to aid installation:

GROUTING •

Prepare the foundation surface. Chip latence off, exposing aggregate. Remove loose material, grease, and water.

•

Level the baseplate using jackbolts bearing on jackplates (Photos 1 and 2). Jackplates should have rounded corners. It’s easier to slice sections from round stock than to cut plate.

•

Remove pump and motor before installation; it’s easier to level the baseplate and pour grout.

•

Check the baseplate bottom for cleanliness. Verify that each compartment has grout and vent holes. Drill holes before lifting the baseplate onto the foundation.

•

Don’t grout around anchor bolts. Baseplates are grouted to provide uniform load distribution. Anchor bolts hold the pump down. To keep anchor bolts free to stretch, install sleeves around bolts.

•

Install the baseplate, establishing correct elevation (within 1/8 in.) and pedestal level (within 0.002 in./ft). Some contractors like to put pumps back on the baseplate to shoot the nozzle elevation.

3. Protect the pump from abuse. Check the pump over. To prevent baseplate distortion, place it level and out of traffic. See that all cover plates are bolted on. Be sure no auxiliary piping or components were lost or damaged in transit; replacing a part may delay start-up. Bearing housings should be filled with oil to the bottom of the shaft and rotated periodically to keep bearings coated. Document who turns it and when. Pumps stored longterm with oil mist lubrication should be hooked to a portable mist generator. Verify that the mechanical seal sleeve locking collar is tight and that the shaft turns freely. Stored drivers may require extra care. Heaters on electric motors should be energized to keep windings dry. The Pump Handbook Series

the DSBE set with the motor rotor in its magnetic center.

This is unnecessary and may distort the baseplate. •

Coat forms with furniture paste wax to ease removal. Fix forms to the foundation block at different elevations to avoid fracture lines from anchor studs. Drilled holes with screws look better after removing forms and eliminate potential impact cracks from hammering nails or using charged drivers.

•

Tape or grease machined mounting surfaces for protection.

•

Ensure that grout flows into all compartments by using a head box and vent tubes. The head box can be six-inch sonotubes RTV’d to the baseplate surrounding the pour holes. Vents can be plastic pipe. These should be at least six inches high to provide enough head to get all voids under the baseplate.

•

•

•

Grout between 60 and 90°F (Photo 3). Cooler temperatures don’t allow curing. Higher temperatures may cause fast curing and heat cracks. Grout should harden in 24 hours. As soon as the grout firms (not hardens), remove vent pipes and head boxes. Grout consistency should be like hard rubber, making it easy to trim. Forms can usually be removed after 48 hours. Remove jackbolts from baseplate and fill holes with RTV.

MOUNTING/ALIGNMENT •

Set pump on its pedestals, center bolts in their holes, and snug. This allows movement if side-toside motor adjustment can’t achieve alignment.

•

Mount motor with a minimum of 1/8 in. stainless shims under the feet using the required distance between shaft ends (DBSE). This is usually found on the general arrangement or coupling drawing. With sleeve bearing motors, the magnetic center of the motor with respect to the stator must be determined and

•

Make sure the mechanical seal drive collar locking screws are tight, then roll locating cams out of the drive collar. Lock cams out of the way or remove them. Remember that future work will require cams to reset seal compression—don’t loose them. The shaft should turn freely.

•

Align motor to pump, free of any piping, using, as a minimum, the reverse indicator alignment method. To avoid soft foot, minimize shims under each support area. When alignment is achieved, tighten holddown bolts and recheck.

•

To check for soft foot, place a dial indicator on each mounting foot, then loosen the hold-down bolt. If the reading changes more than .001 in., reshim.

PIPING Care in fabricating and aligning piping avoids problems that may require recutting, fitting, rewelding, and retesting the pipe or lead to premature pump failure. Good system design supports piping loads and forces along spring hangers and bracing that don’t have to be removed during normal maintenance. The system should be fabricated starting at the pump flanges, working toward the pipe rack, using temporary braces/supports to avoid pump strain. The most common piping fabrication error, producing the largest piping strain, is nonparallel flange faces. A feeler gauge helps detect this. If you see a difference in two facing flange planes, piping strain will result. For example, during installation of circulating water pumps in a refinery, suction piping was forced to the pump flange without checking for non-parallel faces. The resulting strain distorted the casing to the point where the shaft and impeller would not turn. Fortunately, no serious damage occurred. The cases were reclaimed after the piping was aligned and supported.

The Pump Handbook Series

•

After fabrication and pipe testing, remove temporary bracing and lock-pins from spring hangers and check strain.

•

Remove flange covers and inspect the pump for debris. Clean out the case. Bring the piping to the pump flanges. Flange holes should drop through with no binding.

•

Place dial indicators to monitor vertical and horizontal movement of pump shaft relative to driver shaft. Make up suction and discharge flanges separately, continuously observing indicator readings. If movement exceeds 0.001 in., piping strain is excessive. Readjust pipe, retighten, and retest.

PREOPERATIONAL CHECKS The period from installation until full operation may be the most important phase of pump life. It’s filled with activity and riddled with pitfalls that can complicate start-up and prevent establishing a reliable system. The rules above also apply here: 1.

Keep bearings lubricated.

2.

Keep moisture out of the case.

3.

Protect the pump.

Drain and flush bearing housings with clean oil. Oil rings may have moved during handling, so look through the vent caps to verify that they’re in position. On oil mist installations see that mist reclassifiers have been installed correctly. Directed oil mist fittings have a “V” at the orifice. This must be pointed towards the bearing. Insure that all mist lines are sloped so no low points cause liquid buildup and block flow. Greased bearings should be repacked with the correct grease. Make sure all old grease is displaced by new. Different greases (lithiumvs. soda-based) have incompatible additives. Mixing two greases can give an inferior blend. Bump check motors for proper rotation. Do not attempt this while the motor is coupled to the pump. Reverse rotation can cause the impeller to loosen or come off the

47

charge piping. If there are shaft. If rotation is correct, run the leaks, return to a safe situation motor alone for at least one half hour. and repair them. If leaks occur Monitor bearing temperatures and around the shaft, determine if motor bearing housing vibration. This seal faces are leaking or if the should reveal any major problems. leak is under the seal sleeve. Most others will not be revealed until Stop leaks between sleeve and it is loaded and generating heat. shaft by adjusting the drive colInstall the coupling spacer and guard lar. Stop leaks around the seal and verify smooth assembly rotation. flange by retorquing the bolting Don’t overlook small steam turto clamp the stabine drivers. Verify rotationary gasketing. tion direction by inspecting It’s important to nozzle orientation. As soon know where the as steam is available, Good system leak is before check the operation of the pulling the pump governor and overspeed design apart. systems. Run the turbine supports solo at least one half hour. • If no leaks are Mechanical seals and seen, open the piping loads bearings are easily inlet valve 100%. destroyed during initial and forces Vent areas of the start-up on hot pumps system that don’t where water is circulated along spring self-vent. Crack for cooling. Install pressure the discharge hangers and gauges, temperature indivalve. Start-up cators, and valves so water bracing that horsepower is flow can be regulated and minimum to the adjusted. Throttling valves don’t have to left hand side of are typically installed on the pump curve. be removed parallel outlet lines to On systems pumpadjust flow to each pump ing higher specific during normal skid area. It’s a good idea gravity liquids at to flow most cooling water maintenance. start-up than durthrough the seal and bearing normal operaing coolers initially. tion (typical of Debris in the pump cold start-ups), the and seals is a problem durdischarge system may have to ing initial start-up. A welding rod be throttled to avoid motor lodged in an impeller eye can seize overloading. Throttling to 50% the pump. To prevent this, use sucBEP is acceptable in most cases, tion screens. Insert temporary stainbut more than that may cause ers if they’re not built in. Pressure seal problems. gauges on both sides of the strainer • Start the pump. Slowly open indicate when it is plugging. the discharge valve. If a disINITIAL OPERATION charge control valve is installed The electronic equipment adage, and on automatic, the control “If it works the first hundred hours it valve will be wide open until should work a lifetime,” also applies the block valve opens enough to centrifugal pumps. Knowledge of for the control system to take the equipment and the system it will over. If the pump cavitates, operate in are key to successful startthere may be too much flow. up. While each installation is differStart to pinch down on the disent, this general procedure will help charge valving, preferentially prevent problems: using the control valve. Most systems have a flow meter. • Close case drains and vents. Flow can sometimes be deterSlowly open the suction line. mined from the meter directly, Look for leaks at the case and the differential across the flanges, seal area, seal piping, pump can be determined using drain piping, and inlet dis-

48

The Pump Handbook Series

pressure gauges on the pump. Using flow and differential head, determine where the pump is operating on the curve. Low flow, high head may indicate running too far back, leading to bearing or seal problems. High flow, low differential head means the pump is running out on its curve and could cavitate. Check differential across the inlet screen and use a spare pump before low suction pressure causes cavitation. •

Where flow can’t be measured directly with a meter, estimate it using motor current, horsepower requirement of the pump, and plotting that point straight up to the performance curve. The intersection of the vertical line from the horsepower curve to the performance curve should be the capacity point as long as specific gravity is similar to horsepower curve specific gravity. Differential head on the pump should be similar to differential head on the curve at the capacity point determined from the horsepower calculation.

From Figure 1: Example point 1.

P2 – P1 TDH = 2.31 —————— = 188 ft S.G. where P1 = 3 psig P2 = 60 psig S.G. = 0.7 √3xIxVxη M.H.O. = ——————— = 26.5 Hp 746

where M.H.O. = motor horsepower output V = motor voltage = 460 V (30 Hp motor, 3φ, 460 V [line to line])

FIGURE 1 250 Point 2 Point 1

150 100

40 30 20

50

BHP

Total Head (ft.)

200

John Dufour has more than 20 years of experience working with mechanical equipment. He is Chief Engineer, Mechanical Equipment Services, for Amoco Oil Co, and is responsible for specification, selection, installation, and consultation for rotating equipment throughout the company’s refinery, pipeline, and marketing operations. He holds bachelor’s degrees in metallurgical engineering and engineering administration from Michigan Technological University. Mr. Dufour also serves on the Pumps and Systems Editorial Advisory Board.

10 0 0

100

200

300 Flow (gpm)

400

500

0 600

Head and BHP vs. Flow. Operating point 1: using the BHP vs. flow curve with horsepower calculation derived from measurement of current with voltage assumed to be 460 V, flow is found to be 405 gpm. The calculated 188 ft based on pressure differential confirms flow to be 405 gpm. Operating point 2: similar calculations for horsepower and head at operating point 2 also confirm the calculation method. See text under “Initial Operation” for calculations.

η = motor efficiency = 90% (2pole motor, 90% efficiency)

V = motor voltage = 460 V (30 Hp motor, 3φ, 460 V [line to line])

I = phase amp measurement = 27.2 amps

η = motor efficiency = 90% (2pole motor, 90% efficiency) I = phase amp measurement = 27.2 amps

Example point 2. Throttling pump discharge

P2 – P1 TDH = 2.31 —————— = 214.5 ft S.G. where P1 = 3 psig P2 = 68 psig I = 20.8 amps

Lynn Fulton is a professional engineer registered in Indiana and Illinois. He has been with Whiting Engineering more than ten years, in mechanical services and maintenance. He has a bachelor’s degree in mechanical engineering from the University of Illinois at Chicago and is chairman of the Chicago chapter of the Vibration Institute.

•

Check motor and pump vibration. Vibration levels should be below 0.15 in./sec. Most new equipment vibrates less than 0.1 in./sec. true peak.

•

Compile documents for each pump and file them for reference. ■

√3xIxVxη M.H.O. = ——————— = 20 Hp 746

where M.H.O. = motor horsepower output

The Pump Handbook Series

49

CENTRIFUGAL PUMPS HANDBOOK

Upgrading Utility and Process Pumps Improvements to make your equipment better than new. BY KURT SCHUMANN pgrade (up + grade, v.): to raise the grade of; to raise the quality of a manufactured product (Webster’s Third New International Dictionary). A pump upgrade (also called a revamp or retrofit) involves changing mechanical or hydraulic design or materials to solve a problem or increase reliable run time. An upgrade is different than a repair, which attempts to duplicate original construction and design, whereas an upgrade improves the design beyond the original. Rerates are a type of hydraulic upgrade, usually involving a change in pump head capacity. Repowering may involve repairs and/or upgrades. Philosophically, repowering is different from normal pump maintenance because the plant being repowered has decided to spend capital monies to extend the plant’s useful life. Plants being repowered are candidates for pump upgrades because they are expected to run reliably with high capacity factors and can justify the additional cost (above and beyond normal repairs) to upgrade pumps. Pump upgrade goals include: • decreasing plant operations and maintenance expenses • increasing mean time between failures (MTBF) • increasing pump and plant availability • increasing pump efficiency • complying with the latest legislative mandates (such as the Clean

U

50

FIGURE 1

Original residual heat removal pumps in safety service at nuclear power plants. The design has high maintenance hours and exposure dosage due to short mechanical seal life and an overhung shaft design that makes seal maintenance difficult.

Air Act Amendments of 1990) • minimizing the risk of fire or other safety hazards • eliminating hazardous materials Pump upgrades can be divided into major categories: • mechanical design • hydraulic design • material • ancillary/system

The Pump Handbook Series

THE UPGRADE PROCESS To identify upgrade candidates, pump users should review maintenance records to see which pumps were responsible for a disproportionate share of expenses or caused safety or reliability concerns. Once these pumps are identified, work with the upgrade supplier to identify and evaluate upgrades available for your particular pump. Provide the supplier

- replace nickel-aluminum-bronze parts with austenitic stainless steel - use other special alloys for critical parts - install non-metallic bearings Ancillary/systems upgrades: • install vibration monitoring and recording instrumentation • improve lube oil system and instrumentation • modify seal injection • perform pump intake scale model testing The following are examples of upgrades to improve pump operation:

FIGURE 2

RHR PUMP COUPLING MODIFICATION

An upgrade of Figure 1. Spacer coupling between pump and motor allows seal access without disassembling the pump. A bearing above the seal limits shaft deflection. Conversion from non-cartridge mechanical seal to cartridgetype seal eases assembly. Benefits: increased seal life, decreased leakage, and decreased personnel exposure while changing seals. with a maintenance history so problem areas can be addressed. Plan for a future outage where the plant or process will be shut down long enough to complete design and hardware changes. This is important—upgrades take time to engineer and implement, and they must be planned in advance based on repair or outage schedules. This approach can also save money; upgrading worn out parts during the normal repair cycle (instead of replacing components that still have life left) minimizes incremental cost. The following are upgrade examples from each of the areas above: Mechanical design upgrades: • install a stiffer shaft/rotor to reduce vibration • modify structural elements to remove natural frequencies from the range of pump forcing frequencies (rotational frequency, blade pass frequency, etc.) • eliminate threads (a source of breakage) on pump shafts • modify components to make assembly/disassembly easier • convert mechanical seals to further restrict or eliminate leakage

Hydraulic design upgrades: • redesign first stage impellers to reduce cavitation damage • redesign impellers to lower vibration for part load/peaking operation • control “A” and “B” gaps to reduce pressure pulsations and vibration • improve efficiency • optimize blade number to reduce pressure pulsations and vibration • increase pump head capacity to meet system requirements Material upgrades: • eliminate asbestos, an environmental hazard • install impellers made of cavitation-resistant materials for longer life • use hardened wear parts to increase MTBF • eliminate leaded bronzes because of environmental problems with lead • improve product-lubricated bearing materials • change materials for seawater use - replace Monel components with austenitic stainless steel The Pump Handbook Series

These pumps are residual heat removal pumps, close coupled design in safety service at nuclear power plants (Figure 1). The original design resulted in considerable time, expense, and man-rem exposure for normal maintenance activities like seal change-out and motor thrust bearing replacement. There was also a high risk of damaged equipment from the difficulties of rigging in cramped quarters. Pump upgrade kits add a bearing and a spacer coupling. Installation of these kits allow seal removal without pump disassembly. The original design frequently resulted in seal change-out times longer than the 72 hours permitted by most plant safety evaluations. The upgrade easily accommodates a seal or motor bearing change in 72 hours, without the high man-rem dosage involved in pump casing disassembly. This reduction in personnel exposure is an important benefit in any case, but it is especially so given the increased industry focus on the issue. These coupling modifications have been supplied to several utility companies as bolt-on hardware kits installed during short outages (1 or 2 weeks). Other items like oil drain location and mechanical seal venting have also been improved in the design.

BOILER FEED PUMPS Boiler feed pumps are at the heart of most power plants, and economical plant operation depends on reliable pump operation. Many pumps from the utility building boom of the 1950s–70s had larger capacities

51

FIGURE 3 6 5

7 4 8 3

2

9

1

10 Boiler feed pump upgrades. See text under “Boiler Feed Pumps” for details. and more horsepower than previously supplied in order to meet increased plant size. As pump energy levels increased, so did the failure rate. The Electric Power Research Institute (EPRI) studied feed pump problems and proposed corrective actions, including ”A“ and ”B“ gap control, developed by Dr. Elemer Makay. Additional progress has been made through pump manufacturers’ efforts to address first stage impeller design, materials, and rotor dynamics. As a result, good reliability of high-energy feed pumps is attainable. For example, a power plant converted from nuclear to fossil cogeneration recently installed 70,000 Hp boiler feed pumps. These pumps benefitted from upgrades made to utility pumps over the last 10 years. Impeller life was increased by hydraulic

52

redesign and a material upgrade to reduce cavitation damage. The suction ring was modified to prevent fatigue cracking. Stuffing box seals were switched to labyrinth-type seals, desensitizing the seal area to upsets. The impeller to diffusor A and B gaps were modified to current standards to reduce vibration and pressure pulsations. The bearing design was changed from a plain journal to a more stable tri-land design. All of these upgrades have been incorporated into new pump design philosophy. Some of the common upgrades for boiler feed pumps include (Figure 3): 1. The thrust collar nut can be modified to reduce shaft bending stresses and the possibility of shaft fatigue failure. This also minimizes run-out on the critical face of the thrust collar. The Pump Handbook Series

2.

3.

4.

5.

The bearing housing can be modified to spring-load the thrust bearing. This opens the critical balance drum to balance sleeve clearance at low speeds and reduces the potential for contact and seizure. Several sealing options are available, depending on the application. Floating seal rings, serrated bushings, or mechanical seals are often recommended to minimize first cost while maximizing reliability and thermal efficiency. Non-asbestos spiral-wound gaskets replace original asbestos gaskets. Pumps with copper or iron gaskets can be modified to accept non-asbestos spiral wound gaskets to prevent assembly and leakage problems. ”A“ gaps, ”B“ gaps, and ”C“

overlap can be modified to current standards. By evaluating the cost of these modifications in light of expected benefits, users can choose the modification required to meet process requirements with the least cost. 6. Improved first-stage impeller inlet designs expand the stable operating range to lower flow rates without cavitation damage, vibration, or pressure pulsations. Material upgrades for first stage impeller service resist cavitation damage while maintaining ductility, corrosion resistance, and weld repairability. 7. Dry couplings (like the flexible disc or diaphragm-type) eliminate the need for periodic lubrication and the associated chance of failure. They are often lighter than gear couplings, resulting in lower vibration, and they tolerate more misalignment than geartype couplings. 8. Instrumentation can be added to monitor and protect pumps. Possibilities range from simple vibration/temperature switches to complete monitoring of all key operation variables, including remote monitoring, diagnostics, etc. The most common items monitored include: a. shaft or bearing cap vibration b. lube oil temperature and pressure c. axial and radial bearing temperature d. casing (or barrel) temperature Additional items include: e. pump suction condition (pressure, temperature) f. pump discharge condition (pressure, temperature) g. pump flow rate h. horsepower, efficiency i. balance drum leak off (temperature, flow rate) j. seal (drain temperature, mechanical seal face temperature, stuffing box temperature, etc.) Monitoring can be stand-alone or can feed into the plant’s control system. 9. Some older pumps have open vane diffusors. Vanes can fatigue

FIGURE 4

due to unsteady hydraulic loads. Using a shrouded diffusor eliminates breakage problems and allows improved alignment. 10. Improved bearing designs are available, including ”high stability“ designs to eliminate half frequency (”oil whirl“) problems. Special attention is paid to individual plant operating modes (low-speed operation, turning gear, etc.) in recommending a particular bearing design.

9 1

CIRCULATING WATER PUMPS Circulating water pump maintenance requirements vary greatly, depending on whether the pumps are used in freshwater or seawater. For most freshwater applications, typical problems requiring pump maintenance are excessive vibration and premature bearing wear. Vibrations can be analyzed using modal analysis or standard spectrum analysis techniques to identify the root cause of the vibration. If necessary, a finite element analysis (FEA) model of the pump can be built and correlated to the field data to verify the cause. It can also be used to help redesign the pump. To improve bearing and sleeve life, upgrades are available to increase wear resistance through material selection and hardcoating. Circulating water pumps in seawater face additional problems due to corrosion. Material selection is critical, and the selection process must consider general corrosion as well as velocity effects, galvanic compatibility, and pitting resistance, plus manufacturability and cost. The cost difference between materials can be significant because these pumps are large; care must be taken not to over-specify materials and inflate the price of equipment for marginal benefits. In some cases, lower-cost material may be more reliable. For example, 316 stainless steel has better pitting resistance than Monel in seawater, yet Monel is more expensive. Upgrades for circulating water pumps include (Figure 4): 1.

An inner column stop on pullout style pumps to hold down the pump element during startThe Pump Handbook Series

2

8

3

7 4

6 5

Circulating water pump upgrades. See text under ”Circulating Water Pumps“ for details.

2.

ing, stopping, unit trips, and other transients. A flanged inner column with rabbet fits replaces screwed inner columns, resulting in better bearing alignment and easier disassembly.

53

8.

FIGURE 5 New Design

Original Design Leak-Off to Low Pressure Heater (Deaerator)

Keyed shaft coupling improves shaft alignment and eliminates problems associated with removing threaded couplings. Shafts can usually be remachined and re-used. 9. Rabbet-fit drive couplings replace body-bound bolt couplings and improve alignment repeatability.

BOILER CIRCULATING PUMPS Auxiliary Fill Connection

Injection from discharge of boiler feed pump Throttle Bushing

Right: Original boiler circulating pump design. The throttle bushing and throttle sleeve wear quickly, reducing floating seal ring life and shortening service intervals between pump rebuilds. Left: The upgraded design. The retrofit incorporates a water-lubricated carbon bearing and redesigned floating seals.

3. 4.

5.

6.

7.

54

Bearing spiders provide stiffer bearing support. Shroud metallurgy upgraded in high-velocity areas eliminates erosion and corrosion damage and extends efficient pump life. Inlet bell modifications lower the required submergence and reduce vortexing. Intake studies can be performed to correct vortexing and other inlet problems and give uniform flows to the pump, resulting in stable operation. Modified impellers optimize cooling water flow, increase plant output, or save pumping horsepower. Upgraded impeller materials resist erosion, corrosion, and cavitation damage. Erosion in iron casing vanes can be repaired. Coatings can be applied to extend casing life.

Boiler circulating pumps (BCPs) are in a particularly severe duty, handling 600°F water at over 3,000 psig. Many of the original pumps supplied in the 1950s and 60s exhibited less-than-desirable life spans. In light of this, an upgrade program was developed that: • adds a graphite-impregnated bearing • improves the primary sealing device • incorporates other reliability ”lessons learned“ (Figure 5) To date, more than 200 BCPs have been upgraded to this new design, and the resulting MTBF is typically two to three times that of the pumps before upgrading.

API PROCESS PUMPS

Another option, if extremely low levels of emissions are required (for instance, pumping benzene, a carcinogen), is to use sealless (magnetic drive) technology. This can be accomplished by repowering (reusing the casing, bedplate, and driver, along with upgraded pump internals) or replacing the whole pump.

SUMMARY These descriptions cover some typical upgrades. This article focused on specific types of pumps, but upgrades are available for most models and sizes. Pump companies are useful resources for aid in problem solving. They are usually anxious to apply new technology and gain field experience with new designs and materials. Most upgrade suppliers can customize upgrades for individual users. Review your maintenance problems and discuss them with your pump supplier. Pump upgrades are a cost-effective way to improve plant performance within budget constraints. When upgrades are properly performed, an upgraded pump may well be ”better than new.“ ■ Kurt Schumann is Manager of Pump Upgrades for the Engineered Pump Group of the Ingersoll-Dresser Pump Company, located in Phillipsburg, NJ. He has 18 years of experience in design engineering and field service of utility and process pumps.

Process pumps may handle hazardous materials, and as a result seal leakage is critical. Industry standards (API 610 7th Edition), as well as federal legislation like the Clean Air Act, address mechanical seal reliability and pump maintenance. There are upFIGURE 6 grades available for A-Line 7th Edition Upgrades process pumps to Large Seal Chamber improve seal reliabili- Reduced Shaft TIR Steel Bearing Cartridge Seal ty and reduce emisHousing sions (Figure 6), including: • using a heavyduty rotor (shaft and bearings) • enlarging the stuffing box bore Cast Iron Bearing for better seal Non-Cartridge Housing cooling Seal Small Seal Chamber • using a heavyExisting A-Line (6th Edition) duty cartridge seal (single, double, or Pump modification kits upgrade API 5th and 6th edition process pumps with the features of the 7th edition tandem) The Pump Handbook Series

CENTRIFUGAL PUMPS HANDBOOK

Variable Speed Pumping Variable speed pumping can save you money if you select and use systems wisely. BY STEPHEN MURPHY ost users operate their centrifugal pumps at a fixed speed and accomplish any required changes of flow by using a throttling valve. This practice is much like driving an automobile with the accelerator fully depressed and changing speed by stepping on the brake! There is a better way to drive an automobile and there is a better way to accomplish variable flow for a centrifugal pump. Variable speed motors and associated electronic drives can be used to adjust pump speed to produce exactly the desired flow and head. By varying the speed of the pump, users can enhance performance, save energy, eliminate the need for throttling valves and reduce inputs of heat to the pumped liquid. But to achieve these advantages, you must properly select the components of a variable speed system. And proper selection requires a thorough understanding of pump, motor and driver designs for variable speed operation.

M

BEHAVIOR OF VARIABLE SPEED PUMPS A good place to begin a discussion of variable speed pumping is the interaction between variable speed pumps and the fluid handling system. These interactions are different from those of a fixed speed pump. For a fixed speed pump with flow controlled by a throttling valve, process demand depends on system back pressure and piping resistance, as shown by a fixed system curve (Figure 1). Pump performance is also

For a variable speed pump, flow is changed by varying speed. The variable speed pump retains its characteristic performance curve shape, changing flow and head in accordance with the well-known affinity laws (Figure 2). With varying speeds, pumps have wide rangeability and thus any headflow combination within the envelope can be achieved. And with appropriate precautions, pumps can be operated at even higher or lower speeds than those shown on the curve. The shape of the system curve influences the amount that flow will change with a change in speed. Flow FIGURE 1 is proportional to speed if no static lift exists but not proportional to speed if static lift exists (Figure 3). In systems with static lift, a minimum speed exists below which the pump will produce no flow. Such behavior does not violate the affinity laws. It simply reflects the interaction of the shape of the system curve with those laws. In fact, it’s this interaction that makes variable speed pumping advantaFixed speed centrifugal pump operation geous (which also illustrates that users must understand these interactions). is delivered to the system. The additional head (H3 - H2) is wasted across the valve in the form of heat and noise. represented by a fixed curve. With the discharge throttling valve fully opened, the pump seeks equilibrium with the system (point 1 in Figure 1: flow = Q1 and head = H1). To change the flow to Q 2, the throttling valve is partially closed, changing the steepness of the system curve as seen at a point between the pump and the valve (at B-B in Figure 1). Closing the valve causes the pump to “run back” on its curve to point 2, producing flow Q 2 as desired. The pump, which can only operate on its fixed curve, produces head H3 at point B-B. The pump thus produces H3 at Q2 but only H2 at Q2

The Pump Handbook Series

55

FIGURE 2

Variable speed centrifugal pump

BENEFITS OF VARIABLE SPEED PUMPING Because variable speed pumps can produce a desired head and flow over a broad range of hydraulic conditions, users do not have to be as certain of required flow when they select a pump. Instead of finding the exact fixed speed pump for the job, they can install a variable speed pump and adjust the speed to produce the exact conditions they require. For example, one user required Pump A to produce 125 gpm flow at 2500 ft head in an upset condition and 100 gpm at 1500 ft under normal con-ditions and Pump B for a 125 gpm flow at 1500 ft head under normal conditions. The user needed an installed spare for each pump, for a total of four pumps. But by specifying variable speed pumps, the user required only three pumps: one for each duty level and a single spare which was valved to allow operation under either condition. Further savings were achieved for the main pumps since identical pumps were used (desired conditions were met by varying the speed). Parts were interchangeable and significantly less energy was required when running Pump A at the normal (i.e., low-head) condition. In addition to covering a wide range of conditions, variable speed pumping can also eliminate the need for multiple stages. With

56

increased speed, centrifugal pumps produce increased head and flow. As mentioned above, variable speed pumping can also eliminate the need for a throttling valve. Also, bypass valves may no longer be necessary since minimal flow requirements for stable operation decrease with speed. Elimination of valves can reduce capital expense, maintenance costs, risk of leakage and pressure losses (pressure drop across the valve often accounts for 10 percent of total pressure operation rise required). One user saved $20,000 by converting to variable speed pumping in an application involving injection of water into the combustion chamber of gas turbine engines. Since the system curve had relatively little static lift, the pump could be slowed to produce only the desired flow and head and still maintain good efficiency. A change from a fixed speed pump with throttling valve and bypass valve to variable speed eliminated the two valves, reduced the power requirement of the system from 100 hp to 75 hp and made the assembled skid of equipment smaller. Dramatic power savings are available because of reduced head and

FIGURE 3

flow points due to changing speed rather than by dis-charge throttling (Figure 4). For instance, by achieving 60 percent of design flow and head through variable speed, users can save 50 to 80 percent on energy costs compared to fixed speed pumping with a throttling valve. Another advantage variable speed pumping offers is reduced heat to the pumped fluid. At constant speed, efficiency falls with reduced flow rate. The result of hydraulic inefficiencies is heat rise in the fluid. But variable speed pumps remain efficient at low flows (i.e., low speeds). Furthermore, horsepower levels are lower at low speeds, which means that heat input to the fluid is kept minimal. Variable speed pumping can thus be advantageous for light hydrocarbon and other volatile fluid applications.

SELECTING THE RIGHT SIZE PUMP Like any pumping application, variable speed pumping requires proper sizing of pumps. But unlike constant speed pumps, variable speed pumps are not selected for a single design point. To select the correct size pump, you should construct the desired head versus flow range for all anticipated specific gravities. Then be sure to specify a pump that can cover that range (Figure 5 shows a pump that cannot reach point B).

FIGURE 4

FLOW PROPORTIONAL TO SPEED - NO STATIC LIFT

Effects of changing speed of a centrifugal pump

The Pump Handbook Series

Hydraulic HP savings for a centrifugal pump

MOTOR-VARIABLE FREQUENCY DRIVE BEHAVIOR One of the most common methods of changing motor speed is the AC Variable Frequency Drive (VFD). VFDs are designed to take advantage of the fact that speed, torque and horsepower of an AC motor are all related to the frequency and voltage of the electric power supply: Nominal speed

2 x hz x 60 # of Poles

Torque Capability = F(volts/hz) HP Capability = f(Torque x Speed) VFDs convert incoming AC electrical power to DC then invert the DC power into variable frequency and voltage AC power. A number of technologies are available to switch the DC power through semiconductors to achieve the desired voltage or current pulses. The technologies differ in their ability to create optimal waveforms. Because the motor’s torque and torque ripple are determined by the current, the VFD affect motor and pump operation. Thus, by knowing the characteristics of the VFD output, you can select a VFD suitable for your pump. Most VFDs produce a constant volt/hz ratio, thus constant motor torque capability up to name-plate frequency (typically 60 hz or 3550 rpm for a two-pole motor — see Figure 8). Horsepower capability therefore rises from zero at zero speed to full horsepower at nameplate speed. Above nameplate speed,

the VFD cannot provide FIGURE 5 increasing voltage, so torque 120 falls due to the falling volts/hz 100% SPEED - P UM ratio. Horsepower capability, P1 however, remains constant 100 "A" DESIRED BEP since speed is increasing. RANGE OF LIMITS OF HEAD & FLOW Electrically, induction motors CAPABILITY 80 PUMP1 can be run at approximately CAN'T 90 hz in this configuration. DO 60 But mechanical constraints "B" may limit the safe running 40 speed to well below 90 hz. VFDs can be used to pro20 vide extra motor horsepower above 60 hz. Recall that motor 0 torque capability is propor0 20 40 60 80 100 120 140 tional to the volts/hz ratio. If a FLOW% motor is designed for a given PUMP 1 SIZED FOR "A"-UNABLE TO DO "B" volts/hz ratio, and that ratio can be maintained at a higher Improper sizing to meet required duty points speed, torque capability will be constant. VFDs. High efficiency is not a This technique can frequently be requirement, but the extra copper used with standard motors which are and other features are advantacommonly wound for either 230 V or geous for VFD use. 460 V at 60 hz. By connecting for 230 Increased heat can lead to enviV at 60 hz and operating to 460 V at ronmental hazards. Motors pro120 hz, both motor and horsepower posed for use in hazardous (e.g., capability and speed are doubled. Be explosive) environments must be sure to check with the motor manudesigned differently or derated. facturer before using this technique. The skin temperature of a standard The motor may not have the thermal motor operating on a VFD could capacity or mechanical integrity to exceed an area gas autoignition run at speeds considerably above 60 temperature at nameplate horsehz. Also, the motor may not be proppower. Motors nameplated for use erly matched electrically to the VFD. in Class I, Division I, Groups C and SELECTING THE MOTOR D environments, for example, are VFDs are most frequently used with available for VFD use but must the familiar NEMA B squirrel cage AC generally be purchased with a induction motors. Some special “matched” VFD from a single supconsiderations for selecting motors plier. for use with VFDs include cooling, SELECTING A VFD efficiency and operation in hazImportant factors for selecting ardous (e.g., explosive) environVFDs include power supply voltage ments. and frequency, amperage requireMotors operated on VFDs operments, torque requirements and ate at higher temperatures due to motor and load characteristics. the irregular shape of the electrical VFDs must be selected to match waveforms produced by the VFD. the power supply and frequency. To ensure that the motor will not Many VFDs are switch selectable overheat, the motors are typically for a number of voltage/frequency derated at full load from 3 to 10 combinations. percent, depending on the type of You can determine the amperVFD used. age requirement of a motor using This additional heat makes the equation: motors operated on VFDs less efficient than when operated across the line. Thus, many users specify high efficiency motors for use with HEAD%

You may need to specify a “fictitious” 100% speed point to ensure the pump has adequate range (Figure 6). You must also ensure that NPSHA and motor horsepower are adequate for all combinations of flow and speed. NPSHR and efficiency vary approximately as the square of the speed (Figure 7). Since NPSHR increases with speed, in-ducers may be required to reduce NPSHR to available levels. Bearing loads and other pump characteristics must also be carefully examined.

The Pump Handbook Series

57

FIGURE 6 120 "A"

100 DESIRED RANGE OF HEAD & FLOW

"C"

D-P U

MP

60 68% SP EED

40

2 "B"

PUM

P2

LIMITS OF CAPABILITY PUMP2

20

0

0

20

40

60

80

100

120

140

FLOW% PUMP 2 SIZED FOR "C"-UNABLE TO DO "A" & "B"

Proper sizing to meet required duty points

HP x 746 Volts x 1.732 x Motor Efficiency x Motor Power Factor

Nominal horsepower ratings are usually given by the VFD vendors but in some instances a VFD will only produce the stated nominal horsepower if a high efficiency motor is used. Unlike motors, VFDs generally have no continuous service factor. Momentary overloads, however, are permitted. VFDs generally exceed 97 percent efficiency at full load. VFDs are designated constant torque or variable torque, depending on their current overload capacity. Variable torque VFDs can produce 110 percent of rated current for one minute. Constant torque VFDs can produce 150 percent of full load current for one minute and even more for shorter periods. Variable torque VFDs are generally used for centrifugal pumps. VFDs must be matched to the load and motor characteristics. Certain VFDs, known as Current Source Inverters or CSIs, may require addition or deletion of capacitor banks to match the load and motor

58

Be sure the motor will be capable of delivering enough torque to the pump. Motor torque capability (including breakaway or start-up torque) must exceed pump ENVIRONMENTAL CONSIDERATIONS torque required at every speed. Generally, if the motor and VFD are To avoid potential problems in properly sized for 100 percent speed, your application of VFDs, you they will be adequate at lower must take a few precautions speeds. However, in certain regarding their environment. instances, such as applications with Locate VFDs indoors. Units can high suction pressure, motor and be placed outdoors with the propVFD sizing may be governed by er enclosure, but the cost of the start-up conditions. VFDs on positive enclosure can run into thousands displacement pumps must routinely of dollars. Fortunately, the VFD be oversized to provide sufficient can be up to several hundred feet start-up torque. Avoid lateral FIGURE 7 critical speeds. As 60 an example, API SPEED % D 80 EE EED Specification 610 P SP % 50 states that depending on the unbal40 anced response 200 amplification factor, a pump may NPS 150 H@1 20 % S P E E D not be operated between 85 percent NPS H @ 10 0 % SPEEED 100 and 105 percent of NPSH @ 80% SPEED its critical speed. 50 Adherence to these rules can block out 0 a large portion of 0 25 50 75 100 125 150 the allowable perFLOW - % OF 100% SPEED FLOW formance envelope of the variable Centrifugal pump NPSHR and efficiency vs. speed speed pump (Figure NPSHR

AMPS =

APPLICATION CONSIDERATIONS

The Pump Handbook Series

EFFICIENCY

78% SPE E

EFF I CI I CI EN E FF E N C CY Y ICI EN @ 1 @ 00 CY @ % 12 S 0

HEAD%

80

9). Fortunately, many pumps are of a stiff staff design and will operate below their first lateral critical speed. A vendor may be able to change the mechanical design to raise or lower the critical speed to provide full range speed adjustment. Be aware of torsional critical speed. Torsional critical speeds are resonant frequencies at which motor and driven equipment shafts can begin to oscillate with angular displacement as a result of torsional excitation. VFDs can cause torsional excitation problems known as torque ripple. For example, rather than delivering a continuous 295 ft-lb of torque, a VFD-driven, 200 HP motor may deliver torque cycling between 250 and 340 ft-lb at some 21,000 cycles per minute. This oscillation could be damaging. Clearly, careful analysis and selection of the VFD, motor, coupling and pump train are needed to avoid torsional problems.

E FF

100% SPEED -PUMP 2

characteristics. The more commonly used Pulse Width Modulation and Six Step VFDs do not require this matching. They are suitable for a wide variety of motors. Most VFDs operate on 480 V input and produce a maximum of 480 V output. If a higher voltage motor is desired, you can install a step-up transformer between the VFD and the motor or use a higher voltage VFD.

FIGURE 8

460

T

100%

HP

instrument signals if they are fed from the same supply transformer as the VFD.

IS IT WORTH IT? %T, %HP

Despite the list of precautions, variHP 230 50% POSSIBLE able speed pumping THERMAL can save you DERATING money. As shown, you can eliminate 0 30 60 90 120 0 30 60 90 120 the need for throtHZ HZ tling valves. You T = f (V/HZ) may be able to use HP = f (TXSPEED) one variable speed pump in place of two Performance of conventional variable speed motor fixed speed pumps. VFDs also eliminate the need for a from the motor. So it can be motor starter. Variable speed indoors even if the motor is outpumping often reduces power doors. requirements. And some electrical Derate for high temperatures and utilities provide rebates for compahigh elevations. If operated above nies that use energy saving devices 104o F, VFDs must be derated. They such as VFDs. Rebates can be up to must also be derated if used at elevaone-third the purchase price of the tions above 3300 ft. device. Other cost savings come Be cautious of power supply. through better process control due VFDs are sensitive to stiffness and to lower heat inputs and fluid irregularities in the electrical supshear. ply. You may need to install a line These savings frequently pay reactor or isolation transformer back the costs of utilizing variable between the VFD and supply main speed pumping (such as the cost of if the feed transformer is very stiff the VFD, possibly extra costs for (high KVA). Input line reactors or high-efficiency motors and possibly isolation transformers may also be oversized pumps). Payback periods necessary to prevent the VFD from of as little as one year are typical feeding electrical noise back into the when using variable speed pumpsupply main. Such noise can distort ing. V

T

THE FUTURE

FIGURE 9 120

1 0 0 % D E SIG

N SPEE D

100

HEAD%

80

5% 60

}

15% 40

50% DESIGN S PEED

PREDICTED CRITICAL SPEED CRITICAL SPEED AVOIDANCE BAND

20

0 0

20

40

60

80

FLOW

Avoiding lateral critical speeds

100

120

ment in operating parameters will make VFDs easier to use and integrate into a system. Improved reliability and fault tolerance will make VFDs easier to apply. You can expect manufacturers to add adjustment capabilities of output voltage and current waveforms to optimize motor efficiency and smoothness. Improvements in power semiconductors will provide higher efficiency and smoother output. Sizes of VFDs will diminish as components on circuit boards are integrated into chips. Reduced size and improved efficiency will allow packaging to be more compact and environmentally rugged, which will allow placement even in hazardous environments. Prices will come down, possibly by up to 25 percent over the next five years. Even today, you can achieve greater flexibility, energy savings, equipment savings and extra head and flow through variable speed pumping, provided you take extra care in assembling an appropriate combination of pump, motor and VFD. With improvements in technology, more and more users will begin to take advantage of variable speed pumping. ■ Stephen P. Murphy is Senior Business Development Specialist for Sundstrand Fluid Handling in Arvada, CO.

Variable speed pumping will become more popular as the technology establishes its track record. And as more system and plant engineers design for variable speed operation early in the development cycle, benefits beyond energy conservation will become apparent. Advances in VFD technology will also increase user acceptance. New features such as greater adjust-

The Pump Handbook Series

59

CENTRIFUGAL PUMPS HANDBOOK

Self-Priming Centrifugal Pumps The ability to self-prime can be a cost effective solution for many applications. BY TERRY W. BECHTLER ith greater global competition and increased environmental regulations, modern industrial applications over the years have evolved into sophisticated operations, demanding more control over their liquid handling processes. This is particularly evident on the ”dirty“ liquid side of a plant’s manufacturing process, in the drainage, filtration/pollution control/wastewater areas. Self-priming centrifugal pumps are important in meeting this demanding challenge. Single stage end suction centrifugal pumps may be divided by their designs into conventional or standard centrifugals and self-priming centrifugals. Centrifugal pumps incorporate a simple design with minimum moving parts - impeller, shaft and bearings. They are reliable, durable and rela-

W

tively easy to maintain. To better understand the working principle of a self-priming centrifugal pump, let’s first examine the centrifugal force principle and a standard or conventional centrifugal pump. All centrifugal pumps incorporate the centrifugal force principle, which may be illustrated by a car running on a wet road (Figure 1). The tires pick up water and throw it by centrifugal force against the fender. Centrifugal pumps incorporate the same principle, but the tire is replaced by an impeller with vanes and the fender is replaced by the casing (Figure 1b). The liquid enters the center or eye of the impeller. As the liquid reaches the impeller vane, its velocity is greatly increased. Centrifugal force, created by the impeller blades or vanes, directs the

liquid towards the outside diameter of the impeller. Once the liquid reaches the tip of the impeller vane it leaves the impeller at its greatest velocity. As the liquid leaves the impeller, its direction is controlled by the pump casing (the most common casing shapes are spiral or volute and circular). The spiral or volute casing surrounds the impeller, beginning at the point where the liquid leaves the impeller. The liquid enters the casing and follows the rotation of the impeller to the discharge. Within the casing there is a section called the throat or cutwater. The cutwater, also called the tongue, is a cast section of the volute casing, near the discharge that is positioned close to the maximum impeller diameter. As the liquid

FIGURE 1 A

B

LOW PRESSURE

60

The Pump Handbook Series

reaches the cutwater it is diverted into the pump’s discharge opening (Figure 2).

FIGURE 2

SELF-PRIMING Self-priming centrifugal pumps incorporate all the above standard centrifugal pump design features and add the following internal modifications: •

•

A casing design that surrounds the volute and impeller and enables the pump to retain liquid in a built-in reservoir, or priming chamber. This reservoir is filled during the initial prime of the pump, and when the pump completes a pumping cycle and shuts down, the reservoir retains liquid for the next priming cycle. An internal recirculation channel or port. This channel connects the pump’s discharge cavity back to the suction reservoir internally, allowing the continuous recirculation of liquid from discharge back to suction during the priming, usually to the peripheral portion of the impeller (Figure 2B).

These two internal design features, the priming chamber and internal recirculation channel, are what distinguishes a self-priming centrifugal pump from a standard centrifugal pump. Self-priming can also be accomplished by a diffuser design centrifugal pump that is used primarily for clear liquids.

HOW IT WORKS Self-priming centrifugal pumps can be placed above the liquid level of the source (Figure 3). Only the suction pipe enters the liquid being pumped. The pump is initially primed by adding liquid to the pump casing through a priming port, normally located near the discharge. The liquid fills the discharge reservoir, traveling into the eye of the impeller through the pump’s recirculation channel. The suction line, itself, is not filled. A check valve is usually located just inside the suction reservoir. All connections must be airtight. During initial start-up, the impeller rotation causes the liquid

in the pump reservoir to be directed to the discharge cavity via centrifugal force. Simultaneously, a lower pressure is formed in the suction reservoir. This draws the liquid from the discharge cavity back into the suction reservoir through the pump’s internal recirculation channel. This is a continuing action during the priming cycle. While this is occurring, the air in the suction line is drawn by the lower pressure into the eye of the impeller with the priming liquid and travels through the volute into the discharge cavity. At this point velocities decrease, allowing the air and liquid mixture to separate. The air flows up and is ejected, and the priming liquid recirculates back into the impeller. This process continues to draw all the air from the submerged suction line. In applications where the liquid level is at atmospheric pressure, that pressure on the liquid surface, coupled with the lower pressure in the suction pipe due to the evacuation of air, serves to push the liquid in the sump into the pump. When all air is evacuated liquid pumping automatically begins. Note that the diffuser design selfprime principle incorporates an impeller rotating in a stationary multi-vane diffuser (Figure 4). During priming, the diffuser separates the air from the pumped liquid until priming is completed. This priming action might seem somewhat complicated or mysteriThe Pump Handbook Series

ous, but it is actually a very easy task for a correctly installed self-priming centrifugal pump, and it happens automatically in a relatively short time (20 - 30 seconds for a normal 15 foot suction lift). It’s this feature that differentiates self-priming centrifugal pumps from standard centrifugal models. On a suction lift condition, a standard centrifugal pump, with only air in the casing and having no ability to separate air and liquid to create a vacuum, would have an impeller that simply spins, acting as a fan, because it has no way to lower the suction line pressure. By placing a foot valve on the end of a suction line and filling the pump and suction line with liquid, a standard centrifugal pump can be made to operate and pump in a conventional mode. If the foot valve leaks and air enters the suction, such as under a shutdown condition, a standard centrifugal pump stands the risk of losing its prime and becoming air bound. Under suction lift conditions, selfpriming centrifugal pumps are ideal for unattended use. Standard centrifugal pumps are sometimes fitted with priming systems to fill the pump and suction line with liquid prior to starting. In such cases, a control device tells the pump when all air is evacuated and the unit is liquid filled to start.

STYLES Self-priming centrifugal pumps are usually classified into two groups: basic self-priming pumps and trashhandling self-priming pumps. Basic self-priming pumps usually come with different impeller configurations, including fully enclosed and semi-open. Like all centrifugal pumps, the pressure developed is dependent on the impeller diameter and rpm. • Fully enclosed impellers allow self-priming pumps to develop medium to medium-high discharge pressures, up to about 110 psi or 254 ft total dynamic head (TDH). Normal pump sizes range from 1 in. through 6 in. suction and discharge. Pumps with a fully enclosed impeller have a very limited solids handling capa-

61

FIGURE 2B Discharge Outlet Flap Valve

Volute

Suction Inlet

Ball Bearings Replaceable Wearplate

Removable Coverplate Pressure Relief Valve Cartridge Mechanical Seal

Balanced Impeller

A cut-away view of a self-priming centrifugal pump designed to handle solids-laden liquids and slurries

•

62

bility, with sizes from 1 1/32 in. through 5/8 in. in diameter, depending on the size of the pump. This configuration is excellent for handling clear liquids, including processed hydrocarbons, along with general wash-down pressure applications. Semi-open multi-vane impellers are usually designed for slightly lower head conditions than fully enclosed impellers, but they have greater solids handling capabilities. Pump sizes usually range from 3/4 in. through 12 in. suction and discharge, with capacities to more than 5,500 gpm. Spherical solid sizes range from 3/4 in. through 3 in. in diameter, depending on the size of the pump. Basic self-priming pumps with semi-open impellers

are sometime referred to as general-purpose self-priming pumps. They are excellent for handling dirty, contaminated liquids. Applications include extensive use in industrial filtration operations and a wide range of enginedriven models that serve the construction market. Trash handling self-priming pumps generally use a trash-type, semi-open, two-vane impeller that allows the pump to pass larger spherical solids. •

Trash handling self-priming pumps generate medium discharge pressures in the area of 62 psi or 145 ft TDH on electric motor drives and discharge pressures upwards of 75 psi or 173 ft TDH on engine-driven configuThe Pump Handbook Series

rations, with capacities upwards of 3,400 gpm. Normal pump sizes range from 1-1/2 in. through 10 in. suction and discharge. The impeller design allows for excellent solids handling capability, ranging from l in. to 3 in. spherical solids diameter, depending on the pump size. Trash handling self-priming pumps are often referred to as the workhorse of centrifugal pumps due to their rugged design and large solids handling capabilities. These pumps can be found on some of the most severe pumping applications within plants or on construction sites. A desirable design feature of a trash handling self-priming pump is a removable cover plate, located

ical applications. Alloys available for pump construction also offer the same diversity. Cast iron and ductile iron are used for general purpose and refined hydrocarbons, hardened austempered ductile iron (ADI) is employed for abrasive applications, CD4MCu SS serves in corrosive and abrasive applications, and 316 SS, Alloy 20 SS, Hastalloy B, and Hi-Resin Epoxy Plastic are used for other special chemical applications.

FIGURE 3

APPLICATION GUIDELINES

directly in front of the impeller on the suction side of the pump. Trash handling self-priming pumps may be applied in waste sump applications where they are exposed to various size solids. Any pump may clog trying to pump larger solids than it was designed to pass. The removable cover plate allows quick access to the suction side of the pump, expediting the removal of blockage. Some designs allow removal of the cover plate without disturbing the suction and/or discharge line.

The principal application area for self-priming pumps is where their ability to self-prime is a cost effective solution; and when it is more convenient and desirable to locate a pump ”high and dry“ above the liquid. Some general guidelines are in order: The liquid being pumped should be of low viscosity ( 550SSU or less). Horsepower and efficiency corrections are needed for liquid viscosity above 550 SSU. If subjected to liquid freezing temperatures, the pump

must be protected against freezing to avoid damage. The vapor pressure of the liquid and the presence of high levels of entrained air are serious considerations in suction lift application. The NPSHA (net positive suction head available) must exceed the manufacturer’s published NPSHR (net positive suction head required) by a margin that accounts for the liquid properties. Repriming time increases with suction lift. Suction lifts with water as the liquid at normal ambient temperature should be limited to 15 to 18 ft. best efficiency range, although maximum practical lifts are obtainable to 25 feet. For other liquids or liquid mixtures, the vapor pressure of the liquid or the most volatile components of a mixture must be considered. Reducing the speed of operation (rpm) significantly reduces the NPSHR. Suction line piping should be sized to velocities in the 5 to 7 ft. range at design flow. For self-priming pumps it is recommended that the suction piping should be the same size as the pump’s suction inlet. The self-priming centrifugal pump offers a unique solution to many pumping applications. ■ Terry W. Bechtler has been Manager of Inside Sales for The Gorman-Rupp Co. in Mansfield, OH for four years.

FIGURE 4

SELECTION As discussed, self-priming centrifugal pumps have a broad design range that allows them to serve a wide variety of applications. Many metallurgical choices and shaft seal configurations are available to best serve particular services. Mechanical shaft seals can be single, double, or tandem. They are available as double grease lubricated for general purpose applications, oil lubricated with silicon carbide faces for industrial applications with abrasives, carbon against Ni-Resist faces for clean water or refined hydrocarbon applications, or Teflon fitted with carbon/ ceramic faces for chem-

Volute Priming

The Pump Handbook Series

Diffuser Priming

63

CENTRIFUGAL PUMPS HANDBOOK

Centrifugal Pump Testing Laboratory and on-site testing ensure pumps are up to their tasks.

BY LEO RICHARD

The test lab provides a tightly controlled environment and thereby generates the most accurate data.

64

A

s industry becomes increasingly competitive, pumps are being sized to precisely meet their duty requirements without oversizing. This allows users to maximize efficiency and minimize first capital costs. There is also a small but growing trend to question the economics of in-line spares and large spare parts inventories. These developments make it more critical than ever that rotating equipment precisely meets all hydraulic, material, and safety requirements. This is assured by thorough testing of pertinent parameters by manufacturers prior to shipment and by customers at their job sites. The level of justifiable testing will depend on the nature of the service and significance of the parameter to be measured. For instance, a water transfer application can be served with a stock pump that has undergone the manufacturer’s standard quality and performance checks. However, a corrosive, high pressure, or environmentally hazardous application may justify additional testing for material conformity and quality of construction. In addition to the extent of testing, several other factors must also be considered. The first is location. A shop or laboratory test is typically conducted at the manufacturer’s facility. The test lab provides a tightly controlled environment and thereby generates the most accurate data. In contrast, field tests sacrifice some accuracy, but they provide useful data under the actual conditions of service. The Pump Handbook Series

A reasonable split between the two approaches should be employed, depending on the nature of the evaluation and the user’s ability to conduct on-site testing. Also, the user and manufacturer must agree to a set of guidelines such as those published by the Hydraulic Institute (HI). Among other things, HI standards generally define the methods and acceptable tolerances to be used. However, regardless of the standard employed, good laboratory practice requires that all instrumentation be calibrated prior to the test. For maximum accuracy the instruments should be located after straight runs of pipe where steady flow conditions exist. In addition, the local barometric pressure must be considered, especially in applications requiring suction lift. The data obtained should be recorded in a test log, and each round of evaluations must be identified in this document by the manufacturer’s and user’s serial/ equipment numbers. Also, the question of user representation during testing should be clearly defined. This includes issues such as site location, the amount of advance notice prior to testing, and cost. Usually, the added cost and logistics problems make such witness testing inadvisable—unless the user has very limited experience with the manufacturer. This, as well as any other requirements, must be written into the specification prior to purchase. A brief description of the most common performance and quality evaluations is given below. For simplicity, these tests have been characterized in terms of certifying

existing motors and starters are to be reused. Such tests are typically conducted on water using certified motors. Data are collected at several points, depending on the level specified as part of the performance test. This information can be used to generate both wire to water and hydraulic efficiencies.

MATERIAL CONFORMANCE

A technician attaches a mag drive pump to a test tank. conformance in hydraulic capability, materials, or physical integrity.

HYDRAULIC CAPABILITY The determination of hydraulic performance is the most basic and common category of testing. This typically involves performance, net positive suction head (NPSH), and power evaluation.

PERFORMANCE TESTING The performance test of a specially ordered or job pump typically involves the generation of its headversus-capacity curve at the rated impeller trim. Such pumps are shop tested on water at the manufacturer’s site. If the HI standards are followed, the acceptance level must be defined. Level A requires that seven test points of head, flow, and efficiency be evaluated. Level B testing requires that five test points be checked. Each level of acceptance refers only to the head and capacity as specified by the customer for the service, also known as the rated or guarantee point. The defined tolerances for these parameters will vary depending on the size of the pumps and the level of testing required.

NPSH The NPSH test is basically a measure of the suction head requirement necessary to prevent cavitation. The procedure typically involves

holding the flow constant and reducing the suction head until a defined level of cavitation occurs. The data are used to generate a curve of the cavitation coefficient, Sigma, for the pump at the specified capacity. Sigma is defined as the net positive suction head available divided by the total pump head per stage. According to HI standards, the NPSH requirement of the pump is defined as the point at which a 3% head drop occurs on the Sigma curve. However, this criteria is somewhat controversial. The major issue is that incipient cavitation is well under way prior to the occurrence of the 3% head drop. In fact, some companies are considering an internal specification defining the NPSH requirement as only a 1% head drop on the Sigma curve.

POWER/EFFICIENCY TESTING Power and efficiency testing is becoming increasingly critical as companies are closely evaluating power consumption during the pump selection process. Another relevant issue is the growing trend of retrofitting sealed applications with sealless designs. As is well known, due to magnetic coupling and viscous losses, sealless pumps inherently have slightly greater power requirements than their sealed counterparts. Therefore, a confirmation of the published power requirements may be in order, especially for installations where The Pump Handbook Series

The usual considerations for material conformance testing are corrosion and erosion resistance. Again, the added cost for these procedures must be justified with regard to the particular application, as well as the consequences of process downtime and personnel or environmental exposure.

CERTIFICATE OF MATERIAL CONFORMITY The most basic type of documentation is in the manufacturer’s certificate of material conformity. This is a guarantee that the pump is made of the materials called out in the specification. This certificate is based solely on the standard quality tests performed by the manufacturer.

CHEMICAL ANALYSIS This involves confirmation of the material of construction by chemically testing small samples from the pump. These tests range from sophisticated chemical analysis to a basic screening utilizing chemical test kits.

NUCLEAR ANALYSIS This confirms materials used by means of a nuclear analyzer. This is a nondestructive test involving direct measurement on the surface to be analyzed. The composition of the material is determined by the equipment and matched with its internal database to generate an identification. Due to the high cost of this equipment, many sites utilize sub-contractors for this work.

HARDNESS TESTING Hardness testing may be required by the user, especially for pumps in highly erosive services. Though the type of hardness test can vary, the Brinell hardness test is fairly common.

65

MILL CERTIFICATION An extensive form of evaluation involves mill certification. Basically, the mill certs follow the pump along each step of the manufacturing process. This includes data from the initial pour at the foundry to the final checks of the finished components. The downside of mill certification is that it tends to be costly. Also, because additional data are required from the initial pour, stock pumps may not be used. Some parameters typically measured in mill certs include: •Mechanical Test Certification, which includes tensile strength, proof stress, and elongation. •Analysis certificates detailing the chemical composition. •Intercrystallation corrosion and ultrasonic tests.

PHYSICAL INTEGRITY TESTING As the name implies, this category of testing basically involves a confirmation of the pump’s ability to maintain the liquid boundary under the conditions of service. The chief areas of concern prompting such testing are the integrity of welds and possible porosity of castings.

DYE PENETRANT TESTING Dye penetrate testing involves the use of an extremely low surface tension liquid to detect possible leak paths in cast and welded surfaces. If the dye penetrates the surface, the piece is either rejected or weld repaired. If the component is repaired, the user is notified and the part retested to confirm the integrity of the weld.

RADIOGRAPHY

GAS LEAK DETECTION This involves pressurizing the pump with an inert gas such as arcton to detect any leak paths from the pump. Leaks are typically detected by means of a sniffer or mass spectrometer. This test is extremely sensitive and able to detect the slightest porosity in castings.

HYDROSTATIC TESTING Hydrostatic pressure testing is a standard quality check. The procedure usually involves filling pressurecontaining components with water and pressurizing to 1.5 times the rated working pressure. This pressure is held for a specified time, and the piece is inspected for leaks.

TESTING SEALLESS PUMPS The testing procedures utilized to evaluate standard sealed centrifugals are commonly used for sealless configurations as well. However, due to the unique design of sealless pumps, some additional procedures may be considered. A complete discussion of this topic can be found in the Hydraulic Institute Standard for sealless centrifugal pumps. (HI 5.1–5.6, 1st edition, 1992)

MAGNETIC STRENGTH

The determination of hydraulic performance is the most basic and common category of testing.

Radiographic testing is primarily used to confirm the integrity of welds in pressure-containing components. Procedures depend on the configuration and dimensions of the component, as well as the nature of the equipment being used. The test itself is somewhat costly and may impact delivery. For these rea-

66

sons its use is most often limited to critical applications in the power industry.

The strength of the permanent magnets in a magnetically driven sealless pump can be evaluated with a Gaussmeter. This instrument directly measures the strength of the magnetic field in Gauss or Milligauss. Gauss testing is usually an overkill for new pumps for the following reasons: •

The relative uniformity of production magnets.

•

The high safety factor incorporated into a magnetic coupling’s power transmission capability. (A safety factor of 2.0 under full load conditions is typical.)

The Pump Handbook Series

• The fact that the coupling is already inherently tested during generation of the hydraulic performance curve.

BREAKING TORQUE A “low tech” but effective way of site testing synchronous magnetic couplings is by measuring the breaking torque. Breaking torque is simply the force required to break or decouple the two opposing halves of the magnetic coupling. This is accomplished by anchoring the inner rotating assembly and applying torque to the outer magnet ring (OMR). Force is applied and measured by a torque wrench fitted to the drive shaft of the pump (the drive shaft is mechanically coupled to the OMR). The data generated is then compared to the manufacturer’s standards. As in Gauss testing, this procedure is usually unnecessary for a new pump. However, it is a useful field tool for confirming the strength of the magnetic coupling. This is especially important during a rebuild after a dry run failure. During dry runs, the magnets are exposed to extreme temperatures that may reduce their strength. By utilizing the breaking torque procedure, maintenance personnel can pretest the magnetic coupling prior to reinstallation.

SECONDARY CONTAINMENT TESTING CANNED MOTOR DRIVES The stator housing in canned motor pumps is often used as a secondary containment vessel. Testing typically involves gas leak detection on the finished stators. For designs utilizing potting of the wire leads, confirmation of the integrity of the secondary containment chamber as the equipment ages may be in order. This is especially relevant in services with high temperature cycling, which may damage the potting compound.

MAG DRIVE DESIGNS In some mag drive designs the coupling housing and an inboard magnetic seal are utilized for secondary containment. Testing usually involves a hydrostatic or gas leak detection of this assembly.

• After the pump has achieved steady state, the bearing frame, process, and ambient temperatures should be monitored and recorded. This data will be used as an initial check as well as for future reference. • Proper operation of all protective instrumentation should also be verified and any outputs recorded. For instance, many sealless pumps utilize a thermocouple temperature monitoring system to protect against dry runs. The initial temperature reading should be recorded in the commissioning data sheet.

SUMMARY

An A range pump hooked up for testing.

SITE TESTING One of the most important and often overlooked opportunities for evaluating and documenting pump performance is the initial commissioning. Information gathered at this time is critical in verifying initial performance and providing a benchmark for future diagnostic and troubleshooting efforts. It is suggested that, as a minimum, the following areas be evaluated: • The total differential head generated by the pump. It is strongly recommended that both suction and discharge gauges be installed to facilitate measurement of this parameter. Once determined, it should be noted whether the actual operating point differs from the duty listed in the specification. If so, the user must first confirm proper operation of the pump and process. If these check out, an evaluation of potential problems associated with the new duty point must be evaluated. This includes a possible increase in the NPSH requirement and power consumption. Also note that continuous operation at extremely high or low flows will significantly increase dynamic loading on the

impeller. Such loading can dramatically decrease the mean time between failures for the equipment. • Evaluation of the operating point should be conducted for all conditions the pump will experience. For example, many pumps in transfer applications deliver liquid to various locations and are periodically operated in a recirculation mode. Each of these duty points must be determined and possible problems identified. If necessary, modifications in the pump and/or process should be made. Common corrective actions include resizing orifices, changing valve settings, and adjusting the impeller trim.

There are many options for testing the performance and integrity of centrifugal pumps. The use of such procedures depends on the significance of the service and the nature of the pumpage. Users will find that in most cases the standard compliment of manufacturing testing will be sufficient. However, critical services involving serious environmental or health risks may warrant the added assurance of supplemental testing. In either case, the user and manufacturer must work as partners to achieve the best engineering solution for the particular application. ■ Leo Richard is a Technical Service Manager with the Kontro Company, Inc. Mr. Richard has experience in process and project engineering with General Electric and W.R. Grace.

• The amp draw of the motor should be measured. This is then compared with the manufacturer’s stated requirements to evaluate proper operation. Gross differences between these figures may indicate various conditions such as cavitation, operating to run out or shut in, or mechanical problems. • The vibration level should be measured. This will confirm proper operation and serve as a benchmark for future testing.

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67

CENTRIFUGAL PUMPS HANDBOOK

The Canned Motor vs. Magnetic Drive Debate BY GREGORY ZIMMERMAN erhaps you’ve decided to purchase sealless centrifugal pumps. The arguments are compelling: zero emissions, no need for complicated seal support systems, no need to replace expensive seals periodically. All manufacturers of sealless centrifugal pumps agree on these basic advantages. But their agreement ends there. As the two major camps in the sealless centrifugal debate — canned motor or magnetic drive — try to position their chosen technology as the most reasonable choice, they let loose a flurry of claims and counterclaims. It can get confusing. To help you prepare for the barrage, we present advantages for both types of sealless centrifugal pumps as commonly stated by manufacturers and users. Consider the arguments and decide which are most pertinent to your situation. Then you’ll be better prepared to discuss your specific concerns with manufacturers.

P

THICKNESS OF CONTAINMENT SHELL

68

HIGH TEMPERATURE SERVICE Permanent magnets can tolerate heat better than motor windings can. Thus, magnetic drive pumps can pump hot liquids — up to 750° F with just air cooling. Canned motor pumps can also be used in hot service but need water cooling jackets. Manufacturers of canned motor pumps agree that canned pumps should be water cooled for high temperatures. But, they reply, so should magnetic drive pumps since even rare earth permanent magnets cannot tolerate extremely high temperatures.

counter that internal clearances are designed to accommodate bearing wear, not to accommodate solid particles. The crucial dimension, they say, is bearing clearance and that is the same for both types of sealless pumps. Further, canned motor pumps can effectively handle solids if they are outfitted with external flush or filters to remove particulates from the pumpage before they circulate around the bearings.

HIGH PRESSURE APPLICATIONS One point on which all parties agree is that magnetic drive pumps cannot tolerate as high pressures as canned motors pumps can. A canned motor pump is a pressure vessel

PHOTO COURTESY OF TEIKOKU USA

Magnetic drive pumps can use thicker containment shells since their inner and outer magnetic rings do not have to be as close together as the rotor and stator in a canned motor pump. Manufacturers of mag-drive pumps claim that the thicker shell — up to five times thicker than that of canned motor pumps — vastly reduces the chances of breaching the shell, especially as a result of bearing wear.

Manufacturers of canned motor pumps counter that claim with two arguments. First, the thicker shell of a magnetic drive pump reduces operating efficiency. Second, the shell must be thicker (and the internal clearances wider) because magnetic drive pumps do not contain bearing monitors. Unmonitored bearing wear can cause the inner magnetic ring to contact the shell. A thin shell would be too prone to such damage. Canned motor pumps can use a thinner shell because bearings are closely monitored and bearing wear can be projected from the data.

SOLIDS HANDLING With greater internal clearances and thicker containment shells, magnetic drive pumps can handle solids more easily, their manufacturers say. Canned motor manufacturers The Pump Handbook Series

Sealless canned motor pump designed for hazardous liquids.

PHOTO COURTESY OF GOULDS PUMPS, INC.

ples or loses a bearing, the skin temperature on the drive unit can exceed the autoignition temperature of the explosive compound.

COMPACT DESIGN

Magnetic drive process pump designed for zero leakage services. since the stator windings lend additional mechanical strength.

DIFFICULT-TO-HANDLE FLUIDS According to one manufacturer of magnetic drive pumps, the biggest advantage magnetic drives offer is the ability to use non-metallics. These pumps are thus able to pump highly corrosive materials, solvents, and other difficult fluids. That may be true for some fluids, counter manufacturers of canned motor pumps, but other issues are involved. Hazardous materials require failsafe containment. Canned motor pumps, they point out, offer sealless double containment. If the stator lining blows, a backup shell will contain the materials. Doubly contained magnetic drive pumps rely on a mechanical seal — the very thing we’re trying to avoid, say the manufacturers of canned motor pumps. Canned motor proponents point to another benefit of their technology in hazardous environments: UL listing for the entire unit. Because a canned motor pump integrates the electrical and mechanical portions, the entire pump must be UL listed for use in, say, explosive atmospheres. Sundstrand canned motor pumps, for example, are tested under a procedure in which UL fills the pump with oxygen and ethylene and ignites the gas. The explosion must be contained in the pump with no propagation of flames up or down the discharge piping. Magnetic drive pumps are not UL listed — only the motors need to be. If a magnetic drive decou-

Canned motor pump manufacturers cite compact design as an added advantage. Canned motor pumps not only save space but also require no foundation work. Magnetic drive manufacturers counter that they can make compact pumps by using a close coupled design. Besides, they add, the absolute dimensions aren’t as important as meeting ANSI standard dimensions. ANSI standard dimensions make magnetic drive pumps easier to retrofit, according to their proponents.

ALIGNMENT Canned motor pumps have an integrated single shaft and thus come perfectly aligned from the factory. Alignment of the motor and magnetic coupling can be tricky in a magnetic drive pump.

A USER’S PERSPECTIVE For one user, an engineer at a chemical processing plant in the midwest, UL area classification is the most important reason he prefers canned motor over magnetic drive pumps — in situations where canned motor pumps are optimal. This user also relies on magnetic drive pumps for high temperature applications (e.g., heat transfer fluids), high horsepower requirements and for aqueous hydrochloric acid service (which requires nonmetallic pumps). Another advantage this user states for canned motor pumps is he can predict bearing wear and thus schedule maintenance more easily.

we’re getting comfortable with that technology, too,” he said. This user and another engineer at a major chemical plant report high reliability of both types of pumps. Both report that reliability increased as they gained more experience with sealless pumps. In each facility the major cause of damage to sealless pumps is operator and specification error. And as they learned to size the pumps correctly — to operate at the best efficiency point of the pump — and to avoid operating the pumps off design, mean time between failure increased substantially. “We’ve gotten seven years without failure from some of our sealless pumps,” said one user, “but we’ve also had cases where we replace the pump nearly every month because of dead head operation, running dry or cavitation.” All the above manufacturers agree that pumps must be specified correctly for the application and that operators must be trained adequately. “Users need to make sure we know everything about the application,” says one manufacturer. “We especially need to know temperatures and vapor pressures at startup and shutdown, not just normal operating conditions.” Another key point: don’t simply substitute pump problems for seal problems. In other words, if you’re faced with recurring seal failures, be sure to root out the cause of the failure before you simply bring in sealless pumps. Maybe the fault isn’t the seal. If the problem lies elsewhere in the system, you’ll be left wondering why your sealless pumps failed just like the seals did. ■

FURTHER ADVICE ABOUT GOING SEALLESS “We’ve been using canned motor pumps for some time and are comfortable with the technology,” the user says. “Our electricians are adept at repairing the pumps — both mechanically and electrically. We are bringing in more magnetic drive pumps for certain applications and The Pump Handbook Series

69

CENTRIFUGAL PUMPS HANDBOOK

National Electric Code Impact on Sealless Centrifugal Pumps BY: ROBERT MARTELLI ump users are no different from other users of industrial processing equipment who must comply with several codes and government regulations. It can be a formidable task to keep up to date and appropriately apply rules to specific situations. A greater effort is required to get code and regulations updated and clarified to keep pace with changing technology. Nonetheless, users need to understand the impact of the National Electrical Code on sealless centrifugal pumps and to know what monitoring options are available.

P

What users need to know about the National Electric Code and how monitoring options can help.

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ELECTRICAL DEVICES IN HAZARDOUS LOCATIONS The National Fire Protection Association (NFPA) has produced several codes for reducing the risk of and damage from fires. One of these is the National Electric Code (NEC). Section 500 of this document applies to electrical devices operating in hazardous environments—where flammable/explosive materials are either routinely or may be present in the atmosphere. These materials can be gasses/vapors, liquids, a solid dust or

The Pump Handbook Series

liquid mist. Since electrical devices are present in these areas, the NEC imposes requirements to reduce the risk of fires and explosions. A Division 1 area is where explosive materials are routinely present in the atmosphere (such as the bottom of a spill containment, or a below-grade installation where vapors could collect) and requires U.L.-approved electrical devices. Most sealless pumps, however, are operated in Division 2 areas where explosive materials may occasionally be present in the atmosphere. (In the chemical industry, the vast majority of materials are handled in closed systems.) The two areas are covered in the NEC where they have the potential to form an explosive cloud in the atmosphere. A growing cloud that comes in contact with a source of ignition, such as a hot electrical device, can cause a large explosion and fire. To reduce this risk, the NEC requires that the Auto-Ignition Temperature (AIT) be determined for each stream in the process area as well as its geographical area. Electrical devices intended for operation in hazardous areas are also

FIGURE 1

Temperature monitor on containment shell by liquid exit from magnet area

Mag drive pump cooling circuit flow temperature measurement is made after the fluid has picked up eddy current heat and partial bearing heat. required by the code to have “T ratings.” If users follow this section of the code, these electrical devices will not constitute a potential source of ignition, vastly reducing the chance of an explosion should an explosive cloud ever develop. AITs are a concern for light hydrocarbons including n-butane and acetylene, which have AITs below 600°F; pentane and hexane, which have AITs below 500°F; and diethyl ether and heptane, which have AITs below 400°F.

NONELECTRICAL SOURCES OF IGNITION Ignition by nonelectrical sources— for example, steam, heat transfer lines and reactor vessel walls—are also possible in process areas. The NEC does not address these sources. Another NFPA code covers nonelectrical sources of ignition in section 30, the Flammable and Combustible Liquids Code. Specifically, Chapter 5

requires that you take precautions to prevent the ignition of flammable vapors from nonelectrical sources. Preventive measures are to be determined by an engineer and/or “the authority having jurisdiction.” Since “hot surfaces” are normally present in chemical processing environments, one precaution typically taken is to handle materials in closed systems.

EDDY CURRENT HEAT GENERATION Canned motor and magnetic drive pumps with metallic liner/containment shells generate heat due to eddy current loss. Eddy currents are created by changes in magnetic field strength during pump operation in a given area of a stator liner or containment shell. In most applications, pumps will operate at temperatures well below 400°F (which is below most AITs) because of the cooling effects of the pumpage. In general, the eddy current heat source is not The Pump Handbook Series

regarded as heat produced by an electrical device and, therefore, not clearly addressed in the NEC.

CANNED MOTOR PUMPS IN HAZARDOUS AREAS (DIVISION 2) The “skin” or outside temperature of the canned motor pump is an issue in hazardous areas. These pumps contain a thermal cut-out switch, which is located in the stator winding hotspot and shuts down the motor if its setpoint is exceeded. The user is required to wire this switch into the motor control circuit. If the motor cooling is lost due to some upset or misoperation, the pump will heat up and eventually open the switch and shut off the power, preventing an excessive “skin” temperature on the can. If the pump is in a volatile liquid service, it’s usually destroyed. In most cases, the switch will not protect the pump from dry running—it is there only to meet NEC requirements.

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For canned motor pumps, the NEC currently covers only conduit seals. This is to prevent hazardous pumpage from traveling through the conduit system to the motor starter room in case of a stator liner and primary seal failures.

FIGURE 2 Temperature monitor on containment shell by liquid entrance to magnet area

MAGNETIC DRIVE PUMPS (DIVISION 2 AREAS) Mag drive pumps with metallic containment shells are not typically regarded as electrical devices, despite eddy current generation. (Mag drives with nonmetallic containment shells have insignificant eddy current generation and associated heat-up potential.) These losses are about 17% of the maximum rated horsepower of the drive, which works out to between 2 and 3 KW of power loss for a drive rated for 20 hp. If an upset or misoperation results in dry running, recent tests have shown that the containment shell temperature can reach 800°F to 1200°F in one to two minutes of continued operation. Furthermore, mag drive pumps that operate for several minutes with no cooling provided to the magnet area have straw-blue rings in the areas of the strongest magnetic flux, indicating temperatures of at least 900°F. Even more disturbing, with this type of failure there is a good chance of a spill or release occurring! Currently, there is no requirement to monitor mag drive pumps for an abnormal condition and subsequently shut down the pump. In most cases, containment shell temperature, motor power, or pump flow monitoring with alarm and shutdown capabilities can greatly reduce the possibility of ever reaching unacceptable temperatures. Today these monitoring options are routinely available. Most mag drive manufacturers provide the option of containment shell temperature monitoring. However, there is no widely accepted agreement on the best monitoring method for mag drive pumps. Each method has strong and weak points. Several mag drive pump manufacturers have recently taken steps to isolate the containment shell from the outside atmosphere, eliminating the air cooling used in some designs. This will not necessarily prevent the migration

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Mag drive pump temperature measurement is made on the cooling circuit inlet. Temperature variations will be much smaller here. of flammable vapors or gasses to come in contact with the containment shell. Depending on maintenance procedures, there is also a small possibility that the pumpage will leak undetected past an improperly installed gasket and collect near the bottom of the containment shell, next to the outer magnet assembly. Therefore make sure the containment shell temperatures do not rise above the AIT of the materials present. Some mag drive manufacturers offer a leak monitor option for this section of the pump, partially addressing this concern.

HEAT TRANSFER FLUID PUMPS Mag drive and canned motor pumps with ceramic insulation on the stator windings in heat transfer service may present a problem since in some situations the suction fluid temperature can be in the 600°F range. If this temperature exceeds an AIT for other nearby materials, there is no increase in safety by applying this portion of the NEC to either pump. In this case, the “skin” temperature already exceeds limitations set by the NEC for electrical devices! Current NEC interpretations by several users preclude the use of a canned motor pump with ceramic The Pump Handbook Series

insulation in these instances. The only option for a canned motor pump is a unit with a cooling jacket, the necessary service water lines, and a conventional stator with the appropriate thermal cut-out switch. The current mag drive claim, again, is that there is nothing in the pump that meets the definition of an electrical device; therefore, no special monitoring or shutdown devices are required by code. If the magnets can operate at the required service temperature, no cooling water is required. Although difficult, it may be prudent to revise pump and piping layouts so that no low AIT materials are near the pump.

CODE APPLICATION AND CLARIFICATION A word of caution: The NEC has been adopted by OSHA as a reference standard and you are required to follow it as a minimum. Be careful when interpreting the code and remember that common sense does not always apply! When making interpretations or determinations regarding legal regulations, a team approach is advisable. More informed determinations are made, and mistakes are less likely.

FIGURE 3 Temperature monitor on rear bearing housing at liquid exit from rotor-stator liners

Canned motor pump cooling circuit temperature measurement made at its hottest point. If the NEC panel would clarify the application of the code to mag drive pumps with metallic containment shells, it would help pump users considerably. Specifically, does eddy current generation fall within the code definition of an “electrical device”? And for electrical devices of which the “skin” temperature already exceeds an applicable AIT by nonelectrical sources, does the NEC prevent its use? A National Electric Code change can occur no sooner than 1999, when it’s scheduled for update. Until then, users will have to operate under the current code, taking precautions as they deem appropriate.

can be more prone to sudden breakage and failure due to misoperation; hence, these conditions must be identified and the pump automatically shut down if encountered. Pump monitoring is a relatively new concept for most operations people and not well understood. Yet these workers play a key role in implementing monitoring methods. Everyone involved should have patience in finalizing the alarm and shutdown setpoints for successful implementation of the methods used. Nonetheless most users go through several “false shutdowns” or even a pump failure before determining the proper setpoints.

CONTINUOUS MONITORING: SAVING MORE THAN THE PUMP

TEMPERATURE RISE MONITORING

Although monitoring adds cost, users can take advantage of automatic shutdowns for other abnormal conditions (such as dry running) before the pump is destroyed and provide better assurance that AITs are not exceeded. Monitoring can also improve pump reliability in handling heat sensitive materials. Monitoring is especially important with silicon/tungsten carbide bearings. (Most sealless pump manufacturers offer carbide bearings at least as an option.) Monitoring these bearings requires a different approach than for carbon bearings in order to extend life. Carbide bearings

For a single method, temperature rise monitoring offers the best overall protection against most pump failures, including dead-head/very low flow, dry run operation, and restricted cooling circuit flow in the magnet area. Moderate cavitation and gas entrainment in the pumpage are also involved when they reach the point of upsetting the cooling circuit flow. Two temperature points are required to implement this monitoring. One is on the containment shell of the pump. In all current designs, this point must be located between the magnet assembly and the containment shell flange limiting what temThe Pump Handbook Series

perature can be monitored. Note the direction of the cooling circuit flow next to the temperature measurement point. A more sensitive measurement results by monitoring at the exit point for the cooling circuit flow after it has picked up heat from the magnet area (Figure 1). This flow configuration can find this exit point near the rotating magnets in pumps that use a discharge-to-discharge pressure circulation with a pumping vane near the rotating magnets. On pumps that use a discharge-to-suction pressure configuration to drive cooling circuit flow (that is, where the cooling circuit inlet flows past the containment shell at the measurement point, before the temperature rise takes place) temperature rise monitoring will not be as effective (Figure 2). Canned motor pumps may also have temperature monitoring installed. More sensitive readings can be taken when the monitor point is located after the cooling fluid passes between the rotor and stator liners (Figure 3). With the temperature probe in this location, dry run protection will not be as effective as what can be provided by power monitoring. The second temperature monitoring point is on the suction line or supply vessel, providing suction temperature compensation and takes into account temperature changes from day to night, and seasonal variations. This greatly eliminates false shutdowns and failures. Keep enough distance between this point and the pump to ensure that suction recirculation will not conduct heat from the pump and up the suction line to the measurement point during deadhead operation. Locating this point upstream of a suction basket strainer may provide enough isolation to be effective. If you go to the tank for this temperature, keep in mind that the sun can warm up the suction line and pump unit much faster than it can warm the tank during nonoperation. If this happens and you get an inaccurate measurement, you may shut down the pump on start-up when there is nothing wrong. The temperature rise is determined by the difference between the containment shell and suction temperatures (Figure 4). Pump suppliers can provide an expected “normal”

73

temperature rise. Typical alarm and shutdown points may be 10°C and 20°C above this value. Field experience will be required to finalize these setpoints for each application, since this is a relatively new concept for pump users. A good approach is to find a temperature rise that is sufficiently far away from the normal operating range (with its usual variations), and that still results in liquid in the containment shell, with a few degrees of boiling point margin left in the magnet area. Pump suppliers can help by supplying pump cooling circuit pressure. Knowing the pressure, you can calculate the liquid boiling point in the cooling circuit. The maximum cooling circuit temperature needs to stay below this value. Recent discussion about containment shell temperature rises downplays the effectiveness of containment shell temperature monitoring. Some say the temperature measurement is not sensitive enough for the rapid rise found in dry running. However, temperature rise monitoring does not need a large change in containment shell temperature to be effective. Also, temperature rise monitoring does not protect against motor or outer magnet bearing failure. Periodic vibration monitoring or additional bearing temperature monitoring are two proven ways to protect against these types of failure. Wear of the inner sleeve bearings may be detected by temperature rise monitoring if there is enough wear to alter the cooling circuit flow or the eddy current heat generation. This will depend on the pump used since temperature rise is not always a direct result of bearing wear.

FLOW AND LEVEL MONITORING For processes where the supply tank level or pump flow are already measured and sent into a process con-

74

trol computer, adding these monitoring devices can be relatively inexpensive. The only other hardware required is an output relay in the pump motor control circuit. Then the software is programmed to implement a low level shutdown or a high flow/low flow shutdown for the pump. The low level method is effective against running the tank dry but does not cover other common pump failures. Flow monitoring provides a little better protection because it protects against closed suction and discharge valves in addition to dry running. It can also protect against excessive flow. The narrower the range between the shutdown setpoints and normal operation, the better the protection; however, false shutdowns must be avoided. Neither of these methods protects against mild cavitation, a plugged cooling circuit flow path, or worn inner bearings. In services where these other modes of failure are unlikely, this method can be quite effective.

POWER MONITORING In this case, motor power (kilowatt draw) is monitored, usually in the pump starter room. It has the advantage of not requiring any process connections to install, making it one of the easiest to incorporate into existing processes. This can eliminate corrosion/erosion concerns in slurry or acid service where exotic materials of construction are required. It is easier to establish shutdown setpoints if the pump is operating in the range of 60% to 90% of its BEP. As with any electronic device in an operating plant, it must not be affected by radio frequency interference (e.g., portable radios). Power monitoring has the same advantages and disadvantages as flow monitoring in protecting against previously described failures. Current monitoring of the motor

The Pump Handbook Series

FIGURE 4 ALARM TDI 1942

SHUTDOWN

YS 1942 TI 1942

TI 1942

PG

PG

Temperature rise is calculated in a process control computer and compared to alarm and shut-down setpoints for appropriate action. amp draw can also be effective as long as the horsepower draw is near the motor’s nameplate rating. Otherwise, the amp draw versus pump curve becomes flatter, and it’s more difficult to determine realistic shutdown setpoints.

CONCLUSIONS Until the National Electric Code is either revised or clarified, users will need to make their own determinations of the potential hazards of pump operations and choose suitable means of reducing the resulting risks. Efforts are underway at several pump manufacturers to improve continuous pump monitoring. A more universally accepted method should result. ■ Robert H. (Bob) Martelli is Engineering Specialist in Facilities Engineering, Dow Corning Corporation, Midland, MI.

CENTRIFUGAL PUMPS HANDBOOK

Pumping Hydrofluoric Acid Consider proper metallurgies, compatible bearing materials, and hydraulic and pump configurations when pumping this acid.

BY: JOHN V. HERONEMA ydrofluoric acid has touched all of our lives because so many industries use it in their manufacturing processes. For example, a beryllium-shafted golf club and a coffee mug with an etched design have been manufactured using hydrofluoric acid. It has also been used as a catalyst in the manufacture of ozone-friendly refrigerants. Yet hydrofluoric acid is a potentially dangerous chemical. Acid leaks can yield devastating effects, ranging from toxic fume inhalation to severe chemical burns, injuring people and damaging equipment. Many plants pump hydrofluoric acid using traditional seal technology. Of course, mechanical seals can leak. Because of the potential danger involved, hydrofluoric acid leaks are not tolerated. One way to reduce the threat of leakage is to use a sealless technology. Consider several key factors when selecting a sealless pump for hydrofluoric acid applications, including proper metallurgies, compatible bearing materials, and hydraulic and pump configurations.

H

METALLURGIES Proper pump metallurgy is critical for pumping hydrofluoric acid. Two primary variables dictate what metallurgies are necessary under pump operating conditions. The first is temperature. Hydrofluoric acid is similar to many other acids in that as temperature increases, so does the aggressive nature of the fluid. The

either alloy, stress-corrosion cracking second variable is the percent conmay be inevitable if water or oxygen centration. Table 1 and Figure 1 are present, and in that case, corrodefine the most suitable metallurgies sion and cracking would be widefor given applications. When chemispread and not localized.) Both of cal process industries use hydrofluothese metallurgies are excellent ric acid, its nature is generally choices for handling hydrofluoric aggressive. Consequently, worst case acid. scenarios have more significance in decision-making choices. BEARINGS Temperature and concentration are not the only variables that impact Bearing material is every bit as corrosion rates. Factors such as crucial to a pump’s mechanical stabilvelocity, aeration and other contamiity as its overall metallurgical componates play equally important roles in sition because the bearings are metallurgical corrosion. exposed to the acid. This is particuFive metallurgies (Table 1 and larly important in canned motor techFigure 1) are suitable for any given nology because the pumping process condition. Silver, gold is responsible for the and platinum are among cooling and lubrication the metals most resisof the bearings. Temperature tant to hydrofluoric acid What is the proper corrosion. Two other bearing material? What and percent metallurgies are more will hold up under the affordable and provide unforgiving corrosiveconcentration excellent results, mainness of hydrofluoric dictate what taining corrosion at less acid? The answer is than 20 mils per year alpha grade silmetallurgies are 100% (mpy) during adverse icon carbide, which is a conditions. One is 66Ni necessary under pressureless sintered sil32Cu (Monel 400), and icon carbide. Bearings operating the other is 54Ni 15Cr made of this material 16Mo (Hastelloy C-276®). can withstand high temconditions. There are some pitfalls peratures and maintain in the composition of dominating resistance to 54Ni 15Cr 16Mo. This strong acids. Alpha alloy is less resistant to corrosion grade has better resistance to wear than 66Ni 32Cu, especially if oxygen and abrasion than the beta version of is present; whereas 66Ni 32Cu is silicon carbide. However, both are generally corrosive resistant, even to pressureless sintered, or self-sintemperatures up to 300°F. (With tered, silicon carbide products.

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speed, meaning poor hydraulic efficiency. C Specific speed is a dimensionless number that relates the hydraulic perfor250 121 BOILING mance of centrifugal POINT pumps to the shape and 225 physical properties of its impeller. The equation to calculate specific speed is 200 93 shown in Figure 2. Where low flow and high head are 4 requirements, use a partial 175 emission pump with an 6 open or closed radial vane 66 150 impeller. A standard guideline for pumping a fluid as 3 5 volatile as hydrofluoric acid 125 is to keep the specific speed above 200. This ensures the 100 pump will maintain a rea38 sonable hydraulic efficiency (25% to 30%). 75 Specific speed can be 7 2 1 easily manipulated by increasing, gallon by gallon, 10 20 30 40 50 60 70 80 the flow of a pump until CONCENTRATION HF,% the desired N s value is Zone definition for common metallurgies achieved. Another way to impact specific speed includes increasing the rotative Do not use reaction-bound silispeed. This technique is sometimes con carbides for hydrofluoric acid difficult because many motors have processes. These forms of silicon carfixed rotating speeds. To manipulate bide contain free silicon or graphite speed, a variable frequency drive because reaction-bound silicon carmust be used. A variable frequency bides require silicon as a sintering drive can increase the speed at which aid. Free silicon is subject to the a motor runs while maintaining a attack of corrosive acids, resulting in constant voltage. However, these bearing breakdown. In alpha and devices can be expensive. Regardless, beta grades of silicon carbide, no sinthe results are the same—increased tering aids are used, giving both hydraulic efficiency. grades almost complete chemical Hydraulic efficiency is important inertness. The bottom line is that when pumping hydrofluoric acid there is little difference between the because it has a steep vapor pressure alpha and beta grades of silicon carcurve. Unproductive energy, which is bide. Most of the difference lies witha direct byproduct of inefficiency, is in the processing of the final lost in the form of heat. This added products. Nonetheless, alpha grade heat must not be allowed to localize silicon carbide is the preferred materin the suction zone of the pump case. ial for chemical processes that use If it does, and suction pressure is not hydrofluoric acid. Both alpha or beta great enough to suppress vaporizagrades of silicon carbide should tion, the pump may fail. The ability exceed bearing expectations. to carry the heat away is directly related to the specific heat of the HYDRAULIC CONFIGURATIONS fluid. Specific heat is the ratio of a For hydrofluoric acid applicafluid’s thermal capacity to that of tions, the same challenges arise again water at 15°C; in other words, a and again: low NPSHA, low flow and fluid’s ability to carry away energy in high head. In centrifugal pumps, low the form of heat. Unfortunately, this flow and high head yield low specific thermodynamic property of hydrofluTEMPERATURE

FIGURE 1

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The Pump Handbook Series

oric acid is fairly poor. If adequate NPSHA existed for most hydrofluoric acid applications, the ability of the process to dissipate heat would not be crucial. However, NPSHA is often lacking. NPSHA is the net pressure of a process fluid at the suction of a pump. Having adequate NPSHA is important when pumping hydrofluoric acid because of the volatility of the process. The graph in Figure 3 demonstrates the relationship of tem-

TABLE 1. CODE FOR HYDROFLUORIC ACID GRAPH Materials in shaded zone have repeated corrosion rate of <20 mpy Zone 1

Zone 4

20Cr 30Ni 25Cr 20Ni Steel 70Cu 30Ni1 66Ni 32Cu1 54Ni 15Cr 16Mo Copper1 Gold Lead1 Nickel1 Nickel Cast Iron Platinum Silver

70Cu 30Ni 66Ni 32Cu1 54Ni 15Cr 16Mo Copper1 Gold Lead1 Platinum Silver

Zone 2 20Cr 30Ni 70Cu 30Ni1 54Ni 15Cr 16Mo 66Ni 32Cu1 Copper1 Gold Lead1 Nickel1 Platinum Silver Zone 3 20Cr 30Ni 70Cu 30Ni 54Ni 15Cr 16Mo 66Ni 32Cu1 Copper1 Gold Lead1 Platinum Silver

Zone 5 70Cu 30Ni1 66Ni 32Cu1 54Ni 15Cr 16Mo Gold Lead1 Platinum Silver Zone 6 66Ni 32Cu1 54Ni 15Cr 16Mo Gold Platinum Silver Zone 7 66Ni 32Cu1 54Ni 15Cr 16Mo Carbon Steel Gold Platinum Silver

1

= No air

17 PSIA – 36 PSIA 75°F – 115°F

FIGURE 2 1 Ns = NQ ⁄2 H 3⁄4

Ns N Q H

= = = =

(

Example A 1 Ns = 3550 (15 ⁄2) 900 3⁄4 Ns = 83.7

)

Specific speed Revolutions per minute Capacity, at best efficiency, in gpm Total head developed by maximum diameter impeller at best efficiency, in feet

Equation to calculate specific speed perature to pressure. As the temperature rises, the required pressure to maintain the acid in a liquid phase increases, and the vapor pressure curve becomes dramatically steeper at higher temperatures. Any point left of the curve means the process is liquid; conversely, any point right of the curve means the process is vapor. If hydrofluoric acid is being pumped at 100°F, the NPSH must be equal to or greater than 27 psi. If not, the process will flash, resulting in a heavily cavitated or dry running pump. To ensure adequate NPSHA, the heat input from the pump must be considered. Hydraulic temperature rise can be calculated. The equation (Figure 4) considers three variables: hydraulic efficiency, head, and the specific heat value of a process. By using this equation and considering the vapor pressure versus temperature rise curve, you can predict if adequate NPSHA is provided. The following is an example.

PUMPING SPECIFICS Fluid Pumped = HF acid Head (H) = 790 feet Flow (Q) = 20 gpm NPSHA = 7 feet, Mechanical NPSHR = 6 feet Temperature (P) = 95°F Vapor Pressure (P.T. PSIA) = 25 Specific Heat (BTU/lb°F) = 0.78 Pump Hydraulic Efficiency (n) = 15% Specific Gravity (sp gr) = 0.92 1.

Solve for the slope of the vapor pressure curve. Pick one temperature/PSIA point below the design temperature and another 20°F above the operating point. Convert data into PSIA per °F.

=

19 PSIA 40°F = 0.475 PSIA/°F 2. Solve for the maximum allowable temperature rise that can occur before the HF flashes. (NPSHA – NPSHR) sp gr = PSI 2.31 0.475 PSI per °F = maximum allowable temperature rise °F (Actual) (7 – 6) 0.92 = 0.39 PSI 2.31 0.475 PSI per °F

In this example adequate NPSHA is not being supplied. If this data is plotted on a vapor pressure versus temperature curve, the end result is obvious—the HF is vapor (Figure 5). To calculate how much NPSHA is necessary to keep the HF from vaporizing: 1. Total hydraulic temperature rise = 7.37°F 2.

= 0.83°F allowable temperature rise °F 3.

7.37°F – 0.83°F (Allowable Temperature Rise) = 6.54°F 6.54°F x 0.475 = 3.1 PSI Convert 3.1 PSIA to feet 2.31 x 3.1 PSIA = 7.78 feet sp gr (0.92)

Calculate hydraulic temperature rise due to inefficiencies.

H (1 – n) = Temperature Rise 778 x n x Cp Where n = hydraulic efficiency Cp = Specific heat

Convert 7.37°F to PSIA using calculated vapor pressure curve slope and consider allowable temperature rise (0.83°F)

3.

Thus, 7.78 feet in addition to current NPSHA must be provided.

Current = 7 feet+7.78 feet Newly Calculated = 14.78 feet Total NPSHR These calculations are conservative because they assume that the total temperature rise will take place at the suction of the pump. This tends to be valid at minimum flows, but is conservative at design flow. These examples do not encompass every possible scenario that could be experienced, but they are effective

(Actual) 790 (1 – 0.15) = 7.37°F 778 x 0.15 x 0.78 Conclusion: Slope of curve = 0.475 PSIA/°F Max allowable temperature rise allowed = 0.83°F Total hydraulic temperature rise = 7.37°F

FIGURE 3 PSI 140 120 100

STARTING POINT LESS HEAT INPUT FROM PUMP

80

VAPOR

LIQUID STATE

60 AFTER HYDRAULIC TEMPERATURE RISE

40 20 0 -50

0

50

100

150

200

250

DEGREES F PSI Series 1 Graph demonstrating the relationship of temperature to pressure for hydrofluoric acid

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77

guidelines in the determination of adequate NPSHA. In addition to these calculations, always multiply your final calculated NPSHR by a 1.3 safety factor to ensure a successful pump application. Clearly, increasing NPSHA can be an expensive proposition. However, it may be lower in cost than reinvesting money into a problem pump caused by borderline NPSHA versus NPSHR margins. NPSHA, flow and head play equally important roles when selecting a pump configuration.

sure in the motor to keep the acid liquid. These configurations are often referred to as pressurized or reverse circulation designs. Magnetically coupled pumps are also ideally suited to the handling of hydrofluoric acids. Mag-drive pumps with metallic containment shrouds (sometimes called cans) also produce eddy current losses that transmit heat to the pumpage. Some manufacturers offer different internal circulation paths, including rear-mounted impellers to compensate for pressure drop and temperature rise. PUMP CONFIGURATIONS Several mag-drives are available with nonmetallic shrouds. In these no Several effective pump configuheat is produced due to eddy current rations exist for handling hydrofluoric losses, which increases overall pump acid. Before selecting a configuration, efficiency and decreases motor accurately evaluate the NPSHA verrequirements in most cases. Shroud sus NPSHR, flow, head, efficiencies, materials include ceramics, silicon and temperature rise. carbide, PEEK and reinforced fluoroCanned motor pumps offer two plastics. Remember, factors such as designs that are extremely effective inlet temperature, NPSH, contamifor pumping hydrofluoric acid. These nates and system requirements must designs can pressurize the fluid in the be taken into account with either motor to increase vapor pressure canned motor or magnetically coumargins or to reverse the motor flow pled pumps. (internal circulation) direction, routRegardless of the pump style, ing the heated process to the suction several auxiliary items can smooth tank rather than the pump. These the path toward safe and effective design capabilities are important due operation. If low NPSHA is a factor, to the temperature gained from visan inducer can lower a pump’s cous drag, eddy current losses, and NPSHR. This can sometimes alleviate motor inefficiencies. Although differcostly system cha-nges. Over and ent, they fundamentally achieve the under current measuring relays effecsame end result, keeping the process tively protect against dry running. from flashing in the motor. The basic These simple devices enable the premise of both designs is to increase pump operator to set a minimum the pressure in the motor so that, amperage draw based on the specific even though the process temperature functional curve amperage draw. If is rising, there is still adequate presthe current drops below the set point, FIGURE 5 the pump will autoPSI matically shut 140 down. 120 Thermowells 100 STARTING POINT LESS HEAT INPUT FROM PUMP and temperature 80 VAPOR LIQUID STATE switches are also 60 AFTER effective in detect40 HYDRAULIC TEMPERATURE ing overheating of 20 RISE the process within 0 -50 0 50 100 150 200 250 the pump. Often DEGREES F when pumps or systems are experiPSI encing disturSeries 1 bances, process temperatures Vapor pressure versus temperature curve

78

The Pump Handbook Series

FIGURE 4 H (1-n) = Hydraulic temperature rise due to 778nCp inefficiencies of pump performance H = Rated head in feet at design flow n = Rated efficiency at design point Cp = Specific heat of a process fluid defined as BTU/lb°F

Equation for hydraulic temperature rise increase. These devices signal a possible problem and allow for review of the pump and system before extreme damage occurs. However, they are not independent of process temperature fluctuations and may be effective only in constant temperature applications. Bearing monitors are also important because they can detect problems, such as fracturing with silicon carbide bearing systems. Some of the most effective bearing wear monitors detect axial and radial wear. These monitors are important in scheduling proactive maintenance versus reactive maintenance, which is critical when unscheduled downtime can mean lost revenue. If pump metallurgies, bearing materials, and hydraulic and pump configurations are approached properly, pumping hydrofluoric acid can become as routine as brushing your teeth. ■

ACKNOWLEDGMENTS Table 1 and Figure 1 are © 1993 by Nace International. All rights reserved by Nace. Reprinted by permission. John V. Heronema has been with Sundstrand Fluid Handling for six years. He has held positions in manufacturing and quality engineering, and is currently a Product Engineer with Sundyne Canned Motor Pumps, Arvada, CO.

CENTRIFUGAL PUMPS HANDBOOK

User Perspective: When to Apply Mag Drive Pumps Making the move for the right reasons.

BY: MAURICE G. JACKSON agnetic drive centrifugal pumps offer an advantage over normal single mechanical seal centrifugal pumps by preventing fugitive emissions from leaking to the atmosphere. Given proper application and operating procedures, these pumps can perform for years without failure. Rather than discuss the design of these pumps - a subject that has already been thoroughly addressed in articles, papers and presentations - let’s review the justification for installing mag drives and provide installation keys to insure reliability of the investment.

M

KNOW SECONDARY CONTAINMENT OR CONTROL REQUIREMENTS Secondary containment and secondary control are important terms to understand when selecting your mag drive pump. Secondary containment insures the fluid will be contained if the primary can fails. Some mag drive suppliers accomplish containment by installing a secondary can around the primary unit. If the primary can develops a leak secondary control insures the leakage will be controlled to a defined amount, but not contained. In selecting a mag drive pump be sure to know whether secondary contain-

ment or secondary control is specified for your application.

DON’T INSTALL MAG DRIVES TO OVERCOME SYSTEM PROBLEMS

CONSIDER LIFE CYCLE COSTS

Magnetic drive pumps should not be installed to solve a maintenance problem, such as a troublesome mechanical seal, without first determining the real reason for the problem. Once the problem has been identified, insure that installation of mag drive pumps will not create a ripple effect. Typical pump and system problems to watch for are: • cavitation

Magnetic drive pumps are often the only alternative to meet government hazardous materials and safety regulations, such as OSHA 1910, requiring stringent levels of containment or control. In addition, many companies now have policies, odorfree imperatives for example, requiring strict control of emissions. However, for some zero emissions applications tandem seal pumps offer a viable alternative to mag drives and life cycle cost must be considered in the selection criteria. Table 1 shows calculations of life cycle costs of tandem seal versus mag drive pumps for two specified applications. In the first example, the mag drive pump has a significantly lower initial cost and operating costs only slightly higher than the tandem seal. In this application, because mag drives offer more reliable containment most users would select the mag drive. In the second example, however, the mag drive proves to be more costly in terms of both investment and operating cost, and use of the mag drive can not be justified in terms of cost alone.

The Pump Handbook Series

•

operating too far from best efficiency point (BEP)

•

Net positive suction head available (NPSHA) too low

•

slurries

•

pump operating without liquid in the unit

COMMUNICATE WITH YOUR VENDOR A user of magnetic drive pumps should also be aware of potential problems and communicate with the vendor to insure they are avoided. For example, the drive motor should always be sized smaller than the magnets to prevent decoupling if the impeller is overpowered. Decoupling of the magnets will cre-

79

ate excessive heat build-up in the fluid. Decoupling may also result in a locked rotor due to failure of the pump bearings and can. In addition, because magnetic drive pumps are often used to pump low – less than 1.0 – specific gravity fluids and are generally sized for such applications, an operator should be aware that employing the pump for water or higher specific gravity applications may overpower the motor or magnets. Failure to communicate fluid properties may lead to additional problems. A vendor will need to know more than fluid viscosity and specific gravity to size your magnetic drive pump. Users should also specify vapor pressure vs. temperature data and specific heat, as well as size and percentage of solids for the fluid being pumped.

USE PROTECTIVE INSTRUMENTATION TO INSURE RELIABILITY To insure the reliability of mag drive pumps, protective instrumentation is recommended. Listed below are some typical instrumentation available and their features. •

Power meter – monitors power to the motor driving the mag drive pump. The meter can be used to prevent dry and dead headed operation. The power meter is probably the best choice if you are limited to the selection of one type of monitoring instrumentation.

•

Can thermocouple – mounted on the can, it senses dry running and bearing problems.

•

Bearing wear detector – is used to sense the position of the shaft or rotor. It can provide an indication of the condition of the pump’s bearings.

•

80

Level detector – is placed in the suction or discharge piping to insure liquid flow to the pump.

TABLE 1 Based on one year1 operation Application One: 400 gpm; 120 ft head Pump Type

Tandem Seal

Magnetic Drive

Initial Cost Basic hp Required Seal hp Required Total hp Required kW•h Power Cost Product Loss Maintenance Cost Total operating cost per year

$13800 17.5 1.5 19 129400 $6470 $300 $600

$113502 20.4 NA 20.4 138900 $69453 $04 $6005

$7370

$75456

Application Two: 100 gpm; 240 ft head Initial Cost Basic hp Required Seal hp Required Total hp Required kW•h Power Cost Product Loss Maintenance Cost Total Operating Cost per year

$7000 11.4 1.5 12.9 84426 $4220 $300 $600

$80002 14 NA 14 95320 $47763 $04 $6005

$5120

$53766

Table data based on operation 350 days per year, 24 hours per day. Material of construction is 316 stainless steel. 3 Electric power calculated at $50 per 1000 kW. Electric motor efficiency of 92% assumed in calculation of kW usage. 4 Product loss calculated at $50 per pound. 5 Maintenance costs assumed a failure once every three years. The failure modes are assumed to be seal failure for the tandem seal and bearing failure for the mag drive. 6 Figures do not include initial cost. 1 2

Life cycle cost calculations for Tandem Seal Vs. Magnetic Drive Pumps

CONCLUSION Recent developments in magnetic drive pumps, harboring many functional and maintenance advantages for pump users, are testimony to an exciting future for magnetic drive centrifugal pumps. ■

The Pump Handbook Series

Maurice G. Jackson is a Engineering Associate in the Engineering Construction Division of Tennessee Eastman Division of Eastman Chemical Company, Kingsport, Tennessee. He has 25 years of experience in pump operation, maintenance and engineering.

CENTRIFUGAL PUMPS HANDBOOK

Interpreting Sealless Pump Failures The causes of part failures in sealless centrifugals may determine system and operational problems. ealless pump failures can highlight system or operational problems once taken for granted or blamed on mechanical seals. Once a pump has failed, it should be taken apart to identify the broken part or problem area. Frequently, a broken part can indicate the cause of failure. By establishing and remedying the origin of the failure, pump service life can be extended and future failures minimized. The following describes part failures and their causes that indicate system and operational problems.

S

DAMAGED THRUST SURFACES (FRONT OR REAR) Cause - operation below the acceptable minimum flow rate Many pump users think of minimum flow relative to temperature rise and bearing wear problems. However, extreme low-flow operation in a centrifugal pump can also create hydraulic imbalance of the impeller, generating thrusting and vibration. Because sealless pumps do not have the shaft overhang typical of sealed pumps, an imbalance can cause extreme axial shuttling of the rotating assembly which may break thrust surfaces. When a pump’s desired operating point is at a very low flow rate, check with the pump manufacturer for the minimum rate. If the desired flow rate is below the recommended minimum, add a recirculation loop to increase throughput and prevent hydraulic imbalance. Cause - insufficient net positive suction head (NPSH) available A sealless pump may require more NPSH to insure that the hydraulic balance is maintained and the bearing system contains enough fluid. NPSH problems can

result in shuttling of the rotating element, making it bang against thrust surfaces, and this can lead to rupture of the containment shell or liner. Repipe the system to reduce suction piping friction losses. Removing unnecessary valving or changing the pump elevation will solve the problem. Cause - water hammer With sealless pumps, water hammer can manifest itself by causing failure of the thrust surfaces as the rotating element is slammed against them. To solve the problem, review valve operating sequences and piping arrangements. Slow down valve closing speeds or change valve types to reduce water hammer.

INTERNAL SLEEVE BEARING FAILURE Cause - operating the pump dry Sealless pumps require fluid to cool the bearings. Lack of fluid passing through the bearings causes thermal expansion of the bearing or journal, depending on the particular design. This expansion constricts passages, increasing friction and heat, and thereby causing the pump rotating element to lock up and cease operating. Alternatively, if the dry running operation is short, the bearings may heat up enough to fracture due to thermal shock when fresh fluid is introduced. The first solution is to avoid running sealless pumps dry. If running

dry is possible due to the desired method of operation, install a recirculation line around the pump to insure that fluid will always be running through it. Cause - low fluid vapor pressure Pumping fluids at temperatures close to their vapor pressure can create problems. In a sealless pump, fluids close to vapor pressure can flash as the fluid, in passing, picks up heat from the containment shell or bearings. This additional temperature rise brings liquids closer to their vapor pressure, and only a small amount of additional heat from the bearings may increase the liquid temperature above the vapor pressure limits, causing the liquid to flash and preventing it from cooling the bearings. Bearing failure in a sealless pump requires prompt attention to minimize the cost of the repair and prevent external leakage of the fluid. If the fluid is close to the vapor pressure before it enters the pump, or if it has characteristics that suggest that a small temperature change will produce a large change in the vapor pressure, ask your sealless pump manufacturer to predict the expected temperature rise in order to verify that flashing will not occur in the bearings. To prevent flashing, some pump designs incorporate sec-

Magnetic Drive Pump The Pump Handbook Series

81

ondary pumping devices to increase pressure as fluid moves into the bearings. Other designs offer secondary cooling to solve the problem.

CONTAINMENT SHELL FAILURE Cause – drive magnet contacting the outer surface When the antifriction bearings supporting the drive magnet in a magnetic drive pump fail, the magnet may contact and rupture the containment shell. Such contact is indicated by grooves and rub marks on the external surface of the containment shell. If the pump design has safety rub rings installed, check clearances to insure they are correct. Replace the rings when repairing the antifriction bearings. Cause – internal pressure higher than containment shell design limits Water hammer can burst the containment shell. The sudden increase in pressure can drive rotating elements against thrust surfaces and put increased shock into the containment shell. Alternatively, the increased pressure alone can distort and burst

the containment shell. In this case, the containment shell will show signs of expansion or distortion from the inside out. Repipe the system to reduce suction piping friction loss. Removing unnecessary valving or changing the pump elevation will solve the problem. Cause – pump hydrotest pressure was above design limits Caution should be used when hydrotesting assembled pumps that have nonmetallic containment shells. Shells using a fiber fill for strength may not rupture the first time they are exposed to pressure, but the fibers inside the material may be broken. If so, the next time pressure hits them, the shells may burst due to ineffective fiber reinforcement. Review instruction manuals and technical data and do not hydrotest nonmetallic shells above recommended pressures. Note - Hydraulic thrust balance Magnetic drive pumps and canned motor pumps frequently have specific clearances that hydraulically control the amount of thrust which the rotating elements experience. When thrust surfaces or bearings fail, the subsequent internal rubbing that takes place can increase

wear on balancing surfaces and reduce the effectiveness of the hydraulic thrust control. If only the failed part is replaced, the next failure may then result from lack of hydraulic thrust control. When in doubt, always replace a hydraulic thrust surface part that is worn.

SUMMARY The best solution to pump failures is always prevention. Pump products should be properly applied at all times. Don’t hesitate to contact the manufacturer or his representative to ask for help, and be sure to describe the application, installation and operating conditions for the pump thoroughly. Also, save all parts. An examination of them may provide invaluable clues to the origins of pump failure, offering keys to overcoming systems or operational defects. ■

ABOUT THE AUTHOR: Charles A. Myers, Director of Sales and Marketing at IWAKI WALCHEM Corporation, Holliston, MA., has been working with sealless pumps for 14 years. He is active on ANSI and API sealless pump’s standards committees.

Rotor Assembly Stator Liner Rear Thrust Surface

Stator (Motor Windings)

Process Lubricated Sleeve Bearing

Forward Thrust Surface

Canned Motor Pump

82

Process Lubricated Sleeve Bearing

The Pump Handbook Series

CENTRIFUGAL PUMPS HANDBOOK

Magnetic Couplings for Sealless Pumps Elimination of seals ends leakage concerns.

BY RONALD P. SMITH BACKGROUND Electric-motor-driven pumps have been around for about 75 years, and so has the nagging problem of the shaft packing or seal. Because water was the common fluid pumped, it rarely became a dangerous problem. However, as the chemical industry developed, leakage became a major concern, and better seals were needed and developed. Industries are now under scrutiny for hazardous emissions of all types, and must comply with clean air and water regulations dictated by Congress and implemented by the EPA, OSHA, and other government agencies. Currently, any leakage of liquid or gas is a problem and must be minimized or eliminated. The state of the art for mechanical seals is in the range of 500 ppm leakage, with some releasing as little as 100 ppm. By using secondary seals with drainage and control instrumentation, levels closer to zero can be accomplished at increased cost to the user.

THE BASIC PROBLEM If we can accept that contacting surfaces with relative motion between them will eventually wear, then we can conclude that in the case of mechanical pump seals, leakage will ultimately occur. So it is desirable to do away with any shaft seal. By not penetrating the pump housing with a shaft, the seal is eliminated,

along with any potential leakage. This can be done by totally enclosing the impeller/pump assembly and isolating it from the prime mover. The question is, how do you drive the pump with no direct connection to a prime mover (motor)?

A SOLUTION Fortunately, we have a natural force, magnetism, that can be used to our benefit. As children, we experienced the magnetic force of two magnets operating through a table top or pane of glass. One magnet would follow the other until the gap between them became too large and reduced the force. That basic idea is used in synchronous magnetic couplings. There are two basic styles of magnetic couplings in use. Figure 1 shows a face-face coupling and Figure 2 illustrates a co-axial design. Magnetic couplings can be made to develop almost unlimited forces, based on choice of material and scale. Coupling designs for hundreds of footpounds of torque are available. One of the most fascinating aspects of permanent magnet couplings is that although they exhibit powerful forces of attraction and repulsion, they require no outside sources of power. If properly used, they last indefinitely.

COUPLING CHARACTERISTICS In any synchronous coupling, torque is developed in the same

The Pump Handbook Series

way. The maximum attractive force between the poles occurs when the poles are aligned in opposite polarity. Maximum repulsive force occurs when the same polarity is aligned. In both instances, the transverse force (torque) is at a null (zero). The latter position is the least stable. The maximum transverse force (torque) occurs between the two positions where the normal force is zero. Stable positions occur only once per pole pair, so in the case of a 10-pole coupling, there would be five stable positions. The proper application of a permanent magnet coupling requires knowledge of the maximum torque produced by the motor. This is typically twice the amount produced at the rated horsepower. Running torque = (Rated Horsepower x 5,250) /rpm (ft-lbs) In the case of a 5 Hp motor at 1,800 rpm with no load, the running speed with about 3% slip is 1,750 rpm and running torque is: (5 Hp x 5,250) /1,750 = 15 ft-lbs. However, the motor will develop about 30 ft-lbs peak torque during line start, and a magnetic coupling must have a peak torque rating at least that high to prevent loss of coupling. Figure 3 displays the relationship of peak to running torque. The amount of safety factor for the appli-

83

UNCOUPLING, SPECIAL CASES

FIGURE 1

INPUT SHAFT

STEEL

MAGNET

NONMAGNETIC MATERIAL MAGNET

STEEL

OUTPUT SHAFT

A face-face magnetic coupling cation will determine the exact design point on the curve. Slow start conditions can reduce the amount of peak torque required in the coupling and provide overload protection for impellers in case of a mechanical jam. In a coaxial coupling the radial forces are balanced if all of the magnet segments are of equal strength. Concentricity of the inner and outer assemblies is also required for equal air gap distance. These factors develop “magnetic balance,” which is as significant as physical balance in reducing noise and bearing wear. Face–face couplings develop significant axial forces. When in the aligned, attractive mode, the force is at a maximum. At peak torque, the axial force approaches zero. If slippage occurs, it goes through a maximum in the opposite direction. This coupling design requires proper bidirectional thrust bearings on each member to handle the variable

84

forces. Inadequate bearings will allow air gap variations that cause mechanical noise and can be self destructive. For this reason, the face–face coupling is generally restricted to special applications. The torque of face–face couplings is limited by the allowable maximum diameter of the assembly. A coaxial coupling can be made longer for increased torque once a maximum diameter is reached. The torque is essentially linear with axial length. This benefit makes the coaxial design the one of choice for most applications.

STIFFNESS If rapidly fluctuating loads cause mechanical resonance with the pump couplings, changing the number of poles will modify the stiffness. Relationships of pole spacing to gap length must be taken into account to maintain design efficiency. The Pump Handbook Series

The uncoupling phenomenon limits torque and is very useful. In pumps, it might protect the impeller from damage or detect unacceptable thermal conditions. Obviously, a coupling can be made to exceed the torque of the motor, as in a mechanical coupling, and use motor thermal or electrical overload protection to shut down the system. Most pumps are designed with a coupling that will not slip or uncouple within rated performance and proper motor application. In the unique case where the peak coupling torque is exceeded, slippage or uncoupling results. The impeller then stops, and no fluid is pumped. The seriousness of this situation will depend on the application and coupling design. The system should detect lack of flow and shut down the pump before any major damage occurs. A low-level audible warning may be heard from the coupling. Inertia of the system will not allow “pick up” of the impeller magnet until the motor is stopped. Before restarting the pump, the cause of the uncouple should be determined. Running uncoupled for long periods should be avoided. Because the impeller is not rotating and no fluid is being pumped, no fluid is being circulated through the containment can and no cooling of the coupling occurs. When one magnet element is rotating past the other, a significant amount of energy is converted to heat, and because the inner unit usually has a poor thermal escape path, it will get hot. If the temperature rises past the design point, demagnetization can occur. This is either temporary (recovered when the coupling cools down) or permanent (recovered only by remagnetizing). If a coupling has a nonmetallic barrier such as ceramic or plastic, there will be no additional uncoupling effect. Metallic barriers of stainless steel, Hastelloy, etc., will heat up rapidly due to eddy currents and, depending on the fluid contained, could represent a dangerous condition. If the additional heat raises the

FIGURE 2 Barrier (Flange seal to pump)

Magnets Follower Assembly

Motor Shaft

Pump Shaft

Driver

A co-axial magnetic coupling

temperature of the magnets, a further reduction of force by demagnetization is possible. Either of these cases could affect restart and necessitate a “cool down” before restart. This might not be a concern, because some time should be spent identifying the cause of the high torque requirement.

CONTAINMENT BARRIERS The containment barrier is a key element to sealless pump success. It provides the primary fluid containment and the “window” to couple torque in the system. Like other elements in the system, it usually is connected to the pump with a flange and an O-ring seal. The containment barrier is also a critical part of any permanent magnet coupling design. Its magnetic and electrical characteristics affect the heating and power losses of the system. The wall thickness and associated mechanical gaps determine the magnetic air gap and the amount of magnetic material required for a given torque, and therefore significantly impacts the cost of the coupling. Table 1 displays some of the common barrier materials, along with their benefits and related costs. Permanent magnet couplings can easily handle larger air gaps than

allowed in canned motors. This is a major benefit when handling high viscosity fluids or when suspended particles are in the fluid. Magnetic particles are to be avoided because they may collect between the magnet and barrier. Air gap clearance on either side of the barrier should be as small as possible, but their size depends on allowable bearing wear. If either rotating magnet assembly is allowed to contact the barrier, the pressure vessel may be compromised and failure can occur. Mechanical “rub rings” or proximity detectors can be used to indicate bearing failure. It is usually not practical to make the coupling barrier shell an integral part of the pump housing. This shell should have the thinnest wall possible that satisfies the design pressure requirements. The material must be nonmagnetic and preferably non-

metallic to reduce eddy current heating and associated power losses. There are many barrier designs, the most common being plastic shells for small pumps and stainless steel for large pumps, with pressure requirements up to thousands of pounds per square inch. Chemicals being pumped dictate the choice of material, and frequently the shell is made of the same material as the pump housing. Ceramic shells, coated metal, and laminated metal are used for special cases. In the case of solid metallic barriers, eddy current heating is developed. This is torque transfer loss and can amount to 5–10% of the input power. Generally ignored in small systems, it may be significant with motors over 100 Hp. Most cooling can be through the fluid if a generous flow within the barrier is established. Additional heat dissipation through the pump housing is possible if the barrier shell has a metal–metal contact to the housing. Eddy current heating can be reduced by lowering the speed of the motor, as losses are proportional to speed. Because the containment barrier becomes a pressure vessel, fabrication techniques are important. Designs are guided by ASME standard and manufacturing processes dictated by quantity. Small- to medium-size barriers are usually machined from solid bar stock in small quantities. Spun, hydroformed, or deep-drawn shells may be more economical in large quantities. Welded units, which are feasible in all sizes, require close process control to avoid stress corrosion problems. Pressure testing may be part of part certification.

TABLE 1- BARRIER MATERIAL COMPARISON Material Plastic Ceramic Stainless Steel Hastelloy Titanium

Wall Thickness Medium Thick Thin Thin Thin

Pressure Capability Medium Low/Medium High High High

The Pump Handbook Series

Chemical Resistance High Medium High High High

Eddy Current Heating No No Yes Yes Yes

Relative Cost Low High Medium High High

85

COUPLING DESIGN In concentric couplings, the driving element connected to the motor is usually the outer magnet assembly. This part of the magnetic elements has the highest mass and inertia. It can be made of magnetic iron and painted, plated, or coated as necessary. It is usually not subjected to high corrosion environments and does not require special sheathing. Salient magnet poles have gaps between them. These may be filled with epoxy or other potting compounds to improve cleaning and minimize magnet damage during assembly. The inner magnet assembly is the “follower.” Because it is in the process fluid, special care must be taken to prevent corrosion or contamination of the pumpage. This assembly typically has rare earth magnets mounted on an iron ring. If these

techniques are critical to long-term life, and leak or pressure testing is advised. Bathed in the process fluid, the follower assembly may be affected by temperature extremes. For temperatures over 100°C, efficiencies decrease and designs become more material specific. Factors affecting design are: • magnetic gap length (barrier + clearances) • peak torque required • space available for coupling • form factor desired for complement to pump (diameter x length) • stiffness required • fluid and corrosion concern • maximum operating temperature • running speed

FIGURE 3

Permanent magnet couplings are an integral part of a pump. Due primarily to the need for critical alignment, they are sold as a unit with the motor. Pumps are manufactured to meet industry standards such as those published by ANSI, API, and the Hydraulic Institute. These organizations have recently included standards for magnetic couplings. Standards are based on voluntary compliance and in most cases insure interchangeability of parts among manufacturers. To meet customer needs for quality in specific applications, pump manufacturers have their own rigorous standards. Permanent magnet materials are also produced to industry standards that allow sizeable variation in magnetic properties within material grades. Critical applications require greater control of properties and/or selection of parts for uniformity.

MAINTENANCE

Torque vs. Angle of Rotation Coupling Max. Motor Max.

Torque

Running/Operating Torque

Angle of Displacement The relationship of peak to running torque

parts will be corroded by the fluid, they are sheathed or coated with appropriate materials. Frequently, a stainless hub is used with a stainless sheath welded to it, totally encapsulating the magnet assembly. Welding

86

INDUSTRY STANDARDS

• shaft sizes • barrier material and allowable eddy current heating

The Pump Handbook Series

Because there is no mechanical wear in a magnetic coupling, there should be no need for maintenance. However, bearings do wear, and occasionally a pump will require disassembly or a motor will need to be replaced. The manufacturer will have recommendations for handling this operation. Disassembly of the magnetic coupling should be done only by trained personnel using the proper fixtures. Permanent magnet couplings contain some of the most powerful magnet materials ever made, and they are always energized. In systems of 5 Hp and over, the forces are greater than a person can control by hands alone. Large pump couplings have axial forces in the hundreds of pounds. Fixtures or mechanical means to guide the parts and prevent damage to the containment barrier are required during disassembly and re-assembly. Training is also required to avoid personal injury. Magnets can be very unforgiving of mistakes. ■ Ronald P. Smith is Manager of Engineering for the Magnetic Materials Division of Dexter Corporation.

CENTRIFUGAL PUMPS HANDBOOK

Suction Side Problems Gas Entrainment BY: JAMES H. INGRAM

M

ENTRAINMENT VERSUS CAVITATION The audible pump noise from noncondensable entrained gas will produce a crackling similar to cavitation or impeller recirculation. However, cavitation is produced by a vapor phase of the liquid which is condensable, while noncondensable entrained gas must enter and exit the pump with the liquid stream. To test for gas entrainment over mild cavitation, run the pump back upon the curve by slowly closing the discharge valve. The noise will diminish if it originated from cavitation and the pump is not prone to suction recirculation. In contrast, with entrained gas, continued performance at this portion of the curve will choke off or gas-bind the pump, causing unusually quiet operation or low flow.

A pump in this gas bound state, will not re-prime itself, and the gas, with some portion of the liquid, must be vented for a restart against a discharge head. The effort to restart a gas bound impeller depends on

GAS BOUND IMPELLERS

As a process stream containing entrained gas nears the impeller, the liquid pre-rotating from the impeller tends to centrifuge the gas from the process stream. Gas not passing into the impeller accumulates near the impeller FIGURE 1. ENCLOSED IMPELLER-ENTRAINED eye. As entrained gas GAS HANDLING PERFORMANCE flow continues to increase, the accumulatThe LaBour Company, Inc. Effect on head and capacity of ing groups of bubbles varying quantities of air with water being pumped. are pulled through the impeller into the dis160 charge vane area where NO AIR HEAD they initiate a fall in 2% 140 flow performance. The bubble choking effect at the impeller eye pro5% 120 duces a further reduction of Net Positive 8% 100 Suction Head Available (NPSHA). At this stage long term damage to the 80 10% pump from handling 12% entrained gases is gener60 ally negligible when 15% compared with the 50 damage due to cavita0 200 400 600 800 1000 1200 tion. If the process Capacity in U.S. Gallons per Minute stream gas volume increases, however, furSize: no. 55; Type: SQ. Speed: 1750 ther bubble build-up Impeller Diameter: 11” will occur, blocking off Air quantities given are in terms of free air at atmospheric the impeller eye and pressure referred to % of total volume of fluid being handled. stopping flow (Ref. 1). Head in Feet

ultiple symptoms associated with noncondensable suction side gas entrainment, such as loss of pump head, noisy operation, and erratic performance, often mislead the pump operator. As a result, entrained gas is generally diagnosed by eliminating other possible sources of performance problems. To adequately control gas entrainment a user should first be aware of systems most likely to produce gas, and then employ methods or designs to eliminate entrainment into these pumping systems.

The Pump Handbook Series

87

discusses open impeller pump modifications.)

FIGURE 2. OPEN IMPELLER-ENTRAINED GAS HANDLING PERFORMANCE Gould’s Pumps, Inc. Approximate Characteristic Curves of Centrifugal Pump

SYSTEMS PRODUCING ENTRAINED GAS

350

The most common conditions or mechanisms for introducing gas into the suction line are: 1. Vortexing

200 150

250

Efficiency %

250 Brake Horse Power (Bhp)

Head in Feet

300

80 70 60 50

Bhp@ sp gr=1.0

200

0% 2% 4% 6% 0% 2% 4% 6% 0% 6%

Previously flashed process liquid conveying flashed gas into the suction piping.

3.

Injection of gas, which does not go into solution, into the pumpage.

4.

Vacuum systems, valves, seals, flanges, or other equipment in a suction lift application allowing air to leak into the pumpage stream.

5.

Gas evolution from an incomplete or gas producing chemical reaction.

150

0

500

1000 1500 2000 2500 Gallons per Minute

3000

If a particular application produces entrained gas or has the potential to do so, the best solution is to eliminate as much entrainment as possible by applying corrective pump system design and/or a gas handling pump. If liquid gas mixing is desired, employ a static mixer on the dis-

Size: 6x8-18 Speed: 1780 rpm Impeller Diameter: 17 1/4”

impeller position, type and valving arrangement, among other variables. Degassing is easier to accomplish with a variable speed driver, such as a steam turbine, than with a constant speed electric motor drive. In addition, a recycle line to the suction vessel vapor space is often an effective method for degassing an impeller, since with this arrangement the pump is not required to work against a discharge head. (Ref. 1 describes methods for venting gas on modified pumps that are gas bound.) As a rule, if the probability of entrained gas exists from a chemical reaction, the inlet piping design should incorporate a means to vent the vapor back to the suction vessel’s vapor space or to some other source.

EFFECTS OF ENTRAINED GAS ON PUMP PERFORMANCE Figures 1 and 2 illustrate the effect of entrained gas on a LaBour enclosed impeller and a Gould’s paper stock open impeller. As illustrated by the figures, 2% entrained gas does not produce a significant head curve drop. Note that while the LaBour impeller experiences a 22%

88

2.

head loss at 5% gas volume, the Gould’s open impeller experiences a 12% head loss at this volume. Some open impeller paper stock designs can actually handle up to 10% entrained FIGURE 3. DEVELOPMENT OF A VORTEX gas because clearance between the case and impeller vanes allows more turbulence in the (a) process fluid, which tends to break up gas accumulation more efficiently than an enclosed impeller with wear rings. In addition, other designs, such (d) as a recessed im(b) peller pump, may A. Incoherent surface swirl handle up to 18% entrained gas. In B. Surface dimple with coherent fact, most standard surface swirl centrifugal pumps handle up to 3% C. Vortex pulling air bubbles to entrained gas volintake ume at suction conditions without D. Fully developed vortex with air AIR difficulty. (Ref. 2 core to nozzle outlet (c)

The Pump Handbook Series

FIGURE 4. “HAT” TYPE VORTEX BREAKER

charge of the pump. In addition, an anticipated drop in pump head due to an entrained gas situation may be offset by oversizing the impeller. Of the five aforementioned mechanisms, vortexing is the most common source of entrained gas. Therefore, a user should be especially cautious employing mechanical equipment, such as tangential flash gas separators and column bottoms re-boilers, likely to produce a strong vortex.

VORTEX BREAKER DESIGN The extent of gas entrainment in the pumped fluid as the result of vortex formation depends on the strength of the vortex, the submergence to pump suction outlet, and the liquid velocity in the pump suction nozzle outlet. Vortices form not only through gravity draining vessel applications, but also in steady state draining vessels, and in vessels under pressure or with submerged pump suction inlets. Vortex formation follows conservation of angular momentum. As fluid moves toward the vessel outlet, the tangential velocity component in the fluid increases as the radius from the outlet decreases. Figure 3 shows various stages of vortex development. The first phase is a surface dimple. This dimple must sense a high enough exit velocity to extend from the surface and form a vortex. (For experimental observations regarding vortex formation see Refs. 3, 4.) The most effective method to

FIGURE 5. “CROSS” TYPE VORTEX BREAKER

eliminate entrained gas in pump suction piping is to prevent vortex formation either by avoiding vortex introducing mechanisms or by employing an appropriate vortex breaker at the vessel outlet. A ”hat” type vortex breaker, illustrated in Figure 4, covers the vessel outlet nozzle to reduce the effective outlet velocity. This design doesn’t allow a vortex to stabilize because the fluid surface senses only the annular velocity at the hat outside diameter (OD). In addition, the vanes supporting the hat introduce a shear in the vicinity of the outlet to further inhibit vortex formation. An annular velocity of 1/2 ft/sec at the hat OD produces a viable solution. Variations in hat diameters from 4d to 5d and hat annular openings of d/2 to d/3 are acceptable when annular velocity criteria are met. Annular design velocities of more than 1 ft/sec are not recommended. ”Cross” type breakers, installed above or inserted in vessel nozzle outlets as shown in Figure 5, work for some applications by providing additional shear to inhibit a mild vortex from feeding gas into a nozzle outlet (providing enough submergence is available). However, this design will not stop a strong vortex and will decrease NPSHA. A user should be aware of these limitations.

COLUMN VORTEXING If a column draw-off pump is erratic and/or nearly uncontrollable, a vortex may be feeding gas into the

The Pump Handbook Series

draw-off nozzle of the pump as illustrated by Figure 6a. It may be difficult to understand how a pump with 60 ft of vertical suction could be affected by entrained gas, but in this real case example Murphy’s law applied twice. First, since the pump system in question has a NPSHA greater than 50 ft, the piping designer employed a smaller suction pipe with a liquid velocity of 10 ft/sec. Second, the column draw-off nozzle was sized according to normal fluid velocity practice. As a result, the tray liquid had an exit velocity of 5 ft/sec with a liquid level 6-in. above the top of the draw-off nozzle and a vortex formed, feeding gas into the draw-off nozzle. As in the above example, due to a lack of proper submergence, gas is carried into the pump suction piping as a high liquid downward velocity exceeds the upward velocity of a gas bubble. Many draw-off vortexing problems may be eliminated by proper pump system design or by one of two vortex breaker designs illustrated by Figures 6b and c. The selection of the breaker design may depend on the downcomer arrangement and space limitations. The most effective vortex breaker is the slotted pipe design shown in Figure 6c. Application of these corrective pump systems designs or installation of an appropriate gas handling pump can solve suction side gas entrainment problems, resulting in a smoother process operation. ■

89

FIGURE 6. DESIGN MODIFICATIONS FOR A SYSTEM EXHIBITING A LACK OF ADEQUATE SUBMERGENCE AND PROHIBITIVELY HIGH EXIT VELOCITY DOWNSPOUT OR DOWNCOMER FROM TRAY ABOVE

6”

BUBBLE CAP

REFERENCES 1.

Doolin, John H., ”Centrifugal Pumps and Entrained-Air Problem,” Chemical Engineering, pp.103-106 (1963)

2.

Cappellino, C.A., Roll, R. and Wilson, George, ”Centrifugal Pump Design Considerations and Application Guidelines for Pumping Liquids with Entrained Gas,” 9th Texas A&M Pump Symposium 1992

3.

Patterson, F.M., ”Vortexing can be Prevented in Process Vessels and Tanks,” Oil and Gas Journal, pp. 118-120 (1969)

4.

Patterson, F.M., and Springer, E.K., ”Experimental Investigation of Critical Submergence for Vortexing in a Vertical Cylinder Tank,” ASME Paper 69-FE-49 (1969)

5.

Kern, Robert, ”How to Design Piping for Pump Suction Conditions,” Chemical Engineering, pp.119-126 (1975)

10”

10’/sec. Figure 6a. Tray take off nozzle with vortex from lack of correct submergence and too high exit velocity. EXTEND PLATE FROM VESSEL WALL. CHECK VELOCITY AT PLATE EDGE ≤ 1/2’/sec.

James H. Ingram is an Engineering Technologist with Sterling Chemicals in Texas City.

1/2’/sec.

Figure 6b. Plate extension over outlet nozzle lowers high outlet velocity. AREA OF SLOTS—3X PIPE CROSS SECTION AREA. CHECK VELOCITY INTO SLOT AREA ≤ 1’/sec.

Figure 6c. Slotted pipe vortex breaker.

90

The Pump Handbook Series

CENTRIFUGAL PUMPS HANDBOOK

Nozzle Loading – Who Sets the Standards? Or, to what extent should the pump be used as a piping anchor?

BY: KIMBERLY FORTIER, ASSISTANT EDITOR his past year’s Texas A&M International Pump Users Symposium at the George R. Brown convention center in Houston, TX included a discussion group entitled Nozzle Loading and Pump Operability co-coordinated by John Joseph of Amoco Oil and Willie Eickmann of Houston Lighting and Power. According to Gary Glidden, also a discussion leader for this group, the two day discussion was a ”standing room only” affair. Clearly, nozzle loading is a subject of concern to pump users.

T

ESTABLISHED LOADING STANDARDS QUESTIONED Much of the discussion focused on the difficulty of establishing standards for allowable nozzle loads. Although the current API 610 7th edition standard for centrifugal pumps in general refinery service provides values for maximum loads (Table 1, Figure 1), many pump users believe the API allowable loads are too high—especially for use as specifications for installation designs which fail to recognize the possibility of ”unplanned” stresses on the piping, such as those produced by foundation settling. However, as noted by James E. Steiger in his paper, API 610 Baseplate and Nozzle Loading Criteria, ”Before the 6th Edition of API 610 was published, there were no indus-

TABLE 1. API ALLOWABLE NOZZLE LOADS Note: Each value shown below indicates a range from minus that value to plus that value; for example, 160 indicates a range from -160 to +160. Nominal Size of Nozzle Flange (inches) Force/Moment*

2

3

4

6

8

10

12

14

16

Each top nozzle FX FY FZ FR

160 200 130 290

240 300 200 430

320 400 260 570

560 700 460 1010

850 1100 700 1560

1200 1500 1000 2200

1500 1800 1200 2600

1600 2000 1300 2900

1900 2300 1500 3300

Each side nozzle FX FY FZ FR

160 130 200 290

240 200 300 430

320 260 400 570

560 460 700 1010

850 700 1100 1560

1200 1000 1500 2200

1500 1200 1800 2600

1600 1300 2000 2900

1900 1500 2300 3300

Each end nozzle FX FY FZ FR

200 130 160 290

300 200 240 430

400 260 320 570

700 460 560 1010

1100 700 850 1560

1500 1000 1200 2200

1800 1200 1500 2600

2000 1300 1600 2900

2300 1500 1900 3300

Each nozzle MX MY MZ MR

340 260 170 460

700 530 350 950

980 740 500 1330

1700 1300 870 2310

2600 1900 1300 3500

3700 2800 1800 5000

4500 3400 2200 6100

4700 3500 2300 6300

5400 4000 2700 7200

*F = force, in pounds; M = moment, in foot-pounds; R = resultant

try accepted standards for allowable piping loads acting on centrifugal pumps.” Moreover, when these piping load standards are absent or not universally accepted by pump users, The Pump Handbook Series

manufacturers and piping engineers, these groups tend to set independent, often contrary standards, further complicating the design process. In an attempt to overcome these

91

FIGURE 1. COORDINATE SYSTEM FOR THE API FORCES AND MOMENTS Shaft Centerline

Y Z

Y

Y X

Shaft Centerline

X Z

Z

Y Z

X

Vertical In-Line Pumps

complications, one user developed a standard operating procedure based on measured changes in pump alignment to be applied universally throughout their plant. Changes in alignment subsequent to connection of suction and discharge lines indicate shaft deflection. This user set the maximum shaft deflection at 0.002” regardless of pump size or configuration. However, because this procedure relies on establishing a baseline alignment before the lines are connected, this standard cannot be applied to all pumps. For example, the feedwater pumps Glidden operates at Houston Lighting and Power employ welded nozzles, which don’t allow the pump to be isolated in order to determine the ”zero-load” alignment, as opposed to flanged nozzles. And, since this procedure depends on measuring alignment rather than forces and moments as for the API specifications, making a correlation between the two standards is ”almost apples and oranges,” says Glidden. Figure 2 illustrates a common consequence of nozzle loading on a pump. While the case bows in one direction due to piping loads, the shaft sags in the opposite direction as a result of thermal deformation. Pump operators witness the end results of overloading a pump nozzle in misalignment, vibration, bearing

92

X Pedestal Centerline

Horizontal Pumps with Side Suction and Side Discharge Nozzles or coupling failure, and shortened seal life; but, according to some, the established standard fails to address the correlation between loading levels and these failure modes. And, these users are concerned that relatively slight levels of nozzle loading, even those within API specifications, may have costly ramifications, in terms of downtime and pump life, in the long run. In fact, according to Joseph, the discussion at the Pump Users Symposium quickly progressed beyond the question, ”How much (loading) is too much,” to whether the pump should ”even be considered an anchor for the piping.” Joseph concludes, ”The piping (should) exert as little force as practically possible during operation.”

CAREFUL PIPING DESIGN CAN REDUCE STRESSES But how much is ”practically possible”? Joseph recommends ”shooting for 10% or less of API (allowable nozzle loading specifications) during running conditions.” He also points out, however, that piping stresses can be reduced to zero, ”My personal preference under hot conditions is that the piping exert nearly zero forces and moments.” In most cases to obtain zero stress under hot conditions requires exerting some stress on the piping in the cold condition. These stresses will then relax The Pump Handbook Series

with the thermal growth that occurs during hot operation. Joseph favors careful calculations in the design phase to insure that ”the spring hanger forces and the deflection of the beam they’re supported from matches the weight and growth of the piping when it’s full of liquid at temperature.” For example, a spring hanger supporting a 20’ straight vertical piping section might relax a full 1/4-1/2” due entirely to thermal growth in the vertical direction. Add this growth to the pull of the processliquid weight and the result is, the piping stress and strains during hot running differ drastically from those prior to start-up. One operator actually measured a 0.150” horizontal movement of the pump. ”You’ve got to think, what is it (the pump) going to look like with a hot flow of liquid and then back calculate to the cold, empty position you want that pipe at,” says Joseph. To obtain minimum loading during the running condition, the pipe should be supported in a position requiring it to be pulled down to the pump. During hot operation, the thermal growth of the piping and the weight of the liquid will then depress the piping into the relaxed position.

ECONOMICS Piping engineers counter these arguments for low to zero piping loads, claiming, as Steiger notes, that ”the pump manufacturers and rotating equipment engineers are too conservative and the higher piping loads do not usually lead to significant operability problems.” The larger piping loads are desirable because they result in simpler and significantly less expensive piping configurations. Yet, the 1985 Pressure Vessel Research Committee (PVRC) Pump-Piping Interaction Experience Survey indicates that there is ”a significant pump-piping interaction problem and that it has an annual impact on the order of one halfbillion dollars” (Ref. 1). ”Economics plays a big role in these decisions,” adds Glidden, ”However, if you do a bad job up front, this will compound, resulting in a terrible-running pump.”

CONFRONTING THE ROOT CAUSE Understanding how the base plate and piping design relate provides one key to maintaining shaft alignment and thereby pump reliability. However, as Steiger maintains, ”The pump-baseplate assembly represents a complex structure whose response to piping loads is difficult to predict with a high degree of certainty.” Joseph agrees, ”While some pump cases and base plate foundation designs can take very high loads, there are others for which simply tightening a nut one flat at a time changes alignment significantly.” If the pump case and base plate construction is reasonably rigid, higher forces may be applied with little deflection at the coupling. ”However,” Joseph warns, ”if users depend on the pump case and the base plate to provide the rigidity, and the piping is significantly high in stress application, then they’re just covering one problem with another solution. They’re not getting at the root cause—the piping strain.” Joseph recommends proper warm-up procedures and piping supports, in addition to good piping design, as the primary remedy for piping strain. His recommendations are outlined as follows: • Warm-up procedures Large, hot (and, as some users have pointed out, expensive) pumps are especially affected by nozzle loading, and for these pumps proper warm-up is essential. API recommends that the warm-up procedure be very well thought out. Without proper warm-up the pump may suffer from uneven thermal growth in the piping, case, shaft, stuffing boxes, and bearing housings. As the pump comes to equilibrium, it will experience transient thermal growth which may put it under considerable stress. The design of the piping is also crucial to adequate warm-up and should allow hot product to flow from the discharge line to the bottom of the pump case. Many operators bring hot product to the pump by employing a bypass to the discharge check valve. This is a small piece of pipe with a block valve which enables the dis-

charge pressure product FIGURE 2. RESULTS OF NOZZLE LOADING, to by-pass the discharge COMPOUNDED BY HOT OPERATION check valve and follow the discharge nozzle, then cross the top of the pump and exit at the Shaft suction nozzle. This method is inadequate because it heats only the top of the pump, resulting in a 150-300°F differential within the Casing pump. This temperature differential creates a very large humping tem deflects, the pump may become tendency since the pump expands the anchor for the piping. much more at the top than the botPlacing a firm anchor on a knee tom. As a result, the bearing brackets brace right at the pump base may might shift down at both ends of the provide a good solution. The knee pump. The rotor bows in the same brace, which should also be supportdirection. In fact, the rotor will actued by spring hangers, forces the pipally roll over by hand about 70° until ing up and away from the pump, so the wear rings rub. A few minutes that when the load changes as the later, the heat in the top of the rotor process liquid flows into the piping, it will allow it to roll about 70° further. will relax down to the pump, and the To assure that hot product is disspring hangers will bear the full load, pensed to the bottom of the pump, allowing the pump to operate with the product should be evenly distribvery low stresses. uted in all suction and discharge cavities at the drain connections. Even CONCLUSION distribution provides the best opportunity for thorough heat delivery to The response to the discussion the pump case, rotor, and group at the Pump Users Symposium discharge/suction piping, prior to indicates a real need for these kinds pushing the start button. Slow rolling of practical solutions to nozzle loadthe pump will also aid warm-up. Esing. The pump operators present at tablish the slow roll at 50-100 rpm the Symposium recognized that an and then initiate the warm-up flow to inexpensive piping design can be the pump with hot product. costly down the road. And, many are hopeful that the 8th edition of the • Piping supports API 610 standard, currently in Bad installation, deteriorating progress, will advance one step furhangers and foundation settlement ther toward an agreeable solution for are some of the most common causes more uniform nozzle loading pracof piping strain. Piping should be tices. ■ well supported by spring hangers, anchors, expansion loops or compresREFERENCES sion spring cans. In addition, all piping supports should have adjustment 1. Steiger, James E. ”API 610 capability to enable repositioning in Baseplate and Nozzle Loading response to deterioration and settleCriteria,” Proceedings of the Third ment. Glidden envisions a monitoring International Pump Symposium, system, which would examine align(1985). pp. 113-129. ment once a year or so, to test for 2. American Petroleum Institute, changes due to deterioration over ”Centrifugal Pumps for General extended periods. Refinery Services,” API Standard The hangers and other supports, 610, Seventh Edition (1989). including the beams that support these hangers, should be designed to support the weight of the piping and Kimberly Fortier is Assistant Editor the process-liquid. If the support sysfor Pumps and Systems. The Pump Handbook Series

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CENTRIFUGAL PUMPS HANDBOOK

Low Flow Options This service range demands an innovative approach. BY PUMPS AND SYSTEMS STAFF rocess requirements often demand capacities below those achievable with a conventional centrifugal pump. Figure 1 illustrates the range of service conditions considered to be low flow. The minimum continuous stable flow of a typical 1”x2”x7” overhung pump at 1800 rpm is approximately 7 gpm, while at 3550 rpm the minimum continuous stable flow is about 13 gpm. A pump of this size will produce about 240 ft. of head. As the head requirement increases to 5000 ft., the minimum continuous stable flow will increase to about 190 gpm. Conventional centrifugal pumps will not handle these low capacities very well for two main reasons:

P

Suction recirculation The minimum continuous stable flow is usually set by the pump manufacturer to avoid suction recirculation. Suction recirculation results in increased vibration and imparts continuous axial movement to the shaft, decreasing the life of bearings and mechanical seals. The point at which suction recirculation begins may be calculated as described by Dr. S. Gopalakrishnan in his presentation at the 5th International Pump Users Symposium in 1988. The pump manufacturer should perform these calculations and set the pump minimum continuous stable flow at a capacity greater than the calculated capacity. Temperature rise The ultimate limitation on low capacity is minimum continuous

94

FIGURE 1

thermal flow. Temperature rise through a pump determines the minimum flow rate. The maximum safe temperature rise through a pump should be limited to 10°F. The formula for determining thermal rise through a pump is: δT =

H x 1 778Cp (Eff - 1) H = total head in feet Cp = specific heat of the liquid in Btu x °F lb 788 ft-lbs = the energy to raise the temperature of one pound of water by 1°F

PARTIAL EMISSION PUMPS The type of pump most frequently applied to fulfill low flow

The Pump Handbook Series

requirements is a single port diffuser pump with a ”Barske” straight vane impeller close coupled to an electric motor, also known as a partial emission pump (Figure 2). Theoretically, in this kind of pump, the only liquid discharged as each chamber passes the diffuser port is the liquid between the impeller vanes. In reality, however, due to the clearance between the case and impeller, some additional liquid also gets swept out the diffuser port. Unfortunately, this pump has a head capacity that droops at shutoff which inhibits the ability to control the pump capacity by increasing pressure with decreasing flow (Figure 3). As a result, installation of a flowmeter is necessary to effectively control this type of pump.

FIGURE 2. BARSKE STRAIGHT VANE IMPELLER WITH SINGLE PORT DIFFUSER PARTIAL EMISSION PUMP

Impeller Pump Casing

Conical Diffuser

Diffuser Throat

FIGURE 3. TYPICAL CURVE SHAPES “Barske” Impeller-With Single Port Diffuser

120

4 Vane Impeller Volute Case

Head-Capacity Curves

100 90 80 Total Head In Feet

80

4 Vane / Volute

Efficiency %

70 60

60

50 Efficiency % Curves

40

40

30 20

20

Barske Impeller Single Port Diffuser

10 0

0 0

20

40

60 80 100 Gallons Per Minute

120

140

Because the characteristic curve for the ”Barske” impeller, also referred to as a high solidity impeller, is exactly the opposite of a centrifugal pump (where increasing the number of impeller vanes will flatten the curve and eventually cause a droop), the droop of the head capacity curve towards shutoff can be minimized in a single port diffuser pump by increasing the number of impeller vanes. Another method of eliminating head capacity droop is to install a discharge orifice. Since the friction across an orifice increases as the flow increases, the pressure of a discharge orifice will increase the pump curve slope so that the pump can be pressure controlled. Unfortunately, a discharge orifice decreases the pump efficiency. The partial emission pump is also available with an integral gear (either single or double increaser) to produce a higher pressure head than a single stage pump. Since high head application with the integral gear may call for speeds up to 20,000 rpm, an axial flow inducer is often employed in conjunction with this gear to lower the net positive suction head (NPSH). Another means to achieve low flow combined with high head requirements is to drive the pump with a special motor capable of high speed. Application of a variable frequency drive will produce speeds nearing 7200 rpm. With this type of construction a partial emission pump may also be coupled to a canned motor for sealless pump construction. One manufacturer builds the partial emission type pump with an in-line configuration, giving the pump its own bearing frame. In this configuration the pump is flexibly coupled to a standard vertical solid shaft motor. This same manufacturer also builds this pump in a horizontal centerline-mounted configuration.

FLOW RESTRICTION DEVICES Conventional centrifugal pumps can handle low flow conditions with the incorporation of a restriction

The Pump Handbook Series

95

device on the discharge to shift the best efficiency point (BEP) capacity back toward shutoff and increase the pump curve slope. Unlike the partial emission pumps which employ a construction requiring removal of a motor with a special shaft extension for mounting the impeller, or in the case of high speed applications removal of the motor and gear, the application of flow restriction devices on conventional API or ANSI pumps provides the benefit of an easily maintained single stage pump. Even though a restriction device reduces the efficiency by a considerable amount, low flow pumps are generally low horsepower machines, so consuming a little more horsepower to obtain a steep curve rise previous to shutoff is a small price to pay for the more desirable performance. Moreover, the required motor horsepower for the restricted pump is less than that for a non-restricted pump, as the restriction will not allow the pump to run to the extended portion of the curve. The use of an orifice to restrict flow will produce the desired performance. However, if the orifice diameter is considerably smaller than the pump discharge and discharge piping, cavitation and noise may occur on the downstream side of the orifice. For this reason one pump manufacturer incorporates a venturi to modify pump performance. The advantage of the venturi is that the gradual taper down to the required hole size then back up to the discharge pipe size effectively eliminates the cavitation, noise and vibration. Pumps equipped with venturi have been observed to run smoother and quieter as they approach shutoff.

tions the pump will produce heads up to about 1000 ft. A primary advantage of the vertical pump is the ability to stack many stages so that a low capacity impeller of fairly good efficiency will produce a high head. These pumps usually incorporate two, sometimes three or four, impeller designs of various capacities. Mixing impellers will result in a rated point capacity very near BEP. There is a limit, of course, to how many stages a vertical can pump may have. The limiting factors are shaft diameter size required to transmit the horsepower and torque and the availability of shafting in long sections (usually 20 feet). Another limitation is dependent on the machining tolerances of the register fittings of the bowl assembly. Since the tolerances are additive as the bowl is assembled, they may cause shaft binding if they are not tight enough. •

OTHER PUMP OPTIONS •

96

Another type of centrifugal pump that will operate effectively in the low flow range is a vertical can pump. A 5-6 in. diameter bowl assembly will experience its BEP capacity in the 60-120 gpm range at 3600 rpm. At these operating condi-

•

Regenerative turbine pumps will also fulfill low flow requirements. These pumps, available in single and multistage construction, have a very steep head characteristic and will operate on pressure control. The regenerative turbine does not demonstrate any apparent problems with minimum continuous stable flow, so the only limiting factor to set minimum flow is temperature rise. The formula for temperature rise through these pumps is identical to that for centrifugal pumps. The disadvantage of a regenerative turbine is the close internal clearances required to produce the pumping action. To accommodate this close clearance, the pumpage must be very clean. Gear or other rotary positive displacement pumps also will operate in the low flow range without difficulty. These pumps do not, however, operate well in low viscosity services.

The Pump Handbook Series

•

Controlled volume metering pumps can be applied for low flow services and are one of the few types of pumps that will operate at flow rates below 1 gpm. The disadvantages of using a metering pump are the inherent pulsations which may damage downstream piping and instruments. Pulsation dampeners help to smooth out pulsations but never entirely eliminate them.

CASTING LIMITATIONS Development of a truly efficient low capacity centrifugal pump requires prohibitively small liquid passages. These small passages are troublesome to produce in the casting process because the sand mold is prone to collapse at such small sizes and small interior passages are difficult to clean to the degree required for good efficiency in operation. A semi-open impeller is easier to cast and clean. This design is, however, in violation of API 610, which calls for an enclosed impeller cast in one piece. If sufficient advantages of the semi-open configuration are demonstrated, this standard might be changed. Very small impellers might even be machined from billet stock (similar to some centrifugal compressor impellers), thus eliminating all of the casting problems. Similarly, the casing of a low flow pump is difficult to cast and clean, requiring very small passageways which must have a smooth surface in order to produce good efficiency. This obstacle to producing a low flow pump case might be overcome by eliminating the need for a case casting in favor of a machined and fabricated construction.

EVALUATING HYDRAULIC FIT The fact of the matter is most manufacturers usually make little profit on their small model pumps. To convince manufacturers that quality low flow pumps are actually in demand, users must let them know that their quotations for pumps in the low flow area are being

evaluated for hydraulic fit. One method of evaluating hydraulic fit is shown in Figure 4. This evaluating tool adds a penalty, as a percentage multiplier, to the pump price for rated capacity to the left of BEP capacity. This tool is based on the fact that a higher suction specific speed correlates with a smaller stable window of operation. Applying this tool consistently and sending it along with your request for quotations will convince pump manufacturers that low flow performance is an area of hydraulic design that needs to be addressed. ■

FIGURE 4

REFERENCES F. H. Fraser, Recirculation in Centrifugal Pumps, presented at the ASME Winter Annual Meeting (1981). S. Gopalakrishnan, A New Method for Computing Minimum Flow, presented at 5th International Pump Users Symposium (1988). U. M. Barske, Design of Open Impeller Centrifugal Pumps, Royal Aircraft Establishment, Farnborough, Technical Note No. RPD 77 (January, 1953). Trygve Dahl, Centrifugal Pump Hydraulics for Low Specific Speed Applications, presented at 6th International Pump Users Symposium (1989).

ACKNOWLEDGMENTS Pumps and Systems would like to thank the members of its User Advisory Team for their assistance in preparing this article.

The Pump Handbook Series

97

CENTRIFUGAL PUMPS HANDBOOK

Pump Design Changes Improve Lubrication Quantifying the benefits of modifications.

BY LEV NELIK

P

roper lubrication is a key to long, trouble-free life of centrifugal pump bearings. In recent years the issue of lubrication has received renewed attention from pump users in chemical plants, pulp and paper mills, refineries, and other industries. Budgetary pressures have forced many plants to reduce maintenance capital. Many knowledgeable maintenance workers have been laid off. Not surprisingly, the ability to maintain pumping equipment properly is reduced, resulting in increased outages, lost production, and rising maintenance costs. Users have started to look to pump manufacturers to pick up the slack and help solve pump reliability problems, extend component life, and increase mean time between between failure (MTBF) and mean time between scheduled maintenance (MTBSM). Statistics show (Ref. 1) that most pump failures are related to bearings and seals. In this article we will look at bearings, analyzing how design changes affect bearing life in a quantifiable way. The need for improved pump reliability and increased MTBF led to a new design, introduced by Goulds in 1990/1991. Figure 1 shows cross sections of two single-stage, end-suction ANSI pumps. Both have identical wet ends (impeller and casing), but the power end and the seal chambers are different. Improvements in the

98

seal chamber are signifiFIGURE 1 cant. The new design has a larger chamber to ensure better heat transfer and cooler operation of the mechanical seals. The previous design incorporated a tight stuffing box. The new power end design in Figure 1 features approximately three times the volume of the oil sump (I), an oil level sight glass (II) to assure the proper oil level versus the constant level oiler (III), improved cooling via a finned cooler insert (IV) versus bottom cooling pockets (V), labyrinth oilframe seals (VI) versus lip seals (VII), and stiffer footing (VIII) for reduced vibrations. A testing program has been conducted to compare the two designs under extremely adverse operating conditions, such as running endurance testing at Cross sectional views of old and new power overspeed and below end designs. minimum flow. This program was conducted at the R&D lab of the Feedback from users comparing two Technology Center at Goulds, resultdesigns was also obtained , specificaling in quantifiable correlations ly in relation to the operating temperbetween changes in pump design ature of the bearing frame surfaces. and their effect on life extension. The Pump Handbook Series

FIGURE 2

approximately 13% longer life.

INCREASED OIL SUMP DEPTH A deeper sump allows contaminants to settle farther from moving parts, resulting in a cleaner layer of oil near the ball bearings (Figure 2). Contamination of the Larger sump results in reduced concentration bearing races and the of contaminants, which settle to the bottom. balls is the cause of microscopic deterioration of load surfaces, ANALYSIS OF THE POWER END leading to failure. Statistics show (Ref. With regard to the power end 2, 4) that a cleaner oil operation can (Figure 1), the belief that “the bigger increase bearing life by nearly 2.1 the better” is not uncommon in the times (Ref. 3). Similarly, due to pumping community. This idea has decreased air concentration, the oil some merit, but manufacturers often overlook the importance of quantifying the benefits of a particular design or FIGURE 3 modification. Frequently, little information is given as to how much life extension can be obtained by, say, having a deeper sump, or how much added value and savings can be realized from the increased bearing frame heat transfer surface. It is clear that a systematic approach to identify, measure, and improve pump component design is impossible without a proper balance of theory, experimentation, user feedback, and data from real world installations. Theory and experimentation should be balanced by clear communication between manufacturers and users.

oxidation rate by air is reduced for the larger sump. Again, for the type of pump studied in this work, this results in a 2% extension in bearing life (Ref. 3).

LABYRINTH OIL VERSUS LIP SEALS The effects of oil contamination are further reduced by improved oil seals. A proprietary labyrinth seal design was tested against the lip seal. Both pumps were sprayed with water from a hose, simulating plant washdown. The spray was directed at various angles to the frame at the oil and seals area. The oil was then analyzed for water content. It was found that the previous design equipped with lip seals contained 3% water after 30 minutes of spraying, while the new design, with labyrinth seals, showed no water at all. Also, lip seals may

INCREASED FRAME OUTSIDE HEAT TRANSFER SURFACE Heat is transferred from the pump bearings to the oil and through the housing frame walls to the outside air. Some of the heat is also conducted through the casing to and from the pumpage, depending on the temperature of each. Typically, the difference in temperatures is small for the pumping conditions of chemical plants, and the effects are omitted for simplicity. Our investigation has shown (Ref. 3) that the larger surface area can result in a nearly 40°F reduction in bearing operating temperature. The cooler bearings, in this case, result in

Comparison between bearing submergence in oil, operating temperature, and bearing life for the old and new designs.

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99

cause wear and leakage after approximately only 2,000 operating hours.

OIL LEVEL SIGHT GLASS VERSUS CONSTANT LEVEL OILER A large sight glass allows direct visual observation to ensure proper oil level. It is standard in the new design, although a constant level oiler option is available. A constant level oiler is preferred by many users. When properly installed and maintained it can result in satisfactory operation. However, because operation of the oiler is “blind,” depending solely on strict conformance to correct (and nontrivial) oil filling and maintenance procedures, it may lead to an incorrect oil level inside the frame. This can lead to hot operation and premature failure. Another problem is known as the oiler “burping” effect, resulting in a higher actual oil level than perceived (Ref. 5). Obviously, the new pump design can be equipped with both the sight glass and oiler if they are desired by the user. Such improvements in design can be combined because they benefit pump reliability independently. Based on this research, when all are added together, an improvement in pump life of up to 125% may be obtained. Even longer life may be realized due to other design upgrades, such as providing the pump with a more rigid foot, reducing vibration, improving (finned) cooling, and creating larger chambers for mechanical seals. For brevity, these effects are not included in this analysis, but they can be accounted for in the references (Ref. 3, 5).

TEST PROGRAM To support and validate the theoretical derivations and assumptions as outlined below, a testing program was conducted, including lab testing and field data analysis. Figure 3 shows a comparison between operating temperatures and bearing life for the old

100

and new designs. Tests were conducted with oil covering different levels of the lower ball of the bearings. The proper design level (marked 50% on Figure 3) corresponds to oil at the middle of the lower ball of the bearing. At design setting, the new frame ran 40°F cooler, with a corresponding predicted life extension of approximately 6,000 hours.

ECONOMIC BENEFITS Having measured the life increases resulting from these improvements, it is not difficult to assess the economic benefits of the new design. Assuming the average value for MTBF of two (2) years for the old design, a 125% improvement results in a four and a half (4.5) year bearing life for the new design. The reciprocals of these numbers (1/2 years = 0.5; 1/4.5 years = 0.22) give an approximate number of failures or scheduled maintenance per year. The difference, 0.28, when multiplied by the average cost of repair of, say, $260 in parts and labor and 3,000 pumps per plant results in a yearly plant savings of: 0.28 x $260 x 3,000 = $218,400 In addition, savings resulting from increased uptime and a reduction of lost production at approximately $500 per off-line hour, assuming an average four hours per repair for off-line time, would be: 0.28 x ($500 x 4) x 3,000 = $1,680,000 The total, $1.9 million, is annual plant maintenance savings. Obviously, these numbers are approximate and can best be determined by individual maintenance departments using their operating specifics, but the savings potential due to improved design is clear. CONCLUSIONS AND RECOMMENDATIONS Our study demonstrated that substantial savings can be realized

The Pump Handbook Series

through improvements in pump design. To gauge such improvements systematically, it is imperative to quantify the benefits of each pump enhancement. It is also important to maintain a proper balance between the solid theoretical foundations used for the analysis and the laboratory work, field testing, and data supporting such theory. Users should seek quantitative data demonstrating improvements from pump manufacturers, including improvements in MTBF and MTBSM, enabling them to determine added value and other economic benefits. This approach will improve communication between manufacturers and users, and lay the ground work for the next step: further improvements in pump reliability. REFERENCES 1. H. Bloch. PRIME I and II, Pump Seminar Series, 1992/1993. 2. SKF General Catalog 4000 US (bearings), 1991. 3. L. Nelik. Value Added and Life Extension with Regard to Reliability of X-Series 3196 ANSI Pump. Goulds Pumps, Inc. Internal Report, 1993. 4. CRC Handbook of Lubrication, Vol. 1, CRC Press, R. Booser, 1983. 5. L. Nelik. Goulds Technology Video Seminar. Constant level oilers versus sight glass, Series 0693-01.

For more information on these references please call (315) 568-2811. ■ Lev Nelik is Manager of Pump Technology for Goulds Pumps. His responsibilities include developmental work in various aspects of centrifugal pump technology, developing new products, and improving the reliability of existing products. Dr. Nelik has authored publications on centrifugal pumps and hydraulic power recovery turbines, fluid mechanics, heat transfer, and FEA CAD/CAM applications.

CENTRIFUGAL PUMPS HANDBOOK

CPI Pumping Increase reliability and reduce emissions through pump selection. BY RICHARD BLONG AND BOB MANION oday chemical manufacturers and users are faced with global competition and pending environmental restrictions that threaten to reduce profitability. The need to reduce overall operating costs has driven pump users at chemical plants to focus on improving reliability and eliminating or reducing fugitive emissions.

T

ucts while reducing emissions to well below 500 ppm. With this in mind, consider the following: 1. Enclosed impellers are prone to plugging and premature wear in the above services due to small wear surface area. (Performance and efficiency cannot be renewed without replacing wear rings.)

SEALED PUMPS

2.

Open or semi-open impellers are reliable in these services and are standard for ANSI pumps. (Simple external impeller adjustments allow easy maintenance of performance and efficiency, and there are no wear rings to replace, yielding long-term energy savings.)

3.

The small internal passageways in sealless pumps are subject to plugging while handling liquids with only small amounts (5%) of solids. Viscosity handling is also limited.

4.

Design solutions separate the pump end from the drive end to allow sealless pumps to handle these services, but these modifications can be expensive and may not be cost effective.

The mechanically sealed chemical process pump, which meets ASME/ANSI B73.1M standards, is the workhorse of chemical processing industries. It will continue to be used on a wide range of process applications—such as liquids containing significant amounts of solids (sodium chlorate, alum, sodium carbonate, chemical wastewater), light slurries (silver nitrate and acetone slurries), viscous liquids (above 150 cP, including black liquor and titanium dioxide), and stringy materials where sealless pumps may not be economical to use. In addition to its ability to handle tough services, the flexibility of the design—along with improved low-emission mechanical seals—continues to make ANSI pumps the standard in this field. To elaborate on why sealless pumps are not economical to handle the above materials, we must note that they use enclosed impellers to reduce the axial thrust and increase reliability. (Although several manufacturers have tried using open or semi-open impellers in sealless designs, many of these have not been reliable at two-pole speeds.) Also, standard sealless pumps have small internal passageways to circulate liquid for bearing lubrication and drive-end cooling, and mechanical seal manufacturers are rapidly improving the reliability of their prod-

realities of applying pumps on a multitude of services and making them last. These efforts have produced the features listed below that many major ANSI pump manufacturers have incorporated (Figure 1). At a minimum, users should purchase ANSI pumps with features that best meet their application needs. However, most new designs incorporate features systematically to provide reliable products. Compromising designs to save money or add standard plant features—substituting a vendor’s standard labyrinth seal with the plant’s standard oil seal, for example—may not be advisable. New ANSI pump features include the following: 1.

Labyrinth oil seals are designed to prevent premature bearing failure from lubricant contamination or oil loss. These non-contacting seals have replaced Buna-rubber lip seals, whose useful life was three to six months under normal conditions. Materials of construction include carbon-filled Teflon, bronze, or stainless steel.

2.

Increased oil sump capacity provides better heat transfer for more effective oil cooling. Bearings operating at lower temperatures contribute to longer life.

3.

A rigid frame foot reduces the effect of pipe loads on shaft alignment. Misalignment won’t exceed 0.002 in. under load, and pump and driver alignment is better maintained.

4.

Bull’s eye sight glasses insure proper oil level, which is critical to bearing life. Level oilers have often been misused, leading to

Considering all the facts, it’s understandable that mechanically sealed ANSI pumps are the more economical choice to handle these types of liquids.

ANSI RELIABILITY IMPROVEMENTS To meet emissions regulations and improve reliability, process industries have pushed ANSI pump manufacturers to improve performance. Some manufacturers have formed alliances with users to share technology and improve standard designs. By working together, the theoretical has been combined with the The Pump Handbook Series

101

FIGURE 1

ANSI pump improvements. over- or under-filling sumps, both of which contribute to bearing failure. Sight glasses are also convenient for checking the oil condition visually to determine if a change is necessary. Constantlevel oiler manufacturers are just now introducing oilers that eliminate the potential for improper oil level settings while providing a sight glass, combining the best features of both methods. 5.

Mounting flanges accommodate an optional adapter that simplifies pump/motor shaft alignment, saving the user time and money during installation.

6.

Condition monitoring bosses on power ends provide consistent measurement points for temperature and vibration sensors. Many users report increased pump life from using predictive maintenance to identify and correct problems early. Taking measurements at the same point aids in proper interpretation of readings and allows personnel to move through the plant more quickly on inspections.

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7.

Engineered large seal chambers, specifically designed for today’s mechanical seals, increase seal life through improved lubrication, cooling, air venting, and solids handling. The chambers allow seal manufacturers to engineer and apply more reliable designs, including cartridge seals.

These developments extend pump and seal life and reduce emissions at the same time. Experience shows that one cannot be accomplished without the other. For example, a mechanical seal with emissions in excess of regulations has already failed in its application. Another benefit of these features is that several manufacturers and seal suppliers are extending unconditional warranties to as long as three years, helping to further lower operating costs.

SPECIALTY PUMPS FOR IMPROVED RELIABILITY Many diversified chemical producers are moving production of commodity chemicals to the Asia-Pacific and Latin American regions to take The Pump Handbook Series

advantage of lower labor and production costs. As a result, chemical production in the United States is being driven toward manufacturing specialty chemicals typically produced in small runs or batches. Examples include methylisobutyl- ketone (MIBK) and paratertiarybutyl- phenol (PTBP). Pump applications for these batch-type processes are usually low flow, in the range of 0 to 100 gpm. Traditionally, users install standard process pumps and throttle the discharge valves to obtain low-flow performance. However, these pumps are not designed to operate continuously in this range (Figure 2). Higher radial loads and increased shaft deflection lead to premature bearing and seal failure. Costly downtime and maintenance expenses result. For low-flow operation, users should specify a pump designed to meet specific service conditions (Figure 3). ANSI pumps designed for low-flow operation are available to increase pump and plant reliability. Improvements come from a casing and impeller designed for low-flow operation. Low-flow designs use concentric volutes and radial vane impellers to reduce radial loads, eliminating hydraulic and mechanical problems from throttled low flows (Figure 4). Some designs reduce radial loads as much as 85% compared to end-suction expanding volute pumps in this service (Figure 5). Shaft deflections from high radial loads are minimized, optimizing bearing, mechanical seal, and overall pump life. A disadvantage of low-flow ANSI pumps is that they sacrifice some efficiency to reliably handle viscous and solids-containing liquids. Another approach to low flow–high head applications is the regenerative turbine pump. This design directs liquid by a passageway so that it circulates in and out of the impeller many times on its way from pump inlet to outlet. Both centrifugal and shearing action work together to efficiently develop relatively high heads at low flows. Regenerative turbines also use concentric volutes and radial vaned impellers to obtain the reliability benefits discussed above. One drawback is that this type of pump utilizes close running clear-

two items that fail most often in pumps. These failures are often System Curve -ActualThrottled Operation directly related to improper application and installaRated Performance tion, poor operating practices or lack of maintenance, pipe strain, or misalignment. All of these again lead to high bearing loads, shaft deflection, and bearing and seal failure. FLOW Magnetic drive TYPICAL END SUCTION PUMP CURVE pumps have neither a mechanical Off-design (throttled) operation range (darker gray) seal that can fail and recommended operation range (gray). nor a driven shaft that can be subjected to pipe strain or misalignment. The driven ances to keep efficiency high and it is shaft is separated from the drive shaft therefore normally used on clean liqby a magnetic coupling, eliminating uid applications. the two major causes of pump failure. TOTAL DYNAMIC HEAD

FIGURE 2

With the implementation of the Clean Air Act, sealless pumps offer a dynamic solution to controlling emissions. Not only should sealless pumps be strongly considered to control emissions of the 149 volatile organic compounds identified by the Environmental Protection Agency, but they should be viewed as solutions to many difficult applications encountered in CPI plants today. For example, if users are experiencing sealing problems because of the pumped product’s poor lubricity (typical of acidic products in the range of 0–3 pH, such as sulfuric or hydrochloric acids), difficulty with product crystallization at seal faces (usually with caustic products in the range of 10–14 pH, such as sodium hydroxide and potassium hydroxide) or are frustrated with sophisticated auxiliary piping plans to provide clean, cool flush liquid to mechanical seal faces, sealless pumps may be the answer.

IMPROVED RELIABILITY WITH MAG DRIVES

CRITICAL MAG DRIVE FEATURES Reliable magnetic drive pumps must address two critical concerns:

proper lubrication of the journal bearings

2.

removal of heat generated by eddy currents in the recirculation circuit

The design must deliver liquid to lubricate the bearings—it should not be flashing or have risen in temperature, which decreases lubricity, prevents proper cooling, and leads to bearing failure. Proper journal bearing lubrication directs cooling liquid to the bearings, then to the magnets. Dual path designs provide lubrication to these areas separately. Both approaches prevent flashing at the bearings, a leading cause of failure. Another typical mag drive pump failure is liquid flashing at the impeller eye after being circulated through the drive end to remove eddy current heat. The result is a vapor-bound pump. New mag drive designs have virtually eliminated this problem by creating a constant pressurized circulation circuit that prevents flashing of cooling liquid and the associated failures (Figure 6). Not all new designs use pressurized circulation, and because most regulated liquids are volatile, this feature is necessary to achieve extended life in these services.

FIGURE 3 System Curve -ActualThrottled Operation

TOTAL DYNAMIC HEAD

SEALLESS PUMPS

1.

FLOW

Pump curve for a low-flow ANSI pump.

It is well recognized that mechanical seals and bearings are the The Pump Handbook Series

103

Regardless of the design features and modifications available from manufacturers, users are responsible for providing suppliers with as much data as possible on fluid and operating conditions. To apply sealless pumps properly, many factors must be considered: • Is the flow continuous or intermittent? •

Upon shut-down, what reaction (if any) will the process fluid have to residual heat? Chemicals like butadiene and formaldehyde may polymerize, leaving deposits inside the drive section and on the bearings.

•

Can the process shut down automatically, resulting in the pump operating at shut-off condition?

•

Conversely, can the system allow the pump to operate at the extreme right of the pump curve, which can adversely affect NPSHR and cause motor overload or excessive thrust?

•

What are the fluid characteristics, including vapor pressure curves, specific heat, viscosity over the process temperature range, and the effects of heating and cooling on the process fluid? Benzene freezes at 42°F (depending on the installation location, address the possibility of exposure to low temperatures), and toluene diisocyanate freezes at 72°F and begins to polymerize at 127°F (again, protect the installation or use jacketing if necessary). Maleic anhydride freezes at 130°F (use heating jackets or temperature control).

•

What about the customer’s practical knowledge of the corrosive nature of the chemical? Sometimes the standard corrosion charts don’t give the whole story.

ABRASIVES When pumping fluids containing particles, the traditional solution is to use very hard bearings (silicon carbide) operating against a hard or coated journal. The application of sealless pumps should go beyond this seem-

104

FIGURE 4

rience and test data. Solids may be formed by reactions to moisture (titanium tetrachloride), temperature (butadiene or formaldehyde), or a catalyst (any process that uses a catalyst that may vary in quantity or is subject to upsets). When the fluid is understood, it may be best to use modifications, including: backflushing to keep particles out of the drive section, heating or cooling jackets, heat exchangers in flush lines, filters or specially designed units that utilize isolation chambers, built-in seals, and precision back-flushing to reduce process stream dilution, if economical. Otherwise a mechanically sealed ANSI pump may be the best solution.

MAG DRIVE CONDITION MONITORING

An expanding volute pump (top) and a circular volute pump with a radial vane impeller (bottom). ingly easy approach. Consideration must be given to: • the abrasiveness of the solids •

the size of the particles

•

the quantity of particles

•

whether they can agglomerate

•

what creates the particles (reaction, catalyst, temperature)

The size of the particle that can be handled is usually determined by the impeller design and the clearances in the fluid passages. The effects of the quantity of particles are usually predicted from previous expeThe Pump Handbook Series

Magnetic drive pump reliability is also affected by operating practices. Condition monitoring devices can be applied to shut pumps down before a critical failure. Maintenance can then be performed, or operator errors corrected, before the pump is put back into service. Temperature detection and power monitoring together provide the best basic protection. Temperature detection indicates internal pump problems such as plugged recirculation paths, while power monitoring prevents dryrun failure. Other devices available include low amp relays, leak detection indicators, and package control systems.

INSTALLATION The effort involved in selecting the right pump for a given CPI application can be nullified by poor installation. As much effort, if not more, should be put into installation design to insure expected performance is achieved. (To understand how proper procedures improve equipment reliability see “Installation and Start-Up Troubleshooting,” Pumps and Systems, November 1993.) Important steps include: 1.

Lay out suction piping to provide NPSH available to the pump in excess of NPSH required. A common recommendation: NPSHA > NPSHR + 2–5 ft. See “Pump Suction Conditions,” Pumps and

ers’ reps and rely on their expertise, but be informed, as well, and together you can apply pumps properly in your facilities. ■

FIGURE 5

Increasing Radial Load

Expanding Volute

85% Reduction

Circular Volute

0

50

100

150

200

Low Flow Operating Range–GPM

Radial load curves. Systems, May 1993 and “How Much NPSH Is Enough?” September 1993. 2.

Provide a straight run twice the length of the pipe diameter (2D) to the pump suction flange to prevent added turbulence at the impeller eye, which could lead to premature (incipient) cavitation.

3.

Install conventional or cartridge mechanical seals according to manufacturer recommendations.

4.

Meet seal flush requirements by providing an external flush at the necessary pressure and temperature, or add auxiliary piping for flushing on the pump.

5.

Prepare the foundation before grouting the baseplate.

6.

Select grout that will meet installation requirements.

7.

Select a baseplate to maximize pump, seal, and motor reliability. Many vendors offer baseplates with enhancements such as .002 in./ft flatness, leveling screws, motor alignment screws, continuous drip rims, and other features designed to ease installation and alignment and increase pump life.

8.

Align equipment according to manufacturer specifications.

9.

Select and install condition monitoring devices for sealless pumps.

Rich Blong is product manager for chemical pump development for Goulds Pumps Inc. Previously, he was a senior applications engineer responsible for applying chemical pumps for many different processes. He also worked as a pump systems engineer with Union Carbide’s Linde Division, now PRAXAIR. He has a bachelor’s degree in chemical engineering from the University of Buffalo. Bob Manion is product manager for magnetic drive and non-metallic pumps for Goulds Pumps Inc. He has held marketing and sales management positions related to developing, selling, applying, and servicing centrifugal pumps for 13 years. Mr. Manion holds a bachelor’s degree in marketing from the Rochester Institute of Technology.

CONCLUSION Selecting a pump to improve reliability will reduce emissions and operating costs at the same time. Neither a mechanically sealed ANSI pump nor a sealless pump can be universally applied on every process application. FIGURE 6 Make an informed decision based on specific service conditions and total cost (initial + maintenance + operating costs). To insure a return on investment, as much time and effort must be expended on the design of equipment installation as on pump selection. Although selecting equipment for increased reliability and reduced emissions may seem expensive in the short term, it saves money in the long run. Work with manufactur- Recirculation circuit.

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Canned Motor Pumps When the canned-motor pump is the choice to solve a specific pumping problem and control costs, the following points must be considered to achieve satisfactory results:

CHEMICAL CHARACTERISTICS

• • • • • •

There is seemingly never enough information available on the chemicals to be handled. The supplier must depend on the customer to provide this information, but it is also very important that the supplier and the customer exhaust their resources in an attempt to anticipate what a chemical will do inside the sealless pump. Will it cause corrosion, boil, decompose, freeze, or polymerize? Any of these properties can result in rapid failure unless anticipated.

MAINTENANCE

APPLICATION AND METHOD OF OPERATION Will the pump be used for transfer, condensate return, reboiler, or batch operation? Will it be running continuously or intermittently? Will the location be remote, exposed to the elements, or in a hazardous location? How will the pump be operated and what will the process demand? Can the flow range over the complete curve? Is it close to shut-off, which may require a by-pass orifice? Or, conversely, will it occasionally pump at the extreme right of the curve? This can result in cavitation and subsequent failure if allowed to continue. All of these factors, combined with the knowledge of the fluid pumped, will determine the proper selection and modifications necessary for successful pump operation.

DIAGNOSTICS AND CONTROL Once the above factors have been determined, the user and supplier should agree on the type of diagnostic devices and process control that will assure a successful installation. Diagnostics available include:

106

bearing wear monitors rotation indicators motor diagnostic devices bearing temperature sensors leak sensors flow sensors Any of the above may be recommended. The pump and motor can also be fitted with a control device such as: • a water or steam jacket • a water-cooled heat exchanger • a heat exchanger in the circulation line • complete jacketing of the pump and motor (Consider if the insulation will create motor heat problems.) The final consideration must be maintenance. Does the user have a planned maintenance program? Does the user’s and supplier’s experience indicate more frequent maintenance intervals than normal with the chemical product in this particular mode of operation? Proper maintenance and replacement of less expensive bearings and gaskets can prevent a major failure and yield increased savings.

CONCLUSION Using a sealless pump can be easily justified due to the elimination of leakage and emissions because the value of the chemical lost using a sealed pump can be calculated. But there are many other factors that are more difficult to quantify, including housekeeping costs, safety, odor, and public and employee relations. The major elements leading to long-term savings using sealless pumps is the upfront analysis of the application and the supplier’s knowledge of his product. ■ Joe Cleary is the Vice President (retired) of Sales for Crane Co., Chempump Division.

The Pump Handbook Series

CENTRIFUGAL PUMPS HANDBOOK

Pump Buying Strategies BY: J.T. MCGUIRE

A

To get the right pump, all you have to do is decide

•

what you want, state those requirements clearly and

• •

place your order with a

•

capable manufacturer. •

Sounds easy, doesn’t it?

•

It may sound simple at first, but it’s not. For example, some pump purchasers may not know the volume of liquid their system handles. Reliability data is hard to come by, too. With so many factors affecting pump life, mean time between overhauls and replacement may not be known. And the goals are conflicting. Increased service life may also increase energy consumption and purchase price. How do you sort through these factors? How do you determine what you need from a pump, develop a meaningful specification for those requirements and finally buy the right pump? The answer is to take it one step at a time and follow a disciplined approach to pump specification and purchasing.

FIGURE 1 P2 SYSTEM

P1

4

1 3 2

HYDRAULIC GRADIENT 3 4

PUMPING ENERGY 1

12

VAPOR PRESSURE

NPSH

ENERGY LEVELS 1

- EXIT FROM SUCTION SOURCE

2

- PUMP SUCTION

3

- PUMP DISCHARGE

4

- DISCHARGE POINT

t first glance, pump requirements don’t really seem that complicated. After all, a pump only needs to: move a specified volume of liquid through a given system be energy efficient comply with any applicable laws regarding leakage achieve certain mean time between overhauls and replacement be delivered on time with complete documentation and all at minimal cost

DETERMINING PUMP REQUIREMENTS To write requirements for a pump, you should review the basics of pumps. Pumps are designed to move liquid against a hydraulic gradient; in other words, to move liquid from the suction reservoir to the discharge reservoir, which differ in elevation and/or pressure (Fig. 1). You can see immediately from the figure that the pump must supply adequate energy to overcome the difference in elevation and pressure along with the friction losses in the conduits on both sides of the pump. It’s obvious that the pump, as the sole source of energy in the system, must supply all the needed energy. The Pump Handbook Series

Thus it’s no wonder that severe consequences await those who overlook that simple fact! Another, perhaps not so obvious, fact is that the energy available at the suction side must provide a certain net margin over the liquid’s vapor pressure at the pump suction. This net margin, called Net Positive Suction Head Available (NPSHA), is necessary to prevent cavitation — the boiling of liquid in the system. Cavitation impairs pump performance and shortens the service life of the pump. An excessive amount of boiled-off vapor impairs the machine’s hydraulic performance. In addition, the subsequent collapse of the vapor bubbles as they move to regions of higher pressure can cause cavitation erosion. To prevent these problems, you must specify total system head accurately. In most applications, you can determine the normal pump flow and the static components of the total head associated with ideal operation of the plant or process at its design output. Add estimated piping friction losses and control valve pressure drop (if applicable) to find the total system head for that capacity. (Remember friction head varies as the square of the flow ratio.) Normal pump flow and static components aren’t the whole story of operating conditions. You must also factor in the range of operating conditions your pump will be called on to perform under. Changes in operating conditions can be caused by: • process unit downturn • flow swing to cover upset or transient • change in static head as vessel levels or pressures in both change with time • change in friction head as system fouls or scales or as discharge vessel fills • pump wear You use these data to compute the rated flow — the flow under which your pump will need to operate. You then match the performance

107

Formats for the specifications data (power, NPSHA, speed) quoted range from a very simple functional by the manufacturer against the rated spec to a very elaborate functional flow. design and manufacture. The simple You can set the rated flow to the functional spec states only what the maximum rate at which your pump pump will be called on to do. You will be called on to operate. But if the give complete freedom to the manuflow range is very wide (and you plan facturer for designing the pump. An to use a centrifugal pump), you might advantage to such a specification is set the rated flow to the most frethat you can get very interesting quent or efficient flow rate. In either designs for unique pumping probcase, once you set the rated flow and lems. But, such specs required operating flow can be difficult to write range, you will need to and you’ll need to evalulook at the NPSHA for Stated NPSHA ate the engineering the pump at these flows. should reflect a behind the bids carefulIn addition to the ly. rated flow, you must value normally To avoid the backconsider the range of end expenses of a simple flow. Pumps cannot available, not functional specification operate across the entire (and since most pumprange of flow from maxsome possible ing requirements are relimum flow to zero flow. minimum value atively straightforward), With the exception of most pump buyers write direct acting steam with a hefty detailed functional pumps, no pump has an requirements and manuinfinite range. Centrihidden margin. facture specifications. fugal pumps, the most As a start, these common pump in use specifications must address: today, can operate under a wide • operating environment range of flows if they are designed • liquid to be pumped appropriately. Thus, set the flow mar• pump performance and life gin to allow for process transients and • materials of construction pump wear, but don’t set it larger • extent of supply than necessary. And be sure the statMany other items related to ed NPSHA reflects a value normally function, design and manufacture can available, not some possible minibe addressed (Table 1). The number mum value with a hefty hidden marof requirements you choose to gin. Otherwise, you’ll find yourself include will depend on the pumping with an unsuitable pump — an overapplication and your confidence in sized pump or one designed for the potential pump manufacturers. abnormally low NPSHA. Items you may want to pay special attention to include: DEVELOPING A MEANINGFUL SPECIFI• Degree of redundancy (Item 4). CATION This item refers to the proportion Once you determine the specifiof spare capacity the pumping cations for your pump, you need to arrangement has to provide in communicate those specs to the manthe event one pump is lost. ufacturer in concise terms. Typical values are 100, 50, 25 or For the manufacturer to under0 percent. Most purchasers have stand what you really want, your design standards relating degree specification must state, in an orderly of redundancy to the type of manner, all the requirements you service involved. have for the pump. But that doesn’t • Type of pump (Item 7). This mean you should strive for a thick issue is complex, determined by specification document. The value of the hydraulic duty, the degree of the specification is not proportional to flow regulation required, and the volume or weight. If anything, the nature of the pumped liquid. inverse is true. Overly long specifica• Number of pumps (Item 7). This tion documents often fail to state item incorporates the required what the purchaser really wants. degree of redundancy (Item 4)

108

The Pump Handbook Series

along with the total flow to be handled, the total head to be developed, or the total power absorbed. It can also incorporate the physical size of the pump (often related to type of pump) and the required turndown in flow when rated flow is high. • Service lives of the pump and various components (Items 12 through 15). These requirements are usually expressed as mean time between failure. As noted earlier, data on the life of various pump components are meager. Thus, these requirements are often not specified. Generally, antifriction bearings and the first stage impeller of high energy centrifugal pumps are the only components for which minimum service lives are commonly specified. • Materials of construction (Item 16). You’ll have to handle this item since the pump manufacturer does not control the pumped liquid. If you have little or no experience pumping the particular liquid, manufacturers will suggest possible materials. But the only guarantee a pump manufacturer makes for materials is that they will conform to their producing specification. • Extent of supply is an essential issue. (Items 7, 8, 22, 23, 25 and 26). When faced with increasing complexity and extent of specifications, many pump purchasers find it beneficial to summarize extent of supply (also known as terminal points or battery limits) in a list or diagram. That’s a good idea. It helps you state more clearly what you need. To simplify the specification, you should note all technical elements on a basic data sheet. And remember that the basic data sheet should be just that — a sheet. Multi-page data sheets are unwieldy. If your system is complicated, cite and add supplementary sheets rather than cluttering the basic data sheet. Instead of building custom specs, some purchasers in particular industries use general specifications issued by that industry (for example, ANSI B73.1M-1991, which addresses hori-

zontal end suction pumps for chemical process and API-610, 7th edition, which addresses centrifugal pumps for petroleum refining). Some buyers use these general specifications verbatim, others use them as a base and add supplement covering changes they wish to incorporate. A cardinal rule for any meaningful specification, whether homegrown or based on an industry standard, is to avoid multiple tiered references to other specifications. With more than one tier of references, such specifications become too complicated to be meaningful. For example, when addressing government regulations, be sure to identify and specify the exact rules and regulations the equipment has to meet.

The old catchall, “comply with applicable local, state, and federal rules and regulations,” doesn’t add anything to the specification. Beyond technical requirements, the specification also must address the proposed terms of purchase or commercial terms and conditions. Although these items are generally the province of the purchasing department, you, as a specifying engineer, should be aware of what is involved. The major items covered in the terms and conditions are: • delivery period or date • point of delivery • liquidated damages • terms of payment • warranty • default

TABLE 1. ELEMENTS OF A PUMP SPECIFICATION TECHNICAL Item 1. 2. 3. 4. 5. 6. 7. 8. 9. 10. 11. 12. 13. 14. 15. 16. 17. 18. 19. 20. 21. 22. 23. 24. 25. 26. 27. 28. 29. 30. 31. 32.

Function Location and environment Liquid pumped and properties Hydraulic duty Redundancy in pump arrangement Future performance margin Application margins Type and number of pumps Driver and arrangement Minimum tolerable piping loads Allowable seal leakage Allowable noise Minimum pump life Mean seal life Bearing life and basis Mean period between overhauls Materials of construction Rotor design requirements Hydraulic design requirements Allowable stress Type of shaft seal Type of bearings and lubrication Type of coupling Type of base Piping: systems required and construction Auxiliary systems: specification Instrumentation Material tests Welding procedures approval Inspection during manufacture Component and equipment tests Painting and inhibiting Documentation

Design

Manufacture

X X X X X X X X X X X X X X X X X X X X X X X X X X X X X X X X

The Pump Handbook Series

• cancellation • bankruptcy Delivery period and terms of payment should be of special interest to you. For example, if the delivery period is too short, there arises the risk that somewhere in the manufacturing of your pump, a shortcut or two will be taken resulting in a pump that will not function adequately in the field. Also of interest to an engineer is method of payment. By tying payment to achieved manufacturing milestones, you can expedite the manufacturing of your pump, and thereby help to ensure on-time delivery.

BUYING THE PUMP Once you’ve clearly specified the pump, it’s time to place the order. This process can go smoothly, if you: • double check that the pump you’re ordering is really the pump you want • order the pump in time to allow for orderly manufacture • have a post-award meeting, within one month of ordering, to ensure the order is clear and started • don’t change the order unless safety or a major performance problem is involved The final three steps are self-explanatory, but the first two deserve some explanation. Double checking your order is especially important for complex units with an extensive specification. As a check, hold a pre-award meeting with the manufacturer to clarify the bid. If your unit includes major auxiliary equipment or systems, review and settle the basic unit plot plan at this meeting. Selecting a manufacturer can be done in one of two ways: specify-andevaluate or partnership-purchasing. Under the specify-and-evaluate method, you prepare a very detailed specification and issue inquiries with extensive data requirements, then thoroughly evaluate the data in the resulting bids and purchase based on the numerical results of the evaluation. The evaluation generally takes

109

the form of a weighted matrix which includes: • energy consumption • maintenance cost • risk of lost production • purchase price • delivery Manufacturers are free to bid whichever pump they feel meets your specifications. That leaves you, the purchaser, to make the final determination of whether a pump meets your requirements. Thus, you’ll need to build a rigorous inspection regime into your selection process. Over time, you can build a list of acceptable bidders to help narrow the field. Partnership-purchasing avoids the cost of preparing an elaborate specification, issuing inquiries, and evaluating bids. Under this method, you select one manufacturer to work with and provide just a minimal specification. The manufacturer then chooses the best pump for your needs.

110

To select the best manufacturer to work with you, you need to assess the caliber of the various manufacturers that make the class of pump you’ve chosen. Your assessment should cover each manufacturer’s: • order engineering and manufacturing processes • emphasis on quality as an inherent facet of all processes • product design philosophy • detail designs for and experience with the class of pump required After choosing the best manufacturer to work with, you can negotiate prices for equipment according to some fixed relationship to published price lists. While you will incur some costs in assessing manufacturers, this process is likely to be less expensive for a major product or a period of two or three years between assessments than open bidding would be. Which approach is better? For innovation in design and reliability, I’ve found that partnership-purchasing yields distinctly better results than specify-and-evaluate has.

The Pump Handbook Series

Engineers are not surprised by this; they know that technical endeavors proceed best in a cooperative arrangement. The specify-and-evaluate method might help you find a company that will furnish equipment nominally capable of the same function for less money (even when factoring in the cost of writing the specs and evaluating the bids). But the issue isn’t just cost. The real value of partnershippurchasing is innovation in design and reliability of products. Partnership-purchasing is actually the way the pump industry used to operate before competitive bidding became so popular. As an industry, we, the suppliers and purchasers, would do well to resurrect it. ■ J.T. McGuire is Director of Marketing for the Huntington Park Operations Division of the IngersollDresser Pump Company.

CENTRIFUGAL PUMPS HANDBOOK

A Common Sense Approach to Combating Corrosion and Abrasion BY JOHN RINARD

Photo 1. New and corroded centrifugal pump impellers.

A

single look at Photo 1, above, should be enough to convince anyone of the destructive nature of corrosion and abrasion on pumps, and lead to the question of how to prevent this from happening. This article should serve as either a primer or a reminder of factors involved in properly selecting or troubleshooting a pump in corrosive and/or abrasive service. Historically, pump selection has consisted of finding a pump that will, “pump stuff from here to there,” or that will “deliver so many gpm at such-and-such a head.” A greater degree of sophistication leads to “and that will hold up in acid,” or “and that will pump solids.” Obviously, the more that is known about the solution being pumped, the more appropriate the pump selection will be. An interrelationship exists where the chemical and physical properties of the pumpage determines the materials of construction, which dictates pump design, which affects pump performance, which in turn determines the proper pump selection. The more diffi-

cult selections involve services pumping corrosives and/or abrasives, and these are the major factors governing which pump is chosen. You don’t need to be a rocket scientist to select a pump, and you don’t need a PhD in metallurgy to make some basic materials selections and understand the reasoning behind them. We all know that water will “rust” iron, acids “corrode” certain materials that come into contact with them, and solids “wear” when rubbed together; conversely, we know that “stainless” steel is corrosion resistant and that either a hard or soft “rubber” material will resist abrasion or wear. These simple facts lead us to a closer examination of the mechanisms of corrosive and abrasive attack. Corrosion is the wearing away or deterioration of a material by chemical or electrolytic action or attack. Abrasion is the wearing away of a material caused by a solid rubbing or impinging on another. Abrasion

The Pump Handbook Series

caused by the velocity of a liquid or gas is commonly called erosion. Corrosion-abrasion is a combination of both corrosion and abrasion that results in an accelerated attack on material. It is generally more severe than either corrosion or abrasion alone, due to the severe wear caused by the continuous abrasive destruction of the passive protective film built up by corrosion. Table 1 shows the basic types of corrosion. Corrosion and abrasion take many forms, and numerous combinations of these forms exist. Detailed analysis of these combinations can be quite complex and goes beyond the scope of this discussion.

MATERIALS There is no material that will withstand attack from all combinations of liquids and solids found in pumped solutions. However, a basic knowledge of material categories will give us a general idea of what materials will and will not work in certain environments, and then we can zero in on the right pump for a given job.

111

It becomes obvious with examination of Table 2 that the mechanical properties of a material determine the design of a pump. Pumps constructed of hard materials are more difficult to design (flanges, stack tolerances, and clearances), cast (sharp angles and complex shapes), and machine (drill, tap, and finish surfaces); non-metallics may need to be reinforced, supported, or protected with metal armor; and thin or highly stressed components must be made of strong materials. Figure 1 shows a typical configuration of both a chemical (corrosion resistant) pump and a slurry (abrasion resistant) pump. One can readily see that the slurry pump’s hard metal materials of construction dictate the use of through-bolt construction rather than drilled and tapped holes. Less apparent are the facts that slurry pumps are generally more massive than chemical pumps; are designed with open clearances, blunt edges, and looser tolerances due to “as cast” hard metal surfaces and the need to handle solids; and are commonly designed with metallic or nonmetallic liners. As a result of these design constraints, slurry pump efficiencies suffer, and in most cases are lower than chemical process pump efficiencies. Identification of materials that can handle the liquid to be pumped does not necessarily complete the material selection process; quite often this step leads to other considerations. Options and compromises almost always present themselves with either chemical or slurry pumps when comparing service life and wear with cost and availability.

FIGURE 1

CHEMICAL PUMPS

Pump design comparison. On the top is a hard iron slurry pump with side suction. The bottom is a stainless ANSI B73.1 chemical pump with an expeller-type seal.

112

The Pump Handbook Series

Wear. The chemical process industry generally considers that any corrosion rate equal to or less than 20 mils per year is acceptable wear. This, however, may be considered excessive depending on either pump design (pump impellers with relatively thin vanes and shrouds effectively see double this wear rate because they are totally immersed in the liquid and therefore exposed to attack from both sides) or a need for extended service life for pumps in critical services and inaccessible or remote locations.

Cost. Some material costs TABLE 1. TYPES OF CORROSION may be prohibitively high Type Characteristics and therefore lead to General Uniform attack over entire exposed surface selection of less corrosion Erosion-Corrosion Corrosion accelerated by erosive action resistant alternatives or a of fluid or slurry vortex lined rather than a solid Crevice Localized attack at crevices or material pump. stagnant areas Availability. While mateGalvanic Occurs when two dissimilar metals are rials such as 316 stainless immersed in a corrosive or conductive steel, CD4 MCu, and solution Alloy 20 are commonly Intergranular Grain boundary attack stocked and available for chemical pumps, alloys such as monel and Cavitiation Pitting on high pressure areas such as Hastelloy are more likely impeller vane tips and/or low pressure to be special orders. “Standard” materials of areas such as eye of impeller vanes or construction vary from trailing edge of impeller vanes manufacturer to manuPitting Localized accelerating attack by chlorides; facturer. Depending on associated with stagnant conditions the pump type and the Selective Leaching Dissolves one component of an alloy manufacturer, material availability can vary anywhere from being in stock to needing up to several months sive and abrasive, an acid sludge, for lead time. example, presents the greatest chalLined or coated pumps and nonlenge in pump selection. Many materimetallics offer possible solutions to the als are essentially suitable to either high cost and long lead times of noncorrosion or abrasion, but not to both; stock special metallic materials. titanium, for example, is a very strong, Lining pumps is where nonmetallics corrosion-resistant material, but it is really shine, and they are less expenunsuitable for slurries because of its sive and more readily available than softness, and white iron is a very hard, special metallics. However, just as abrasion-resistant material that is not there is no single metallic that is good practical for corrosive conditions. Nonfor handling every solution pumped, metallic elastomers, on the other there is also no single nonmetallic for hand, may be used in a service that is all services. Each must be carefully both corrosive and abrasive. When selected to fit the service. selecting elastomers, consideration must be given to solids size and configSLURRY PUMPS uration, temperature, and a pump The selection of abrasion-resisdesign that must generally preclude tant materials for slurry pumps, liquid contact with any metallic armor much the same as corrosion-resistant or reinforcing. chemical pump material selection, PUMP PERFORMANCE also involves consideration of service life (wear), cost, and availability. The efficiency of a pump as well Abrasion, unlike corrosion, is generalas the location of the operating point ly combated by the use of either very on the pump performance curve is hard materials or soft, resilient elasoften overlooked or ignored during tomeric materials. Hard materials are pump selection. The location of the generally used for slurries with large operating point is overlooked more or sharp solids. Soft, resilient elasoften than the pump’s efficiency. tomeric materials are used for small This alone will contribute as much as or blunt solids. Once again, here we any other factor to pump failure find that non-metallic elastomers lend when abrasion or corrosion-abrasion themselves for use as pump liners. are present. Efficiency is a measure-

CHEMICAL-SLURRY PUMPS It was mentioned earlier that a pumped solution that is both corro-

ment of smooth flow—and therefore reduced turbulence and recirculation—within the pump. Turbulence The Pump Handbook Series

Remarks Most common type of corrosion

Commonly found at gasketed or flanged surfaces

Weld decay is a type of intergranular corrosion occurring in areas adjacent to a weld

Common in 304 and 316 SS, and A20 Zinc removed from brass or bronze is called dezincification; when grey cast iron is attacked, graphite is left undisturbed

and recirculation result in increased liquid and solids contact with wetted pump surfaces, as well as unpredictable angles of impingement. Proper pump selection, therefore, dictates selection at or near the best efficiency point of the pump. Selection just to the left of best efficiency is considered good practice, as illustrated by Figure 2. Analysis can often trace the cause of pump problems to operating the pump at or too close to shut off (to the far left of best efficiency) because the pump is oversized. Intentional oversizing may occur through the use of system design safety factors, selection for future increases in performance, and using an existing pump without consideration of size. Unintentional oversizing may occur because of miscalculations or changes over time in the process or the piping system. The end result is the same; the pump is operating too far away from the best efficiency point. When analyzing pump performance, we must think in terms of a pumping system rather than just the pump. A system consists of the pump and all the related piping, valves, and process equipment on both the suction and discharge sides of the pump. All of these items directly affect the pump performance in that the “system curve” (which can be analytically derived from the pressure drop/resis-

113

FIGURE 2 OPTIMUM SELECTION BEST EFFICIENCY POINTS

PUMP SELECTION

ACKNOWLEDGMENT

The final selection decision is made by the pump user. This decision may be more subjective than analytical, but should include such factors as:

My thanks to Dr. George Calboreanu, Chief Metallurgist, Western Foundry, a division of A.R. Wilfley and Sons, for his materials expertise and assistance in preparing this paper. ■

Head

•

Flow

Optimum pump selection results in a pump operating just to the left of its best efficiency point. tance to flow across various in-line hardware) dictates where the pump will operate on its curve, the pump point of rating. Less sophisticated considerations include rules of thumb such as: •

•

Keep the suction piping as short and straight as possible. Slope the suction piping toward the pump suction when handling slurries.

Centrifugal pumps tend to become unstable the closer they approach either shut off (zero flow) or maximum flow. This instability may be manifested in cavitation, recirculation, and turbulence. Recirculation and turbulence can result in a liquid temperature rise in the pump that can cause accelerated corrosion as well as erosion-corrosion.

114

Availability (of both pump and parts) • Maintainability • Reliability • Service life • Standardization • Cost There are always trade offs. The user ultimately makes a selection based on the priorities that best meet the process needs. This paper has presented an overview of corrosion and abrasion factors that should be a part of that selection process.

John W. Rinard holds a bachelor’s degree in industrial engineering from Texas A&M University. His experience includes positions in Sales Engineering and Management with the Buffalo Forge Company and the Duriron Company. He is presently with A.R. Wilfley and Sons.

TABLE 2. MATERIALS COMPARISON Category

Subcategory

Material

Metallics

Ferrous

Steel Ductile Cast Iron 27% Chrome 304 SS 316 SS CD4MCu A20 Hastelloy B/C Brass Bronze Aluminum Titanium (pure) Zirconium (pure) Rubber (gum) Neoprene Urethane Teflon (PTFE) Epoxy (cast) Polypropylene Silicon Carbide Aluminum Oxide

Stainless

Copper base Miscellaneous

Non-Metallics

Elastomers

Plastics

Ceramics

Typical Mechanical Properties Tensile Elongation Strength (Min % in 2") (Min, psi) 150 Brinell 70,000 22 160 Brinell 60,000 18 600 Brinell 80,000 Nil 150 Brinell 70,000 35 150 Brinell 70,000 30 225 Brinell 100,000 16 125 Brinell 60,000 35 225 Brinell 75,000 20-25 60 Brinell 37,000 30 65 Brinell 35,000 18 130 Brinell 65,000 8 200 Brinell 80,000 18 210 Brinell 55,000 12 35 Durometer A 3,500 500 55 Durometer A 3,000 650-850 75-95 Durometer A 4,500-7,500 250-900 50-65 Shore D 3,000-4,000 200-400 M75-110 Rockwell 2,000-12,000 Nil R85-95 Rockwell 5,000 500-700 2,500 Knoop 44,500 Nil 1,000-1,500 Knoop 22,000-45,000 Nil Typical Hardness*

There are numerous alloys, formulations, and compounds of metallics and non-metallics; those shown are typical and are not to be considered all-inclusive. *Conversion relationships of hardness scales/numbers are discussed in ASTM E140 and Metalcaster's Reference and Guide, 2nd Edition, 1989, The American Foundrymen's Society, Inc., (for metals), and ASTM D2000 (for rubber). The Pump Handbook Series

CENTRIFUGAL PUMPS HANDBOOK

Recommendations For Vertical Pump Intakes BY: HERMAN GREUTINK LOCATION A vertical turbine, mixed flow or axial flow pump’s location in a sump is critical to good performance. Figures 1 and 2 provide good design criteria for sump layout. These criteria are based on a maximum bell entrance velocity of 6 ft/s. However, because bell diameters vary from manufacturer to manufacturer, these ratios must be adjusted to accommodate the differences. According to the U.S. Army Corps of Engineer’s design guide, ”For satisfactory pump performance based on research and prototype experience, recommended submergence, S, should be 1.25 D or greater, and the dimensionless flow ratio through the individual pump should not exceed a value of 0.40 for: Q/ √gD5

FIGURE 1

1D

2D

2D

.8 D

1D

Velocity preferred to be 1 ft/sec

where D

Q= discharge, cfs

.5 D

D= pump bell diameter, ft g= acceleration due to gravity, 32.2 ft/s2 Submergences that are less than, and flow rates that exceed the above limits were investigated, and more complex designs were required for satisfactory hydraulic performance.” The recommendation of 1.25 D minimum submergence is suitable for storm water and flood control pumps (provided a vortex supressor beam is used as illustrated by Figure 2); however, for continuous service pumps a submergence of 1.75 D is recommended. If the submergence is less than these values, the bell diameter must be enlarged. For instance, to meet a 1.25 D submergence value, the bell diameter should be enlarged to produce

an average entrance velocity of 3.3 ft/sec. This velocity may be a bit conservative, but the cost of enlarging the diameter is low and the benefits are tangible.

VORTICES If a vortex still occurs after you have followed the above guidelines, it is not generally difficult to alleviate. It takes very little energy to form a vortex; therefore, it takes very little energy to get rid of it! Submerged vortices, however, can be troublesome. These vortices will touch the floor and/or wall of an intake. They are the result of swirling masses of water next to or under the pump and are not continuous. Although submerged vortices sound like cavitation due to the lack of net The Pump Handbook Series

positive suction head (NPSH), the noise created by a vortex comes and goes as the vortex comes and goes. To mitigate submerged vortex formation, apply the following strategies: • Place a cone under the bell. •

Employ splitters.

•

Fill-in intake corners.

•

Use diffuser screens.

HIGH VELOCITY As a rule, high velocity to a pump in the intake and/or at the bell leads to reduced life of the pump. For a given head and capacity, today’s pumps operate at approximately double the speed of the pumps in use before the

115

FIGURE 2 A 1D (TYP)

1D

0.25D

Rounded

135°

➤

R=2D 135°

2D

6D

Divider Walls

Pump Bay

Vortex Suppressor Beam

2D

➤

Curved (wing) Wall

1960’s. The net result of these higher speeds is a drastically increased frequency of pump repairs. Slowing down continuous service pump speeds may be more expensive initially but the long-run savings on maintenance will more than compensate for the increased pump costs. A high velocity stream aimed at or near the pump could also be a source of premature failure. A fluid force of this nature should be diffused by piling, screens or walls in front of the conduit outlet. Figure 3 provides a simplified depiction of distances required to diffuse a high velocity flow out of a conduit. The breakup of jet streams can be achieved by baffles as shown in Figure 4. This configuration also promotes better distribution to multiple pumps.

DIVIDING WALLS

A W

45°(wing) Wall

PLAN S = Submergence D = Pump Bell Diameter

1.5D

0.5D

Minimum Water Level

Because short dividing walls are not recommended, they are not pictured in any of the figures. (Figure 1 shows no dividing wall while Figures 2 and 4 show long dividing walls.) With multiple pump stations, the front of the short walls can propagate vortices when one or more pumps are out of service. So it is better to have no walls than short walls. Long walls provide easy support for the pumps, as well as drainage for individual pump sumps when stop logs are used.

0.25D

116

0.5D 1.0D

S

1.25D

1.0D

INTAKE TESTS

Section A-A

The Pump Handbook Series

When guidelines such as those published by the Hydraulic Institute and the British Hydro-mechanics Research Association (BHRA) cannot be followed, model intake tests should be performed, especially for pumps larger than 50,000 gpm. ■

FIGURE 3

REFERENCES Area I - Potential Core Area II - Transition Area III - Similar Velocity Profiles Area IV - Jet Center Line Wanders

•

4-5d I d

U.S. Army Corps of Engineers, Engineering Technical Letter No. 1110-2-313 ”Hydraulic Design Guidance for Rectangular Sumps of Small Pumping Stations with Vertical Pumps and Ponded Approaches.”

2.

Prosser, M. J. ”The Hydraulic Design of Pump Sumps and Intakes.” British Hydromechanics Research Association.

3.

Hydraulic Institute Standards, 14th Edition

IV

30d 10d

1.

III

II Vm

V

Herman Greutink is vice president and technical director for Johnston Pump Company in Brookshire, TX.

x Up to about 30 times diameter d, the formula Vm d ___ = 6.5 x ___ is used to determine Vm x V V and Vm (ft/sec.), d and x (ft.)

FIGURE 4

Suction Bells

Trashrack Flow Baffles

Pump Bay

Forebay

Suction Bells

D Plan D = Suction bell diameter, ft. U.S. Army Corps of Engineer’s Design Instructions for Flood Control Pumps The Pump Handbook Series

117

CENTRIFUGAL PUMPS HANDBOOK

Hydraulic Instabilities and Cavitation Causes, Effects and Solutions

ydraulic excitation forces and pressure pulsations created by excessive flow deceleration at partial load have a profound impact on the possible failure of a variety of pump components. These forces and pulsations are the result of flow recirculation in the impeller inlet, diffuser or volute sections of the pump. Some degree of recirculation is present in every centrifugal pump below a specific flow rate representing the ”onset of recirculation.” In fact, recirculation is of minor concern for the majority of pump designs. On the other hand, excessive recirculation can be extremely harmful and destructive. Consequently, the onset of damaging recirculation is of greater concern to pump operators than the onset of recirculation itself.

H

IMPELLER INLET RECIRCULATION Three physical mechanisms trigger flow recirculation during partial load at the impeller inlet: 1. deceleration of the velocity upstream to the impeller relative to the velocity in the impeller throat 2.

pressure gradients perpendicu-

lar to the direction of main through-flow 3.

excessive incidence (i.e., difference between blade angle and flow angle at the impeller vane leading edges).

The primary geometrical parameters impacting the above phenomena are: • impeller throat area •

angle of approaching flow

•

impeller vane angles

•

ratio of impeller eye diameter to hub diameter

•

ratio of vane tip diameter to hub diameter

•

impeller shroud curvature

•

impeller leading edge position (in planar view and meridional section).

However, no simple general relationship exists between the onset or amount of recirculation and the geometry of the impeller. Relationships have been derived that are valid only for particular families of impellers (Ref. 1). Applying these relationships to impellers developed

BY: J.F. GUELICH AND T.H. MCCLOSKEY 118

The Pump Handbook Series

according to design rules differing from those underlying the correlation would be very misleading.

IMPELLER OUTLET RECIRCULATION Downstream to the impeller the flow may be decelerated in a stationary component of the casing. This deceleration can occur in a diffuser, a volute, an annular casing or a combination thereof. The physical mechanisms of downstream deceleration are quite similar to those previously mentioned for upstream recirculation. They are: 1.

deceleration of the absolute velocity from the impeller outlet to the throat of the casing

2.

incidence at the diffuser vanes or volute cutwater

3.

pressure gradients perpendicular to the direction of the main flow (particularly for semi-axial or axial pumps).

The main geometrical parameters impacting flow recirculation at the outlet are: •

velocity distribution at the impeller outlet as determined by the geometry of the impeller

TABLE 1. MEANS TO UNDERSTAND AND TO MODIFY THE SHAPE OF THE HEAD/CAPACITY CHARACTERISTIC (HCC) Possible causes or mechanisms

Symptoms Q-H-curve

Axial thrust, β/ω = f (Q)

Internal pressures ψp

• Insufficient recirculation at impeller inlet (insufficient centrifugal head increase at low flow)

1. HCC drooping towards shut-off Hp, ψρ flat towards Q=0 Hc

Q

• Insufficient recirculation at impeller outlet (insufficient exchange of momentum between impeller and diffuser at low flow)

β 0.5

Typical for nq < 30

Q Hc, flat or drooping towards Q=0

H

2. Excessive shut-off head and/or excessive shut-off power

ψp

Q • Excessive recirculation at impeller inlet

Rise of ψp towards Q=0

β

Rise of Hc towards Q=0

• Excessive recirculation at impeller outlet

0.5

Q Typical for nq > 35

H = total dynamic head Q = net flow through pump

Hc

H

nq < 60

Q Q

H

4. HCC too steep at high flow rates

• increase a3 or s8 • reduce b3 / b2 • reduce b2 • increase δTE

δTE

H

• Excessive flow accelera- • Increase diffuser or volute throat area tion in diffuser or volute throat • Cavitation in diffuser or volute

NPSH3% H

Q

Hp

η

Q

Q

• reduce d1i (increase d1 eff) • cut back return vane (increase cou) • reduce d1a reduce • increase hub dia. d1 eff • inlet ring

• Flow separation in diffu- • Axial thrust fluctuations and senser (volute) but not yet sitivity to rotor position / axial fully developed recirculstage stacking tolerances can be ation eliminated by reducing gap A and introducing proper overlap S • Shifting of flow patterns Q (zones of recirculation / • To remove saddle type instability Axial thrust excursions flow separation) detailed flow analysis and / or testing often required • Very sensitive to manu0.5 facturing tolerances of • Differences in stage geometry to diffuser and impeller get onset of flow separation in different stages at different flow • Very sensitive to impeller rates inlet flow conditions and S B impeller inlet geometric S = outlet recirculation on parameters shroud • Sensitive to axial rotor B = outlet recirculation on position: HCC and Fax hub Fax

30< nq < 60

Q

• reduce a3 or s8 • increase b3 / b2 • reduce gap A, increase overlap δTE • increase δTE • increase b2 • reduce gap B (with small nq) but beware of increased pressure pulsations

Q Sudden decrease of β due Q to outlet recirculation ψp = static head rise of F = axial thrust towards ax 2 impeller / (u2 /2g) β suction Hc = head rise in casing ω = ratio fluid/shroud rotation

3. HCC with saddle or flat position

H

• advance impeller vane leading edge (reduce d1 eff) • advance return vane trailing edge (reduce cou) • increase d1a (increase ∆ d1 eff) • reduce hub dia. (increase ∆ d1 eff)

}

Q Hc

Possible remedies NOTE: any modifications may have sideeffects which should be carefully assessed

Hc

Q Steep rise of NPSH little or not at all affected by impeller inlet

P

Q Q The Pump Handbook Series

119

itself as well as the velocity distribution at the impeller inlet •

diffuser or volute throat area

•

diffuser vane or cutwater angles

•

ratio of impeller outer diameter at tip to hub (oblique cut of radial, semi-axial or axial impellers).

INCREASING SHUT-OFF HEAD Ample evidence suggests that increasing the recirculation at the impeller inlet and/or outlet increases the head. A number of geometrical parameters can be altered to increase the head in this manner. If the flow versus pressure head (Q-H) curve is drooping towards shut-off, invoking recirculation may be appropriate, especially since the shut-off head is particularly affected by an increase in recirculation. Table 1 presents typical shapes of Q-H curves, explains the physical mechanisms responsible for these curve shapes and suggests possible remedies and geometric parameters by which the shape of the Q-H curve can be corrected. However, the reader should be aware that these remedies may produce undesirable side effects (e.g., reduction in head or efficiency at best efficiency point (BEP)).

DAMAGING RECIRCULATION For every type of pump there exists a range of optimum recirculating flow, and operating a pump within it avoids the risk of unstable Q-H-curves on one side and the risk of damaging levels of recirculation on the other. This range is illustrated qualitatively by the Figure 1 graph. Unfortunately, there is no established method to predict exactly the onset of damaging recirculation. An unacceptable level of recirculation can be determined indirectly, however, in individual cases by applying the following strategies: •

Measure cavitation noise to assess the risk of cavitation erosion.

•

Test for vibration.

•

Monitor shut-off head or shut-off power. If excessive, this may indicate inordinate recirculation.

•

Measure the radial or axial hydraulic excitation forces. A sudden rise in these forces as the flow rate is reduced and/or consistently excessive levels of these forces could indicate an unacceptable degree of recirculation.

•

Measure pressure pulsations. A sudden rise in pressure pulsa-

FIGURE 1. OPTIMUM AMOUNT OF CIRCULATION CNL

Optimum Design Lcav

Pressure pulsations / noise Hydraulic excitation forces

HMax-Ho

Hydraulic excitation forces and pressure pulsations may be responsible for a number of possible component failures. Table 2 details the root causes and mechanisms of such failures along with possible remedies. Partial load flow phenomena also strongly influence pump vibration. Observed vibration phenomena causes and mechanisms, along with possible remedies, are given in Table 3.

CAVITATION EROSION If, as a result of recirculation, the local pressure at the impeller inlet drops below the saturation pressure of the pumped liquid, vapor bubbles are generated and are then swept by the flow into zones of higher pressure, where they implode and may cause erosion of the impeller. To eliminate or reduce cavitation damage, the following remedies are available: • Change operation procedures if damage occured at partial load or overload. •

Increase net positive suction head available (NPSHA).

•

Reduce speed or use a varible speed drive if partial load is required.

•

Increase cavitation resistance of material.

•

Modify geometry of impeller (profiling of blades, impeller redesign).

•

Improve inlet flow conditions by geometric modifications.

•

Increase gas contents.

Low freq. pulsations

H=f(Q) unstable

QRec

H=f(Q) stable

120

tions as the flow rate is reduced and/or consistently excessive pulsation levels could indicate damaging flow recirculation. However, a sudden rise of pressure pulsations might also result from standing wave resonance. For this reason, to avoid the possibility of misinterpretation of high loads of pressure pulsation, careful testing and data analysis is imperative when diagnosing the true nature of pulsation.

The Pump Handbook Series

In addition, Table 4 outlines cavitation damage mechanisms and offers correlating remedies. As illustrated

TABLE 2. EFFECT OF HYDRAULIC EXCITATION ON COMPONENT FAILURE Failure / incident

Possible hydraulic causes or mechanisms

Possible remedies

Possible non-hydraulic causes / remarks

1. Fracture of impeller blades at outlet, diffuser vanes at inlet, tie bolts, instrument piping, or other components

• High dynamic stresses induced by pressure pulsations (impingement of wake flow from impeller blade trailing edge on diffuser vanes or volute cutwater)

• Increase gap B by cutting back

• There are a number of other failure mechanisms related to design, material selection and quality Remark

(1) diffuser vanes if diffuser throat does not increase by more than 3% (2) impeller blade trailing edge (head of pump will be reduced unless speed cannot be adapted)

• Reduce excitation at part load by modifying hydraulic components (careful analysis and redesign)

Pressure pulsations and dynamic stresses are expected to decrease with a power of -0.77 of gap B. For example to achieve half of the original level gap B must be increased by a factor of about 2.5

2. Side plate breakage

• High dynamic stresses induced by pressure pulsations • Impeller side plate resonance if z3 - z2 = 2 and z3 n/60 close to impeller side plate natural frequency

• Increase gap B (see previous item) • Change z3 / z2 combination • Modify natural frequency • Reduce exciation at part load by modifying hydraulic components

• Insufficient quality of impeller casting and / or finish (notch effect) • Insufficient thickness of impeller side plates

3. Mechanical seals

• High pressure pulsations caused by wake flow or recirculation / separation • High frequency pressure pulsation due to cavitation • Shaft vibrations

• (see above) • Reduce cavity volume by redesign of impeller and / or inlet • see table 3

• There are a number of other failure mechanisms related to design, material selection and quality

4. Excessive labyrinth wear

• Excessive radial thrust

• Reduce flow asymmetries around impeller by - double volute in case of single volutes - analyzing / eliminating cause of asymmetry (casting tolerances, differences in resistance in channels of double volutes, discharge and suction nozzle,...) • see table 3

• Thermal deformations of casing and rotor

• Excessive vibration

5. Failure of radial bearings

• Excessive radial thrust • Excessive vibration

• see above item • see table 3

• Mechanical / design

6. Failure of axial bearings

• Axial thrust excursions • Excessive labyrinth wear (high leakage increases rotation on shroud; reduces rotation on hub with multistage-pumps)

• see table 1, item 3 • Replace wear rings

• Mechanical / design • Transients

by the following case study, geometric modification of the impeller is frequently the only feasible solution.

CASE HISTORY After a boiler feed pump had operated for more than 50,000 hours with no trace of cavitation on the suction impellers, the load demand of the process changed, requiring prolonged partial load operation. The pump operated about 1000 hours at 60% load and 1100 hours at 80% load before cavitation damage was discovered on the pressure side of the impeller blades. The attack varied between 2-4mm from blade to blade. Since the cavitation damage occurred

at partial load on the pressure side of the blades, flow recirculation was identified as the most probable cause. To improve the partial load range and thereby increase the impeller life, an inlet ring was designed and installed in the pump. Figures 2 and 3 show the fluid-borne and solid-borne noise prior to and after this modification. Prior to modification the noise recorded at 100% and 80% flow is virtually equal. Since no erosion occurred during more than 50,000 hours of operation at full load, this evidence suggests that the operation at 60% load was entirely responsible for the damage. As illustrated by the figures, the modification of the pump The Pump Handbook Series

decreased the noise at 60% flow to the unmodified 100% flow noise level, and the erosion problems were solved. ■

ACKNOWLEDGEMENTS This article summarizes the results of investigations on hydraulic instabilities and cavitation erosion sponsored by the Electric Power Research Institute (EPRI), Palo Alto, CA and conducted under EPRI RP 1884-10. The authors are grateful to R. Egger, W. Handloser and A. Roesch, who carried out the extensive test program.

121

REFERENCES 1.

2.

Editorial Advisory Board. He is a fellow of ASME and a member of the Hydraulic Institute’s technical committee on pump intakes.

W.H. Fraser. ”Recirculation in Centrifugal Pumps.“ ASME Winter Annual Meeting. 1981. J.F. Guelich et. al. ”Feed Pump Operation and Design Guidelines.“ EPRI Final Report TR-102102. June, 1993.

NOMENCLATURE

D1

impeller eye diameter

D2

impeller outer diameter

d1

D1/D2

d1eff

impeller vane inlet diameter where flow enters the impeller

Fax

axial thrust towards suction

f

frequency

fn

rotational frequency

H

head per stage of pump

Hc

head rise in casing

A

amplitute

a3

diffuser throat width

b2

impeller exit width

b3

diffuser inlet width

Com

absolute velocity at meridonal inlet point

CNL

cavitation noise level

Hp

static head rise of impeller

cou

absolute velocity upstream of impeller

Lcav

cavity length

nq

pump specific speed (metric convention)

Q

flow rate

QSL

flow at shockless entry

S

outlet recirculation on shroud

80,000

S8

volute throat area

60,000

u1

circumferential velocity

z2

number of impeller vanes

z3

number of diffuser vanes

ß

angular velocity of liquid

δTE

trailing edge angle

σ

slip factor

σu1

cavitation coefficient (2gNPSH/(u1)2)

ψp

static pressure rise of impeller

ω

angular velocity of impeller

J. F. Guelich is manager of hydraulic pump design for Sulzer Pump Division in Winterthur, Switzerland. T. H. McCloskey is manager of turbo-machinery at the Electric Power Research Institute in Palo Alto, CA and a member of Pumps and Systems

FIGURE 2. FLUID-BORNE NOISE IN A FEED PUMP (1 TO 180 kHz) NL

N/m

2

100,000 X

original

40,000

X

X X

20,000

X

X

modified

X

X X

0 0

10

20

30

40

50

60

70

80

90 100 110 % Flow

FIGURE 3. SOLID-BORNE NOISE AT A FEED PUMP (10 TO 180 kHz) CV

m/s2

120

Subscripts: original

100 X

80 60 40

XX X X

20

modified

X X X X

X

0 0

122

10

20

30

40

50

60

70

80

90 100 110 % Flow

The Pump Handbook Series

av

available

BEP

best efficiency point

rec

recirculating flow

TABLE 3. INTERACTION BETWEEN FLOW PHENOMENA AND VIBRATION Observed vibration Spectral component

Q/QBEP

1. Subsynchronous peak close to fn

< 1.0

A

fn

Instability

> 1.0

f

Subsynchronous vibrations increase with time

all

2. Synchronous vibration

all

Possible hydraulic causes and mechanisms

Possible remedies

Major non-hydraulic causes / remarks

• Increased labyrinth preswirl ⇒ reduced rotor damping • Unloading of bearings due to change in radial thrust with flow • Rotating stall • Labyrinth wear ⇒increased leakage ⇒increased preswirl ⇒reduced rotor damping • Hydraulic unbalance due to various impeller tolerances

• Increase rotor stiffness and damping by introduction of plain labyrinths or shallow serrations only • Reduce preswirl to labyrinths by swirl brake • Impeller or diffuser redesign

• Labyrinth design • Thrust balancing device does not provide sufficient rotor damping • Bearing design / bearing unloaded Remark: Instability may be recognized by very steep increase in amplitude with increasing speed

• Pressure pulsations caused by wakes from the impeller blade trailing edge • Harmonics other than blade passing frequency are due to impeller casting tolerances (pitch of the blades) • z3 - z2 = +/- 1 resulting in non-zero radial blade force component at z2 fn

• Peaks nearly always present. If excessive: - Increase gap B (see table 2, item 1) • Harmonics other than blade passing frequency: reduce impeller casting tolerances • Change number of impeller or diffuser vanes • Reduce excitation force by proper staggering of impellers on shaft

fn

A

• Reduce casting / manufacturing tolerances of impeller (precision casting, ceramic core procedures, manufacturing) and implement more stringent question/answer procedures

f

3. Supersynchronous peaks A

fn

Z2fn

all

2Z2fn

f

A f fn z2

= amplitude = frequency = rotational frequency = number of impeller blades

4. Broad band shaft vibrations fn

A

f

5. Structural resonances below frequency of shaft rotation excited by broad band hydraulic forces (e.g. bearing housing, bed plates, piping, ...) A

typically • Flow recirculation at impeller below inlet and outlet 50% of • Some broad band vibrations BEP flow are unavoidable. If excitation is excessive this can be due to oversized throat areas of diffuser or volute, oversized impeller eye or excessive incidence • Fluctuating cavities

• If excessive: reduce diffuser or volute throat area; reduce impeller eye (careful review of hydraulic design required) • As a cure of the symptoms the rotor damping might be increased (swirl brakes, labyrinth redesign, see above) • Reduce cavitation extension (higher NPSHav, redesign of impeller or inlet)

Remark: It is typical that structural resonances excited by broad band forces do not depend on the speed of the shaft

fn

f

6. Rotating stall

typically • Stall cells in diffuser or impeller • Analysis and redesign of below rotating with a frequency below hydraulic components 90% of the frequency of rotation • Increase rotor damping (swirl BEP flow brakes, labyrinth redesign)

7. Surge-like strong pulsations

low

Remark: The peaks are expected to be proportional to the rotor speed

• Vapour core forming in the • Structures upstream of imsuction pipe at low NPSHav due peller to avoid formation of core (flow straightener, cross, inlet to strong part load recirculation ring, hub diameter, “back-flow catcher”) • Impeller / inducer redesign • Air admission (if possible)

The Pump Handbook Series

123

TABLE 4. CAVITATION DAMAGE MECHANISMS AND REMEDIES Type of cavitation / damage pattern

Flow mechanisms likely to induce damage

1. Suction side of blade, starting close to leading edge of blade

Sheet cavitation on suction side of blade at Q < Q SL • Damage near shroud • Damage near hub

2. Suction side of blade, within channel

Sometimes erosion is also observed on the shroud and / or the hub or on pressure side of the blade

Possible causes

Possible remedies

• High incidence • Unfavorable leading edge profile

• • • • •

Increase flow rate Reduce blade inlet angles Improve leading edge profile Reduce incidence by inlet ring Reduce impeller eye diameter if above optimum range • Increase pre-rotation

• Outer blade angle β1a too large • Inner blade angle β1i too large • High incidence at low σ (typically σu1 av = 0.15 to 0.3) • Insufficient NPSHav • Insufficient cavitation resistance of material

• Reduce impeller eye diameter if above optimum • Reduce blade inlet angle • Increase NPSH available • Increase cavitation resistance of material

• Negative incidence due to excessive flow • Unfavorable leading edge profile • Excessive (run-out) flow

• Reduce flow rate • Increase blade inlet angle (but beware of partload cavitation) • Improve leading edge profile if damage close to leading edge

• Excessive flow deceleration (excessive impeller eye diameter, excessive impeller throat area, excessive blade angles)

• Increase flow rate • Impeller redesign (decrease eye diameter / throat area / blade angles) • Inlet ring at the impeller entrance (reduce deceleration, reduce shear flow effects)

• Negative incidence near hub due to partload recirculation

• Increase blade angle at hub • Reduce recirculation • Improve leading edge profile to reduce flow separation near hub under recirculation • Reduce preswirl (vanes, ribs, backflow catching elements) of recirculating flow

• Blade angles not properly matched to the flow • Fillet radii too large

• Impeller redesign (adaption of blade angles) • Required fillet radii

Vortex cavitation on suction side of blade at Q < Q SL at low σ. Vortexes developing downstream of a long, thick cavity attached to the blades. Bubbles created in the vortexes are swept away by pulsating flow and can implode anywhere in the channel. QSL = flow at shockless entry

3. Pressure side of blade, any location, starting close to leading edge of blade

Sheet cavitation on pressure side at Q > Q SL 4. Pressure side of blade, damage at outer half of impeller width starting close to leading edge Bubbles in free stream generated by shear flow due to partload recirculation. Bubbles impinge on pressure side of vane. 5. Pressure side of blade, damage near hub. Difficult to distinguish from item ③ unless it can be determined whether pump has operated at partload or above BEP

Lcav

l

tua

in

ω

m 0Q

ac

BEP

ω0 Q oretical ω 0 Qmin the

COM

U1

Excessive partload recirculation creates a negative incidence near hub 6. Damage on hub or shroud or in fillet radii

Comer vortex cavitation often combined with high incidence

124

The Pump Handbook Series

CENTRIFUGAL PUMPS HANDBOOK

High Speed/Low Flow Pumps: Top 10 Issues The authors answer your top ten questions about high speed/low flow pumps that perhaps you were afraid to ask.

he September 1994 issue of Pumps and Systems featured an article entitled ”Low Flow Options.” Among other things, it discussed the single port diffuser pump design credited to Dr. U. M. Barske. A high-speed version of this design was popularized in 1959 with its application as a water injection pump for the Boeing 707 airliner. An industrial version of this concept was first marketed in 1962 and tens of thousands have now been placed worldwide. The single port, or partial emission, pump has been developed commercially by a number of manufacturers. The designs accommodate both two and four pole electric motor drives and are most often applied for low specific speed services. Figure 1 shows the basic geometry of an open impeller, ”partial emission” diffuser, design. The high-speed version of this design has garnered its share of ”wives’ tales” over the 30-plus years of commercial manufacture. Following are the ”Top 10” notions and real issues that generate continuous discussion among users of these products. Although many of the topics are worthy of individual papers, we are limited here to an overview along with some helpful hints.

10. EFFICIENCY: HERE TODAY, GONE TOMORROW? Partial emission pumps use open impellers and therefore do not rely

PHOTO COURTESY OF KOP-FLEX, INC.

T

Photo 1. A centrifugal pump performing high speed/low flow duties in a Gulf Coast plant. on wear ring clearances to maintain hydraulic efficiencies. Although this design allows more recirculation back to the suction than an enclosed impeller design, efficiency definitely depends on the disk friction component. Disk friction is the drag loss between the body of rotating fluid being carried by the impeller and the stationary walls of the chamber (Figure 2). In clean, non-corrosive

streams, the finish on the chamber walls remains in ”as new” condition resulting in steady, long term hydraulic efficiency. Typical finishes are machined to a 62 (micro inch) RMS value. Thus, a surface finish of 250 can reduce the hydraulic efficiency by as much as 10 points. In extreme cases, total power consumption has been known to double that of the ”clean” rating. Of the

BY: DAVE CARR AND ROBERT LINDEN The Pump Handbook Series

125

rienced in contaminated butane streams has been negated by upgrading to 316 stainless steel construction. Either of these effects can be verified by measuring the motor amperage or the liquid’s temperature rise. The latter indicator increases proportionately to the decrease in hydraulic efficiency, i.e. by the equation:

FIGURE 1. OPEN IMPELLER

∆T={H(1-η)/778Cpη}

Impeller Bowl Surface

Back Cover Surface

two chamber surfaces, the finish of the backing plate is actually of greater significance than the impeller bowl in its effect on efficiency. Machined skin cuts at depths of 0.005 to 0.010 inch have been shown to efficiently restore the pump’s original performance by reconditioning the surfaces to the material’s original finish. In addition to wear of the bowl material, there is also potential for foreign material build up on the surfaces. The corresponding skin, or film, similarly changes the effective finish and may also degrade the efficiency. The pump’s sensitivity to this condition increases with tip speed and operation with lighter liquids. For this reason, wetted materials may be selected to shed any build up or to resist the potential for accumulations to occur. For example, salts have an affinity for carbon steel materials. Build-up of the type commonly expe-

126

9. THE SUCTION SPECIFIC SPEED IS WHAT? Quoting from Lobanoff & Ross, ”The inducer is basically a high specific speed, axial flow pumping device...that is series mounted preceding a radial stage to provide overall system suction advantage.” This relationship is paramount to evaluating suction specific speed with high-speed pumps. Figure 3 shows a portion of a well-rounded inducer family. An industry rule of thumb is to limit suction specific speed to a value of 11,000 (English units) in heavy duty process pumps. It is based on the premise that pumps operating at greater suction specific speeds have

FIGURE 3

Inducer C

Inducer B

Suction Specific Speed

FIGURE 2. PARTIAL EMISSION CHAMBER SURFACES

Units: T is in °F, H is head in feet, Cp is specific heat in BTU/LB-°F and η is efficiency expressed as a decimal. Comment: Practically speaking, temperature rise analysis can be difficult because it typically ranges between 2 and 10°F.

demonstrated a shorter service life, most likely due to rougher off design operation.. As conventional impellers are capable of suction specific speed values greater than 11,000, via oversized impeller eye configurations, this may be a good guide. To provide a useful net positive suction head (NPSH) performance guide for highspeed pumps, however, will require a change in the reference point! High-speed pumps inherently require inducers to achieve competitive performance. Inducer designs have advanced to the point of providing reliable, cavitation-free operation with suction specific values of approximately 24,000. Efforts to desensitize inducer operating ranges have been addressed by optimizing blade number, angle and passage areas as well as using various inlet bypass designs. One design that resists cavitation surge, referred to as a ”backflow recirculator,” is reported to improve inducer turndown to a near shut off condition. Suction vortex breakers are also used but are less effective than flow stabilizers. Nonetheless, they can improve low flow stability, with high-speed/inducer style pumps, by as much as 25 to 35%. In general, inducer use is discouraged with conventional API pumps, but high-speed pumps clearly require them. Several leaders in the Hydrocarbon Processing Industry

GPM / RPM The Pump Handbook Series

Inducer A

FIGURE 4. PRESSURE-SPEED LIMITS AS A FUNCTION OF SEAL FACE MATERIALS 25,000

20,000

PUMP SPEED (RPM)

1. 15,000

2. 10,000

3. 4. 5. 6.

5,000

7.

SIC. = SILICON CARBIDE T.C. = TUNGSTEN CARBIDE C. = CARBON

0 0

500 1,000 SEAL DIFFERENTIAL PRESSURE (PSIG)

1. 1 1/4", 1 1/2", & 2" SIC. vs. C.

5. 1 1/2 IN. T.C. vs. SIC.

2. 1 1/4 IN. T.C. vs. C.

6. 2 IN. T.C. vs. C.

3. 1 1/2 IN. T.C. vs. C.

7. 2 IN. T.C. vs. SIC.

1,500

4. 1 1/4 IN. T.C. vs. SIC.

(HPI) have recognized the unique position that this equipment occupies in the marketplace and have explicitly exempted such designs from the 11,000 suction specific speed limit. Care still must be exercised, however, in matching particular inducer configurations with the pump’s operating flow range, and interaction with manufacturers that recognize cavitation erosion limits within their application guidelines.

8. HOW CAN MECHANICAL SEALS HANDLE THAT SPEED? Mechanical sealing of high-speed pumps presents numerous obstacles including the potential for vibration, high sliding speeds, heat generation and high sealing pressures. Much of this is the direct result of rotational speeds that reach a maximum of 25,000 rpm, which is more common with compressors than with pumps. These conditions place high-speed

pump manufacturers in the unique position to accept ”seal success ownership.” The bulk of the issues are addressed by reversing the conventional seal configuration. Vibration control dictates the spring-loaded component as the stationary part with the hard face as the rotating member. Small seals, typically with either 1-1/4” or 1-1/2” diameters, are used to minimize sliding speeds. A 11/4” seal at 15,000 rpm has a face sliding speed of approximately 82 feet per second. This equates to a 5-1/2” seal operating at a traditional 3,550 rpm (Figure 4.). When combined with the fact that open wheel impellers present seal pressures nearer to suction than to discharge (i.e., approximately 10% rise over the suction pressure) manageable pV values are seen at the seal faces. This allows most applications to be handled by conventional carbon vs. tungsten carbide material combinations. Silicon The Pump Handbook Series

carbide also may be used to raise the operational limits. Most high-speed pump seal problems are not caused by high speed, but by a lack of understanding or information regarding the fluid’s properties. The most common problems are rust and scale, inadequate vapor pressure margin and a build up of solute at the atmospheric interface. Proper application of typical API seal flush plans, e.g., -31 (flush through a separator), -13 (reverse flush) and -54 (quench), normally negate these concerns and promote good seal reliability with high speed pumps.

7. PREPARING THE PUMP FOR STARTUP. High-speed pumps are typically accorded extra care during the startup process due to respect for the technology. The lubricant level, oil cooler venting, fresh oil filter installation (if applicable) and driver rotation are inspection points which are consistently addressed. Seal piping, however, can be quite another story. API Standard 610 dictates, and industry practice provides, that all seal ports be plugged prior to factory shipment. In the field, the permanent appearance of some of these plugs frequently leads to failure to remove them. The pump case tags, engineering drawings and the instruction manual all provide a critical definition of the ports’ functions and the corresponding auxiliary piping requirements, and must be followed. Unfortunately, failure to properly configure the pump often will not cause any immediate problems. A common occurrence is an improperly vented port that can direct accumulated process seal leakage toward the back side of the gearbox mechanical seal, or bearing seal. The potential for lubricant contamination exists when an atmospheric drain is connected to a flare header that experiences significant upsets. Care should be exercised to ensure that these actions do not occur since the livelihood of the gearbox depends upon limiting this pressure to an absolute maximum of 10 psig.

6. SUBTLE TRUTHS ABOUT NPSH. Net positive suction head (NPSH) is a subject worthy of its own full

127

128

High-speed centrifugal pumps are often installed in applications designed for positive displacement (PD) pumps. This is due to the inherent ability of both pumps to deliver a high differential pressure. Unfortunately, the two designs must operate under significantly different control schemes. This fact must be recognized when retrofitting from one configuration to another. Figure 5 shows the theoretical characteristics of PDs and centrifugals with regard to flow and head capabilities. It is evident that the positive displacement design is limited by head (pressure) and the centrifugal by flow. Consequently, the PD uses a pressure relief valve to prevent over pressurization and to bleed off excess capacity. In practice, centrifugal pumps exhibit only a moderate head rise across their operating region. The radial vaned centrifugal, in particular, demonstrates a 5-10% head rise from the best efficiency point (B.E.P.) to the peak of the head versus capacity curve. This margin does not facilitate the use of a pressure relief system for control purposes. Further, the relief valve scheme can result in wasted power when the pump is allowed to run out to the extreme right of the B.E.P., under ”low load” conditions. The Pump Handbook Series

4. WHAT ABOUT UNCONTROLLED FLOW OPERATION? The effect of low flow operation on centrifugal pumps is commonly discussed, but rarely is the opposite end of the performance curve considered. Regardless of the hydraulic hardware, NPSHR generally increases with a pump’s operation at greater than its rated flow. This topic has been mentioned within issue #6 and therefore will not be discussed further. Single volute and diffuser style pumps experience increased radial loads when applied at greater than

FIGURE 5. THEORETICAL HEAD RISE Radial Vaned Centrifugal Positive Displacement Backward Lean Centrifugal

Flow Coefficient

FIGURE 6. CENTRIFUGAL PUMP RADIAL LOAD CHARACTERISTICS Single Volute

Design

5. CENTRIFUGAL PUMPS IN A POSITIVE DISPLACEMENT WORLD.

The recommended high-speed pump control system is with the use of flow control. At first glance, this concept can be intimidating but is essentially synonymous with level and mass control schemes that are typical within process systems. Some processes do demand strict pressure control. When that is the case, a pressure controlled throttle or bypass source may be required.

Head Coefficient

prohibited by company specifications. In general, increasing the NPSH margin by an additional 4 to 6 feet is appropriate with light gravity liquids, i.e., less than 0.7 specific gravity. Also of concern is the fact that the NPSHR value typically increases beyond the best efficiency point of the machine. Pump startups are commonly uncontrolled and result in operation at the end of the curve due to a lack of sufficient back pressure. The practical result of this situation is the fact that the pump may then run at too high an NPSH requirement, and could promote a disconnect between the impeller/inducer and the liquid stream. Therefore, the control valve’s initial trim position and manual venting of the pump system should be anticipated in preparation for startup.

Load

fledged article, and has been covered previously in Pumps and Systems. It should be emphasized, however, that a centrifugal pump’s NPSH performance is established based upon the breakdown of the standard head versus capacity curve. When inducers are used, the total pump NPSH performance is measured, not just the inducer’s performance. The most commonly accepted parameters are based on the Hydraulic Institute’s 3% head suppression criterion. By definition, however, the pump will cavitate when conditions at the suction flange meet those test, or predicted curve, conditions. High-speed inducers (in essence axial pumps) can develop as much as 25 to 100 feet of head. This energy, however, cannot be included within the overall pump requirements due to corresponding inlet eye losses at the centrifugal impeller. The net result is that the centrifugal portion of the pump still must be sized for the full design head rating. The subtle aspect of NPSH revolves around the pumped fluid’s properties and its potential to flash. Cavitation damage is a function of a liquid’s propensity to release vapor. Water pumps typically will produce rated head and flow, albeit with the potential for some material damage, despite close proximity to the 3% head suppression value. This may be attributed to the high surface tension characteristic of water. The same trait makes it aggressive toward cavitation damage but the pump generally works! Difficulty occurs at the other end of the spectrum where high speed pumps are used with low specific gravity services. Using the 3% suppression value for light fluids, plus the industry rule of thumb for an additional 2 to 3 feet safety factor, the NPSH margin may not be adequate with low specific gravity fluids. One example of this situation is realized when a slight heating of the fluid occurs on the suction side of the pump, particularly with aboveground supply pipes from storage vessels to transfer pumps. This can result in off-gassing and failure of the pump to hold prime. Ironically, this situation contradicts the API hydrocarbon offset factors that are typically

0

Diffuser

25 50 75 100 125 150 Percent of Design Flow

FIGURE 7. POINT EMISSION PUMP PERFORMANCE CHARACTERISTICS

Head

Knee of Curve

Potential Zone of Discharge Cavitation

50 100 120 Percent of Design Flow

FIGURE 8 B

Resistance

V S F

S'

Flow B = System Curve V = "Variable" Component F = "Fixed" Component

design flow rate (Figure 6). It is generally understood that sufficient bearing over capacity, and/or control limits, must be applied to the pump to account for these events. Highspeed centrifugal pumps share these same basic design needs, but also must address a phenomenon often referred to as discharge cavitation. Discharge cavitation occurs within single divergent conical (point emission) diffusers when the pump is operated at rates to the right of knee of the performance curve (Figure 7). Under such conditions, a low pressure zone forms on the trailing edge of the impeller’s blades. Vapor bubbles are formed and subsequently collapse in a manner consistent with the standard definition of cavitation. The result is impeller blade pitting, increased pump vibration and the classic ”rock pumping” noise associat-

ed with suction performance problems. Such an occurrence is common with systems where the pump has been oversized on head, inadequate control is afforded to maintain operation within a maximum flow limit, or simply during a process startup where the system is being filled. A good rule of thumb is that this type of pump should be controlled to a maximum flow of 120% of the pump’s rating. System requirements exceeding this value require a conversion to a larger diffuser throat to accommodate the actual process demand. Use of the design’s conversion capabilities is always preferable to grossly oversizing the machine for future requirements.

3. A CONFLICT BETWEEN CURVES! Some, but not all, high-speed pumps produce flat or ”drooping” curves. This characteristic, which is common to low specific speed pumps, has elicited much discussion regarding their inherent ability to be controlled. Such pumps, however, have been successfully applied in tens of thousands of applications in spite of these concerns. Unfortunately, however, many times this has been accomplished at the expense of energy by oversizing the pump’s rated head and then employing a discharge orifice to artificially steepen the curve as it runs back toward its minimum flow point. The application key is to understand the pump and the system curves. It is a common belief that drooping curves are difficult to control because the pump has two flow points associated with a single head (pressure) point. The result is a tendency for the pump to ”hunt” between the two flows. What is often overlooked, however, is that the pump merely reacts to what the system presents it with (i.e., it operates at the exact point where the system and pump curves intersect). A process system’s characteristic resistance curve typically is made up of two components. The first is referred to as the ”fixed” element and is associated with the system’s static component, e.g., the operating pressure of a process tower. This element The Pump Handbook Series

is considered to be constant with respect to the flow rate of the system. Conversely, the ”variable” component may be simply thought of as the frictional element that is related to the pumped flow rate. These two factors are shown in Figure 8. When combined, they become the basic system curve. Figure 9 shows a typical head versus capacity curve (ABC) with a drooping characteristic. Point B signifies the best efficiency point (B.E.P.), point A the cutoff flow and point C is the stonewall condition. Superimposed on this curve is the basic system curve (SB) which was derived from the previous discussion. Without supplemental pump control, the system will demand a flow rate equal to XB. The system head curve can be modified with changes to the piping system or by regulating the pressure drop across a control valve. The latter approach is a typical means of controlling centrifugal pumps and yields new system curves as indicated by (SD) and (SA). It is seen that the pump’s head capability is equal at the XA and XB flow points yet successful pump operation is accomplished as a result of the modified system curve. We would be remiss in not pointing out that this type of curve does have its shortcomings. First, pressure control normally is not practical due to the relatively small head rise that occurs between the rated and maximum head points, typically on the order of 5-10%. This fact favors the use of flow control methods. Second, systems that are comprised predominantly of the ”fixed” component, i.e., those that exhibit little influence as a result of the system’s demand flow,

FIGURE 9 D A

B

S C XA

XD

XB

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may result in an undersized pump if an active appraisal of the pump/system interaction is not performed. Regardless of the centrifugal pump type, a discrete intersection between the pump and system curves will always complement pump stability and controllability. Conscientious attention to the interaction between pumps and systems can tame both so that they work in harmony.

2. WHY WON’T THIS PUMP WORK ON WATER? Pump manufacturers typically use water as a performance test medium for safety and convenience reasons. Many pumps, however, are sold for process liquids that vary between a 0.4 and 0.8 specific gravity. Factory tests often use one-half speed motors, or other speed changes to compensate for the increased power that results from water’s density difference to the contract liquid. Field operation, however, commonly uses an initial run-in on water to flush and prove the system. Two typical repercussions of this action are an expected overload of the driver or an unexpected overload of a highspeed gear or bearings. The highspeed pump is particularly vulnerable to this off-design operation as a result of its common use in light hydrocarbon processing applications and as a result of small, custom matched, components which capitalize on the specific application’s needs. Throttling the pump may or may not meet the pump’s needs because bearing loading may be violated in extreme cases. The moral of the story is to check with the pump vendor before proceeding with off-spec tests. The optimal approach is to advise him of your complete run-in conditions prior to placing a pump order to ensure that it will meet all of the intended uses.

1. WHERE IS THAT CONTROL VALVE? Centrifugal pump designers expect throttle valves to be very near the pump discharge, while system designers prefer a location near the demand point. This issue becomes more than one of aesthetics when it involves high energy pumps. A surge phenomenon may occur with pumps with either continuously rising or drooping head versus flow

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curve shape attributes. It is distinguished by fluctuations in head capacity at the pump’s low flow conditions. The combination of highspeed technology with relatively low design flows introduces unique challenges to the pump designer and user. Empirical testing shows that this pump design’s low flow stability is directly influenced by the system within which it operates. The single point emission diffuser and a Barske impeller may be simplistically characterized as discharging flow each time a blade passes by the throat. It is reasonable to envision a void between the time in which one blade passes and the next arrives to distribute its supply. This interim period represents the opportunity for the discharge control valve’s location to influence the pump’s low flow stability. The valve’s interactive process may be visualized in the context of a simple spring mass system. The liquid in the system acts as the mass and all storage devices within the system, e.g., piping, vessels, etc., provide the spring medium. Excitation of this system may be initiated from a number of sources but often may be related to the blade passing frequency. The greater the energy that is accumulated within this system, i.e., the spring, the greater the propensity for it to disrupt the pump’s stability. The momentary flow reversals cause surge circulation between the blades. This type of system is sometimes referred to as ”soft” or ”spongy” since it reinforces the amplitude of the theoretical spring force. The destructive energy of this situation is greatest with an increasing mass of liquid, when the throttle valve is remotely located, and with increasing pump power. This phenomenon is minimized when the system can be described as ”hard.” This is accomplished by placing a control valve, or orifice, near the pump’s discharge flange. In effect, this change reduces the liquid mass, thereby minimizing the amplitude of the spring’s movement, and the flow oscillations. The valve/orifice also disrupts the frequency of the excitation force and further improves the pump’s low flow stability. The ”hard” system should always be the goal since it discourages the formation of a potentially dangerous energy source that can damage piping

The Pump Handbook Series

and induce mechanical vibration into the high-speed pump. Good rules of thumb are that transmitted power levels of less than 25 horsepower are minimally affected by this phenomenon and the optimum control valve location is within 5 feet of the pump’s discharge flange. Failure to address this situation can reduce a 200 horsepower pump’s minimum recommended continuous flow rate from 40 to 65% of the B.E.P., based upon the valve’s placement 25 feet, rather than 5 feet, from the pump’s discharge flange. ■

REFERENCES Val S. Lobanoff and Robert R. Ross, Centrifugal Pumps Design & Application, Gulf Publishing Company, Houston, TX, 1985. Donald P. Sloteman, Paul Cooper and Jules L. Dussord, Control of Backflow at the Inlets of Centrifugal Pumps and Inducers, presented at the First International Pump Users Symposium (1984). Robert Linden is the director of Sundyne and Sunflo products for Sundstrand Fluid Handling, Arvada, CO. Dave Carr is a senior marketing specialist with Sundstrand Fluid Handling.

CENTRIFUGAL PUMPS HANDBOOK

Pump Ratings Vital When Pressure’s On Though a relatively simple subject, pump ratings can generate much disagreement. Using objective evaluation procedures, however, we can shed light on the topic without generating heat and pressure.

ll pumps are pressure rated. The rating is the maximum pressure a pump casing can safely contain, often termed maximum allowable working pressure or MAWP, at a given temperature (Figure 1). Temperature affects the rating because the strength and stiffness of the materials used for casings – mostly metals – vary with temperature. From that simple definition, the subject becomes more complicated as we add definitions for the pump’s maximum discharge pressure, the means of correcting for different temperatures, and the question of what to do with the casings of pumps that develop high differential pressures. The maximum allowable working pressure (MAWP) of a pump’s casing is a function of its geometry, the material from which it is fabricated and the intended service temperature. Before delving into numbers, let’s address a fundamental aspect of casing geometry, namely the casing joint.

A

THE CASE FOR CASING JOINTS Pump casings must have a joint — either at right angles to the shaft

axis (radially split) or parallel to it (axially split) — to allow the pump to be assembled and dismantled. Most pump casings are radially split: in small overhung pumps, either horizontal or vertical, because of their inherently lower cost; in medium size and large vertical pumps for both lower cost and ease of maintenance; and in high pressure pumps, either horizontal (Photo 1) or vertical, because it’s the more cost effective solution. In between these extremes, axially split casings are preferred for horizontal single stage double suction and multistage pumps (Photo 2) for lower cost and ease of maintenance. They are, however, limited in the pressure rating they can economically achieve. API-610, 7th Edition1, recommends using radially split casings for hydrocarbon service when the maximum discharge pressure is above 1,000 psig (70 bar), the pumping temperature above 450ºF (235ºC), or the liquid specific gravity below 0.7. These conservative limits reflect the difficulty various refiners have had maintaining pressure tightness in axially split casings. There are, however, many examples of axially split

casings being used successfully in hydrocarbon service beyond these limits, and the forthcoming 8th Edition of API-610 will recognize this by raising the recommended pressure limit to 1,450 psig (100 bar). In water injection and boiler feed applications, axially split casings are regularly used to pressures of 2,500-2,750 psig (175-190 bar). Higher pressures are possible, but the cost of the casing can become prohibitive, and maintenance of the split joint gasket a major concern. The radially split casing in Photo 1 is one piece with a cover or head at the outboard end. This design has one high pressure seal, and the pump can be dismantled without breaking its suction or discharge connections or moving its driver. There is an alternative form of radially split casing, known variously as “ring section” or “segmental ring,” composed of many pieces clamped together with tie bolts (Photo 3). This design achieves low cost at the expense of maintenance. It has many casing joints, does not comply with API-610 and was wisely dropped from use in the US in the mid 1930s for all but small industrial

By J. T. (“Terry”) McGuire The Pump Handbook Series

131

Photo 2: Axially split multistage pump casing

boiler feed pumps.

DETERMINING MAWP

Photo 1: High pressure radially split casing

For a given casing geometry, a pump’s MAWP is determined by allowable stress unless strain (deflection) at critical sealing surfaces dictates a lower stress. The allowable stress may be that from ASME Section VIII, Division 12, as required by API-610, or some other similar limit. Designs using the ASME stress limits also include a casting integrity factor, which is 0.8 unless volumetric NDE of the castings allows a higher factor. Note that the allowable stress from ASME Section VIII, Division 1, includes a large design or “ignorance” factor to account for design using simple means of estimating the stress. When more sophisticated means of estimating the stress are employed, such as finite element analysis (FEA), using a higher allowable stress is justified because the local stress values are

Photo 3: Segmental ring casing

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The Pump Handbook Series

now known with quite high accuracy. In such cases, the allowable stress from ASME Section VIII, Division 2, can be used provided that the FEA model has been verified for accuracy and that appropriate material quality is used.

CALCULATING MAXIMUM ALLOWABLE DISCHARGE PRESSURE In operation, the maximum discharge pressure developed by a centrifugal pump is equal to the sum of the maximum suction pressure it can be exposed to plus the maximum differential pressure it can develop. In a simple world, that definition would be sufficient, but the world’s not so simple. Pumps are purchased with various reserves, margins and tolerances in head and rotative speed that complicate the definition of maximum discharge pressure. Following the requirements of API-610, 7th Edition, produces two definitions, one for fixed speed pumps, the other for variable speed. With Figure 2 as a reference, the definitions are: Fixed speed pump Pd, max = Ps, max + 1.05(∆ Pmax)fH

(1)

Variable speed pump Pd, max = Ps, max + ( ∆Pmax)fH(fN)2

(2)

Where: Pd, max is the maximum discharge pressure Ps, max is the maximum suction pressure ∆ Pmax is the maximum differential pressure at rated specific gravity fH is a factor to account for the allowable tolerance on shut-off head fN is a factor to account for the allowable overspeed to trip. API-610’s requirement for variable speed pumps covers the head reserve in the allowable positive head

tolerance, and it assumes the pump will be operated as if fixed speed, i.e. its flow will be controlled by throttling. Boiler feed pumps for central stations are often variable speed drive to avoid the losses associated with control by throttling. As such, their head always corresponds to the system resistance at any given capacity, and it’s therefore appropriate to use a different definition for the maximum discharge pressure. Referring to Figure 3, it is: Pd, max = Ps, max + PCMR (3) Where: PCMR is the differential pressure at the pump’s continuous maximum rating (CMR).

fA is a factor to account for the accumulation pressure. PUMP RATING VERSUS REQUIRED For a pump’s pressure rating to be acceptable, the MAWP of its casing, at the intended service temperature (Figure 1), must be at least equal to the pump’s maximum discharge pressure, as calculated by the applicable equation above. Depending upon the design status of the pump being considered, there are two approaches to achieving this. For existing designs, which is the usual case, the MAWP of the pump’s casing at the design temperature, generally 100ºF (40ºC), is corrected to that for the intended service temperature, T, by the equation: MWAPT = MWAPD(σT/ σD)

With this definition, the 10-15% higher discharge pressure that could be developed in the event the driver went to overspeed while the pump was blocked in is classified as a momentary excursion into the margin provided by the casing’s hydrotest pressure. Positive displacement pumps, unlike centrifugal pumps, will develop pressure equal to the resistance they encounter, up to the mechanical limit of the pump or its drive. This is obviously an extremely dangerous possibility, and so the cardinal rule in the application of positive displacement pumps is the provision of a full capacity relief valve at their discharge, upstream of any possible obstruction. The accumulation pressure of the relief valve, the additional pressure drop across the valve to discharge its rated capacity, should be no more than say 20%. With this provision, the pump’s maximum discharge pressure is (Figure 4): Pd, max = PsetfA Where: Pset is the relief valve set pressure

(4)

(5)

Where: σT is the allowable stress at the intended service temperature σD is the allowable stress at the design temperature Unless specifically stated otherwise in the engineering specification, the intended service temperature for casings is taken as the pump’s normal operating temperature. The rationale for this is that the maximum temperature generally represents a possible short term transient condition. When the casing is being designed specifically for the application, a common practice with fabricated casings, the pump’s maximum discharge pressure is used to calculate a minimum design pressure, MDP, by the equation: MDP = Pd, max(σD/ σT)

(6)

The MDP is generally rounded up to the nearest common increment, 25 psi being the ASME practice, or to a minimum pressure required by the engineering specification or a connected flange.

MINIMUM CASING DESIGN PRESSURE The Pump Handbook Series

133

FIGURE 1

FIGURE 3

MAWP

H @ rated speed

Total Head

Pressure

CMR

H @ minimum speed

System head

Bypass Pumping Temperature

Flow

Variation of MAWP with pumping temperature

Variable speed boiler feed pump

FIGURE 2

FIGURE 4

Allowable test curve with maximum positive tolerances

Pressure

Total Head

Slip

Proposal curve Rated point

Relief valve accumulation Relief valve set pressure

System resistance Suction

Flow Flow

Pressures in positive displacement pump applications

Allowable actual pump head curve, API-610, 7th Edition

FOR SHOP TESTING When designing a pump to handle a liquid of low specific gravity, it may be necessary to raise the MDP calculated from equation (6) to accommodate the pressure that will be developed during shop testing with cold water. The equation for the maximum discharge pressure on test is: Pd, test = Ps, test + (Hmax/2.31)fH

(7)

Where: Ps, test is the maximum suction pressure during the test Hmax is the highest head the pump will develop during the test, in ft.

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The maximum pump head developed during the shop test is usually at shutoff, except in the case of a pump with a drooping head characteristic, or high energy multistage pumps, which are tested only down to their minimum continuous flow. The head tolerance factor, fH, is that for shutoff.

CONTROVERSY OVER DUAL PRESSURE CASINGS Multistage centrifugal pumps develop large pressure differentials. This means various regions of their The Pump Handbook Series

casing normally are subjected to distinctly different pressures. Once the pump exceeds a certain size and pressure rating (3-inch discharge and ANSI 1,500 # flanges are a good starting point) it becomes more economical to design the casing for two pressures, a convenient low pressure for the regions subjected to suction pressure, maximum discharge pressure for the remainder of the casing (Figure 5). This practice is normal in the utility industry and is recognized by API-610, the standard for the process industry. API-610 also recognizes this is a controversial topic within the process industry, so it includes the option of specifying that the entire casing be designed for maximum discharge pressure. Designing the normal low pressure regions of a barrel pump casing for discharge pressure has two negative effects that need to be taken into account. The first is that the shaft seal design is often compromised. Hot charge pumps, for example, are best equipped with metal bellows type shaft seals. These cannot withstand static pressures of more than 350-400 psig (24.0-27.5 bar). Therefore, pusher type seals have to be used. Pusher seals rely on some form of cooling to preserve their elastomer dynamic gasket so the design has become more complicated and potentially less reliable. The second drawback is that the flange of cartridge mounted seals becomes so large and heavy that installing and removing it presents a serious handling problem. The controversy over dual pressure casings appears to have its origin in operating practices and the static pressure tightness of mechanical seals. In the utility industry and about half of the process industry, standard isolation practice is to block in the discharge, then open a drain before blocking in the suction. When this sequence is followed, the entire casing and suction piping back to the

FIGURE 5

A

B

A

A = suction pressure + 75 psi B = maximum discharge pressure Regions of barrel pump casing subjected to different pressures.

suction block valve cannot be accidentally subjected to discharge pressure by a small leak past the discharge valve. The alternative sequence, blocking in the suction before opening a drain, when carried out on pumps with mechanical seals, does risk subjecting the entire pump casing and suction piping downstream of the suction block valve to discharge pressure. There have been instances where doing this has ruptured the suction piping or the pump casing. It does seem that a judiciously placed burst disc or relief valve could practically eliminate the risk of misadventure without compromising the design of the shaft seals.

HYDROSTATIC TEST PRESSURE A casing, or the various regions of a dual pressure casing, is hydrostatically tested at Phydro = 1.5 Pdesign

(8)

Pdesign is either MAWPD for pre-engineered casings (see Equation 5), or the greater of MDP from Equations 6 or 7 for engineered casings.

Casings with the highest pressure ratings are radially split. There are actual examples of these designed for 14,000 psig. Axially split casings are preferred for large axis single stage double suction pumps and horizontal multistage pumps because of ease of maintenance. They cannot always be used, however, because their pressure rating is currently limited to 1,000 psig by API-610, 7th Edition, for hydrocarbon service.

This is due to be increased to 1,450 psig effective with the 8th Edition of API-610, although there are examples of their operation at 2,500 to 2,700 psig for boiler feed and water injection services. Higher pressures are possible, but the cost of the casing can become prohibitive, and the maintenance of the split joint gasket a major concern. Single stage, single suction pumps are generally radially split – for economics in the smaller sizes, both horizontal and vertical, and for ease of maintenance in the very large vertical axis designs. When a methodical approach is taken, the issue of pump pressure ratings is not too difficult. This evaluation and selection process can be further simplified by following the evaluation procedure outlined in the box accompanying this article. ■

REFERENCES: [1] API-610, 7th Edition, Standard for Centrifugal Pumps for General Refinery Service, American Petroleum Institute, Washington, DC, 1989. [2] ASME Boiler and Pressure Vessel Code, Section VIII, Divisions 1

Evaluation Procedure Answering the following questions in sequence will help avoid errors: Is the preferred form of casing joint suitable for the intended pressure, temperature and liquid specific gravity? Is the maximum discharge pressure, calculated as required by the applicable industry standard, less than the casing’s MAWP at the intended service temperature? Is there a minimum industry or code pressure rating for this class of pump? Do the casing flanges or nozzles have a pressure rating at least equal to that of the casing region to which they are connected? Is the minimum design pressure of the casing determined by the pressure developed during shop testing with water? Does the purchaser’s specification require all regions of multistage pump casings to have a MAWP equal to or greater than the maximum discharge pressure?

REVIEWING THE OPTIONS The Pump Handbook Series

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CENTRIFUGAL PUMPS HANDBOOK

Communicating Your Pump Needs Purchasing and installing a new pump requires a team effort between customer and supplier. By John Bertucci any industry pump problems are not caused by improper operation or faulty maintenance, although these are the primary focus of most reliability improvement efforts. In reality, most problems are traceable to improper initial application or changed operating conditions. The pump is just not right for the job. Or it once was, but is no more. Misapplication can be avoided, however, by observing three important principles: communication, communication and communication. In the past, and sometimes even today, pump users assume an adversarial relationship with their suppliers. Mutual suspicion is the order of the day, with each side trying to “win” or gain an advantage over the other. On the other hand, wise pump users are treating suppliers as valuable resources. They seek win-win outcomes where everyone benefits. Pump manufacturers, too, have found that making today’s sale is not as important as building a long term relationship with customers. Out of this industry-wide shift in attitude has come a new opportunity for suppliers and users to work together to put the best pump in a given application. With all of today’s corporate downsizing and re-engineering, pump users cannot afford to ignore the wealth of help available from pump suppliers. After all, suppliers can draw from a broad spectrum of industry experience to help a user solve a particular problem. This information exchange benefits users by taking solutions developed in one industry segment and introducing it

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to other sectors. Pump suppliers benefit by getting valuable feedback on their designs so they can improve their products and broaden their applications. And as this article explains, communication is the key to making it happen. There are two primary areas of pump user/supplier interaction. The first is when the user initially purchases a new pump. The second is when a user attempts to improve the performance of an existing pump. These two activities are different yet have much in common. The most important of these is the need for clear, open and honest communication between users and suppliers.

NEW PUMP PURCHASES Purchasing a new pump requires a team effort between customer and supplier. Much effort has been expended over the years to develop specifications and industry standards such as API and ANSI. These standards form the foundation of a new pump purchase. Following are some additional ideas that will help you select and purchase the best pump for a given application.

PROCESS DATA Accurate process data is needed to achieve a successful pump application. Without good information, the pump supplier is already fighting with one hand tied behind his back, and the battle hasn’t even started. A whole book could be written on the subject of properly sizing pumps, but the calculation and engineering aspects are beyond the scope of this article. Good communication is the key to getting accurate process data onto The Pump Handbook Series

the pump data sheet. Typically, the flow rate is established by the process for which the pump is being selected. A calculation then determines the discharge pressure required to move this amount of flow through the piping system. This process is fairly straightforward, but pitfalls exist even at this elementary stage. Many process design engineers do not understand that a typical centrifugal pump has an operating range of only 40–120% of Best Efficiency Point (BEP) flow rate. This may lead them to set the rated flow rate of the pump higher than required, possibly to allow for future expansion of the unit. Unfortunately, this future expansion either is many years away or never happens. As a consequence, the process unit is left with a pump that is operating at, or at much less than, the lower limit of its operating range. This low flow operation is the root cause of many pump reliability problems. Good communication during the sizing process can help avoid this and other sizing errors. The plant’s rotating machinery group should be brought in early to work with the process designers. Working together, they can explore options other than pump oversizing. The rotating equipment group also can bring pump manufacturer expertise to the sizing exercise. Pump manufacturers frequently offer other choices such as an upgradable pump, variable speed drive or other means to improve operational flexibility.

SITE/INSTALLATION DATA It is important to communicate the location and type of installation both internally and to the pump sup-

INFORMATION ITEMS Suction Specific Speed

POTENTIAL USES 1. Determining Stable Flow Range. 2. Re-rates, especially to lower rates.

Number of Impeller Vanes

Vibration Analysis, especially vibrations caused by low flow.

Seal Flush Flow Rate Calculations and Stuffing Box Pressure

1. Determine the cause of seal failures. 2. Change seal design.

Stable Flow Range

1. Determine the minimum or maximum allowable flow. 2. Determine if pump is in a low or high flow condition.

Thermal Growth

1. Determine if pump is distorting due to thermal forces. 2. Aid in getting good alignment.

Table 1 INFORMATION ITEMS Wear Ring Clearances

POTENTIAL USES 1. Checking existing clearances to determine if wear ring replacement is necessary. 2. Determine if performance problem was due to excessive wear ring clearance.

Bearing Number (Rolling Element Bearings)

1. Vibration Analysis - determine ball pass frequencies. 2. Analyze potential bearing upgrades.

Bearing Clearance (Hydrodynamic Bearings)

1. Analyze vibration problems. 2. Set alarm limits on probe type vibration monitors. 3. Determine bearing replacement needs.

Stuffing Box Dimensions

Mechanical seal upgrades and changes.

Materials of Construction

1. Determine repair methods. 2. Emergency fabrication of replacement parts.

Table 2 DOCUMENT Installation, Operation & Maintenance Manual Cross Sectional Drawing with Parts Identified Dimensional Outline Drawing Spare Parts List Performance Curve Curve Family Completed Data Sheet Test Data (if applicable) Driver Data

REVIEW BEFORE PURCHASE NO YES YES YES YES YES YES NO YES

COPY AFTER PURCHASE YES YES YES YES YES NO YES YES YES

Table 3

plier. Teamwork in this area is especially vital in vertical pit pump installations such as cooling tower pumps. Errors in designing the pit and suction approach to the pumps are easy to correct when the pit consists of lines on paper. But they are extremely expensive to correct once concrete is poured.

USER RESTRICTIONS AND PREFERENCES Every plant has certain things such as seal type, maximum suction specific speed and bearing type that they like and dislike in their pumps. These preferences and restrictions are usually based on years of experiThe Pump Handbook Series

ence in solving that plant’s particular pump problems. They must be explained well internally so that the pump data sheet accurately conveys them to the suppliers. Unfortunately, these preferences usually are in the heads of the plant rotating equipment group while a lot of the pump selection is done by a project engineering group or outside engineering design firm. One way to assure that these preferences are given due consideration is to require review of the pump data sheet by the rotating equipment group. Another good way is to place a member of the rotating equipment group on the project design team where his/her knowledge can be tapped by the project design engineers. A third way of accomplishing this is to develop a local specification that contains the various preferences and restrictions from the rotating equipment group.

COMMUNICATION WITH THE PUMP SUPPLIER Communication with the pump supplier should be a two-way street. The information should flow freely back and forth between the user and supplier. This should happen even in situations where competitive bidding will determine the ultimate supplier of a new pump. The only difference in the competitive bid situation is that all of the data sent from the user to the supplier should go to all suppliers equally, with no favoritism shown. Of course, all commercial information (prices, delivery details, etc.) should be kept confidential.

INFORMATION OBTAINED FROM PUMP SUPPLIER Much information can be obtained from the pump supplier that will aid in future pump maintenance and troubleshooting. Some of the less obvious or often forgotten items are listed here.

TROUBLESHOOTING AND MAINTENANCE Table 1 shows information that can be obtained from the pump’s manufacturer for future use in troubleshooting. This is information that the manufacturer generally has and will provide upon request. Table 2 lists pump supplier information that can be very useful in future pump maintenance.

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OTHER DOCUMENTATION Table 3 shows what can be considered a good minimum requirement for a documentation package.

PERFORMANCE IMPROVEMENTS Pump performance improvements generally come in two flavors: upgrade of an existing pump that is performing well, and correction of a problem pump. In either case, the pump’s original supplier can be an extremely valuable resource. Be Open to New Ideas. When your friendly pump supplier calls and asks for an appointment, make time for him (even if it’s not lunch time). Suppliers are constantly coming up with new and better ways to do things. Examples include A and B gap modifications, new overlay materials for severe service and new impeller designs. You can’t consider these and other potential solutions to your problems if you don’t take time to learn about them from the experts. Invite Your Supplier to Participate. The manufacturer’s local representative should be a regular part of your maintenance resources. You may have a problem that your competition solved years ago. You’ll never find this out from the competition, but the pump supplier may already know the solution. However,

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he can’t help if he doesn’t know about your problem. In addition, the pump supplier will be intimately familiar with the design issues of a specific pump. Visit Supplier’s Repair Facility or Factory. A visit to the supplier’s repair shop can give you much valuable information: It enables you to evaluate the shop’s capabilities, in case you ever need them. It allows you to develop face-toface contacts with the people who repair these pumps every day. A good relationship with these folks may help you in the future when you need information. Also, the information you get there is not filtered through a salesman. A visit to the factory can be even better. It allows you to meet with the people who know the most about the design of your pump. Factory contacts can get you information in a hurry. They can also help expedite shipment of desperately needed parts. Go to Conferences. There are many good conferences and symposia that have pumps as a primary theme. One highly recommended conference is the International Pump Users Symposium held in early March each year in Houston. Sponsored by Texas A&M University’s Turbomachinery Laboratory, it includes short courses, tutorials, discussion groups and hun-

The Pump Handbook Series

dreds of pump manufacturer and related industry displays. These conferences are like having many pump factories under one roof. You can go from booth to booth and meet representatives of many companies. The discussion groups are also a good way to interface with other users as well as suppliers. Share Knowledge/Experience with the Pump Supplier. Users accrue valuable practical experience with pumps over their years of operation. This experience should be shared with the pump supplier. It will help the supplier improve his products and eventually benefit all users. Also, it is only fair that knowledge go both ways in any relationship. In summary, your pump supplier will never know or understand your needs if you don’t take the time and effort to develop a mutually beneficial relationship. Most pump suppliers place a high priority on meeting their customers’ needs, but they need all the help that they can get. ■ John Bertucci is a Mechanical Equipment Engineer with the Norco Manufacturing complex of Shell Oil Products Co.

CENTRIFUGAL PUMPS HANDBOOK

Impellers and Volutes: Power with Control A review of centrifugal pump impeller and volute design applications can help you optimize both power and control. By Robert R. Ross the volute throat.

ll centrifugal pumps have two major components: the rotating element and the pump case. Together they establish how much head will be generated, best efficiency point (BEP) capacity, the slope of the head capacity curve and net positive suction head required (NPSHR), Figure 1. The rotating element consists of a shaft and one or more impellers whose function is to convert mechanical energy into high velocity kinetic energy. The pump case directs liquid to the impeller from the suction nozzle, collects the liquid discharging from the impeller, and then converts the kinetic energy into pressure by controlled deceleration in the diffusion chamber immediately following

A

TYPES OF VOLUTES Single volute pumps have only one volute throat and one diffusion chamber, and because of their simple casting geometry, they can be produced at lower cost than the more complex double volute designs. Figure 2 shows a single stage version with the diffusion leading directly to pump discharge. Figure 3 is a multistage version in which the liquid, following diffusion, is directed through the crossover to the next stage impeller. Pressure distribution around the impeller is non-uniform and produces a radial load on the impeller which, depending on the developed head, may deflect the pump shaft and cause wear at impeller wear rings and seal faces. Single volutes are used routinely

FIGURE 1

HR

NPS

40

HEAD-CAP

30

90 80

450

70 B.E.P.

F. %

400

E. F.

350

60

300

R. 1.0

250

3HP SP. G

200

200

100

150 100 0

50

400

300

BRAKE H.P.

HEAD IN FEET

500

0 400

800

Pump performance curve

1200 1600 2000 CAPACITY GPM

2400

2800

THE PUMP HANDBOOK SERIES

40 30

EFFICIENCY %

550

NPSHR FT

60 50

in slurry and sewage pumps to minimize plugging at the throat, and on low head pumps where radial loads are nominal. They are also used on low specific speed pumps where the throat area and hydraulic passages are too small to cast as a double volute. Double volute casings were introduced to minimize the radial thrust problems of single volute pumps. They are actually two single volute designs 180° apart with a total throat area equivalent to a comparable single volute design. The nonuniform pressure distributions are opposed, thereby greatly reducing radial loads (Figure 4). Double volute pumps are the preferred choice on higher head pumps.

TYPES OF IMPELLERS Pumps can be built with a single impeller or, in the case of high pressure applications, with two or more impellers. They will be either single entry or double entry type, more commonly known as single or double suction (Figure 5). Because double suction impellers have a greater total eye area, velocity of the liquid entering the eye is reduced, producing a lower NPSHR. The shape of the impeller depends on specific speed (Ns), which should only be calculated at BEP with maximum diameter impeller. In U.S. units this is: Ns=RPM x GPM.5 (Head/Stage feet).75

20 10 0

Figure 6 shows the change in shape from the low Ns radial flow impellers to the high N s axial flow types. Whereas the suction geometry is selected to reduce inlet losses for low

139

NPSHR, the discharge geometry is selected to satisfy the required head, slope of the head capacity curve and BEP capacity.

FIGURE 2 Diffusion Chamber

HEAD CAPACITY CURVE SLOPE ANALYSIS

Impeller

Volute Throat

FIGURE 3

FIGURE 4

The desired slope is determined during system analysis with the percentage rise to shutoff (zero flow) from rated head often determined by system limits. When the pump is started against a closed discharge valve, the pressure up to the valve will be the pump differential head at shutoff plus suction pressure. Because this should not exceed the safe working pressure of the system, the rise to shutoff can be critical and is controlled by impeller discharge geometry. An evaluation of the system head curve is needed to determine if the pump should have a flat or a steep curve. This is a graphical plot of the total static head and friction losses for various flow rates. For any desired flow rate, the head to be generated by the pump is at the intersection of the head capacity curve and system head curve. In a simple pump application in which system head is due entirely to friction loss, a flat head capacity curve with 10 to 20% rise to shutoff from the head at rated capacity would satisfy the application and minimize shutoff pressure on the system (Figure 7). Where system head consists of both friction and static head — that is, where there is a change in elevation — a flat curve also would be appropriate if little or

no change was anticipated in the static head (Figure 8). Figure 9 represents an application in which changes in static head caused by changes in the suction tank level result in a range of system head curves. For illustration, two pump head capacity curves have been superimposed — one flat, the other steep. The advantages and disadvantages of both must be considered in deciding which pump curve is more suitable for the application. Advantages of the flat curve are low shutoff pressure and relatively small differences in operating pressure as the system head moves. The disadvantage is a larger variation in flow rates. Advantages of the steep curve are smaller variations in flow rates and additional head margin to accommodate potential increases in static head. Disadvantages are high shutoff pressure and larger variations in head.

IMPELLER DISCHARGE GEOMETRY Various methods are used to modify the impeller so the percentage rise to shutoff will match the slope resulting from system head analysis. Among these are changes in the number of vanes, changes in the vane discharge angle and changes in the exit width b2. The curve can be changed with variations in vane number and discharge angle, while BEP capacity is held constant by changing b2. Another method is to use a constant discharge angle and b2 with changes in the number of vanes only.

FIGURE 6 Radial Flow

Semi-Radial Flow

Mixed Flow

Axial Flow

3000

6000

12000

FIGURE 5

SINGLE SUCTION IMPELLER

DOUBLE SUCTION IMPELLER

1000

SPECIFIC SPEED Ns

Centrifugal impellers

140

The Pump Handbook Series

FIGURE 9

FIGURE 7 SYSTEM HEAD CURVE, ALL FRICTION

EVALUATING PUMP CURVES AGAINST VARIABLE SYSTEM HEAD CURVES

Capacity

Head

S

NO

SU

T UC

RM

CT

IO

N

AL

IO

N

T

T

K AN

LO

W

O TI RA E O P FL AT

K AN

FU

LL

N

ST EE P

FIGURE 8 SYSTEM HEAD CURVE, STATIC HEAD PLUS FRICTION

Capacity

Robert R. Ross is Director of Engineering for BW/IP International and a member of the Pumps and Systems Editorial Advisory Team.

Capacity

BEP CAPACITY The impeller discharge geometry and volute throat area establish BEP capacity. By adjusting the ratio of liquid velocity leaving the impeller to liquid velocity entering the volute throat,

BEP can be increased or decreased. Modifications of this type are used to upsize or downsize existing pumps hydraulically, moving BEP to the normal operating capacity for optimum efficiency and generating significant savings in the cost of power. ■

THE PUMP HANDBOOK SERIES

Editor’s Note: Some text and figures for this article have been excerpted with publisher’s permission from Lobanoff, V. S. and Ross, R.R., Centrifugal Pumps: Design and Application, 2nd Edition, Gulf Publishing Company, Houston, TX 1992.

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CENTRIFUGAL PUMPS HANDBOOK

Fully Lined Slurry Pump for FCCU Bottoms Use By replacing conventional API pumps with fully lined slurry pumps in FCCU bottoms applications, refineries are improving production and profitability. By Dan Clark and Julio R. Cayro he Fluid Catalytic Cracking Unit (FCCU) is one of several processes critical to a refinery’s productivity. Its long term, safe operation translates into increased production and profitability. Yet prior to introducing fully lined pump technology for use in FCC (Fluid Catalytic Cracking) main column bottoms applications, refineries replaced or repaired conventional double volute API process pumps several times a year. This caused serious safety risks and a process shutdown frequency unacceptable in today’s production environment. The fully lined slurry pump has thus emerged as the pump technology of choice for providing 3 – 5 years of maintenance-free operation for FCCU refinery bottoms applications. This particular application involves pumping a highly erosive high temperature (350 – 800°F) slurry at flows to 12,000 gpm, pressures to 600 psig and heads from 90 – 900 feet. Catalysts used in FCC processes are also extremely erosive, and they are applied in varying concentrations depending on the process unit configuration and/or upset operating conditions. The fully lined slurry pump design (Figure 1) is an engineered approach to providing long term, reliable pump performance in this severeduty application. To understand how this new pump technology meets such demanding requirements, we will take a closer look at FCC processes and the design of fully lined slurry pumps.

T

FCCU PROCESSES 142

FIGURE 1

Fully lined slurry pump design Conventional Fluid Catalytic Cracking. A typical fluid catalytic cracking unit (Figure 2) consists of a reactor, catalyst regenerator and fractionator column. This process converts straight run heavy gas oil from the crude distilling unit, and flasher tops from the vacuum flasher unit, into high octane gasoline, light fuel oils and olefin-rich light gases. In the vertical reactor vessel, vaporized oil contacts fluidized catalyst particles, causing a reaction that yields lighter hydrocarbon and coke. During the reaction, carbon (coke) is deposited on the catalyst, rendering it The Pump Handbook Series

inactive. This inactive catalyst is recirculated from the cyclones at the top of the reactor back to the regenerator where the coke is combusted, rejuvenating the catalyst. Vaporized cracked products flow through the cyclones at the top of the reactor into the vapor line that feeds the bottom of the main fractionator column. The cyclones operate at less than 100% efficiency so that some coke and catalyst particles continuously reach the fractionator. In this type of FCC process, the main column bottoms pumps must pump the bottom oil at a high rate through the heat exchanger and over

FULLY LINED SLURRY PUMP TECHNOLOGY Unlike conventional API process pumps designed for maximum efficiency with clean liquids, fully lined slurry pumps are engineered to provide maximum reliability when handling abrasive hydrocarbon slurries. All pumps are selected to operate in the optimum hydraulic fit (80 – 110% BEP) for specified flow, head and erosive characteristics of the slurry. Design considerations such as using larger diameter impellers at lower speeds (870 – 1,770 rpm), selecting proper construction materials and maximizing appropriate mechanical seal designs all help optimize life cycle cost. Liners. The fully lined slurry pump uses replaceable, abrasionresistant 28% chrome iron liners to protect the pressure casing, providing 5 – 6 times the life of diffusion coated CA6nm components (Figure 3). Abrasion-resistant liners are machined, toleranced components that form the hydraulic wet end of the

PHOTO 1

PHOTO COURTESY OF LAWRENCE PUMPS

a vapor contact section within the fractionator tower. This desuperheats and scrubs the fine particles of catalyst from the reactor vapors entering the fractionator without causing oil coking. A small concentration of alumina-based catalyst particles that have been scrubbed out of the vapors is continuously circulated with the main bottoms product, causing erosion of the pump internals. New FCC Processes. Unlike conventional FCC processes that recycle part of the main column bottoms directly back to the reactor vessel, newer FCC processes such as the UOP process continually circulate the main column bottoms in a closed loop through heat exchangers and back to the fractionator tower. With the closed loop design, catalyst concentrations run as high as 2% by weight, compared to conventional systems where the catalyst concentration is 0.25 – 0.5%. With conventional FCC processes, catalyst levels reached 1 – 1.5% only during prolonged upset operating conditions, destroying conventional API process pump internals in a matter of days or weeks (Photo 1). Today, a 2% concentration of catalyst continuously circulating with the main column bottoms is considered a normal operating condition.

Conventional API process pump internals can be destroyed in a matter of days during prolonged upset operating conditions.

FIGURE 2

UOP “Straight-Riser” Fluid Catalytic Cracking Unit GAS & GASOLINE TO GAS-CONCENTRATION PLANT

FLUE GAS TO CO BOILER

REACTOR PRESSURE REDUCING CHAMBER

LIGHT-CYCLE GAS OIL

CATALYST STRIPPER FRACTIONATOR

HEAVY-CYCLE GAS OIL

STEAM

CLARIFIED OIL REGENERATOR

SLURRY SETTLER

AIR

CHARGE

pump. They are easily replaced individually or as a set. Flow stream turbulence is reduced by using a 125 rms machine finish on the liner surfaces. Rotating Elements. Like the liners, impellers are constructed of abrasion resistant 28% chrome iron. These high efficiency enclosed impellers (Photo 2) use front and back repelling vanes to reduce slurry recirculation. Repelling vanes eliminate the clean oil flush required by The Pump Handbook Series

conventional wear ring impeller designs. Large open passages reduce frictional losses and allow maximum solids-passing capability to avoid clogging due to coke buildup in the fractionator column. Several impeller mounting configurations are available depending on horsepower and catalyst slurry properties. Impellers can be fastened to the shaft using a tapered polygon or straight bore. Each is locked in

143

MECHANICAL SEAL CONFIGURATIONS AND OPTIONS All fully lined slurry pumps operating in FCCU bottoms applications feature a removable seal chamber that allows pressure testing of the seal before installation. This design gives the user the ability to switch seal types easily to accommodate catalyst slurry changes and/or to comply with environmental regulations. Seal chambers are designed to accommodate single, tandem and double mechanical seal arrangements. A replaceable throat bushing mechanically fastened to the seal chamber rides on the impeller hub. The shaft sleeve is 316 stainless steel, and its straight design allows easy adjustment of impeller clearances without disturbing the seal setting. A Ringfeder locking collar replaces the conventional locking collar with set screws. The Ringfeder eliminates

144

FIGURE 3

500 450 400 350 300 COATED API PUMP

250 200 150 100

FULLY LINED API PUMP

50 0 0

6

12

18

24

30

the need for set screws to drive the shaft sleeve. It also seals the cavity between the shaft and sleeve by compressing the sleeve onto the shaft. This clamping arrangement prevents slippage of the sleeve on the shaft under severe slurry upset conditions, and it eliminates the need for grafoil packing by providing

36

42

48

54

60

66

72

a high temperature, high pressure metal-to-metal seal between the shaft and sleeve.

STARTUP AND OPERATION ISSUES The startup procedure for fully lined slurry pumps operating in the FCCU bottoms application is critical to the pump’s long term performance.

PHOTO 2

PHOTO COURTESY OF LAWRENCE PUMPS

place with an enclosed impeller nut and then secured with a locking bolt. As shown in Figure 1, fully lined slurry pumps use large diameter shafts that meet the stiffness criteria set by API-610. Each shaft is engineered to provide space for single, double or tandem seal arrangements and a conservative L3/D4 deflection index. Stiff shaft designs limit deflection, maximize mechanical seal and bearing life, and minimize vibration. A heavy duty bearing assembly employs 7300 Series bearings with slight preloads to support the shaft. The anti-friction bearings provide a minimum L-10 rated life of 100,000 hours at the rated pump condition. Thrust bearings are duplex, angular contact type mounted back to back. The radial bearing is either the antifriction ball or spherical roller arrangement, depending on radial loads and rotative speeds. Bearings are lubricated using ring oil, flood oil, oil mist and forced feed arrangements. Oil mist is used to minimize friction, but it is not acceptable for cooling bearings that are subject to heat transferred from an external source such as the shaft. Thrust and radial bearing covers are equipped with isolators that have a deep grooved labyrinth which prevents oil from escaping from the bearing frame.

Enclosed impeller with front and back repelling vanes to reduce circulation of the slurry. The Pump Handbook Series

To avoid thermally shocking the hard metal liners and impeller, hot oil is injected into the pump casing to preheat these internal components gradually — at a rate not exceeding 150°F per hour. Initially, the oil steam is introduced at less than 250°F through either the casing drain or seal flush connection at a pressure higher than the downstream discharge pressure.. The steam is then allowed to flow through the casing. Operators are asked to maintain a constant preheat rate until the pump is heated to within 150°F of its actual operating temperature and allowed to soak for one hour at the maximum preheat temperature before startup. Providing the proper cooling to the bearing frames and pedestals, and flush oil to the mechanical seals, becomes very important once the pump reaches its startup temperature. To ensure long term successful operation of a fully lined slurry pump in the FCCU bottoms application, the following steps are recommended. • Provide dual strainers in front of the pump to allow cleaning the strainer without shutting down the pump. • Use slurry impeller designs to allow for larger mesh openings in the suction strainers and increase the cycle time between cleaning by 300%. • Equip the FCCU system with a minimum flow bypass line so the

pump can be operated continuously at its BEP. • Use mechanical seals in the cartridge canister arrangement to allow performance testing during seal design stages and hydrotesting before installing the seal. • Consider using double seal configurations for added safety and provide for inboard seal faces to run continually on a clean liquid.

MAINTENANCE ISSUES Fully lined slurry pumps operating in the FCCU bottoms application require maintenance, repair or replacement of the casing liners and impeller every 3 – 5 years. This maintenance is simplified by a back pull-out arrangement allowing access to the liners and impeller without disturbing suction and discharge piping. After assembly, impeller clearances can be adjusted easily by moving the thrust bearing cartridge relative to the bearing frame. Shims are then placed between the cartridge and bearing frame to lock the shaft and impeller into position. The large oil reservoir for the bearings ensures a continuous source of clean oil to lubricate the bearings. Oil level is monitored using a 2” bulls eye in the side of the bearing frame. The entire pump can be rebuilt by mechanical craftsmen using standard shop tools. All wear components are replaceable, and all fits and clearances are standard.

The Pump Handbook Series

LONG TERM PERFORMANCE Several hundred fully lined slurry pumps are operating in FCCU bottoms applications in refineries around the world. More specifically, a fully lined slurry pump installed in a UOP FCCU unit has been running continuously since 1991, requiring only scheduled maintenance. Fully lined slurry pump technologies continue to evolve to meet the increasingly demanding performance criteria of the FCCU bottoms application. Refineries that have installed fully lined slurry pumps in their cracking units have eliminated the expense of replacing conventional API process pumps several times a year. More important, they’ve eliminated the safety risk of catastrophic pump failure and the prohibitive cost of shutting down the FCCU for several days. ■ Dan Clark has been with Lawrence Pumps Inc. since 1973 and has overseen the development of the FCCU bottoms fully lined slurry pump. Julio R. Cayro is a Mechanical Equipment Consultant and owner of Cayro Engineering Company in Houston, TX.

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CENTRIFUGAL PUMPS HANDBOOK

Options for Sealless Centrifugals

I

1. DOUBLE CONTAINMENT This is an important consideration in services that pose extreme health and/or safety concerns. Several manufacturers of canned motor pumps (CMP’s) offer doublecontainment (i.e. welded primary containment, hermetically sealed secondary containment). Because of efficiency issues, few magnetic drive pump (MDP’s) manufacturers have approached secondary containment. Most MDP’s can provide secondary “control” by utilizing mechanical seals on the OMR shaft penetration. Additionally, an MDP provides a thicker containment shell, and thus more resistance to penetration by corrosive or mechanical failure. Typical CMP primary liner thicknesses are 0.022–0.35″, while an MDP containment shell is 0.029–0.060″. Bearing monitoring plays a role, however, because bearing and internal rotor positions are easier to monitor in CMP’s, by design, than in MDP’s. This enables the CMP to detect extreme bearing wear prior to containment shell contact or violation.

2. SOLIDS/SLURRY HANDLING Both the canned motor and magnetic drive designs will handle moderate amounts of solids, and optional designs for both will handle higher concentrations of slurries. For CMP’s, if low concentrations of

146

solids are present (about 2%) and their size is relatively small (25µm maximum), hard bearings running against a hardened journal work well. CMP’s can handle increased solids if they are outfitted with external flush or filters to remove particulates from the pumpage before they circulate around the bearings. In fact, CMP’s with a slurry modification are able to handle slurries in the concentration ranges handled by a standard centrifugal pump. An MDP cannot be isolated as easily as the CMP, and therefore would require a completely different internal bearing-mag coupling flow path than is conventionally offered.

3. HEAT INPUT Both sealless centrifugal pump designs add heat by hydraulic and drive inefficiencies. For a CMP design a “high efficiency” motor will be 80–85% efficient. But because of the ease in isolating the motor area, CMP’s can offer optional configurations to control the fluid temperature, pressure, or both, to prevent product vaporization. However, “heat soak” can occur. In this situation the process fluid will be heated to higher temperatures at potentially lower pressures than during normal operation. This is because a CMP motor tends to be a large insulated mass once the unit is shut down. This may result in flashing of the contained fluid with a potential of vapor locking if restarted.

and must be evaluated on a case-bycase basis. Several things must be considered on MDP designs. A synchronous drive is more efficient than an eddy current drive. A non-metallic containment shell is more efficient than a metallic shell. The MDP lacks the large insulated mass around the containment shell and is less susceptible to “heat soak.”

5. JACKETING Either design can accept steam or hot oil jackets added to the pump to maintain the proper temperature of the product with a high freezing point. This insures that the pumpage remains a liquid during pump operation and shut-down. CMP designs allow jacketing of the pump case, stator and rear bearing housing. Users must be careful not to exceed the thermal limits of the stator insulation with the heating media. Most MDP’s can jacket the casing and add some heat in the area of the containment shell without fully encapsulating it.

6. HIGH SUCTION PRESSURE Typically, the CMP is more efficient than the MDP in high pressure applications because of the increased primary containment thickness. A canned motor pump is a pressure vessel since the stator windings lend additional mechanical strength. For applications in which the suction pressure and maximum allowable working pressure require-

4. COOLING REQUIREMENTS Both the CMP and MDP can be configured for operation at high temperatures. However, permanent magnets can tolerate heat better than motor windings can, so MDP’s are able to pump hot liquids—up to 750°F—with just air cooling. Canned motor pumps require water cooling jackets for high temperature service. For example, some CMP designs can operate up to 1,000°F with water cooling. Various designs require temperature limitations by component The Pump Handbook Series

PHOTO COURTESY SUNSTRAND FLUID HANDLING

n recent years, the case for using sealless centrifugal pumps has centered mainly on zero emissions—and the fact that they do not require seal support systems and periodic mechanical seal replacement. But this is only part of the story. Design modifications and accessories are expanding the performance ranges of both canned motor and magnetic drive pumps. Here is a generic comparison of what these pumps offer when it comes to hydraulic application features and options:

Cutaway of a Kontro A-range ANSI sealless magnet drive pump used in chemical processing services.

ments exceed the standard pressure design capability of either a CMP or MDP, optional designs are available for both. In fact, some CMP’s are available that can have as high as 5,000 psi system pressure design. Modifications in the CMP design for high pressure applications include the use of primary containment shell backing rings, thicker secondary containment shells, additional pressure-containing bolting and high pressure terminal plates. For MDP’s high pressure application modifications include usage of a thicker containment shell and additional bolting. In addition to application features, users should be aware of certain hydraulic and design differences between canned motor and magnetic drive pumps. Here again, both offer advantages.

7. ANSI While a few manufacturers of canned motor pumps offer units with ANSI dimensions and/or hydraulics, most do not. On the other hand, the majority of mag drive pump suppliers do offer ANSI dimensions and hydraulics. In addition, most manufacturers of sealed and sealless ANSI pumps offer interchangeability between their pumps’ wet ends and bearing frames.

cient in low flow/high head hydraulics. A Francis design impeller (enclosed with backswept vanes and an increasing-radius volute) is more efficient at moderate to high flows with low to medium heads. You need to evaluate “wire-towater” efficiency to get a truly accurate picture of efficiency.

toring features vary by manufacturer and care must be exercised when selecting a sealless pump vendor to insure that the desired monitoring features are provided.

11. SPACE CONSIDERATIONS In general, a CMP (integral pump and motor) occupies less of a foot print than a comparable MDP (pump, coupling and motor). However, closecoupled MDP’s are available, and these may require the same or less space than a CMP.

9. INTERNAL CLEARANCES While clearances between bearing ID and mating surfaces are typically the same (0.003–0.007″), differences occur between other rotating parts. Typical CMP clearances between the rotor and stator liners vary by manufacturer between 0.018–0.044″ radially. Typical MDP clearances between the inner magnetic ring (MR) and containment shell range from 0.030–0.045″ radially. This larger clearance gives MDP’s the advantage of allowing more bearing wear to occur prior to containment shell contact.

12. COUPLING ALIGNMENT AND VENTING

10. BEARING MONITORING While different monitoring methods are available for both MDP’s and CMP’s, the latter design lends itself more to real bearing monitoring. A CMP bearing monitor can provide axial, radial and liner corrosive wear indications. However, bearing moni-

CMP wire-to-water efficiency is defined as hydraulic times motor efficiency. As mentioned in the section on heat input, a typical CMP motor will be 80–85% efficient. MDP wireto-water efficiency is defined as hydraulic efficiency times motor efficiency times coupling efficiency. Again, the magnetic coupling is 80–85% efficient. However, containment shell metallurgy (or lack thereof) and magnetic coupling type play a big role in coupling efficiency. Further complicating the efficiency discussion is wet end hardware. Impeller and casing geometry play a vital role in hydraulic efficiency. For example, a Barske design (open radial blade impeller and diffuser discharge) is usually more effi-

PHOTO COURTESY CRANE CO., CHEMPUMP DIVISION

8. EFFICIENCY

Cutaway view of a canned motor pump. These pumps are used on a wide variety of fluids at temperatures from cryogenic service to 1000° F and at system pressures up to 5,000 PSI. The Pump Handbook Series

Coupling alignment is not required for CMP installations because the pump impeller is directly mounted on the motor shaft inside of the containment area, so no coupling exists. Typical MDP installations utilize frame-mounted motors which require coupling alignment. Some MDP suppliers, however, offer close-coupled designs which eliminate coupling alignment. Both MDP and CMP designs are usually self-venting back into the process piping and do not require additional external lines. In addition to the issues already discussed, some manufacturers offer canned motor and magnetic drive centrifugal pumps with options such as the following: two-phase flow designs; tachometers; open impellers (nonshroud) with isolated motor sections for solids handling; diagnostics that indicate rotor position, stator liner rupture, temperature and pressure; vibration pads, and redundant systems to indicate breach and contain fluid. The best thing you can do if you’re considering options for your sealless centrifugal pumps is to make sure your supplier(s) know everything about your application, particularly temperatures and vapor pressure at startup and shutdown, not just normal operating conditions. ■ This article was developed with the assistance of Steven A. Jaskiewicz, of Crane Chempump (Warrington, PA) and David Carr, of Sundstrand Fluid Handling (Arvada, CO).

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CENTRIFUGAL PUMPS HANDBOOK

Tips for Selecting ANSI Process Pumps Versatility is the key to ANSI process pump applications. By Charles Cappellino and Richard Blong

148

entrained gases. Open impellers – particularly those running at 3600 rpm – must be carefully engineered to control axial thrust, seal chamber pressure and mechanical integrity. Closed impellers are typically employed in less corrosive environ-

PHOTO COURTESY OF GOULDS PUMPS, INC.

A

• ANSI/ASME B73.1M - 1991, Specification for Horizontal End Suction Chemical Pumps for Chemical Process (Photo 1) • ANSI/ASME B73.2M - 1991, Specification for Vertical In-Line Centrifugal Pumps for Chemical Process (Photo 2) • ANSI/ASME B73.3M - (in process), Specification for Sealless Horizontal End Suction Centrifugal Pumps for Chemical Process • ANSI/ASME B73.5M - (in process), Specification for Thermoplastic and Thermoset Polymer Material Horizontal End Suction Centrifugal Pumps for Chemical Process

PRINCIPLES OF OPERATION The pumps covered by ANSI/ ASME B73.1M are classified as end suction centrifugal pumps. Centrifugal pumps use an impeller and angled vanes to impart velocity to the liquid entering the pump. The liquid leaving the impeller is collected by the pump casing, an action that converts a portion of the fluid velocity into pressure. Either mechanical packing in a stuffing box or a mechanical seal in a seal chamber is used to seal the rotating shaft. Figure 1 shows the basic components of an ANSI pump. Impellers. Centrifugal impeller designs are of two basic types, the open style and the closed style, as shown in Figure 2. Most ANSI pumps employ some type of open impeller and axial adjustment feature. This allows critical operating clearances within the pump to be maintained, which is important for maximizing hydraulic performance. The open impeller is also better at handling solids and pumping liquids with The Pump Handbook Series

Photo 1. Example of a horizontal metal ANSI process pump

PHOTO COURTESY OF GOULDS PUMPS, INC.

revolution swept through the chemical process pump industry more than 30 years ago. It was the beginning of chemical pump design and dimension standardization. Before the 1960s, chemical process pump manufacturers offered a proliferation of designs. Each manufacturer had its own design and dimensional envelope. Industrial users faced significant piping, baseplate design and potential foundation changes if existing pumps had to be replaced. This very expensive possibility became the driving force behind the development of what industry today refers to as an "ANSI" pump. During the 1960s and 1970s, the American Voluntary Standard (AVS pump) served as the chemical process pump standard. In 1974, the American National Standards Institute (ANSI) used the AVS standard as the foundation for its B73.1 specification covering chemical process pumps. After several revisions today’s ANSI/ASME B73.1M - 1991 Specification for Horizontal End Suction Centrifugal Pumps for Chemical Process serves as the industry standard. It covers dimensional interchangeability requirements for 20 pump sizes. This includes mounting dimensions, suction and discharge flange size and location, input shaft size, baseplates and foundation bolt holes. It also addresses many mechanical design features such as pressure limits, temperature limits, drain and gauge connections and seal chamber dimensions. This enables today’s pump user to replace a pump with one from a different manufacturer easily. Chemical process pump specifications developed through the B73 committee include:

Photo 2. Example of vertical in-line ANSI centrifugal pump for chemical process service

ments such as light duty chemical, petrochemical and utility applications. This is because closed impellers utilize renewable wear rings to maintain performance-sensitive running clearances. Renewable wear rings are subject to crevice corrosion and are generally undesirable for corrosive services. Casings. Three basic types of centrifugal pump casing designs are used in chemical process pumps: the circular volute, single volute and double volute (Figure 3). The different casing designs are used to reduce hydraulic radial loads. Pumps designed to operate at extremely low flows normally use circular volutes to minimize hydraulic radial loads. Single volute casings are simple to manufacture and are the most commonly used for ANSI pumps. Larger pump sizes require a double volute to reduce hydraulic radial loads. The casings utilize ANSI/ASME B16.5 Class 150 or Class 300 flanges for suction and discharge connections.

Casings must withstand a hydrostatic pressure test of 1.5 times the maximum design pressure for the material used, and they must employ an 0.125" corrosion allowance by design. Seal Chambers. ANSI/ASME B73.1M provides dimensional guidelines for shaft sealing. The guidelines cover a stuffing box design used for packing, a large diameter cylindrical seal chamber and a self-venting seal chamber used for mechanical face seals. Large radial clearance between the shaft and the inside of the seal chamber is specified due to its importance to mechanical seal face temperatures. The large radial clearance also allows mechanical seal manufacturers to build more robust and reliable designs. Because mechanical seals are one of the most significant causes of pump downtime, most ANSI pump manufacturers have developed new seal chamber designs that enhance the operating environment for mechanical seals. Typical seal chambers offered by manufacturers are shown in Figure

Rotation

4. The most recent self-venting designs incorporate some type of flow modifying ribs or superior performance vanes to control solids and entrained gas. Most chemical pump manufacturers offer several seal chamber designs, as well as selection guidance for various services. Bearing Housings. ANSI/ ASME B73.1M requires a bearing selection that provides 17,500 hours of life for the radial and axial thrust bearings, calculated according to ANSI/AFBMA-9&11. This typically results in a double row thrust bearing being used at the coupling end of the shaft and a deep groove ball bearing at the impeller end. (Photo 3) The pump industry is steadily improving the reliability of pump components. To increase mean time between planned maintenance, many manufacturers have improved bearing housing designs and added features to improve reliability. Some ANSI pump manufacturers offer heavy-duty housings that use angular contact bearing pairs to handle higher hydraulic thrust loads. Most bearing housings are sealed using either lip

Open Impeller

Closed Impeller

Figure 2. Centrifugal impeller designs

Figure 1. Basic components of an ANSI B73 pump

Circular Volute

Single Volute

Double Volute

Figure 3. Circular, single and double volute casing designs The Pump Handbook Series

149

4A

4B

4C

4D

Figure 4. Seal chamber styles A. Standard bore (packed box) is characterized by long, narrow cross section. Originally designed for soft packing, mechanical seals were forced into cavity envelope. Requires an API/CPI flush plant for optimal performance. B. Enlarged bore features increased radial clearances over the standard bore. This chamber design enables optimal seal design. Restriction at bottom of the seal chamber limits fluid interchange. Requires an API/CPI flush plan for best performance. C-D. Tapered bore features increased radial clearances similar to the enlarged bore, except there is no restriction at the bottom of the cavity and is open to the impeller backside. Current designs include vanes or ribs to provide solids and entrained gas handling. Flushing is often not required as design promotes cooler running seals by providing increased circulation over faces.

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seals or higher performance labyrinth seals to prevent lubrication contamination– the number one cause of premature bearing failure. Many manufacturers also have increased the capacity of oil sumps to provide superior heat transfer and cooler running bearings. Most housings have some type of finned cooler that is submerged in the oil sump. This controls oil temperature in hot services. And large diameter (1") sight glasses are incorporated into many designs to provide a means of viewing oil condition and level. Baseplate Designs. ANSI/ASME B73.1M specifies a set of baseplate dimensions covering motor sizes typically required for the full range of ANSI pump sizes. Dimensions specified include the pump and motor mounting surfaces, bedplate footprint and foundation bolt hole locations. Proper baseplate design and installation are necessary to maintain accurate pump and driver alignment. This lengthens the life of bearings and seals, which are sensitive to vibration and correct alignment. Proper baseplate selection is the key to maximizing mean time between planned maintenance. Most pump manufacturers offer a selection of baseplate designs. Camber top cast iron baseplates offer heavy-duty construction with machined pump and motor pads. They also have good vibration damping characteristics. Fabricated steel baseplates provide an economical choice in carbon steel and the option of various metallurgies such as stainless steel. Many ANSI pump manufacturers also offer some type of nonmetallic composite baseplate (such as fiberglass reinforced plastic FRP) for superior corrosion resistance. FRP bases are used with FRP pumps as well as high alloy pumps. A recent addition to the bedplate options is a heavyduty fabricated baseplate with integral adjustment features such as adjustment screws and baseplate leveling screws (Photo 4). Finally, most bedplates can be stilt or spring mounted. These supports raise a pump above the floor for improved cleaning access, and they accommodate piping thermal expansion. Stilt and spring-mounted designs must be carefully engineered to ensure proper rigidity. This will maintain alignment and avoid vibration problems.

CHEMICAL PROCESS PUMP SELECTION The Pump Handbook Series

The ANSI/ASME chemical process pump is the most widely used centrifugal pump in industry. Its wide use is attributed to its adaptability to a wide range of process service pumping conditions. It can be mounted vertically to save installation space, or vertically suspended for use in a sump application. Flexibility in casing and impeller designs enable it to handle extremely low flows, pump solids, move highly corrosive liquids, self-prime or withstand temperatures to 700ºF (371ºC). Recently, the same basic design has been made sealless by eliminating the need for packing or mechanical seals to seal liquid in the pump. These pumps are magnetic drive ANSI/ ASME chemical process pumps. ANSI/ASME chemical process pumps are perhaps the most versatile pumps in the world. To help in applying the various types of chemical process pumps, a selection guide is shown in Figure 5. The pump type is shown vertically, and the service parameter is listed horizontally. ANSI chemical pump solutions are available for nearly any pumping service parameter.

Photo 3. Bearing housing configurations

Photo 4. Enhanced fabricated steel baseplate

Pumpage Corrosives

Pump Type

Moderate

Severe

Pumping Conditions

Solids Non Abras., Fibrous Stringy

Hazardous

(Noxious Explosive Volatile Abrasive Toxic)

Capacity

Low Flow

High Cap.

Installation Considerations

Materials of Construction

Temperature

High Press.

Cryogenic

0-500F.

Non-Metallic

Sumps

Limited Floor Space

ANSI Dim.

No Align. Req’d

PFA Teflon

FRP or PolyTefzel* Propylene

Metallic

Iron

Alloy Steel

High Alloys

Vertical Sump FRP Vert. Sump Inline Process Horizontal Process Teflon Lined FRP Process SelfPriming Low Flow NonClog Horizontal Sealless Process NonMetallic Sealless Process

Figure 5. Process pump selection guide

TOMORROW'S CHEMICAL PROCESS PUMP Having evolved for more than 30 years, the ANSI/ASME chemical process pump now offers users improved reliability, easier installation and broader application flexibility. Its development over the next 30 years will surely produce further improvements in these areas. But instead of 3 years meantime between planned maintenance, the industry will be driving towards 5+ years. Pump emissions will not be acceptable at 1000 ppm. Zero (0) ppm will be the goal. It will not take 2 hours to align a pump and motor to 0.002 TIR. It will take only 15 minutes to align to 0.0005 TIR. And these and other changes will surely take place in less than 30 years. In fact, manufacturers and users both are demanding change now. The result is that current versions of the ANSI/ASME B73.1M and B73.2M are up for revision this year. The ANSI Pump Committee has also developed two new specifications. In addition, a new group called Process

Industry Practices (PIP) has developed two new specifications that supplement ANSI B73 specifications with additional requirements commonly specified in the industry (both horizontal and vertical types).

1. POTENTIAL REVISIONS TO ANSI/ASME B73 SPECIFICATIONS Changes are under way in the ANSI/ASME B73 Pump Committee. The basic B73.1M Horizontal Process Pump Specification will be revised and probably issued in 1997. Areas being addressed to improve pump reliability are: a. Nozzle loading b. Seal cavity dimensions c. Auxiliary connections to glands and seal cavities d. Baseplates e. Additional pump sizes f. Hydraulic Institute Class A performance criteria g. Allowable operating range New Specifications. The ANSI/ ASME Pump Committee recently issued a new B73.5M Specification The Pump Handbook Series

addressing nonmetallic pump designs. In addition, a Canned Motor/Magnetic Drive Specification B73.3M is expected to be approved in 1996.

2. NEW PIP STANDARD The new specifications covering typical B73 pumps are the PIP (Process Industry Practices) RESP73H and RESP73V. Engineering contractors and pump users have formed a Machinery Function Team whose sole task is to develop a set of standards that will eliminate variations in chemical process pumps manufactured to multiple user and contractor pump specifications. This team and its work will greatly minimize the problems associated with multiple specifications. Lower engineering costs and enhanced pump reliability will be the benefits. The PIP RESP73H (Horizontal Chemical Process Pumps ANSI B73.1M Type) and RESP73V (Vertical Chemical Process Pump ANSI B73.2M Type) cover the same pump design areas as the original ANSI specification, but they also address:

151

a. Solid shafts b. Shaft deflection L3/D4 c. Shaft sealing design responsibility d. Bearing lubrication e. Preparation for shipment f. Couplings g. Baseplate design h. Hydraulic performance acceptance criteria The PIP team has the same goals as the ANSI B73 pump team, but it is driving standardization both to the project and local levels. ■ Charles Cappellino is an engineering project manager for the Industrial Products Group of Goulds Pumps Inc. He is a professional engineer with a Bachelor of Science in mechanical engineering from Clarkson University and has been involved in centrifugal pump analysis and design for 15 years. Richard Blong is a product manager for the Industrial Products Group of Goulds Pumps Inc. He has a Bachelor of Science in chemical engineering from the University of Buffalo and has been involved in centrifugal pump applications for 10 years.

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The Pump Handbook Series

CENTRIFUGAL PUMPS HANDBOOK

ANSI Upgrades Require More Than Technology Total program commitment is key to ANSI pump upgrade success. By Joseph Dolniak

erely upgrading technology is not enough to increase pump reliability. A technological upgrade is just one of many factors that must be addressed to achieve maximum reliability. Reilly Industries has been increasing the reliability of its large ANSI pump population since 1990. At year end 1995 the company's total pump repairs were at their lowest levels since 1988. Meanwhile, plant production has tripled due to expansions and improvements in efficiency. We shall examine the procedures Reilly Industries instituted to bring its pumps into compliance with the new ANSI standards. Specific areas covered include new pump installations and orders, inventory, converting old pumps to new standards and future plans. Sealless ANSI pumps are not covered because there are none on site. Testing is currently under way to determine if certain sealless brands will be accepted into the plant. A specialty chemical manufacturer located in Indianapolis, Reilly Industries is about to celebrate its centennial under the leadership of Tom Reilly, Jr., the founder's grandson. The company has grown over the last century to where it currently manufactures more than 100 intermediate and specialty chemicals for a worldwide market. Just over six years ago a pump improvement program was initiated to increase the reliability of the company's more than 800 pumps. In implementing pump upgrade projects, certain steps must be followed

M

to obtain maximum benefit. The steps discussed below can be followed to complete any project successfully. They go by many names but most often are referred to as good engineering practices.

GOOD ENGINEERING PRACTICES Following are some of the quality steps or good engineering practices followed in Reilly's pump reliability upgrade project. 1. Know the current situation 2. Analyze the current situation 3. Formulate a plan 4. Initiate a trial 5. Set up standard and use 6. Train and communicate to work force and others 7. Maintain data 8. Continue making refinements 9. Analyze pump technology 10. Phase in and phase out 11. Stick to the plan 12. Redo poor installations Steps 1 - 8 are good engineering practices overall. Steps 9 - 12 refer more specifically to the pump upgrading project. For positive results in upgrading pumps for improved reliability, the pump, mechanical seal, gland, pump base, pump pad and immediate pump piping all must be addressed. The pumpage and flow rate conditions also must be compatible with the type of pump used. When this project was started in 1990, Reilly Industries had been using a computerized maintenance system for about 5 years. This helped greatly because the maintenance repair and cost history already was on file. The data indiThe Pump Handbook Series

cated that pump failures were increasing at an unacceptable rate. Production levels also were increasing, so there was an urgent need to improve pump reliability. The historical data were only a portion of the information that needed to be analyzed to implement a good upgrade program. Visual inspections of the pumps installed at that time revealed other factors that needed to be addressed. The most obvious was a lack of proper grout. Repair inspections also showed that many pump sites had excessive pipe strain. Indicator reverse, or laser alignment, was almost never done on the ANSI pumps. These deficiencies could be handled in-house through better maintenance practices and training. Other difficulties could not. The most important factor that could not be controlled inhouse was that mechanical seals were operating in stuffing boxes designed for packing. But a newly approved standard known as ASME B73.1M-1991 changed ANSI pump history.

ASME B73.1M-1991 ASME B73.1M-1991 is the "Specification for Horizontal End Suction Centrifugal Pumps for Chemical Process." One important item addressed in this revision was an increase in pump base sizes to add rigidity. Another was additional motor protection. Perhaps even more important, however, was the new designation of a seal chamber versus the old stuffing box. The introduction of the seal chamber allowed the mechanical seal manufacturers freedom to create new technology and mechanical seal designs. The gas barrier seals that

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Photo 1. Recently received group two size pump incorporates motor jacking bolts, grout hole, bearing isolators, bull's eye oil indicator and bearing housing expansion chamber as standard features.

are now on the market are a result of this. Specifics not addressed in this revision were the standardization of the gland bolt circle, shaft size at the mechanical seal, and sealing surface diameter at the gland/seal chamber interface and gland piloting area. Standardizing these areas would allow the end user even greater freedom to reduce me-chanical seal and gland inventory while maintaining competitive bidding from various mechanical seal vendors. With this revision, three basic seal chambers could be used. They included the 4º taper bore, the large bore with a throat restriction and the large bore with no throat restriction. All three designs are beneficial to mechanical seals and can be swapped for the old stuffing box. In each the gland has to be replaced because of the larger static sealing surface and bolt circle diameters.

formulated to address pump reliability. Effort would be directed mainly at ANSI pumps because of their greater population and the fact that many improvements made on them would apply to other types of pumps also. The upgrade plan involved a two-step approach. The first step was to ensure that no more pumps were misinstalled or improperly ordered. This would help reduce maintenance problems from the start. Reilly engineers began by writing standards for proper pump installation procedures. Covered in the standards were such details as pump base dimensions, heights, depths, hold-down bolt designs, pump spacing from one another, pipe strain, grout and alignment. Before the standards were approved, they were tested on four pumps that showed normal repair rates for the plant at that time. After these pumps were reinstalled according to the proposed standard, they ran much quieter and smoother. Looking at the repair frequency and costs for the three years before the reinstallation and after, total repair costs dropped 94%, from $46,470 to $2,908, and total repairs dropped 69%, from 49 repairs to 15 in the same time period. These pumps were still fitted with the standard stuffing box because the seal chamber was not yet on the market.

Shortly after the pumps were reinstalled, however, the standard was accepted as a site engineering standard for ANSI pump installation. At the same time, new standards were written for future ANSI pump orders. These standards incorporated into all new pump orders many of the improvements specified in ASME B73.1M-1991. Included were requirements that all new pumps have 4º taper seal chambers, drain and discharge taps on the pump casing, labyrinth bearing isolators, bull's-eye oil level indicators and bearing housing expansion chambers. On pump-base assemblies the base would conform to ASME B73.1-1991 di-mensions, have a grout hole centered on the base (4" preferred), and have motor jacking bolts for alignment purposes. Some of these upgrades are shown in Photo 1. This would give us a head start as we would soon begin an alignment and pump installation program. The jacking bolts and grout holes saved maintenance time when these programs were under way. The accomplishments up to this time took more than a year to achieve. They helped prevent misinstallation of ANSI pumps and eliminate the process of ordering pumps with old technology. This marked the beginning of the next phase of the reliability program, but in actuality it would have little noticeable effect for some time as the new

PUMP RELIABILITY UPGRADE PLAN With the pump repair history, installation analysis, repair analysis and the new ASME B73.1M-1991 standard in place, a plan could be

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Photo 2. An upgraded second generation pump receives backup as part of the reliability improvement program. The Pump Handbook Series

pumps would be installed only as required through expansion or attrition. The next phase of the upgrade had more immediate impact because the ANSI pumps were upgraded as they were worked on. This step addressed standardizing and consolidating pumps and pump inventory while adding the newtechnology parts to stores. This was also a two-step process. Step one was the reduction in numbers of ANSI pump brands and the consolidation of mechanical seals. The second step consisted of phasing in new parts while the old parts were phased out of stores.

PUMP CONSOLIDATION At least nine specific brands of ANSI pumps had been in use on site. The goal was to reduce that number to three. Some of the factors considered in selecting what brands to retain were past reliability, maintenance shop familiarity, local distributor professionalism, location of the OEM and the size of the current plant population in specific makes. When selection was finished, seven of the nine brands were eliminated, and a new brand was added. The standards list reflecting our preferred vendors was updated, and Reilly is now testing a fourth brand of ANSI pump for possible addition to the list. As for the mechanical seals, we are about 95% committed to one manufacturer. This simplifies the process of consolidating our mechanical seal inventory. The time-consuming pump brand reduction project is just ending after more than four years of phasing pump parts out of inventory. Caution should be taken when adding additional brands of pumps to a plant site once this point is reached because the amount of work needed to add a brand is almost as time-consuming as it is to eliminate one. Also, the system must be allowed a break-in period to determine how well it is working. Continual changes do not allow this to happen. Also, additional shop training is needed. That is why the preliminary work of determining what pumps will best suit the plant is so important. After

finalizing the brands of pumps accepted on the vendor list, work began to phase in the accepted brands and eliminate brands that would no longer be used. The company was careful not to be wasteful in eliminating brands. Repair parts in inventory that were associated with the pump brands to be deleted were classified as POR (purchase on request). This would cause no parts to be ordered when the reorder point came up, but it also would not delete the part from stores. In so doing, most of the repair parts currently in inventory for the brands of pumps to be deleted could be used. If one or two minor parts were needed to complete a repair and the repair parts were not on hand, giving the stores clerk a work order number would allow the clerk to order the POR part. When most of the parts were used up, the part code would be changed to DELETE, and any remaining parts would be pulled from the shelves. Thus, most of the parts could be used while eliminating unwanted inventory. When no parts or not enough major parts remained, a new pump of an accepted brand was ordered to replace the pump that was being repaired. The new pump incorporated all of the upgrades we required when it arrived in the plant. Next month in Part II of this article we will assess how Reilly Industries conformed to the new upgrade plan. We will discuss consolidation and parts inventory changes, show how mechanical seals were upgraded, and reveal some of the training procedures that have been instrumental in helping this project succeed. We will end with a look at future plans. ■

CONFORMANCE TO PLAN It would have been very easy at the conformance stage to deviate from the plan because there were times when the up-front dollar amount involved in purchasing a part being deleted was less than the cost of upgrading the pump technology and purchasing a new pump. Economizing up front, however, The Pump Handbook Series

would have compromised the plan and been costly in terms of overall life cycle cost. This process is proving profitable for the company as our overall pump repairs have declined as production has steadily risen, as shown in Graphs 1 and 2. Total repair costs for all pumps on site appear to be going up slightly. This was expected because we began requiring our work force to do more thorough repairs, which take more time. Maintenance workers now replace parts that were not normally replaced due to lack of proper inspection. More parts are now being found to be out of specification when checked with dial indicators. And some auxiliary parts now cost more because superior materials are being used. Yet if the pump repair costs were corrected for inflation, the costs would be almost level. Also, the total number of pumps has increased. The inhouse maintenance system showed 792 pumps at the end of 1989 and 837 pumps at the end of 1995. Achieving near level repair costs while doing more thorough repairs with superior parts is attainable because we have increased our MTBF (mean time between failure), and thus there are fewer repairs. Another technique helped maintain costs while phasing out specific pump brands. If a pump being removed for an upgrade or taken out of service was one of the brands that was to be kept, and if it was worth rebuilding, it was put in a specified area. If the pump filled the requirements of a pump brand that was to be eliminated, it sometimes was used in lieu of purchasing a new pump. Several important considerations in doing this were the pumpage material and the repair frequency and the costs of the deleted pump brand being removed. Photo 1 is an example of one of these pumps being brought back into service and adding a back-up pump at this site, which before had no back-up pump. This helped determine if replacing old technology with old was appropriate, or if new should be used, and it improved reliability while maintaining costs. To date, the equivalent

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of more than $50,000-worth of pump parts has been removed from inventory records.

SEAL CONSOLIDATION Consolidating the mechanical seal parts was more difficult because we were phasing out standard stuffing boxes and phasing in 4º taper bore seal chambers. This required double inventory of glands and gland gaskets for about two years because the gasket and bolt circle on the seal chambers and their glands were larger. There was no need to duplicate any mechanical seal parts. The mechanical seal stationary insert fit both the old glands used with the stuffing boxes and the new glands used with the seal chambers. Parts relating to the stuffing box backheads have recently been eliminated, and now only seal chambers and the new glands remain in inventory. There was another important factor in keeping the additional inventory, now current inventory, to a minimum. In evaluating what pump brands to keep, we also looked at the shaft diameter at the mechanical seal, the seal chamber gland bolt circle, and the gland gasket and pilot diameters. The three ANSI pump brands that were chosen had the same dimensions for the above items. This was important, and it is why, generally speaking, adding these items to ASME B73.11991 as part of the standard should be beneficial to the end user. By choosing pumps with the same dimensions on the above items, we were able to use only one gland to fit all three brands of pumps, per pump group size designation. Thus, for all of our group one and group two size ANSI pumps, which are the majority of ANSI pumps in the plant, we have only two glands: one for all group one size pumps, and one for all group two size pumps. The three types of component mechanical seals (one OEM) that we use per pump group size all fit the one gland. The glands come with vent, flush and drain. If they are not needed, they are simply plugged. This has enabled us to eliminate at least nine specific glands

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from inventory. This strategy also works for consolidating cartridge mechanical seals. The process took much communication between myself, the mechanical seal engineers and the pump engineers. The second major benefit of consolidating the glands besides the inventory reduction is that there is no confusion to the work force on what gland to use on what pump, since there is only one gland. This entire upgrade process is always evolving and being updated as needed. We are currently

consolidating three other glands into one.

PARTS INVENTORY CHANGES Other changes made at this time were the addition of the new technology parts to inventory. These parts included solid shafts (versus the sleeved shafts), bearing isolators (versus oil seals), bull's-eye oil indicator or column sight glasses (versus the constant level), the new glands and the 4º seal chambers (Photo 2). The old parts were earmarked so

■

1200 —

■

■

■ ■

■ ■

■—REPPMP ■—REPPMPP ■—REPPMPV ■—REPPMPC ■—REPPMPTOT

■

1000 —

800 —

■ ■

■

600 —

■

400 — ■

■

■

■

200 —

■ ■

■ ■

■ ■

1992

1993

1994

■ ■

0— 1987

1988

1989

1990

1995

Graph 1. Total plant pump repairs: REPPMP = repairs pump (one of a kind pumps), REPPMPP = repair positive displacement pumps, REPPMPV = repair vertical pumps, REPPMPC = repair centrifugal pumps, REPPMPTOT = total pump repairs 3.5 — ESTIMATE ■

3—

■

2.5 — ■ ■

2—

■ ■

■

■

1.5 —

■

1 —■

I

I

I

I

I

I

I

I

I

I

1987

1988

1989

1990

1991

1992

1993

1994

1995

1996

Graph 2. Total plant output using 1987 as the base quantity of one unit of output.

The Pump Handbook Series

that they could be phased out of inventory and deleted at the appropriate time. This created a few minor problems because some of the pump OEMs also were modifying various parts because of the new ASME B73.1-1991 standard. It caused us to add parts that were being changed, and this created minor problems with associated parts. The confusion was minimized, however, by keeping communications open among the end user, vendors and OEMs. As mentioned earlier, this phase of the project is currently drawing to a close. Because Reilly has a large number of pumps installed that are two generations old, and were designed when mechanical packing was the norm, they have inherent reliability deficiencies when fitted with mechanical seals. These pumps have long thin shafts and stuffing boxes. Due to their age and generation, the pump OEM will not be manufacturing upgraded parts such as seal chambers for these pumps. Because they represent a large part of the plant pump population, however, it was important to improve their reliability also. Two primary points were addressed: seal environment and shaft stability. After conferring with the OEM, it was agreed that the stuffing box could be bored. A new diameter for the stuffing box was determined, and the stuffing box was bored all the way through, giving us what we call a modified large bore with no throat restriction. Our local vendor for this pump brand sees to it that these parts now come into stores already bored through when they are ordered from the OEM. To increase the shaft stiffness, we opted for a solid sleeveless shaft. This type of part modification is also being applied to some of our larger group two and group three size pumps, which represent a minority of the plant pump population. Because of their limited numbers, we stock only commonly used repair parts (shafts, bearings, mechanical seals). Large-dollar parts such as casings and seal chambers are listed as POR. By boring out the stuffing box and keeping the standard gland, we are able to receive

some of the reliability benefits of the taper, or large bore seal chamber, while holding down repair costs. When there is need for more complete repair, these pumps also will be converted to updated technology.

MECHANICAL SEAL UPGRADES Upgrading the mechanical seals was another important process for us. It went hand in hand with consolidating the mechanical seals. Specific failures were noted on one type of elastomer. Other failures specific to one plant were noticed on one of the seal hard faces. Also, with more stringent regulations approaching, we felt it would be beneficial to upgrade the seal/elastomer combinations used throughout the plant. Because our ANSI pumps generally were not pumping "easy" products, we decided that it would be best to use premium hardfaces and elastomers for all mechanical seals on ANSI pumps. Due to our consolidations, this affects only eight specific seals. The reasoning was that our average pump repair plantwide (ANSI and non-ANSI) was about $1,000 per repair. This covers any action ranging from no repair or minor repair to installing a new pump. Significant failures were noticed on encapsulated o-rings. Converting to an elastomer such as Kelrez added several hundred dollars to the cost of the mechanical seal. Still, if repairs were reduced, costs would decrease in the long run because there would be fewer failures (Graph 1). The new hardface used to replace the hardface experiencing problems in one specific plant would also work throughout the entire Indianapolis site. Therefore, this conversion was completed as well. Through failure analysis, running dry was determined to be one of the major factors contributing to premature seal failure, even after pump upgrades were completed. To retain the gains discussed here, three op-tions were identified to address this problem. The first is to fit pump motors that can be turned off while the pumps are in service with power The Pump Handbook Series

monitors. This pertains primarily to transfer pumps. Use of properly calibrated power monitors has all but eliminated seal failures at specific pump sites within our plant. These sites also have been correctly upgraded, which eliminates earlier root causes of pump failures before the power monitors were installed. The second option involves replacing the 4º taper seal chamber with a restricted throat large bore seal chamber. Swapping seal chambers is considered because of the pumpage properties in certain pumps. There are signs of running dry, which could actually be entrained air, or a phase change of the pumpage at the seal faces. The restricted throat large bore seal chamber with various flush plans allows us to obtain a higher pressure in the seal chamber and thus reduce or eliminate this problem. The third option to prevent pumps from running dry pertains to areas where the pump cannot be turned off automatically because of production needs. For some of these areas, we are installing gas barrier seals designed for the pumping requirement. Until recently the gas barrier seals required that a seal chamber be present, as they would not fit in the common stuffing boxes. This situation is changing and gas barrier seals are becoming available for the use in stuffing boxes (see Pumps and Systems, February 1996, pages 8 and 9). As of yet, we have ordered and received only one gas barrier seal, and we are waiting to install it when the current pump comes out for repair.

TRAINING One aspect not yet mentioned is training for the maintenance work force. Workers were generally concerned about all of the changes. New brands of pumps were being used. Different types of technology were represented. More was expected of workers. All of these were real issues that had to be addressed. Although one-on-one discussions and shop meetings were held, the only way to ensure full communication of all of the changes was through formal training for the

157

maintenance work force. Formal training assures us that all of our workers are up to date on new parts and procedures. Specific training areas have included ANSI pump rebuilding, mechanical seal installation and indicator reverse alignment. Training in ANSI pump rebuilding covered the specific dial indicator checks that must be performed to ensure that the pump components are in the proper operating specification. Mechanical seal installation training noted the proper techniques to install both component and cartridge mechanical seals by the print, and by the stack method, and testing the seals on the bench before installing the pump in the plant. Indicator reverse alignment training covers how to align pumps properly and overcome common alignment problems such as soft foot, bolt binding and shim pack requirements.

OEM AND VENDOR TRAINING Training the work force was only part of the project. Training the vendors and OEMs and maintaining good communication among them were critical parts of the training program. This meant keeping them up to date on our new pump requirements, allowing them to read and understand our plant engineering standards relating to the pumps, and ensuring that they held enough pump part inventory to cover our normal use. This was especially important in the beginning stages because being out of a specific new technology part needed for a repair, especially on a critical pump, would add fuel to the fire of the nay-sayers. This was true more so when we implemented the changes because many of the pump OEMs were not yet willing to stock taper and large bore seal chambers because there "was no demand" at the early stages of their introduction to the market. My reply was that we were demanding them, so please stock them. Because these new parts have proved to be important in increasing pump reliability, and there is indeed demand for them, most vendors now stock these parts. When this program started, one

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of the major components of upgrading the ANSI pumps was the new seal chambers. Reilly first upgraded the bulk of its ANSI pumps to the new ANSI standard. We felt that this alone would increase the pump reliability. When the general population was updated, more specific problems would become clear and could be addressed. Looking at the three options that were available, it was decided that the taper bore seal chamber would best fit the overall needs of the company. Because of the many other factors that needed to be addressed to get the pumps, stores, and the vendors up to date on the pump upgrades being implemented, only one seal chamber was chosen Adding two seal chambers to the system would be detrimental to the entire process, and it would just add confusion until the project was more mature. Now that this general pump upgrade is accomplished, as mentioned earlier, we are looking at certain specific installations where the large bore seal chamber with the restricted throat might be of more benefit. This is primarily because of the makeup of the product, and not because of gland erosion. We have seen only one instance of gland erosion as a result of the cyclonic effect of solids and entrained air in a seal chamber, and the one instance was in clean acid. We have also had our mechanical seal OEM perform tests on various flushing arrangements using the taper bore seal chamber and our own gland (group one size pumps, 1800 rpm) with respect to particulate and entrained air in the seal chamber, and we have had interesting results that we can directly apply to field use. With conversion to the 4º taper bore seal chambers from stuffing boxes well established, if we now want to change from a taper bore to a large bore seal chamber, the gland, seal and other pump parts will fit. We need only change the type of seal chamber used for the specific pump in question.

FUTURE PLANS Our ANSI pump upgrade program has been operating for near-

The Pump Handbook Series

ly three years now. But along with upgrading the pump technology, we also had to address the root causes of failures. Reilly has begun to eliminate pipe strain proactively. Completely reinstalling problem pump sites has been accomplished with good results. Some of these sites incorporated both epoxy bases and epoxy grout, which give superior chemical resistance and vibration dampening (Photo 3). Steps have been taken to improve our grouting techniques, and we are completely phasing in epoxy grout. Compared to cement types of grout, this gives superior vibration dampening, chemical resistance and adhesion when the pump base and pad components are properly prepared. Also, bearings with tighter tolerances than we normally stock are being phased into stores. Because of the long term results, there are now more requests from upper management to apply the pump reliability and upgrading techniques further – to specific problem pumps in each plant. This will include improving our vibration analysis program and reiterating the importance of aligning every pump that is worked on, no matter what the size. As pump reliability continues to improve, new "bad" pumps will make the list of pumps to be addressed as old ones are removed from the list because of improved reliability. This is part of the continuing refinement. By analyzing pump failure data, improvements can continue as long as there is a well thought out improvement process. Looking at the repair graph, it may appear that our reliability is improving at a slow rate. But in looking at the entire picture, our plant has been undergoing extensive growth through expansion and efficiency improvement. Our production has steadily increased, our pump population is increasing, new products are being brought to production while production of older core products are being reduced. These changes have created new challenges, and such conditions can have a dramatic effect on the pump and

Photo 3. Upgraded pump site includes epoxy base and grout for improved reliability.

mechanical seal population of any plant. When all these points are considered, however, there is significant satisfaction with the progress achieved in a pump reliability program of this type. ■ Joseph Dolniak has been the maintenance engineer at Reilly Industries for more than 6 years and is involved in reliability improvements of rotating equipment. In addition to producing alignment, seal installation and pump rebuilding training videos for Reilly, Joe has published several articles on ANSI pump reliability and has lectured at the Pump Users Symposium and the Pump Reliability and Maintenance Conference.

The Pump Handbook Series

159

CENTRIFUGAL PUMPS HANDBOOK

Selecting Mag Drive Pumps Magnetic drive pumps offer irresistible force in sealless pumping. By Robert C. Waterbury, Senior Editor

ndustrial processes involving toxic, hazardous or environment-threatening chemicals often employ the magnetic drive pump (MDP) as a safe, sealless solution. But even though MDPs offer a simple answer to a common need, certain characteristics must be considered to select and apply them cost-effectively. Kaz Ooka, president and founder of Ansimag, Inc., points out that magnetic drive pumps historically developed along two lines: metallic and non-metallic. The metallic designs traditionally were used in process or heavy-duty applications. But non-metallic pumps, once considered only for light duty applications, have moved up in power and size due to development of improved rare earth materials such as samarium-cobalt and neodymium-iron-boron. Synchronous MDPs use rare earth magnets. Because they are affected by high temperatures, they often require special cooling provisions for applications in excess of 400ºF. Eddy current MDPs employ a torque ring that is normally unaffected by temperatures found in hot oil heat transfer systems. They use a rotating assembly sealed by a containment shell. Power is transmitted by permanent magnets mechanically coupled to the driver rather than through motor windings. Sealless MDPs prevent liquid leakage and eliminate common environmental concerns. They are also used to move liquids that crystallize upon contact with air, and the seal flush liquids or gases they employ help avoid contamination of process fluids. Yet they don't solve all seal-

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ROTAN MD SERIES PUMP

Figure 1. Rotan MD series magnetic drive pump

related problems. Two issues that MDPs still must address are minimum flow conditions and dry-running. Minimum flow rate is greatly affected by radial or thrust load on the bearings or pump shaft and the temperature rise. Dry-running is the most common cause of failure in MDPs and results in thermal damage to the metallic containment shell and/or in mechanical or thermal shock to the bearings and shaft. We shall discuss these in more detail.

DESIGN CONSIDERATIONS The inside of a sealless magnetic drive pump reveals a complex internal flow system that is difficult to model. However, the internal design holds the key to cooling the magnet drive and lubricating the bearings effectively and to the safe transport of solids. Engineers at HMD Seal/Less Pumps in East SusThe Pump Handbook Series

sex, England, which is affiliated with Sundstrand Kontro, identify five critical design elements: 1. the liquid end, comprising pump casing and impeller 2. a magnetic drive including an inner and outer magnet assembly and the containment shroud 3. internal support bearings 4. an internal feed system that circulates among 1, 2 and 3 above and is needed to cool the magnetic drive, lubricate the bearings and transport any solids in suspension 5. a power frame that comprises the external bearings supporting the outer magnet and the interface to the prime mover

INTERNAL FEED SYSTEM Of these, HMD considers the internal support bearings and internal feed system the most critical, and yet perhaps least discussed. The

PHOTO COURTESY OF ANSIMAG, INC.

feed system removes heat generated in the drive assembly by eddy current and viscous friction losses, and it lubricates the process lubricated bearings that support the loads exerted upon the rotor assembly. Discharge to suction internal flow. To remove heat produced in the drive assembly, liquid is taken from the discharge of the pump and returned at a lower pressure point within the pump. The simplest and perhaps lowest cost system takes process liquid from the pump discharge and recirculates it to the pump suction end. Typically, process liquid leaves the exit of the impeller and returns to the magnetic drive through a hole in the rear casing plate. With this system the liquid always returns to the suction of the pump at a higher temperature than the bulk suction temperature. Thus, it is necessary to ensure that the hotter temperature liquid does not affect the NPSHR of the pump and vaporize as it returns to the impeller eye. This condition is often overlooked, according to HMD, if suppliers test using only water. Because water has a liquid with a high specific heat and gradual vapor pressure curve, test results using water alone may mask this potential problem. Discharge to discharge internal flow. An alternative is to take the discharge liquid from the pump and return it at a point of pressure higher than suction. The internal feed system is similar to the discharge to suction feed system, except that the liquid is directed to a high pressure area in the casing typically behind the back shroud of the impeller. There are two advantages to this system. First, it eliminates the NPSHR problem. Second, pressure distribution within the drive is related to a high pressure area as opposed to suction and therefore reduces the possibility of vaporization within the drive. The main disadvantage is that the supplier must test to ensure that there is adequate pressure difference under all conditions to force sufficient flow through the drive. If not, low flow and excessive temperature rise will have the

Photo 1. Mag drive pump transfer of 93% sulfuric acid has operated nearly two years without repair

same effect as high flow and excessive pressure drop in causing vaporization.

INTERNAL BEARINGS Michael "Todd" Stevens, senior maintenance engineer at Hoechst Celanese Chemical in Houston, has analyzed bearing failures in magnetic drive pumps and offers some helpful observations. He suggests looking at the lubrication scheme for the roller/ball bearings and the shaft seal that prevents lubricant from entering the drive magnet section of the pump. When using wet sump lubrication, shaft failure will allow an oil level to build in the drive magnet section, causing a heat buildup. Normally under these conditions the pump will begin making noise and vibrate. If the oil level buildup is not detected, the temperature of the drive magnet section will increase and eventually cause a roller/ball bearing failure. Stevens also suggests selecting silicon carbide rather than carbon as a main bearing material. It has better wear properties and will withstand most thermal shock without failure. Some suppliers even diamond coat silicon carbide journal bearings to withstand brief periods of dry-running without bearing failure. Thrust bearings should be engi-

The Pump Handbook Series

neered so that full-face contact is achieved between the bearing and thrust runner. Failure to do so results in point loading, which can damage the bearing. Although Stevens recommends a fixed face bearing, he concedes that a floatingface spherical-seated thrust bearing could be acceptable if there were some way to secure the floating face at all times. Also, open impellers impose higher thrust loads than closed impellers. Thus, thrust loads must be either offset by balancing the impeller or absorbed by the thrust bearings. Lower thrust loads obviously mean increased bearing reliability.

INTERNAL HEAT GENERATION Bearings are not the only mechanisms that generate heat in a mag drive pump. In fact, Stevens points out that most heat is actually caused by eddy current losses between the driven and drive magnets. With a permanent magnetic coupling, the driven rotor turns at the same speed as the drive rotor. The two magnets are separated by a containment shell. Heat generation is thus a function of pump rotation speed and containment shell construction. Containment shell materials such as 316 stainless steel generate more heat than Hastelloy C, Stevens notes, and Hastelloy C generates

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MDP MONITORING/INSTRUMENTATION

PHOTO COURTESY OF ANSIMAG, INC.

Safe, reliable operation of MDPs clearly depends not only on pump selection and installation, but on monitoring pump operating conditions. Following are some of the more common techniques and instrumentation.

Photo 2. A mag drive pump transferring 50% sodium hydroxide at a specialty chemical manufacturer in the southeastern U.S.

more heat than high performance plastics. The temperature of the containment shell directly between the two rotating magnets can easily rise above 750ºF within 30 seconds of the onset of dry-running, and it can eventually reach nearly 1000ºF. So internal fluid flow between the driven magnet rotor and the containment shell is needed to remove the heat. Dry-running heats the shaft. If cool liquid is introduced at this time, however, the shaft and bearings may fail due to thermal shock. Ceramic materials can minimize these effects but offer widely varying thermal shock limits. Ooka says alumina ceramic can withstand only a 200ºF thermal shock while sintered carbide offers resistance up to 600ºF. Silicon carbide offers such high thermal shock resistance because it has a very low coefficient of thermal expansion combined with a very high thermal conductivity. This allows the material to equalize in temperature very quickly while exhibiting minimal thermal strain.

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Temperature monitoring. One way to monitor pump condition is to use a thermocouple or RTD (resistance thermal device). It can be positioned to monitor the temperature of the containment shell or placed in a thermowell to indicate the temperature of the fluid leaving the shell. Either way, it monitors the heat produced by the eddy current losses in the magnetic coupling as well as the bearing friction.The temperature can be used as an absolute or a differential measurement. Used as a differential temperature indicator, it is referenced to the pump suction fluid temperature and directly measures heat input to the fluid by the pump. Flow protection. Minimum flow protection is normally provided by installing a flowmeter in the discharge line of each pump. Stevens points out that the minimum flow to protect against is either the thermal or the stable minimum flow, whichever is greater. It is only reliable, however, if no more than one pump is operating in a two-pump system. Otherwise, each pump must be individually instrumented and protected. Low suction. Low suction vessel protection ensures that a pump will not run dry. According to Stevens, it is used in tank loading and offloading applications to ensure that the pump will not run dry or suffer from inadequate NPSHR. If a tank must be emptied following unloading, then a mag drive pump should not be used. Otherwise cavitation and subsequent failure of the thrust bearings and possibly the sleeve bearings could result.

The Pump Handbook Series

Power monitoring. Technically, a power monitor can help guard against low flow, high flow, magnetic decoupling and dry-running. It measures the power consumed by the motor and thus responds quickly to load changes that could lead to mechanical damage. It obviously applies only to pumps driven by electric motors, however, and whether it is sensitive enough to distinguish between low flow and shutoff is questionable. In such cases a flowmeter or other device may be required as backup. Vibration monitoring. Sleeve bearings typically run so smoothly (less than 0.1 in/sec overall) that periodic vibration monitoring has not proven useful in predicting failure. It has been successful, however, in predicting the failure of the drive magnet support bearings – normally roller or ball bearings. This could help ensure that the drive magnet does not contact the containment shell in case of a support bearing failure.

THE DECISION PROCESS The Clean Air Act targets 179 liquids in setting limits for allowable chemical leakage into the air. The primary concern is for safety of humans and the environment. And magnetic drive pumps eliminate the dangers normally associated with seal leakage in mechanical pumps. But even though MDPs are purchased initially for safety reasons, many users are now specifying them for reasons of improved reliability, extended service life and longer mean time between maintenance and repair. In the long run they may prove more economical even though the initial cost is higher. Once the decision is made to purchase a magnetic drive pump, however, many questions must still be answered as part of the selection process. The following selection guide developed by Ansimag can help users tailor solutions to meet their specific needs. Dimensions and design. Does the manufacturer make a pump in

the design that you require? There are numerous configurations and standards including ANSI, ISO, API, DIN, etc. Solution: Determine the design that you need and consider only those manufacturers that build that type of pump. Vapor pressure of the liquid. Metal magnet drive pumps add heat to the process fluid due to losses in the magnetic coupling. The typical magnetic coupling is anywhere from 70% to 80% efficient. This inefficiency is translated into BTUs that enter the liquid as temperature. As much as 20ºF can be added to the liquid that is lubricating the bushings. If the liquid vaporizes, the bushings will be starved for lubrication, and the pump will fail. Solution: Ask the manufacturer to run a heat balance calculation to determine if the vapor pressure of the liquid will ever exceed the local pressure in the pump. If it does, it is the wrong pump. Also, consider a nonmetallic magnet drive pump with zero losses in the magnetic coupling. This will eliminate the possibility of vaporizing the liquid. Solids. Because the bushings are lubricated by the process fluid, a "clean" liquid is required. Some pumps will handle more solids than others. Solution: Ask manufacturers to state maximum limits and give references of applications handling similar solids content. A general rule of thumb is 5% by weight and 150 microns maximum. Flush systems are available from some manufacturers to increase solids handling capacity. Hydraulic capacity. Some manufacturers have more capability with regard to head and flow than others because they offer more models. Solution: Look at what models the manufacturer has available and ready to ship (not just planned as future products but currently available). Even if the manufacturer has what you need currently, consider also that you may want to add larger

units at a later date. If he does not have them available, you lose commonality of parts and continuity of design. Temperature. Magnet drive pumps have a variety of temperature capabilities. Solution: Determine not only what temperature you will be operating at, but also what maximum temperature the pump could see due to excursions or future design revisions. Also, consider the effects of steam cleaning or heat tracing if you plan on either of these. Look at both the magnet capability as well as the material capability of the pump with regard to maximum temperatures. Simplicity and ruggedness. These two items are critical now that maintenance staffs are slimmer. The less time spent on the pump the better – ruggedness of design is key. When maintenance is required, the simpler the better since the time spent on repair should be minimized and the risk of making an error should be reduced. Solution: Ask the distributor or manufacturer to demonstrate the pump to determine if it is indeed simple and rugged. Viewing the pump in operation is critical because every manufacturer claims to have a simple and rugged design.

CAUTION: RISK AHEAD As Stevens says, process engineering people design systems for normal operation and project engineers then use these flow requirements to purchase pumps. This is a normal method of sizing and purchasing a pump, but it is not always successful in purchasing and sizing a mag drive pump. An MDP is somewhat more sensitive to changes in pumping conditions than perhaps an ANSI design pump. Startup procedures, for example, do not always call for operation using the same fluid, pressure, temperature, specific gravity and viscosity indicated on the data sheet. Furthermore, pumps may be used for more than one operation, or they may be required to The Pump Handbook Series

pump at widely differing flow rates during unit startup, operation and shutdown cycles. The greatest risks are always posed by conditions that fall outside the normal range of operation. The pump system is designed to operate at the Best Efficiency Point (BEP). However, real world conditions demand more from a pump than a single BEP. A pump may be used to transfer fluid from tower to tower before unit startup to achieve a normal tower operating level. Similarly, it may be used to clear the unit in case of a unit trip, or it may be used as a spare for a completely different service via a jumper line. These scenarios must be postulated and the implications explored before installing a magnetic drive pump. Properly done, this exercise will provide an operating window of minimum/maximum values to be considered in the selection process. Improper application is perhaps the most frequent cause of failure. Calculations of available versus required NPSH must be extremely accurate and compatible; otherwise, cavitation and pump failure will follow quickly. In addition, obvious practical considerations such as direction of motor rotation must not be overlooked. In a recent pump startup operation, three of nine initial failures were due to incorrect motor rotation.

SELECTED PRODUCTS AND APPLICATIONS Ansimag. Different versions of the Ansimag K1516 mag drive pump are being used to move an extensive list of hazardous, corrosive and toxic chemicals. Applications noted most frequently in a new Ansimag case history publication involve such chemicals as hydrochloric acid, sulfuric acid, sodium hydroxide and sodium hypochlorite. The main reasons users give for switching to these pumps include zero leakage requirements, safety, elimination of seal problems and increased system uptime. The users include specialty chemical and petrochemical companies, pulp and paper processors, food and pharmaceutical companies,

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and steel, plastics and electronics manufacturers. In its list of user applications, Ansimag records user chemical concentrations ranging from 1 -100%, flows from 1-500 gpm and TDH in feet from 5-250. Process temperatures generally range from 94º to 250ºF. Kontro/HMD. Kontro and HMD Seal/Less pumps are available in a wide range of capabilities designed for specific target applications. The A-Range mag drive pumps feature capacities to 2000 gpm, heads to 700 ft TDH, temperatures to 400ºF and system pressures from full vacuum to 275 psig. Applications include toxic or hazardous liquids, high temperature vacuum distillations and liquids that are expensive or require controlled purity. The H-Range pumps offer capacities to 5000 gpm, heads to 700 ft TDH, temperatures to 750ºF and system pressures from full vacuum to 300 psig. Applications include heat transfer fluids, molten solids and high temperature vacuum distillations. The API pumps range to 5000 gpm capacity, heads to 700 ft TDH, temperatures to 750ºF and pressures to 580 psig. Applications are refinery and petrochemical services. The HSP series fills high system pressure requirements including: capacities to 2000 gpm, heads to 350 ft TDH, temperatures to 750ºF and system pressures to 5000 psig. These pumps are used in nuclear, high pressure densitometer and pipeline detection systems. Finally, the self priming SP pumps are designed for truck and tank car offloading of hazardous chemicals. They accommodate capacities to 350 gpm, heads to 300 ft TDH, temperatures to 400ºF, pressures from full vacuum to 150 psig and suction lift to 15 ft. Micropump. Micropump offers precision fluid pumps and systems. Its Integral Series line is used in hemodialysis, chemical dosing, dispensing and filling, water purification, ink jet printing and laser/electronics cooling. Its distinctive features include motor, variable speed control via external signal, low

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power consumption, pressures to 100 psi and flow rates to 5.8 gpm. The pump head, brushless dc motor and electronic controller are integrated into a single compact unit with no mechanical seals, packing or leakage. The drive system is distinguished from conventional magnetic couplings by sealing the rotor inside the pump and driving it directly by the motor stator. The electronic controller, an integral part of the motor drive, accepts separate 0-5 Vdc or 4-20 mA signals used to adjust speed control. DESMI/Rotan. Mag drive sealless pumps in the MD series offer capacities to 225 gpm, speed to 1750 rpm, differential pressure to 250 psi, temperature range to 500ºF, suction lift to 15" Hg vacuum while priming and 25" Hg while pumping. MD pumps are recognized for their integral pump cooling system, dynamic axial balancing feature that reduces energy consumption and increases MTBM, a thrust control system that maintains correct running clearances and reversible pumping capability through changes in motor rotation. A patented system circulates the pumpage around the magnetic coupling for cooling. A special shaping at the rear of the rotor uses the hydraulic pressure itself to balance the liquid pressure dynamically on the rotor. Maag Pump. MPS pumps from Maag are known for their use in very high pressure applications. Their operating conditions feature temperature ranges to 300ºC, suction pressure to 16 bar vacuum, discharge pressure to 66 bar maximum, differential pressure to 50 bar maximum and viscosity to 1000 m Pa. They are available with either single or double containment shells. Price Pump. The model CD 100MD from Price pump offers flows to 70 gpm and heads to 95 ft TDH. Incorporating 316 stainless steel or higher alloy materials, its standard configuration withstands pressures to 300 psig and temperatures to 350ºF. Factory engineerThe Pump Handbook Series

ing is available for higher temperature and pressure ratings. Three magnet strengths are available for varied load conditions, and insrumentation options include power meter, temperature probe, and vibration, temperature and pressure switches. Roth Pump. Magnetic drive pumps from Roth offer the advantages of regenerative turbine pumps in addition to a low NPSH feature. A floating, self-centered impeller is able to produce any level of differential pressure from 50 to 500 ft TDH. Maximum pressure with 316 stainless steel (standard) is 230 psi, or up to 360 psi with optional Hastelloy C material. A power factor sensor indicates both high and low load and reacts to upsets caused by blocked valves or vapor-bound conditions. Dean Pump. The M300 conforms to the dimensional specifications of ANSI B73.1 and features a one-piece hydroformed containment shell. The seal for this shell is the only o-ring in the pump, and optional flow paths are offered to meet special situations. The RM5000 is a heavy-duty process pump with a centerline mounted refinery type pump wet end. This design offers higher head and capacity ranges along with optional flow paths. The containment shell is gasket sealed with no o-rings. The RMA5000 is a high temperature variation that can be air cooled to pumpage temperatures of 750ºF. It features the external-external flush system and uses no o-rings. Klaus Union. A wide variety of mag drive pumps conforming to ANSI, API, and DIN specifications is available from Klaus Union. They accommodate heads up to 575 ft, temperatures from -300°F to 840°F and design pressures to 5800 psig. Multistage high pressure pumps offer total delivery heads to 3300 ft. A patented double isolation shell is noteworthy among its offerings in addition to a pressure switch for continuos monitoring. Standard products offer designs geared to toxic

duty, high temperature, high pressure and slurry applications. Products are available in single and multistage horizontal, vertical and even screw pump configurations. Additional heating and cooling is provided by jackets or coils, and forged casings and special isolation shells are used in high pressure designs.

PUTTING IT ALL TOGETHER In Michael "Todd" Stevens' experience "...every mag drive pump failure has been the fault of some system design or operation upset. The mag drive pumps, when operated properly, have been very reliable." So how can one ensure the best continued operation of magnetic drive pumps? The answer is not necessarily simple. First, Rob Plummer of Dean Pumps suggests that all mechanical seal problems really need to be resolved before even considering the purchase of a mag drive pump. His reasoning is that most of the materials used in mag drive pumps are the same used in mechanical seals — just in a different manner. Furthermore, a mechanical seal is tolerant of certain design features such as an

elbow on the suction of an end suction pump. A mag drive pump installation, on the other hand, requires a straight pipe with smooth flow. And of course there are the dry-running and eddy current heat problems that we have discussed. Having considered these issues and still electing to install a mag drive pump, one must closely analyze all possible operating conditions (including startup and shutdown) and pass that information on to the supplier. This involves specifying such characteristics as type of chemical(s), specific gravity, viscosity, specific heat, vapor pressure/temperature rise, percentage of solids and dissolved gases. Describe all possible process operating conditions. Protect the pump by using at least a temperature sensor to indicate the fluid heat as it leaves the containment shell and a power monitor to indicate internal operating conditions. Enlist the assistance of maintenance personnel in system design, pump selection, startup, training and general operation. And be sure to check obvious details such as the direction of motor rotation! ■

The Pump Handbook Series

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CENTRIFUGAL PUMPS HANDBOOK

Operation Protection for Mag Drive Pumps Learn how to avoid dry or semi-dry running conditions, which can lead to damage. By Kaz Ooka and Manfred Klein

M

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problems. Sealless pumps are ideal for preventing liquid leakage and mitigating the associated environmental concerns. They also work well in the pumping of liquids that crystallize upon contact with air, and they avoid contamination of process fluids by seal flush liquids or gases. And they are excellent for pumping corrosive liquids. However, several issues need to be addressed, the most important of which are minimum flow conditions and dry-running. While these issues are relevant to traditional sealed pumps as well, this article will concentrate on their effects on mag-drive pumps.

which will not be discussed in this article.

MINIMUM FLOW Two factors determine minimum flow rate: radial or thrust load on the bearings or shaft of a pump and temperature rise. This discussion applies specifically to single stage, low specific speed (400-800 rpm*√gpm/H3/4) mag-drive pumps in the 1-30 hp range. Low flow operations of higher specific speed pumps can lead to additional problems such as suction recirculation,

PHOTO COURTESY OF ANSIMAG, INC.

agnetic drive centrifugal pumps are products of an evolving technology utilizing new materials, stronger magnets and new concepts. Beginning shortly after World War II, the mag-drive concept developed along two paths – namely, metallic and non-metallic pumps. Metallic designs have been utilized primarily as process or heavy duty pumps, notably in Europe. In earlier years, non-metallic mag drives were usually considered applicable in light duty situations only – fish tank pumps drawing 100 watts or less, for example. With the development of rare earth magnet materials such as samarium-cobalt and neodymium-iron-boron, however, the size and power of non-metallic designs have been greatly improved over the last ten years. This rapid increase in magnet strength has allowed for a corresponding reduction in the size and weight of the magnetic coupling. Photo 1 shows an industrial nonmetallic mag-drive pump. This machine is clearly a great improvement over the original fish tank pump. As more magnetic drive pumps have been applied in the process industries to solve increasingly complex problems, some confusion has developed among users accustomed to sealed centrifugal pumps such as those described in the ANSI/ASME B73.1 standard: Specification for Horizontal End Suction Centrifugal Pumps for Chemical Process. Many users have believed that due to its sealless cocnstruction the magnetic drive pump can solve all seal related

Photo 1. Example of an industrial nonmetallic magnetic drive pump The Pump Handbook Series

Radial Load on the Bearings and Shaft When a standard centrifugal pump operates off its best efficiency point, the impeller experiences higher radial loads due to hydraulic unbalance in the casing. The radial load becomes severe when the pump is operated near shut-off. In a standard sealed pump the loads on the bearings are much higher than the load on the impeller. Typically, they are two times higher. This is a consequence of the long overhang distance between the impeller and the first bearing, which is necessary to provide adequate space for the seal. In contrast, the first bearing of a sealless pump is located very close to the impeller. This results in bearing loads only slightly greater than the impeller loads. Figure 1 shows typical mag-drive pump bearing arrangements. The dimension L denotes the span between the impeller and the first bearing. Each layout has its own strengths and weaknesses, but all provide for a bearing close to the impeller. In Type 3 the bearing rotates with the impeller and consequently is very close to the load. With this design the overhang distance, L, can approach zero. The Type 4 design also has the bearing close to the impeller, but it uses a shaft cantilevered from the containment shell. This design is typically used only in small pumps because of the stresses at the shaft-to-containment shell connection.

Basic Construction

Description Type 1

L

Impeller and magnet separate, connected via rotating shaft. Bearings stationary and supported in a bushing support. 1. Front bearing has higher load capability than type 4. 2. High manufacturing cost and more complex than type 3 and 4. 3. Popular in metallic pumps.

Type 2

L

Impeller and magnet separate, connected via rotating shaft. Stationary front bearing behind impeller. Stationary rear bearing in containment shell. 1. Front bearing has higher load capability than type 1. 2. Longer span between bearings than type 1, 3 and 4. 3. More chance of dry running or vapor lock at rear bearing, since difficult to lubricate.

L

Type 3

Impeller and magnet a single unit. Shaft is stationary and supported at both ends. Bearings rotate with impeller. 1. Front bearing has highest load capability of all 4 types. 2. Simple design 3. Lower stress concentration on plastic parts than type 4. 4. Non-clogging fluid paths. 5. Popular in non-metallic pumps.

L

Type 4

Impeller and magnet a single unit. Stationary cantilever shaft. Bearings rotate with impeller. 1. Simple design 2. Non-clog fluid paths. 3. Containment shell (rear casing) supports shaft bending moment. Shaft socket in containment shell must be very strong and resist creep if non-metallic.

• Black squares represent the mounting locations of the stationary bearings (type 1 and 2) or the mounting locations of the stationary shaft (type 3 and 4). • Red rectangles represent product lubricated bearing locations (type 1-4).

A mag-drive pump, therefore, has a significant advantage in terms of impeller deflection, and it is more resistant to the radial loads encountered during low flow operation. For example, an impeller deflection of 0.005" can be a problem for seal life but is not a concern in most sealless pump designs. It is important that mag-drive pumps operating at low flows be designed with bearings able to handle consistently higher radial loads. The design must also provide for adequate cooling and lubrication flow to the bearings at these low flow rates. One reason for the increasing acceptance of mag-drive pumps in low flow rate service is that product lubricated bearings can

be manufactured from materials such as pure sintered silicon carbide that effectively provide zero wear for the life of the pump.

must include all losses. The three types of efficiency loss in magnetic drive pumps are: (1) hydraulic, leakage and friction losses, (2) radial and thrust bearing friction, and (3) eddy currents in the containment shell (rear casing). (1) Hydraulic losses are inherent in all centrifugal pumps. However, the efficiency of impellers is improving with better design, and, with the use of more powerful magnets, the size of the inner magnet assemblies and their resultant fluid friction losses are shrinking. Remember, though, that since the efficiency will always be zero at shut-off, the temperature rise will rapidly increase as the flow rate approaches zero. (2) Radial and thrust bearing friction account for the smallest portion of efficiency losses. For example, a 1 1/2 x1x6 ANSI pump will lose 0.1-0.26 hp to bearing friction. This represents only 3-9% of the shut-off power of the pump. (3) Eddy current losses. In metallic magnetic drive pumps the containment shells are usually made of a nickel alloy (e.g., Hastelloy®) or stainless steel. Both are electrical conductors. These stationary shells are placed between the two sets of rotating magnets within their powerful magnetic fields. When a magnetic field moves past a conductor such as a containment shell, eddy currents are generated. Generally, the eddy current loss for a 0.060”

Temperature Rise The temperature rise (ºF) of the liquid passing through a pump is given by ∆T= H 778˙Cp˙η where H is head in feet, Cp is specific heat in Btu/lb ºF, and h is efficiency, written as a decimal value. This equation assumes that all losses result in heat that remains in the liquid. To predict the temperature rise in the fluid accurately, the efficiency factor in this equation The Pump Handbook Series

Photo 2. A Zirconia containment shell

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168

10

180

9

160

8

140

7

120

6

100

5

80

4

60

3

40

2

20

1

0

Efficiency (% x 10) Power (hp)

TDH (Feet)

200

0 0

50

100

150

Capacity (U.S. Gallons) Figure 2. Performance curves for ANSI 1 1/2 x 1 x 6 size non-metallic magnetic drive pump Temperature Rise in a K1516 Pump @ 3500 rpm

25 20

Water 50% Sodium Hydroxide 37% Hydrochloric Acid

15

98% Sulfuric Acid 10 5

0

0

5

10

15

Flow Rate, gpm Figure 3. Temperature rise in same ANSI pump 500

CP = 1.0 Btu/lb°F (Water)

400

300

TDH (Feet)

Dry and Semi-dry Running Dry running and semi-dry running are the most common causes of failures in magnetic drive pumps. Damage can be caused by excessive heating in metallic containment shells and/or by mechanical or thermal shock at the bearings and shaft. If a metallic containment shell is used and a pump runs dry, the shell will be rapidly heated by eddy currents, with temperatures rising to nearly 1000ºF. This temperature rise is so quick that even jogging the pump (e.g., to check rotation) is not recommended

K1516 Performance Curve 3500 RPM, S.G. 1.00

Temperature Rise, °F

thick nickel alloy containment shell is about 15% of the magnetic coupling rating. If a 10 hp coupling is used, about 1.5 hp or 1.1 kW of power is directly transferred to the liquid at the containment shell. This would be equivalent to equipping the pump with a 1.1 kW heater. Additionally, the heat generated in the containment shell remains essentially constant regardless of pump flow rate. This 1.1kW can increase the temperature of water flowing at 1 gpm by approximately 7.5ºF. Cooling of the containment shell and prevention of flashing are the primary constraining factors for metallic mag-drive pumps at low flow rates. Performance curves for a nonmetallic mag-drive pump are shown i n F i g u r e 2 . The efficiency curve includes all losses. As with all pump performance curves, the efficiency is zero at zero flow. If we take values from the efficiency and TDH curves and insert them into the temperature rise equation, given above, we can calculate the temperature rise for this pump. Figure 3 illustrates the temperature rise for flow rates of 020 gpm. If a maximum temperature rise of 10ºF is specified, then this pump can be operated at 2 gpm for water. Slightly higher minimum flow rates may be required for other liquids due to their lower heat capacities. Figure 4 provides the minimum continuous flow rates for a series of three non-metallic mag-drive pumps. These flows are based on a temperature rise limit of 10ºF for pumping water.

Continous Operation Region

200

100

0 0

2

4

6

8

10

Capacity (U.S. Gallons) Figure 4. Minimum continuous flow for non-metallic magnetic drive pumps The Pump Handbook Series

20

Temperature (F°)

250

SiC Bushing

Carbon Bushing

200 150 100 50 0

5

10

15

20

25

30

35

Time (min) Figure 5. Shaft temperature U.S. time

when there is no liquid in the pump. A few minutes would be long enough to demagnetize the magnets completely and ruin the shell. Non-metallic containment shells generate no eddy currents and therefore no heat, and so do not experience this kind of failure. Some manufacturers provide non-metallic containment shells for their metallic pumps. (Photo 2.) Typical materials of construction are zirconia ceramics or plastics. The critical component with respect to dry running in pumps with non-metallic containment shells are the bearings. The bearings are designed to operate while wetted by the pumpage and will exhibit a very low coefficient of friction in this condition. Some material combinations such as a carbon bushing on sintered silicon carbide will maintain a relatively low coefficient of friction even when dry. Such combinations are thus more forgiving in instances of dry running. Sintered silicon carbide against sintered silicon carbide is the best bearing material when wet, but it will typically show a sudden increase in friction level when bone dry. The resultant increase in shaft temperature is shown in Figure 5. The temperatures were measured using a non-metallic mag-drive in a broken suction application. Since the carbon/graphite bushing has a lower coefficient of friction when dry, the pump shaft temperature increase gradually. The silicon carbide bushing exhibits the same temperature rise in the beginning, but then an almost instantaneous temperature rise will occur. Both material combinations provide plenty of time to shut down the pump before damage

occurs provided proper monitoring equipment has been installed. During dry running the pump bearings and shaft will become hot. If while hot, cool liquid is reintroduced to the pump, the bearings and shaft may fail due to thermal shock. Different ceramic materials have far different thermal shock limits. For example, alumina ceramic can survive only a 200ºF thermal shock while sintered silicon carbide is safe up to 600ºF. Silicon carbide can resist thermal shock because it has a very low coefficient of thermal expansion combined with a very high thermal conductivity. This allows the material to equalize in temperature very quickly while undergoing very small thermal strains. Some manufacturers provide other bushing/shaft combinations. However, in an actual service such as unloading, the coefficient of friction is unpredictable. In a laboratory test, the carbon bushing/silicon carbide shaft combination allows extended periods of dry running if the bushing is new and clean, or if the pump suction and discharge are open and air flows freely to cool the shaft. In practical operation, dry running is not recommended in magnetic drive pumps because a) the majority of magnetic drive pumps are still using metallic containment shells, b) residue on the shaft and bushing from a previous pumping operation can change the friction coefficient, c) abrasive matter in the fluid can alter the bushing and shaft surface finish, d) vibration from other equipment may match natural frequency of the pump rotating parts, and e) piping usually restricts cooling by air flow through pump. Monitoring, which is strongly recommended, can include a pressure switch, flow switch, electrical current monitor or electrical power monitor. A

The Pump Handbook Series

pressure switch or flow switch is reliable as long as liquid is clean and extra electrical wiring in the pump field is feasible. An electrical current monitor or electrical power monitor is very popular since no extra wiring into the pump field is required and the device can be easily placed outside the hazardous area. Exercise caution if the motor is selected for the pump’s maximum requirements but actual operation is at a very low flow rate. When a motor runs at less than 50% of rated load, current monitoring will not be sufficiently sensitive. This is due to the characteristics of a motor. However, motor input power remains sensitive below 50% of rated load. For this reason sensing the performance changes between a pump operating at a low flow condition and the same pump running dry will require a power monitor. These devices can be conveniently installed at the motor starter box. However, power monitoring may not be sufficiently sensitive to detect the changes between very low flow operation and shut-off. These situations will require the use of a flowmeter or differential fluid temperature measurement between the pump inlet and discharge.

CONCLUSION It is essential for pump users to determine if the minimum flow rate specified by the pump manufacturer is constrained by a radial load on an impeller or by heat rise. In the case of a radial load limitation, the specific gravity of the pumped liquid must be taken into account. If heat rise is the major limiting factor, the specific heat of liquid must be taken into consideration. These two factors are unrelated. ■ Kaz Ooka is President and cofounder of Ansimag, Inc. He has a B.S. in Mechanical Engineering from Tokyo Denki University. Manfred Klein is Chief Design Engineer for Ansimag, Inc.. He holds a Bachelors and Masters degree in Mechanical Engineering from Carleton University in Ottawa, Canada.

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CENTRIFUGAL PUMPS HANDBOOK

Sealless Options Optimize Solutions If zero leakage is the goal, sealless pump options can help tailor-make the solution. By Robert C. Waterbury, Senior Editor

eak free? Zero emissions? Hermetic sealing? When environmental protection needs and hazardous substances raise questions, sealless centrifugal pumps often provide the answers. The term "sealless" generally describes a class of pumps that do not allow fluid leakage into the environment. And although this de-scription covers a number of pump types, the two most prominent examples are the canned motor pump (CMP) and the magnetic drive pump (MDP). According to David Carr, marketing specialist at Sundstrand Fluid Handling, neither the CMP nor MDP requires a dynamic shaft seal to contain the pumped fluid. Instead, a stationary containment shroud isolates the pumpage from the ambient environment. In the case of the magnetic drive pump, power is transmitted across the stationary shroud through magnetic lines of flux. These induce rotation in the impeller shaft and thus avoid the leakage that is a normal byproduct of mechanical face seals.

The synchronous type typically uses rare earth magnets. Because these can be adversely affected by temperatures in excess of 400ºF, special auxiliary cooling provisions are often required for such applications. The eddy current designs, according to the Kontro Co. of Sundstrand Fluid Handling, employ a torque ring that is normally unaffected by temperatures experienced in hot oil heat transfer systems. Eddy current sealless pumps feature a rotating assembly that is sealed by a containment shell. Power is transmitted through permanent magnets mechanically coupled to the driver, rather than through motor windings. Cooling water is not required because the outer magnets and antifriction bearings are remotely located. A torque ring integrally connected to the impeller shaft forms the inner rotating element. This assembly is supported by journal bearings that are lubricated using recirculating hot oil. No precooling is required. The recirculated flow also removes the heat generated from the magnetic coupling losses.

MAGNETIC DRIVE PUMPS

CANNED MOTOR PUMPS

Magnetic drive pumps are simply centrifugal pumps with an integral magnetic coupling between the driver and the liquid end. The magnetic coupling replaces the seal chamber or stuffing box so that the liquid end is hermetically sealed. The mechanical seal or packing is eliminated, and the only seal is a stationary gasket or O-ring. The two main subgroups of magnetic drive pumps are the synchronous and eddy current designs.

The canned motor pump consists of an induction motor whose rotor is integral with the impeller shaft. A thin metallic can mechanically separates the rotor from the windings and seals the pumpage from the stator. The rotor is supported by journal bearings that are lubricated by recirculated hot oil. Recirculation is provided by an external tube that feeds hot oil into the back end of the pump. CMP features may include a

L

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The Pump Handbook Series

Photo 1. Chempump NC Series canned motor pumps feature an electronic diagnostic system.

two-bearing single shaft arrangement, dry stator and sealed junction box, both primary fluid containment and secondary leak containment shell, a controlled bearing operating environment with monitor and a minimum of required components. The benefits or advantages of using CMPs include: no shaft seals and no external leak paths, no buffer pots, no buffer or process fluid leakage disposal, no coupling or alignment problems, low noise levels and low maintenance costs. Because high temperatures routinely encountered in hot oil systems normally exceed the limits of the motor, CMPs are designed with an integral cooling water heat exchanger. This exchanger surrounds the outside wall of the stator and removes the heat associated with motor losses. The recirculation

tube also makes a single pass inside the shell of the exchanger. Thus, the recirculated pumpage is precooled to a safe temperature before entering the back end of the pump once again.

CMP AND MDP FEATURES AND OPTIONS The first and most obvious job in selecting a sealless pump is to determine the basic system requirements and operating parameters. Operating criteria such as high or low temperatures, fluid capacities in gpm, system pressures, abrasive slurries and corrosive agents will determine not only the size and capacity of the sealless pump, but also any special construction requirements. Once these criteria are identified, it is possible to look at the comparative features and options offered by canned motor pumps and magnetic drive pumps. Much of the following material is adapted from information compiled by Sundstrand Fluid Handling. Containment. This is a major consideration if there are health and/or safety concerns. Several CMP manufacturers offer some type of double containment. This might

consist of welded primary containment and hermetically sealed secondary containment. For efficiency reasons few MDP manufacturers offer secondary containment. Most MDP manufacturers, however, do offer secondary control in the form of mechanical seals on the OMR shaft penetration. Also, an MDP typically provides a thicker containment shell. This offers more resistance to corrosive or mechanical penetration. CMP primary liner thicknesses normally range from 0.022"-0.035". MDP containment shells range from 0.029"-0.060". These differences are offset somewhat by the ability to monitor bearing and internal rotor positions. CMPs, by design, are easier to monitor than MDPs. This means that they are perhaps more apt to detect extreme bearing wear prior to containment shell contact or penetration. Solids/slurry handling. Standard CMP and MDP products will accommodate moderate amounts of solids. Optional designs for both are capable of handling higher slurry concentrations. With a slurry modification, CMPs can deal with concentrations in the range handled by most standard centrifugal pumps. The MDP, however, cannot be iso®

Optional INSIGHT Bearing

Silicon Carbide Bearing with Carbon Graphite Option

Optional Secondary Containment

Monitor and Protection Systems Replaceable Non-Sparking Bump Ring Synchronous Drive Design Utilizing Samarium Cobalt Magnets

Bearing Isolator

Dual Power- End Pull Out Design Replacable Casing Wear Rings

Fully Contained Magnets Single Casing Gasket

Hastelloy “C” Containment Shell

NOTE: Liquid end internals utilize Kontro’s unique cartridge system allowing fast, easy maintenance and minimum downtime.

Figure 1. Today's generation of magnetic drive pumps offers advanced features and materials of construction, as shown in this example of Kontro's A-Range design.

The Pump Handbook Series

lated as easily as a CMP and thus requires a different internal bearingmag coupling flow path than offered normally. Heat input. Heat is added by hydraulic and drive inefficiencies in either type of sealless pump. For our purposes we can consider the hydraulic efficiencies to be the same for both types. Thus, for CMP design, a high efficiency motor is 80-85% efficient. But due to the ease in isolating the motor area, CMPs can offer optional configurations to control the fluid temperature, pressure or both to prevent product vaporization. However, because a CMP motor tends to be one large insulated mass once the unit is shut down, it allows heat soak to occur. This permits the process fluid to be heated to higher temperatures at potentially lower pressures than during normal operation. The result could lead to flashing of the contained fluid. Also, the pump may vapor lock if restarted. Synchronous MDPs are known to be adversely affected by temperatures above 400ºF. But a synchronous drive is also more efficient than an eddy current drive. A nonmetallic shell is more efficient than a metallic shell. The synchronous drive MDP with a metallic shell has a drive efficiency of 80-85%, which is comparable to the CMP. However, it lacks the insulated mass around the containment shell and is less susceptible to heat soak. Cooling requirements. Both CMP and MDP can be configured for operation at elevated temperatures. Different designs require temperature limitations based upon specific components and therefore must be evaluated on a case-by-case basis. Jacketing. Steam or hot oil jackets can be added to either type of pump to maintain the proper product temperature. This ensures that the pumpage will be liquid at all times. CMP designs allow jacketing of the pump case, stator and rear bearing housing. Most MDPs can have jacketed casings that add heat in the area of the containment shell without fully encapsulating it.

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High suction pressure. Optional designs are available for both CMPs and MDPs in applications where the suction pressure and maximum allowable working pressure requirements exceed standard pressure design capability. Modifications in the CMP design for high pressure applications include the use of primary containment shell backing rings, thicker secondary containment shells, additional pressure containment bolting and high pressure terminal plates. The double containment feature of CMPs is maintained even for high pressure. Modifying the MDP for high pressure applications includes use of a thicker containment shell and additional pressure containment bolting. Typically, the CMP is more efficient than the MDP in high pressure applications due to its increased primary containment thickness.With thicker primary containment, the MDP shows greater hysteresis losses (inefficiency) than the CMP. ANSI conformance. Although some CMPs are available with ANSI dimensions or hydraulics, most are not. Many MDPs, however, offer ANSI dimensions and hydraulics. Further, most manufacturers of sealed and sealless ANSI pumps offer interchangeability between their pumps' wet ends and bearing frames. Efficiency. CMP wire-to-water efficiency is defined as hydraulic efficiency times motor efficiency. As mentioned in the discussion of heat input, a typical CMP motor is 8085% efficient. MDP wire-to-water efficiency is defined as hydraulic efficiency times motor efficiency times coupling efficiency. Again, the magnetic coupling is 80-85% efficient. However, containment shell metallurgy (or lack thereof) and magnetic coupling type play major roles in coupling efficiency. Wet end hardware, however, further complicates the efficiency discussion. Impeller and casing geometry play a vital role in hydraulic efficiency. A Barske design (open radial blade impeller and diffuser

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discharge) is typically more efficient in low flow/high head hydraulics. A Francis design impeller, enclosed with backswept vanes and an increasing radius volute, is more efficient at moderate-to-high flows with low-to-medium heads. The user, therefore, must thoroughly evaluate wire-to-water efficiency for an accurate comparison. Internal clearances. Clearances between bearing ID and mating surfaces are typically the same (0.0030.007"), although differences between other rotating parts do occur. Typical CMP clearances between the rotor and stator liners vary by manufacturer from 0.018-0.044" radially. MDP clearances between the inner magnetic ring and containment shell are usually 0.030-0.045" radially. This larger clearance gives the MDP the advantage of allowing more bearing wear prior to containment shell contact. Bearing monitoring. Although different monitoring methods are available for MDPs and CMPs, the design of the CMP lends itself more readily to real bearing monitoring. A CMP bearing monitor can provide axial, radial and liner corrosive wear indication. Bearing monitoring features vary by manufacturer, and care must be exercised in selecting a sealless pump manufacturer with the desired monitoring features. Bearing material options. Most sealless manufacturers offer a hard bearing – typically silicon carbide – running on silicon carbide or tungsten carbide. Alternatively, a soft bearing is normally carbon graphite running on stainless steel. Because the materials are the same, the mounting and lubrication plans assume greater importance. Number of bearings. CMPs have two bearings; MDPs have six (including the motor). Most experts feel that fewer is better. However, a properly designed, applied and operated MDP will last just as long as a properly designed, applied and operated CMP. Noise. CMPs have no motor fan and thus produce less noise. Space. A CMP with its integral pump and motor occupies less real The Pump Handbook Series

estate (has a smaller footprint) than a comparable MDP with its pump, coupling and motor. However, close coupled MDPs that may require no more space than a CMP are available. Orientation. CMPs can be mounted vertically or even hung on a pipe. Few MDPs can be mounted vertically. Temperature fluctuations. Both CMPs and MDPs can effectively handle wide temperature variations in the pumpage. Manufacturers should be consulted in evaluating specific operating conditions, however. Initial cost. For general duty service (clean, cool, non-volatile) an MDP is usually lower in price than a CMP. As the application becomes more tortuous and difficult, however, pricing reaches parity and may even favor the CMP. Installation cost. A standard CMP or close coupled MDP with no auxiliaries (coolers, flush systems, etc.) will have lower installation costs than a frame mounted MDP. This is due to the smaller footprint and minimal foundation requirements.When optional CMP configurations are considered, the cost of cooling lines, reverse circulation lines and flush lines may equal the cost of an MDP. Instrumentation also adds cost to both types of installation, so situations must be evaluated individually. Simplicity. Although the number of pump parts may vary by manufacturer, CMPs generally have fewer parts. As optional configurations and more parts are added, however, they can easily equal the number of parts in an MDP. Items adding parts and complexity include auxiliary impellers, tilting washers and heat exchangers. While some MDP manufacturers may add such equipment, most offer only two basic varieties: discharge-to-suction recirculation and discharge-to-discharge recirculation. Adding a standard electric induction motor as well helps some users accept the sealless technology. Field serviceability. CMPs have bearing monitors to predict the

need for bearing changeout prior to containment breach. CMP bearings are typically easy to replace. MDPs have thicker containment shells to prevent breach. However, they have limited ability to monitor lubricated bearings to determine when routine maintenance is required. Ease of bearing replacement also varies with manufacturer. If a primary containment shell is breached, even the most serviceable CMP must be sent outside to be "recanned." CMP manufacturers have countered by offering spare rotor/stator sets at 60-80% of the cost of a new pump. A containment shell breach of an MDP usually results in the purchase of a spare process-wetted rotating assembly and a containment shell. Pricing is typically 60-80% of the cost of a new pump. Coupling alignment and venting. The typical CMP design has the impeller mounted directly on the motor shaft inside the containment area. Coupling and coupling alignment problems are therefore nonexistent. Most MDP installations use frame mounted motors that require coupling alignment. Many MDP suppliers, however, offer close-coupled designs that nearly eliminate coupling alignment problems. Both MDP and CMP designs are typically self-venting back to the process piping and do not require additional external lines.

PRODUCT OFFERINGS Because sealless pumps are often used in severe duty applications, monitoring and diagnostic options are important aspects of many installations. Monitoring and diagnostics. A monitoring and diagnostic system called IntelliSense from Crane Chempump displays the position of the entire rotating assembly in its NC Series of canned motor pumps. The system provides precise, realtime wear data with an accuracy of 0.001" radial and 0.002" axial. This information enables users to plan simple parts replacement long before costly failure occurs. The

entire diagnostic system is isolated from the process fluid and is therefore not a sacrificial part that requires replacement. The display unit can be hand held, mounted near the pump at eye level for easy viewing, or located in a remote control station. Teikoku canned motor pumps offer the Teikoku Rotary Guardian (TRG), which is available either as a built-in meter or as a remote panel meter. It monitors the running clearances between the stator and rotor, bearing condition and rate of wear, reverse rotation, loss of phase and short circuit conditions. Sundyne offers a remote mechanical bearing monitor. It allows remote alarm or shutdown of the pump signalling a need for bearing maintenance to protect the motor from damage. A dry operation protection meter is also available. In services such as tank unloading, the meter detects low load in time to shut down the motor to prevent dry operation. A similar system called INsight from Kontro provides radial and axial bearing wear monitoring in its line of magnetic drive pumps. Kontro also offers a port for thermocouple or RTD temperature monitoring, an amperage monitoring system to protect against cavitation or dry running, a sensor port for vibration monitoring as well as liquid and pressure sensing options. Jacketing. Complete or partial jacketing of the pump case, motor stator and rear bearing housing is offered by Sundyne to control temperature when heating or cooling is required. Teikoku and other manufacturers offer certain pump lines with built-in heat exchangers as well as motor cooling jackets for high temperature applications. Crane Chempump offers three lines of high temperature CMPs: • GH – external circulation for fluid temperatures to 650ºF without liquid cooling • CH – internal circulation for fluid temperatures to 650ºF without liquid cooling • GT – motor isolation with exThe Pump Handbook Series

ternal cooling for fluid temperatures to 1000ºF The motors in the GH and CH models incorporate a high-temperature insulation system. The motor in the GT model employs an integral liquid-cooled heat exchanger. Inducers. Teikoku, Sundyne, Crane Chempump and other manufacturers offer a wide selection of inducers to meet net positive suction head (NPSH) requirements. It is not unusual, according to Teikoku, for one of its pumps to operate at a specific suction speed of more than 13,000. Casings and impellers. Single and double volute casings are options used in CMPs from Buffalo Can-O-Matic and other manufacturers, depending upon pump size and service requirements. Likewise, open and closed impeller designs can be specified as options on many models depending upon operating requirements. Bearings. In addition to various manufacturer options, there are special bearing constructions to take into account. Buffalo Can-O-Matic, for instance, features spring-loaded, self-adjusting and self-lubricating tapered carbon graphite motor bearings. These are designed to distribute and automatically compensate for bearing wear. This maintains concentric rotation of the rotorimpeller assembly, prevents mechanical contact between the rotor can and stator can, and improves both operation and maintenance. Segmented bearings of selflubricating carbon gra-phite construction are used on large CanO-Matic pump motors. Again, they are spring loaded to compensate for wear. Other construction options. Most manufacturers offer optional construction materials for certain components. Buffalo Pumps, for instance, provides casings in a choice of ductile iron or 316 stainless steel. Likewise, impellers are available in a choice of cast iron, bronze or 316 stainless steel according to the requirements of the application. Chempump lists hardened rotor

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journals, pressurized circulation systems, sealless junction boxes and explosion-proof CMP designs among the options in its G Series line. And Teikoku provides a large number of adapters to accommodate many different pump and motor combinations.

DRIVING FORCES Environmental protection and zero emissions of hazardous substances continue to drive the use of sealless CMP and MDP technology.

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Add to that, however, the philosophy of continuous process improvement and reduced maintenance, and there is ample reason to expect that sealless pump technology has a bright future. It is appropriate not only for new installations, but also for many retrofit applications. And standardized ANSI dimensions can help make sealless technology a highly cost-effective alternative as well. ■

The Pump Handbook Series

CENTRIFUGAL PUMPS HANDBOOK

Vertical Turbine Pumps Power Petrochemicals Verticals are a popular choice for low NPSH applications, versatility of construction and minimal floorspace requirements. By Herman A. J. Greutink

P

Deepwell pumps (Figure 1) are commonly used to raise water from underground aquifers. Lineshaft pumps are either oil lubricated or water lubricated, and they are built mainly to AWWA standards. Submersible motor driven pumps are also used. Materials of construction are mainly steel or cast iron for heads and bowls and bronze for impellers. Service water pumps (Figures 2 and 3), used to pump from ponds,

lakes, rivers or oceans, are generally larger than deep well pumps. Pumps intended for fresh water intake have steel columns and heads, cast iron bowls and bronze impellers. For pumping brackish or sea water, coated standard materials are normally used. Experience dictates whether coated standard materials will offer acceptable life. Otherwise, stainless steels (316, 316L, duplex stainless) or nickel aluminum bronze may be specified.

PHOTO COURTESY OF JOHNSTON PUMPS, INC.

etrochemical plants need water, whether from local water systems, deepwell pumps, rivers, lakes or oceans. The vertical diffuser pump normally plays a major role in providing service water, cooling water and occasionally fire pump service. This type of pump is also used in oil field production as well as oil field pressurization. Process fluids ranging from crude oil to liquified petroleum gas and other liquids (sulphur, for example) are moved by vertical pumps in one or more stages of extraction or production. Frequently, liquified petroleum gas, propane, butane and anhydrous ammonia are supplied from underground caverns. In-plant pumps with very low NPSHA are also apt to be the vertical turbine type. To economize on length of barrel or can and reduce installation and pump costs, many vertical pumps are supplied with a first stage low NPSHR impeller. Some users require that the suction specific speed be limited to perhaps 11,000. This limitation may be highly important for certain types of pumps, but if a properly designed impeller is used in a vertical turbine pump application, experience shows that the suction specific speed can range up to 15,000 without creating problems. Range of operation over the pump curve and recirculation possibilities must be considered.

SUPPLY WATER PUMPS Following are brief descriptions of various vertical diffuser pumps used in the petrochemical field:

A 500 hp, 1800 rpm product pump operates in a midwest chemical plant. The Pump Handbook Series

175

Driver

Driver Driver

Head Assy.

Column Assy.

Motor Stand and Tube Tension Nut Assy.

Discharge Head Assy.

Column Assembly With Below Ground Discharge

Column Assy.

Water Level

Bowl Assy.

Bowl Assy.

Figure 1. Variation on water lubricated deep well turbine pump

Figure 2. Supply and drainage pump, axial flow (propeller) from 5'-20' of head

Vertical cooling tower and plant supply water pumps (Figure 4) follow fresh water material standards and construction. Lineshaft pumps for moving water are normally built with a packing box for sealing. Lately, however, there have been requests for water pumps built with mechanical seals. Perhaps this is due to packing box maintenance requirements. There is an art to using packing boxes properly (they need to leak some water). Of

course, mechanical seals also need proper maintenance and installation. But if either type of sealing technology is used and applied properly, maintenance labor can be greatly reduced. Vertical turbine pumps used in re-pressurization of oil fields pump produced water at high pressure – up to approximately 4500 psi. Because this water is sometimes very corrosive, stainless steel bowl assemblies are commonly used, and

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The Pump Handbook Series

Bowl Assy.

Figure 3. Mixed flow type service water, plant water, with heads of 20-60' per stage

barrels and heads are made mainly of coated steel (Figure 5). There is an application in which no water will show up at the surface and pressure is throttled off below the surface back to the water supply. This method can be used when pumping from atmospheric or vented bodies of liquids.

PUMPING RAW STOCK AND PRODUCT Most pumps that transport product or raw stock in the petrochemical industry are built to API

Driver Driver

Discharge Head Assy.

Packing Box

Discharge Head Assy. Suction

Discharge

Barrel Water Level Column Assy. Bowl Assy. Bowl Assy.

Figure 4. Service water/cooling tower pump with heads from 50' and up per stage

610. Many users add custom requirements. In general, the specifications are tight, and extra attention to detail must be paid to the pump's proper design and fabrication. Most vertical turbine pumps in these applications are built as barrel or can pumps (Figure 6) with mechanical seals. Pressure containment construction conforms to ASME section VIII. Sealing and bearing clearance specifications also must be followed closely. The mechanical seal

Figure 5. Oil field pressurization pump. High pressure multiple pumps (barrel or can) can be used in series.

configuration depends on the product pumped and in many cases must adhere to strict environmental protection rules. Most of these pumps have motor drives with thrust bearings in the motor, although some requirements call for the European style thrust bearing in the pump. The reason for the latter is better mechanical seal maintenance with less run-out. However, this increases total head and motor assembly, which could aggravate vibration problems. It also increases maintenance difficulty as more parts must be disassembled to get to the mechanical seal. Pipeline pumps may be horizontal multistage pumps, but boosters

The Pump Handbook Series

from the tank farms to the pipeline pump are preferably vertical can or barrel pumps (Figure 7). These offer best utilization of storage capacity. The available NPSH can be very low – putting the first stage at a level where the tankage can be pumped empty to zero NPSH available and at the same time supply the necessary NPSH through the vertical barrel or can type pump to the horizontal pipeline pump. High pressure pipeline pumps also can be built economically using vertical multistage pumps up to approximately 2000 hp. Extremely high or low temperature liquid applications are designed to individual requirements. For instance, sulphur is handled by vertical turbine pumps with steam or hot oil jacketing from top to bottom of the pump. Hot oil circulating pumps (for heat transfer) are built to withstand forces caused by temperature differentials and pipe loading. The pump specifier should remember that material requirements per API 610 are aimed mainly at horizontal centrifugal pumps. The requirement for cast steel is all right for most horizontal pumps, but for turbine pump bowls and impellers it can create problems that are expensive to correct. A better way generally is to make the bowls and impellers out of stainless steel, especially the 300 series and duplex stainless. Practically all vertical turbine pumps are made to order. Off-theshelf items consist only of standard bowls, impellers, cast iron heads and some smaller pieces like threaded shaft couplings. As a result, communications among user, consultant and manufacturer must be very good to get the best equipment for the application.

CONSTRUCTION OPTIONS As mentioned, construction materials vary according to fluid characteristics. Materials are often selected according to the corrosiveness and abrasiveness of the pumped liquid. The following are some construction variations.

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Driver

Driver

Flanged Coupling Discharge Head Assy.

Suction

Mechanical Seal

Mechanical Seal Flanged Column

Discharge

Barrel Suction Barrel with Below Base Suction Bowl Assy.

Bowl Assy.

Figure 6. High pressure LPG pump (barrel or can)

Below-base discharge and/or suction is one option to the more prevalent above base-discharge design. The figures show variations in packing boxes and mechanical seals. Product lubricated lineshaft bearings can be used as an option to oil lubricated or clean water flushed lineshaft bearings. Rubber bearings and bronze bearings are used most often. Rubber is used successfully on hard shafting or hard-faced shafting in water with abrasives such as sand. Do not use rubber bearings in petrochemicals that attack rubber. We also see the use of metal-filled graphite bearings in light ends and

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Figure 7. Pipeline or pipeline booster pump

condensate pumps. Where abrasives are present and rubber cannot be used, hard bearings (from nitronic 60 to carbide) can be used on properly selected shafting such as 17-4 pH, nitronic 50 or hard-faced shafting. Cast iron bearings can be used in oils. It is also possible to coat cast iron bearings for corrosion protection in water pumping applications. Constructions shown in the figures can be altered to suit the requirements of the system in which pumps will be used. Turbine pumps

The Pump Handbook Series

have been built to run upside down, horizontally or at an angle. The variations of driver systems and construction systems are endless. Any type of driver with right angle gears, vertical motors (hollow or solid shaft) and submersibles can be used. Obviously, multi-staging allows one to reach the head required at a given capacity (it's like connecting pumps in series). However, complete pumps can be put in series, too. For instance, driver, construction and shaft limitations of a 12" multistage pump may limit a pump to perhaps 1500 psi. If 4500 psi is needed, just put three pumps in series.

INSTALLATION Installing vertical turbine pumps is a matter of keeping all the pieces or sections in good shape. First, all the pieces that have to fit together to make a vertical turbine pump run properly must be manufactured with three basics in mind: straightness (of shafting), concentricity and parallelism. This is true for all the pieces that fit together to build a total pump. If these considerations have been well addressed, one normally should be able to turn the shafting by hand after the impellers have been lifted off the seat. Once this condition has been reached, the pump must be set to turn in the right direction when the power is applied. Most pumps of the vertical turbine, mixed flow and axial flow type turn counter-clockwise when viewed from the top. But some very old units may turn clockwise – so pay attention. For pumps with threaded lineshaft, the motor/driver rotation must be checked before the driver is hooked up to the pump. In the case of slipfits between motor and shaft, remove the piece that forms the slipfit before checking rotation. Otherwise, the pieces may gall and cause serious and expensive damage. Check the pump's performance, vibration and runout when it is put in service. At times, runout can also be checked before running. In most cases the pump will operate satisfactorily, but sometimes problems that require immediate attention arise.

Excess runout often can be corrected by changing coupling parts 90 or 180º (where bolted couplings are used). It is more difficult to correct if the shaft is bent or some registers are not true or some faces are not flat. Check driver base and shaft and go from top to bottom. Also, dirt between butting faces frequently causes major problems! Vibration. Considerable runout may be accompanied by vibration. If there is no runout problem, however, it is necessary to determine what causes the vibration. First disconnect the driver from the pump and run the driver. If it does not vibrate, then check the pump. If the driver still vibrates, do the following: push the stop button or (in case of gear drive) slow down the engine drive and check vibration while slowing down. If vibration diminishes in line with speed, it is a balancing problem with the driver. Get it rebalanced. If vibration immediately disappears when slowing down, it is probably a critical speed problem, which may have several causes: 1) the design is such that the driver/head assembly has a critical equal or close to the operating speed, 2) pipe loading, 3) foundation loading. In the case of electric motors, an electric unbalance between leads can cause vibration, but this is rare. Try to find out what 2) and 3) are doing by just loosening the anchor bolts a couple of turns. Many pumps exhibit 2) and 3) as

problems. If so, it may be necessary to rework the piping connection. In any case, if the vibration problem is not easily solved, get the pump manufacturer involved. Sometimes a piece of wood or other matter gets into the pump and causes vibration. Maybe it can be backflushed out. And no matter what, the sump and the piping must be clean from the start. When vibration begins to show up after years of running, it's time to pull the pump and replace bearings and repair it as new. Keep in mind the three musts: concentricity, parallelism and straightness. In a particular installation of four identical barrel pumps, one unit persistently vibrated. The motor and even the pump were replaced with still no improvement. The investigation, however, identified that the grout under the barrel flange was totally inadequate for the vibrating pump. In other words, weaker support put the pump in a natural frequency mode equal to the running speed. This has been seen in other installations where pumps were mounted on beams or slabs that happen to lower natural frequencies to the pump speed. When using variable speeds, the pump may be in a natural frequency mode at one of the speeds. If so, it is best to block out that area of speed in the controls. Operation. When pumps are selected, built and installed proper-

The Pump Handbook Series

ly, they will have a reasonable life expectancy. If a pump's life is not up to reasonable expectations, its selection, use, construction and installation must be thoroughly checked. Performance can be changed to suit head capacity requirements; construction and materials can be upgraded as necessary; and bearing systems, including the thrust bearing systems in the driver, can be improved. Many vertical pumps will start in upthrust. So, the bearing system must be built to handle that, and in some cases the pump may be required to run in continuous upthrust. The system can be built for that, too. Sometimes particular pumps continually vibrate or rattle or run out and repairs have to be made frequently. That's when time must be taken to solve the problem properly, whatever the problem is – from flow to the pump, head, capacity and all materials and construction in between!! ■ Herman Greutink is Vice President and Technical Director for Johnston Pump Company (Brookshire, TX). Among other professional affiliations, he is a longstanding member of Texas A&M's International Pump User’s Symposium Advisory Committee and the Hydraulic Institute, and he frequently contributes to Pumps and Systems on the subject of vertical turbine pumps.

179

CENTRIFUGAL PUMPS HANDBOOK

Chopper Pumps Digest The Solids These tough workhorses eliminate plugging problems in heavy-duty applications. By John Hayes

f the hydraulic performance characteristics are right, chopper pumps can be used in any industrial or municipal application that involve pumping solids-laden slurries. They are a cost-effective means of eliminating pump plugging problems and optimizing system perform- ance. Before discussing specific applications for chopper pumps, however, let's look first at the design details that make chopper pumps unique.

I

caught between one of the rotating impeller blades and one of the two stationary shear surfaces that are cast into the cutter bar. These surfaces extend all the way across the intake opening in the cutter bar and divide it into two segments. Positive chopping is required of all solids

DESIGN DETAILS A chopper pump is a centrifugal pump that uses sharpened semi-open centrifugal impeller blades to cut against a stationary bar across the full diameter of the inlet. This bar is known as the "cutter bar," and this style of chopping and pumping is known as "positive chopping." All incoming solids too large to pass through the impeller are chopped prior to entry into the pump, thus eliminating any possibilities of pump clogging. Other items critical to the success of a chopper pump are a severe duty seal-and-bearing system, hardened wear parts and a history of successful pump installations by the manufacturer. Solids cutting by the impeller and cutter bar occurs when: 1. The suction created by the rotating impeller pulls material into the center areas of the pump impeller. To reach the impeller center areas, liquid and the material entrained in the flowstream must pass through the cutter bar openings in the lower suction plate located just below (or ahead of) the impeller. 2. As material passes through the cutter bar openings into the low pressure areas of the impeller, it gets

180

entering the pump prior to pumping. 3. The leading edge of each impeller blade is sharpened and machined flat on the blade face where it runs next to the cutter bar. This forms a cutting edge. The material caught between the impeller blade and cutter bar is severed. The advantage of the chopper is that pumping and chopping are integrated into one efficient system. The chopping is done right where the pumping is done. The material naturally fits through the flow passages of the impeller and casing. If the The Pump Handbook Series

material won't fit through, the pump keeps chopping it until it does. This is an example of positive chopping.

BENEFITS Following are some of the benefits of positive chopping. 1. Positive chopping allows large, troublesome materials to pass through the pump, eliminating downstream plugging of valves, heat exchangers, nozzles or other pumps. 2. A chopper pump often can replace two pieces of equipment, a comminuter (or pregrinder) and a "non-clog" pump. This approach is extremely costeffective because the maintenance costs on comminuters alone can be very high. 3. Chopping material in the pump produces a more homogeneous slurry and reduces pipeline friction. 4. Chopper pumps can handle materials in sumps that no other pump can handle. 5. A severe duty seal-andbearing system that incorporates double row thrust bearings and a mechanical seal reduces down time by handling the heavy workload of chopping and pumping solids reliably. 6. Hard, wear-resistant pump parts hold up to the rigors of chopper pump service. Standard pump impellers and cutter bars made of cast alloy steel heat treated to 550 Brinell provide extended service life in most applications. 7. Selection of a manufacturer with extensive experience in severe duty applications helps assure the user of dependable service. Chopper pumps are not a commodity

item. Every application is unique in its severe nature and duty. Procurement of inexpensive equipment or equipment that is "new" to the market could provide disappointing results. Always ask the supplier about the history of the product and its applications.

WASTEWATER APPLICATIONS Lift Stations. Small residential lift stations are equipped with submersible grinder pumps, and larger lift stations with non-clog pumps. However, lift stations can experience an unusually high concentration of solids such as hair, rags or plastics that cannot be reliably handled by these conventional pumps. If heavy solids loading is anticipated during the engineering stage, many of the larger lift stations include a comminuter ahead of each pump. Chopper pumps have solved many lift station plugging problems and eliminated the need for comminutors. In some instances chopper pumps have directly replaced existing pumps without the need for repiping. Chopper pumps for lift stations can be sized for hydraulic requirements without regard to minimum sphere passing diameters. Generally, non-clog pumps must be sized according to maximum anticipated sphere size, with hydraulic considerations secondary. For example, a requirement of 200 gpm at 130' tdh might normally dictate a 3" pump hydraulically. But a 3" non-clog pump cannot provide the high head and meet a 3" sphere requirement at the same time. Therefore, the engineer and/or manufacturer would be forced to use a 4" pump. This results in higher costs – due not only to the larger pump, but also to lower efficiency and higher power consumption. However, because a chopper pump reduces the size of solids before they enter the pump, sizing of the pump is based mainly on hydraulic requirements, with little consideration given to sphere size. Septage Receiving. Septage is comprised of concentrated solids from septic tanks plus rags and plastics that can plug conventional non-

clog pumps. Chopper pumps eliminate plugging problems by chopping all incoming solids prior to pumping. Septage receiving pit pumping is a very tough challenge. Because conventional non-clog pumps often require oversizing to handle solids, another means of pump protection is to pre-grind all materials with a comminutor upstream of the pump. This is an unnecessary added cost in installation, operation and maintenance when a chopper pump, which is one piece of equipment, does both jobs. Sludge Transfer and Recirculation. One particular problem with sludge pumping is the "roping" of hair and other stringy materials created by prerotation within the piping ahead of non-clog pumps (especially vortex pumps). Chopper pumps eliminate roping by chopping the solids as they enter the pump. Another common problem is pumping grease-and-hair balls or other reformed solids. In digester recirculation, passing a grease-andhair ball from the bottom to the top accomplishes nothing. If the reformed solid is chopped during the recirculation process, then the chopped solids have a high surface to volume ratio and are digested faster. Chopper pumps can handle a higher solids content than conventional non-clog pumps. In normal treatment plant processes this means that sludges can be more concentrated without exceeding the pump's capability. Chopper pumps also reduce downstream pipeline plugging and friction losses associated with high solids content. The chopping of these solids creates smaller solids with sharp edges that tend to scour the inner walls of discharge piping. Friction is reduced due to maintaining full pipe diameter, rather than choking flow with grease build-up on the pipe walls. Chopping and shearing of sludge tend to reduce viscosity, further reducing friction loss. Digester Scum Blankets. Both aerobic and anaerobic digesters tend to form scum blankets when conThe Pump Handbook Series

Photo 1. By slurrying the scum blanket with a recirculating chopper pump, Tony Kucikas of Nut Island WWTP rejuvenated the digester and saved Boston $1.5 million in cleanout costs.

ventional mixing methods are used. These blankets inhibit methane production and eventually require more frequent cleaning of the digester. A vertical chopper pump with a recirculation nozzle and sealed deck plate can be installed through an existing manhole opening in the top of the digester and used to chop and mix the scum blanket. The object is to inject supernatant through an adjustable nozzle into the scum blanket approximately 1 foot below the surface. With the pump located near the periphery and the nozzle aimed at a proper angle to the wall, the action of the nozzle forces the blanket to rotate. Once the blanket is mixed, methane gas production will increase. The pump then can be used intermittently to keep the scum blanket mixed. This can increase digester capacity up to 40% and increase gas production up to 300%. The action of chopping and conditioning material to reduce particle size is of proven value in sewage treatment plants. If digestible material is in smaller particles, then the surface area of these particles is relatively large in comparison to their volume. Bacterial action can then be more effective

181

and rapid. Plants that have used chopper pumps for digester recirculation and/or scum blanket mixing have seen that particle size reduction increases both the rate of decomposition of digestible material and digester gas production. Clarifier Scum. Inherent problems with clarifier scum include plugging and air binding. The chopper pump addresses plugging problems. Air binding, on the other hand, must be addressed indirectly. Inherent recirculation around the chopper pump inlet usually causes enough mixing to keep air binding from occurring in scum pits with short retention times. If a scum pit has a long retention time, then the scum may concentrate and form a blanket on the top. As the pit level is pumped down, this blanket can block the pump suction. In this case, a recirculation nozzle should be used to pre-mix the scum pit prior to pump-out. Also, because chopper pumps have heavy-duty oil bath lubricated bearing and seal systems, the scum pits can be pumped completely down to the pump inlet without damaging the pump bearings or seal due to loss of coolant or lubricant. This allows full scum removal during each pump-down. By adding a motor low current monitoring system in the pump controls, the "OFF" function can be based on low motor current draw rather than an on "OFF" mercury float switch. Once the liquid level drops to the point where air enters the pump inlet, the current draw will drop off, and the low current relay will shut the pump off.

pumped directly into trucks for hauling to feedlots or land application, sometimes first across a separation system to reduce water content. Wood Products. There are many applications within the wood products industry where chopper pumps greatly reduce downtime due to pump clogging. First, raw logs are stacked in sorting yards where rainwater runoff sumps can collect bark, limbs, rocks and large quantities of dirt. This requires a heavy duty pumping system, and is ideal for a chopper pump. Next, the logs might be debarked, leaving a slurry heavily laden with bark, most of which floats. Recirculation nozzle systems added to the chopper pump help to suspend the bark while pumping. Then, the logs are either sawn in a sawmill, or chipped in a pulp mill. Runoff sumps collect the chips and sawdust, requiring further handling of solids too difficult for a standard non-clog pump. Finally, waste pulp or broke is also easily handled by a chopper pump. Hydropulper rejects in wood products recycling are the most

INDUSTRIAL APPLICATIONS Food Processing. Chopper pumps can be applied in almost all food processing waste handling operations. Applications include vegetable waste with whole vegetables and stalks, poultry and turkey parts with feathers and whole birds, beef processing with hair and fleshings, fish carcasses with offal, and any other food processing wastewater. Slurries are generally

182

Photo 2. A vertical chopper pump handles bark, wood, rocks and dirt in a pulp mill log yard stormwater runoff sump.

The Pump Handbook Series

severe pumping applications in all industry. Rejected materials from the recycled bundles can contain wire and plastic strapping, large quantities of plastic sheeting or wrap and wood from pallets – all suspended in a heavy pulp slurry. Due to the heavy nature of these slurries, chopper pumps are generally oversize, and motor horsepower is increased to handle the heavy chopping load. Other Industries. The world of heavy industry contains numerous other special applications for chopper pumps. The steel, chemical, automotive, contractor, mining, sand & gravel, and petrochemical industries all use chopper pumps in applications involving pumping of waste solids with the occasional unknowns. Anytime a waste sump must deal with the unplanned worker's glove, pieces of wood crating, rocks, bottles, glass, cans, plastics or other items larger than the sump pump is designed for, then a chopper pump is applicable.

FINAL CONSIDERATIONS All of the applications discussed here center on the pump's ability to handle solids from a pumping standpoint. However, we must not forget that chopper pumps can also eliminate seal and bearing failures observed in other pumps. Because the chopper design requires heavy shafting, an added benefit is longer life resulting from stronger parts and less vibration. Quite often seal failures in conventional pumps are associated with solids wrapping or binding at the impeller or seal. This can cause severe vibration that is transmitted through the shaft and seals to the lower bearing. This results not only in seal and bearing failure, but also can introduce moisture into a submersible motor. The heavier shafting and short overhang of the chopper pump bearing and seal design addresses this problem and reduces maintenance costs as a result. More industries, municipalities and engineering firms alike are discovering the economics of applying

chopper pumps in applications in which conventional pumps have historically failed. These failures are generally due to plugging or sealand-bearing failure, and all contain hidden costs that must be addressed. The largest fallacy of the "low bid" system is that low initial price does not mean lowest overall cost. More often than not, "low bid" equipment has a higher failure potential than properly specified and purchased equipment. The solution starts with the user's request to obtain equipment that will operate maintenance free, and it ends with the foresight of those with purchasing authority to think toward the future. ■ John Hayes has a BS in Mechanical Engineering Technology from Clemson University, has worked with chopper pumps for over 12 years and is presently the Marketing Manager for Vaughan Co., Inc.

The Pump Handbook Series

183

CENTRIFUGAL PUMPS HANDBOOK

So You Need Pumps For A Revamp! Here are tips for specifying and selecting the right centrifugal pumps.

V

HEAD

SYSTEM HEAD w/out THROTTLING

POWER

➚

POWER

POWER ABSORBED BY THROTTLING

NORMAL

FLOW

Figure 1. Centrifugal pump vs. system with control valve

bility of the pump. In other words, displacement pumps must always be installed with a full capacity relief valve upstream of the first valve in the discharge system. And the relief valve must have an accumulation pressure (rise above cracking pressure to achieve full flow) that keeps the pump's discharge pressure and corresponding power below the maximum allowable.

losses due to conservative assumptions and margins. In most cases these losses are far greater than those resulting from differences in efficiency among various selections for the same duty. Determining the actual system head requires accurate measurement of: • pump flow rate at one condition • pressure at the pump's suction

LEARNING FROM THE EXISTING INSTALLATION The system energy requirements for new units are estimated using various assumptions and margins. When centrifugal pumps are used, the system designer usually relies on a control valve to balance system and pump energy at the desired flow rate. In engineering a revamp, it is possible to determine the actual system characteristic accurately and thus avoid energy The Pump Handbook Series

KINETIC

➞

184

2

SYSTEM HEAD w/ THROTTLING

ENERGY ADDED

➞

A centrifugal pump operates at the capacity determined by the intersection of its head capacity curve and the system's head capacity curve (Figure 1). At this point the energy added by the pump equals the energy required by the system. Note in Figure 1 that the energy required by the system is often increased by throttling across a control valve to regulate the pump's capacity. Also note that this means of flow control is feasible only with kinetic pumps – those that add energy by raising the liquid's velocity. Displacement pumps (Figure 2), deliver essentially a fixed capacity at a given speed and thus add only the energy needed to move that capacity through the system. Care is therefore needed to ensure that this energy never exceeds the mechanical capa-

3

S

PUMP SYSTEM INTERACTION

V

S

S

Z

➚

uccessfully specifying and selecting pumps for a unit revamp requires many of the same disciplines as for a new unit, but with a difference. The difference is that a full-scale working model is available for examination and analysis. Taking advantage of this opportunity can lead to employing pumps that consume less energy and have a longer mean time between repair (MTBR). This, in turn, lowers plant operating costs and can raise plant output, hence revenue, through higher plant availability. In a process unit, pumps move liquid and raise its pressure to allow the process to run, but their role is fundamental and their interaction with the system a critical factor in their performance. This last point, interaction with the system, leads to the first step in specifying and selecting pumps for a revamp.

D

By J. T. ("Terry") McGuire

DISPLACEMENT

@ CONSTANT SPEED

FLOW

Figure 2. Flow regulation of kinetic vs. displacement pumps

and discharge and in the suction and discharge vessels • liquid levels relative to a common reference in both suction and discharge vessels • pressure drop across the control valve, if used, taking care to measure the downstream pressure some 10 diameters from the valve to avoid the influence of any flow distortion To make use of the pressure measurements, it is also necessary to determine the pumped liquid's specific gravity (SG) at each measuring point. This can be determined from liquid temperature as long as the liquid being pumped is known with certainty. Using Figure 3 as a reference, the system head is: Hsystem = (P4 - P1)2.31 + (Hz2 - Hz1) + HL1-4 (1) _________ SG

and the pump's total head is: Hpump= (P3 - P2)2.31 + Hz + HL2-3 + ∆V2 (2) ________ ___ SG 2g

where Hz is the correction for gauge elevation, if any, HL2-3 the friction loss between the suction and dis2 charge pressure gauges, and ∆V /2g the difference in velocity heads at

the points of suction and discharge pressure measurement. The friction loss is significant when there are elbows, valves or reducers between the gauge and the pump. The difference in velocity head usually matters only when the pump head is low and there is a difference of more than one pipe size at the points of pressure measurement. Subtracting the static head components from the pump head, Figure 4, yields the system friction head, HL1-4. When a control valve is used in the system, the head being lost to throttling across the control valve is calculated from the measured valve pressure drop, then subtracted from the total system friction to give the head lost to friction in the piping, including entrance and exit losses. Recognizing that the head lost to friction varies as the square of the flow rate, the equivalent system friction at several other capacities now can be calculated and the system head characteristic plotted (Figure 4). If the static head varies with time, as it often does in a transfer process, then the range of system heads can be plotted after allowing for maximum and minimum liquid levels in the suction and discharge vessels. The other critical aspect of the system to be verified using the measurements already made is the NPSH available at the pump. For measurements at the pump suction,

Hz2

1

➞

Hz1

4

3

➞

➞

P1

➞

P2

DATUM LEVEL

2 Figure 3. Hydraulic gradient

The Pump Handbook Series

again referring to Figure 3, the equation is: NPSHA = (P2 + Pa - Pvap) 2.31 + Hz + Vs2 (3) _________ ___ SG 2g where Pa is atmospheric pressure at site, Pvap is the vapor pressure of the pumped liquid at the pumping temperature, Hz is any correction for gauge elevation to the pump's reference level, and Vs is the fluid velocity at the point of suction pressure measurement. The pump's reference level is the shaft centerline for horizontal machines and the centerline of the suction nozzle for vertical machines. Because NPSH is also equal to: NPSHA = (P1 +Pa) 2.31 + Hz1 - HL1-2 (4) _________ SG

it is possible with the measurements already made to calculate the friction loss in the suction side of the system. And then following the same procedure used for the system head, the system NPSHA characteristic (Figure 4) can be developed. If the liquid level in the suction vessel can vary with time, the range of NPSHA can be plotted in the same manner as the range of system heads. With the net system head now known accurately, the head needed to move the required flow and allow flow control can be minimized. At the same time, an accurate NPSHA characteristic eliminates hidden margins. This means an appropriate NPSHA margin can be set for the application, facilitating selection of an optimum hydraulic design. Keeping the pump head to the minimum necessary lowers energy consumption. And an optimum hydraulic selection can contribute to both lower energy consumption and longer MTBR. Before preparing a pump specification, two more factors must be addressed at the site. The first is suction piping. Many problems are caused by poor suction piping. A unit revamp is an opportunity to correct this. The important features of the suction piping layout are the

185

VALVE

➞ ➞ ➞

➞ ➞ ➞

HEAD

PUMP TOTAL HEAD

➞

SYSTEM HEAD

PIPING FRICTION TOTAL STATIC HEAD ELEVATION

➞

PRESSURE

FLOW

TEST

➞

Figure 4. System head from measurements

TOTAL SUCTION HEAD PRESSURE & STATIC

VAPOR PRESSURE

➞➞

HEAD

➞

NET SUCTION HEAD FRICTION

➞

NPSHA

FLOW

TEST Figure 5. NPSHA characteristic from measurements

➚

➚

➚ a) Eccentric Reducer Flat Side Up

b) Eccentric, Flat Side Down

Figure 6a & b. Correct installation of reducers

orientation of reducers, the proximity of elbows to each other when in different planes, the orientation of the elbow immediately upstream of double suction pumps, suction piping slope and submergence over the vessel outlet. When the suction piping rises to

186

the pump, reducers in horizontal runs must be eccentric and installed flat side up (Figure 6a). In suction piping from above the pump, horizontal reducers can be concentric or eccentric, installed either flat side up or flat side down (Figure 6b). A The Pump Handbook Series

concentric reducer is necessary for end suction pumps of high Ns or S or both. Eccentric reducers installed flat side down are used by many designers to eliminate low points in the piping, which can accumulate dirt. At normal suction piping velocities of 7-8 fps, two elbows in series with their planes 90º apart should be separated by 10 diameters and should have the reducer downstream of the second elbow (Figure 7). Separating the elbows in this manner largely dissipates the flow distortion produced by the first elbow before it reaches the second. This avoids development of a swirl at the outlet of the second elbow. A reducer placed downstream of the second elbow helps dissipate the flow distortion and any swirl that may have developed. The last elbow in the piping to a double suction pump must be in a plane normal (at right angles) to the axis of the pump's shaft (Figure 8a). An elbow in a plane parallel to the pump's shaft axis (Figure 8b) will cause uneven flow into the impeller. This can result in higher power consumption, noise, vibration, premature erosion of one side of the impeller and thrust bearing failure. Moving away from the pump, the suction line must not have any high points that might accumulate air or vapor leading to reduced flow or even cessation of flow from air binding. And back at the suction vessel, the submergence over the vessel outlet must be sufficient to prevent vortexing (Figure 9) or the outlet must have an effective vortex breaker. The second factor to be addressed while at the site is the pump's service history. This can be obtained from the plants' maintenance department. What needs to be looked for is evidence of problems in the pump's application, materials of construction or mechanical design. Evidence of poor application might be indicated by frequent shaft seal and bearing failures, rapid wear at the running clearances, frequent shaft failure, noise and vibration or

S S

>4D1

➞ >10D2

➞

➞

➞ S

CONCENTRIC

S

➞

>10D2

➞

>4D1 ECCENTRIC

b) END SUCTION Ns –> 3,200 (2,750) or S > 11,600 (10,000) SOURCE ABOVE

a) END SUCTION SOURCE BELOW

Figure 7. Suction piping layouts with two elbows in series long radius elbow

path of water suction path of water

a) Desirable

suction

b) Undesirable

Figure 8 a & b. Elbows at the suction of double suction pumps

PUMP OPTIONS FOR UNIT REVAMP ➞

Armed with accurate data on the system head and NPSH available, and knowing whether the suction piping or pump need correction as part of the revamp, it's time to look at what has to be done and how best to do it. First, data developed by the process designer must be checked against the actual system head and NPSH available characteristics and corrected if necessary. As indicated, the questions to be answered at this

➞

S (feet)

➞

V

14121086420-

S

V ➞

premature impeller erosion. These are all symptoms of prolonged operation at low flows. Whether this is the case can be determined by comparing known flow rates with the pump's performance curve to see where it has been operating relative to the pump's best efficiency point (BEP). Improved construction materials are warranted if the pump has a history of component failure due to general corrosion, corrosion-erosion, erosion, fatigue or erosion-fatigue. It is often difficult to differentiate among these causes of component failure, so it may be necessary to consult a metallurgist. In some cases changing materials may not be enough. It may be necessary either to correct a problem in the process, such as lowering the concentration of abrasive solids or bringing the pH closer to neutral, or to change to a more suitable type of pump. Mechanical design is suspect only when the influences of application and the pumped liquid have

been eliminated. (This may be the reverse of common practice, but is the sequence to be followed in troubleshooting pumps.) Strain caused by piping loads is a major cause of mechanical problems. If the pump has had a high incidence of seal, bearing, coupling or shaft failures, the cause may be piping loads. The question then is whether the piping loads are too high or the pump not stiff enough. A computer analysis of the as-built piping is the first step in resolving this question. If the piping loads are reasonable or high but can't be changed, a switch to a pump with higher piping load capacity may be necessary. Short MTBR caused solely by pump mechanical design is rare in modern designs, but not uncommon in designs dating back 30 years or more. The usual difficulties are rotor stiffness, rotor construction, bearing capacity, bearing cooling, bearing housing stiffness and casing and baseplate stiffness. These problems typically manifest themselves as frequent seal, bearing and shaft failures and rapid running clearance wear. Most of these are also symptoms of poor application, so care is needed in sorting out the true cause of the problem.

Application

NPSH Margin % NPSHR3 ________________________________

S ' 2

' 4

' 6

' 8

V (ft/sec) Figure 9. Submergence over suction vessel outlet The Pump Handbook Series

' 10

Water, cold Hydrocarbon Boiler feed, small (3)

10-35 (1)(2) 10 (2) 50

Table 1. Typical NPSH Margins

187

NEED Lower flow Higher flow - small Higher flow - large Corrosion resistance Erosion resistance Better mech design

Rerate ∆ ∆

OPTIONS Replace Materials

Add ∆

point are the following. How much pressure drop does the control valve need to function reliably? Is it more economical to change to a variable speed pump? What NPSH margin is needed to ensure rated pump performance and expected life? The first is a question for the valve designer. Table 1 provides a starting point for the third. Notes: 1. depends on size, higher margin for larger pumps 2. minimum 3 feet 3. up to 2500 hp at 3600 rpm 4. U1 greater than 100 fps New service conditions for the unit revamp can be met by exercising three options: • rerate the existing pump or pumps • buy an additional pump or pumps of the same design • buy pumps of a new design

∆ ∆ ∆ ∆

∆ ∆

188

∆ ∆

therefore, the combined head capacity characteristic is developed by adding capacities at the same head (Figure 10). Two cautionary comments are needed here. First, to share flow reliably, the head of each pump must rise continuously to shutoff, or to the minimum bypass flow in the case of multistage pumps. Second, the increase in flow with each additional pump depends on the steepness of the system head characteristic (Figure 10). When the system head curve is steep, as with reference curve B in Figure 10, the increase in total flow with two pumps is quite small. In this case the flow per pump can be well below BEP, with the result that the MTBR of the pumps is reduced. At the other extreme, if the rating of the pumps is also increased by changing to a larger impeller, the A. FLAT SYSTEM HEAD B. STEEP SYSTEM HEAD

H.

These choices may, in turn, be influenced by the operating history of the existing pumps. Table 2 summarizes the needs developed from investigation of the existing pumps and the usual options for satisfying them. Each of the options is then discussed, starting with hydraulic considerations. Rerating the existing pump is the simplest course. To be successful, the rerate must be designed to meet the new conditions of service and at the same time overcome any deficiencies in the original application, such as being oversized for the normal flow or having a suction specific speed that is too high. Adding a pump, either in parallel or series, is one way to achieve substantially higher flow or head, respectively. To do so successfully requires care. For pumps in parallel the fundamental rule is that the pumps operate at the same head;

Construction

B A

B A

2 PUMPS 1 PUMP POWER NPSHR3

RUNOUT

FLOW

Figure 10. Pumps in parallel H. 'B' 'A'

2 PUMPS 1 PUMP

FLOW

B1

A2 B2

Figure 11. Pumps in series The Pump Handbook Series

runout NPSHR and power of a single pump must be checked to ensure that the pump has enough NPSH margin and that the driver will not be overloaded (Figure 10). Pumps in series operate at the same capacity unless flow is taken from between them. Their combined head capacity characteristic is therefore developed by adding heads at the same capacity (Figure 11). Using series pumps against a system with high static head (curve A in Figure 11) poses the risk of flow cessation if one of the pumps shuts down. This possibility must be taken into account in calculating the degree of redundancy built into the system and in designing the pump control system. Beyond this hydraulic consideration are two mechanical issues relating to pressure containment. First, the casing of the second pump must have a maximum allowable working pressure (MAWP) greater than the maximum discharge pressure developed when both pumps are running at shutoff with maximum suction pressure. Second, the shaft seal(s) of the second pump will, in most designs, be sealing suction pressure. Unless the pumps are separated in elevation, the suction pressure of the second pump is close to the discharge pressure of the first. This must be recognized in selecting the shaft seal. Buying new pumps is the most complex option, but it is necessary when the new service conditions are beyond the capability of the existing design. Regarding mechanics, a hydraulic rerate of older pump designs typically is combined with a mechanical upgrade to raise MTBR. Most manufacturers now have standard upgrades available for pumps ranging from single stage overhung to multistage. If the existing pumps have suffered corrosion or erosion abnormal for the service and the class of pump, a change of materials should be considered. On the other hand, if the corrosion or erosion appears more related to the type of pump than to the service, changing the pump might be more economical in the long run. In rare circumstances it will be clear from an existing pumps' ser-

vice history that it is the wrong configuration for the service. Rerating such a pump to an even higher energy level would simply aggravate this condition. A high energy overhung pump in a severe service is a typical example. So as not to jeopardize the success of the entire project, a pump that is clearly not suitable for its service must be replaced. Once an option is selected, the next step is to prepare the specification. The essential rule for a good specification is to keep it simple. Many a good solution has turned into a purchasing nightmare to the detriment of the revamp project because those preparing the pump specifications forgot this simple rule. The elements of a simple specification are: • a one or two page data sheet • scope of supply summary, supplemented by a terminal point diagram if needed • agreed upon terms and conditions A more detailed discussion of this phase and the equipment purchasing options can be found in the article "Pump Buying Strategies" by J.T. McGuire in the January, 1993 issue of Pumps and Systems. (Available as part of The Pump Handbook Series from the publishers of Pumps and Systems.) ■ J.T. ("Terry") McGuire is director of marketing for the Huntington Park Operations Division of the IngersollDresser Pump Company.

The Pump Handbook Series

189

CENTRIFUGAL PUMPS HANDBOOK

Pumping Hot Stuff – Another Perspective Here are some general considerations for “hot” applications and, specifically, heat transfer fluid services. By Pumps and Systems Staff

190

(PHOTO COURTESY OF GOULDS PUMPS)

M

thermal growth of the pump in high temperature services without impacting the alignment with the driver. A foot mounted pump has only one way to grow – up from the base. The result will often be misalignment to the degree that bearing life will be compromised. Differences in material and flange standards between the product types are also evident. ANSI pumps are commonly available with ductile iron, stainless or other alloy casings and 150lb flat-faced flanges. Steel is sometimes an alternative, but it is rarely available with a short lead-

Photo 1. Example of a horizontal metal ANSI process pump (PHOTO COURTESY OF INGERSOLL DRESSER PUMPS)

any words can mean different things to different people – and equipment manufacturers. This truth is important to appreciate when selecting pumps for your application. Probably because manufacturers are often “industry oriented,” this situation seems to occur regularly in the case of high temperature services. For example, some pump suppliers serve the chemical process industry (CPI) primarily while others sell mainly to petroleum refiners. In the former case, ANSI standard pumps (Photo 1) are the norm while API (American Petroleum Institute, Photo 2) standards are the design of choice in the latter. Per ASME B73, the specification applicable to ANSI models, pumps should be designed to handle temperatures up to 500º F. (Some manufacturers have provisions for higher limits, though non-metallic materials may be limited to relatively low temperatures in the 200º F range.) API pumps are available for much higher temperatures. A number of vendors make hybrid or specialty equipment that does not fit either standard but is appropriate for high temperature services. What are the primary differences between the two broad categories of pumps relative to handling high temperature liquids? The most apparent is the pump mounting method. ANSI pumps are foot mounted. API units are pedestal, or centerline mounted. This means the pump is effectively mounted at the center-line of the shaft, allowing for

Photo 2. Example of an API process pump The Pump Handbook Series

time. On the other hand, steel is the “lowest grade” casing material offered by API pump manufacturers, and 300 lb raised-faced flanges are also typical. In any application where the risk of fire is significant, such as a refinery service, the effect of thermal shocking of the casing material during fire fighting efforts must be taken into account. The last thing petroleum processors would want is to provide additional fuel to a fire as a result of a cracked iron casing. API pumps also have more extensive cooling options – even the mounting pedestals are sometimes cooled. Another major consideration when selecting pumps for high temperature services is the sealing system. Packing may be used for some applications, like boiler feed – in which case a smothering gland would be employed. In the case of mechanical seals, a film of liquid is needed to cool and lubricate the seal faces. While the high temperature liquid being pumped may be appropriate for this purpose, often it is not suitable due to lack of lubricity. This necessitates an external seal flushing arrangement. Consideration must be given when selecting the fluid used to avoid contamination of the process liquid. A combination flush and quench gland might also be needed. The quenching function involves an externally supplied fluid, such as steam, to contain the pumpage that passes the seal faces. If it is important that the pumped liquid not reach the atmosphere, either a multiple seal arrangement –

(PHOTO COURTESY OF DEAN METPRO)

possibly with barrier fluid system – or a sealless pump may be required. This latter category includes canned motor and magnetically driven units. We do not have space here for exhaustive explanations of various types of high temperature services. By exploring one of the most critical pumping applications in process plants, however, we can illustrate the importance of these considerations in the handling of hot liquids. That category of pump services involves heat transfer. If process temperatures are not maintained (often within very narrow ranges), the result can include plant shut-downs and/or substantial losses in product and production. Restarting the process can be an extremely complicated and timeconsuming procedure. The typical method to keep those temperatures where plant operators want them involves pumping one of several heat transfer fluids (HTF’s) in “hot oil” recirulation systems. These fluids have a common trait – they are expensive. Also, many are noxious and even toxic to plant personnel – or potentially detrimental to the environment. Looking at the options available to a buyer of pumps for an HTF service will serve to illustrate some of the complicated issues facing the decision maker. Several factors related to costs and operation will need to be evaluated. For example, if a mechanically sealed pump is selected, the pumped fluid will like-

Photo 3. The Dean R400A mechanically sealed pump is used for HTF service

ly prove to be an HIGH TEMPERATURE excellent seal flush because of its typiEDDY CURRENT DRIVE cally high lubricity and flash point, ROTATING OUTER MAGNET ASSEMBLY although it will probably need to be cooled before being injected at the seal faces to assure reasonable seal life. TORQUE RING This is usually done by taking a small amount of the pumpage off the discharge and piping it through an appropriate cooler. The small amount of fluCONTAINMENT SHELL id that passes the seal faces will evapOuter Magnet Ring is rotated by the motor. The orate into the atmosMagnetic Flux which passes from one pole to the next, passes through the Torque Ring. This creates a flow of phere. Without the EDDY CURRENTS. This power pulls the Torque Ring after use of a quench, forthe Outer Magnet Ring at a slightly slower speed which mation of crystals at is proportional to the power requirement. the seal faces can be a problem resulting Figure 1. Kontro, torque ring design in shortened seal life. coincident cost advantages. The Over time some make-up amount canned motor arrangement is, in will be needed. In the case of a seal effect, an induction motor with its failure, obviously there could be a rotor serving as the pump’s shaft. A significant amount of HTF lost, and thin metal can separates the arrangements to handle and replace pumpage from the stator. The rotor this material need to be made. A bearings are journal type and lubripopular mechanically sealed pump cated by the pumpage. In the case of used for HTF service in recent years HTF, as for the mechanically sealed is the Dean model R400 (photo 3.) pump design, a small portion of the This design has many API pump feapumpage is often cooled – typically tures, such as centerline mounting through an integrally mounted cooland bearing frame cooling provier, but one that in this case is used to sions, but it is priced between ANSI cool the motor windings. A number designs and those that fully comply of manufacturers make canned with API-610. A number of manumotor pumps suitable for heat transfacturers make pumps of this latter fer service. Magnetically driven type. If applicable, corporate policy pumps fall into one of two sub-cateor government regulations relative gories: synchronous and eddy-curto these fugitive emissions may rent. The first group utilizes one of necessitate provisions for monitortwo rare earth materials for both its ing and reporting, as well as specific external and internal magnet assemcontainment and processing proceblies. Neodymium is significantly dures for pumpage that leaks. affected by heat, especially in the Alternatively, and likely berange of heat transfer fluid services, cause of these regulatory controls making this material unsuitable in and corporate mandates, many buyhigh temperature applications. More ers are opting for one of the two genappropriate in this temperature eral categories of sealless pumps. range is Samarium Cobalt, which, Though canned motor and magnetiwhile the more expensive of the two cally driven designs typically have a materials, is impacted by high temsomewhat higher first cost, they perature operation to a significantly offer operational and, in some cases, The Pump Handbook Series

191

lesser degree. Several manufacturers serve the HTF market with appropriate units. The second group – mag drive pumps – isn’t really a group. The Kontro product line manufactured by Sundstrand Fluid Handling offers a unique torque ring design (Figure 1) that is unaffected by heat in the range for this application. As in the typical mag-drive pump, permanent outer magnets are coupled to a driver. Operating outside the fluid, they see reduced temperatures and are not affected by the heat as similar magnets would be in an internal assembly. What makes this design different is that power is transferred not to a set of inner magnets but to a rotating element consisting of the impeller, shaft and integrally mount-

192

ed torque ring (with a coil much like that in an electric motor.) This assembly runs in product lubricated journal bearings. In all sealless options the elimination of noxious emissions and potential leaks from seal failures simplifies or negates the need for internal corporate or external regulatory reporting, as well as provisions for containment and subsequent handling of fugitive emissions. The improved environment for operators is often cited as a significant advantage of using sealless pumps. Make-up requirements of expensive HTF lost through seal operation and occasional seal failures are eliminated. Interestingly, the somewhat lower efficiencies of sealless pumps actually contribute heat to the HTF, reducing the

The Pump Handbook Series

load on hot-oil heaters. This is offset by cooling water requirements needed for many canned motor pumps, though manufacturers do offer optional features, such as ceramic insulation, to allow higher temperature operation without cooling in lower horsepower sizes. Without the need for cooling of the recirculated lubrication fluid, a net heat gain is provided to the system in the case of this latter group of canned motor pumps as well as the mag-drive and torque ring designs. As you can see, there are many factors to consider when selecting a pump for HTF or other high temperature services. First cost or electrical power requirements are two obvious considerations, but often the less apparent issues will have a significant bearing on the purchase decision.■

CENTRIFUGAL PUMPS HANDBOOK

The Power of Speed and Staging The energy consumption of most centrifugal pump applications can be reduced by optimizing the specific speed. This can be done with either staging or shaft speed.

Energy and Efficiency As with all laws of nature, the key factor in minimizing the energy per unit time required to perform work on an on-going basis is the efficiency of the process. The same is true with centrifugal pumps. In U.S. units, the power equation is {HP = H x Q x S.G./ 3960 / η} where HP is horsepower, H is the total dynamic head of the pump in feet, Q is flow rate in gallons per minute, S.G. is the specific gravity of the pumped liquid (compared to water) and 3960

Ns= 2000

1500

1400

1200

1000

800

600

500

200

500

100

C

1 0.9 0.8 0.7 0.6 0.5 0.4 0.3 0.2 0.1 0 50

entrifugal pump decisions naturally require purchasers to judge the suitability of alternative selections. Pumps comprise the second largest segment of rotating equipment, and never have such decisions been more critical. Intense competition dominates nearly all of today’s markets. The minimization of expenditures is a constant focus with organizations intent on survival. Energy costs associated with pump drivers are a routine target in the continuous effort necessary to maintain market viability. Clear recognition of the most efficient pump options to meet process “flow versus head” will arm pump users with essential tools to attain reduced operating and maintenance costs. Gains in centrifugal pumping efficiency can be readily realized through the optimization of speed and staging. We will present a practical approach to understanding these issues using a simple format tabulated over a broad range of flows and heads.

EFFICIENCY

By Dave Carr and Bill Mabe

FLOW GPM

Figure 1. Efficiency as a function of capacity and specific speed

is a constant. The pump’s efficiency is usually represented by the Greek symbol eta (η ) and is expressed as a decimal (i.e., 75% is written as 0.75). To lower consumed power effectively, the efficiency component is the only variable factor in this equation for a given set of process conditions. Options do exist when selecting or beginning the design, however, since efficiency is a function of the pump’s specific speed.

Specific Speed and Efficiency Much has been written about “specific speed” (NS), a term that relates the pump performance to shape or type of impeller. It is not our intent to review the background of the term, but a few basic premises must be explained. The first is that a pump’s best efficiency point (BEP) corresponds to the head, flow, and speed combination matching the impeller’s design specific speed. Thus, the closer a match between a The Pump Handbook Series

pump’s normal operating point and the BEP, the less energy it will consume. The second premise is that, inherent to the definition: NS = N x Q0.5/H0.75

(1)

specific speed is directly proportional to rotating speed (N) in revolutions per minute and the square root of flow (Q) in gallons per minute while inversely proportional to the three quarter power of head (H) in feet. Finally, there are physical constraints to the best efficiency that trend flow – that is, there is increasing efficiency with increasing flow. This relationship is depicted in Figure 1 showing an expected band of achievable efficiency within a flow range representative of the vast majority of pump installations. It will be shown later that the effect of flow is more accurately related to the size of the pump. This is an important concept when consider-

193

Speed Reduces Energy Specific speed is directly proportional to the shaft rotating speed according to Equation 1. As the design speed is increased beyond typical electric motor speeds, the specific speed also increases for a given head and flow. As discussed earlier, efficiency generally increases with an increase in specific speed and flowrate. One could conclude that it is always better to increase the speed of a pump to maximize the specific speed and the efficiency. As we shall see later, however, higher speeds may not always be the best solution. Futhermore, most published data, such as that shown in Figure 1 taken from Karrasik (1985), can be seriously misleading for high speed pumps. High speed pumps are generally smaller than low speed single stage or multistage pumps. Euler’s equation shows us that head rise is proportional to impeller tip speed for a given exit blade angle. Tip speed is the product of the impeller tip diameter and the shaft rotating speed. Once the designer fixes the exit blade angle, the impeller diameter for a fixed design head decreases with increasing speed. Figure 2, established from experience with a large number of high speed rocket engine turbopumps, shows that the efficiency of a small pump is consistently lower than the efficiency of a larger pump. Consequently, it is not accurate to use Figure 1 directly for high speed pumps without correct-

194

10 4

2000

1900

1800

1700

1600

1500

1400

1300

1200

1100

1000

900

800

700

1

600

Staging is one way to increase the efficiency for a given set of hydraulics. By dividing up the head by the number of stages (Z), the effective specific speed is increased by Z3/4. As a result, the overall efficiency approximately increases according to the relationship shown in Figure 1. When referring to this figure, use stage head to calculate the specific speed.

DIAMETER, IN.

1 0.9 0.8 0.7 0.6 0.5 0.4 0.3 0.2 0.1 0 500

Staging Reduces Energy

EFFICIENCY

ing the potential efficiency to be gained by increasing the specific speed by raising the rotating speed.

SPECIFIC SPEED

Figure 2. The effect of size on efficiency

ing the data for size effects. The effect of impeller size on hydraulic efficiency at a fixed specific speed is due to surface finishes, friction factors, vane blockages, and internal leakage clearances that can not be practically scaled down from a large size. As a result, the percentage of hydraulic losses for small pumps is significantly larger than that for larger pumps. For impeller diameters greater than 10 inches, the size effect is small and generally insignificant. For diameters less than four inches, the efficiency penalty is particularly severe. The optimum speed for a given head and flow is sometimes a tradeoff between the efficiency gain from increased specific speed and the efficiency reduction due to smaller impellers. For most commercial pump applications, the gain from increased specific speed generally offsets the loss in efficiency due to size effects at higher design speeds.

Speed or Staging? Tables of average hydraulic efficiencies (Table 1 and Table 2) have been calculated over a range of typical heads and flows for commercial applications using data adapted from Anderson (1980). The tables can be used to compare efficiency at different speeds and numbers of stages. These efficiencies are for common motor speeds of 1800 and 3600 rpm. The tables also include data for shaft speeds of 7200 and 14,400 rpm routinely obtained with

The Pump Handbook Series

commercially available speed increasing gearboxes. The calculations include size effects. Note that increasing flowrates and shaft speeds both tend to produce higher pump efficiencies for a given head rise. At a given flowrate, the efficiency generally increases with increasing shaft speed. It should also be obvious from the tabulated values that reducing the head per stage increases the efficiency. Efficiencies tend to maximize at the higher flowrates with speed and multiple stages having minimal effects. The shaded portions of the table are usually impractical selections that one should avoid. For these applications, choose either multistage pumps, higher speed pumps, or mixed/axial flow impeller designs.

Overall Power The most prevalent pump drivers in use today are two or four pole AC electric motors. Speed in excess of those capabilities, however, requires the use of alternative equipment. Turbines are routinely used in process plants that have low to medium pressure steam to attain speeds as high as 10,000 rpm. Mechanical gearboxes, both integral and free standing designs, are the norm with electric motors, but belt and sheave arrangements can also be used. Speed ratios as high as 10:1 are possible with some gearboxes giving the pump designer more than enough flexibility to optimize specif-

HEAD PER STAGE, FT

HEAD PER STAGE, FT

250

HEAD PER STAGE, FT

500

FLOW, GPM

1800

3600

7200

14,400

1800

3600

7200

14,400

1800

750 SPEED, RPM: 3600 7200

50 75 100 125 150 175 200 225 250 275 300 325 350 375 400 425 450 475 500 525 550 575 600 625 650 675 700 800 900 1000 1250 1500

0.43 0.49 0.54 0.57 0.60 0.62 0.63 0.65 0.66 0.67 0.68 0.69 0.70 0.71 0.71 0.72 0.72 0.73 0.73 0.74 0.74 0.75 0.75 0.75 0.76 0.76 0.76 0.77 0.78 0.79 0.81 0.82

0.55 0.61 0.64 0.67 0.69 0.70 0.72 0.73 0.73 0.74 0.75 0.76 0.76 0.77 0.77 0.77 0.78 0.78 0.79 0.79 0.79 0.79 0.80 0.80 0.80 0.80 0.80 0.81 0.82 0.82 0.83 0.83

0.62 0.66 0.69 0.71 0.72 0.73 0.74 0.75 0.75 0.76 0.76 0.77 0.77 0.77 0.77 0.78 0.78 0.78 0.78 0.78 0.78 0.78 0.79 0.79 0.79 0.79 0.79 0.79 0.79 0.79 0.79 0.79

0.63 0.66 0.68 0.69 0.70 0.71 0.71 0.71 0.72 0.72 0.72 0.72 0.72 0.72 0.72 0.72 0.72 0.72 0.72 0.72 0.72 0.72 0.72 0.72 0.72 0.72 0.72 0.71 0.71 0.71 0.70 0.69

0.27 0.35 0.41 0.44 0.47 0.50 0.52 0.54 0.55 0.57 0.58 0.59 0.60 0.61 0.62 0.63 0.63 0.64 0.65 0.65 0.66 0.66 0.67 0.67 0.68 0.68 0.69 0.70 0.71 0.72 0.75 0.76

0.44 0.51 0.55 0.58 0.61 0.62 0.64 0.66 0.67 0.68 0.69 0.70 0.70 0.71 0.72 0.72 0.73 0.73 0.74 0.74 0.75 0.75 0.75 0.76 0.76 0.76 0.77 0.78 0.78 0.79 0.81 0.82

0.55 0.60 0.64 0.66 0.68 0.69 0.71 0.72 0.73 0.73 0.74 0.75 0.75 0.76 0.76 0.76 0.77 0.77 0.77 0.78 0.78 0.78 0.78 0.78 0.79 0.79 0.79 0.80 0.80 0.80 0.81 0.81

0.60 0.64 0.67 0.68 0.70 0.71 0.72 0.72 0.73 0.73 0.74 0.74 0.74 0.74 0.75 0.75 0.75 0.75 0.75 0.75 0.75 0.76 0.76 0.76 0.76 0.76 0.76 0.76 0.76 0.76 0.76 0.76

0.17 0.26 0.31 0.36 0.39 0.42 0.44 0.46 0.48 0.49 0.51 0.52 0.53 0.54 0.55 0.56 0.57 0.57 0.58 0.59 0.60 0.60 0.61 0.61 0.62 0.62 0.63 0.64 0.66 0.67 0.70 0.72

0.36 0.43 0.48 0.52 0.54 0.57 0.58 0.60 0.61 0.63 0.64 0.65 0.66 0.66 0.67 0.68 0.68 0.69 0.70 0.70 0.71 0.71 0.71 0.72 0.72 0.73 0.73 0.74 0.75 0.76 0.78 0.79

SPEED, RPM: SPEED, RPM:

SPEED, RPM: SPEED, RPM:

14,400 0.57 0.61 0.64 0.67 0.68 0.69 0.70 0.71 0.72 0.73 0.73 0.74 0.74 0.74 0.75 0.75 0.75 0.75 0.76 0.76 0.76 0.76 0.76 0.76 0.77 0.77 0.77 0.77 0.77 0.78 0.78 0.78

CENTRIFUGAL PUMP HYDRAULIC EFFICIENCY

195

250 > Ns > 4000 AND DIAMETER > 15 IN.

0.49 0.55 0.59 0.62 0.64 0.66 0.67 0.69 0.70 0.70 0.71 0.72 0.73 0.73 0.74 0.74 0.75 0.75 0.75 0.76 0.76 0.76 0.77 0.77 0.77 0.77 0.78 0.78 0.79 0.80 0.81 0.81

TABLE 1

The Pump Handbook Series

NOTE:

SPEED, RPM:

196 HEAD PER STAGE, FT

HEAD PER STAGE, FT

1000 1800

50 75 100 125 150 175 200 225 250 275 300 325 350 375 400 425 450 475 500 525 550 575 600 625 650 675 700 800 900 1000 1250 1500

0.09 0.18 0.24 0.29 0.32 0.35 0.38 0.40 0.42 0.43 0.45 0.46 0.47 0.48 0.50 0.50 0.51 0.52 0.53 0.54 0.54 0.55 0.56 0.56 0.57 0.57 0.58 0.60 0.61 0.63 0.66 0.68

NOTE:

3600 0.30 0.37 0.43 0.46 0.49 0.52 0.54 0.56 0.57 0.58 0.60 0.61 0.62 0.62 0.63 0.64 0.65 0.65 0.66 0.67 0.67 0.68 0.68 0.69 0.69 0.69 0.70 0.71 0.72 0.73 0.75 0.77

14,400

1800

0.45 0.51 0.55 0.58 0.61 0.63 0.64 0.66 0.67 0.68 0.69 0.70 0.70 0.71 0.72 0.72 0.73 0.73 0.73 0.74 0.74 0.75 0.75 0.75 0.76 0.76 0.76 0.77 0.78 0.79 0.80 0.81

0.54 0.59 0.62 0.65 0.66 0.68 0.69 0.70 0.71 0.72 0.72 0.73 0.73 0.74 0.74 0.74 0.75 0.75 0.75 0.76 0.76 0.76 0.76 0.76 0.77 0.77 0.77 0.77 0.78 0.78 0.79 0.79

0.03 0.12 0.18 0.23 0.27 0.30 0.32 0.35 0.37 0.38 0.40 0.41 0.43 0.44 0.45 0.46 0.47 0.48 0.49 0.49 0.50 0.51 0.51 0.52 0.53 0.53 0.54 0.56 0.57 0.59 0.62 0.64

0.24 0.33 0.38 0.42 0.45 0.48 0.50 0.52 0.53 0.55 0.56 0.57 0.58 0.59 0.60 0.61 0.62 0.62 0.63 0.64 0.64 0.65 0.65 0.66 0.66 0.67 0.67 0.68 0.70 0.71 0.73 0.75

250 > Ns > 4000 AND DIAMETER > 15 IN.

0.41 0.47 0.52 0.55 0.58 0.60 0.62 0.63 0.64 0.65 0.66 0.67 0.68 0.69 0.69 0.70 0.71 0.71 0.72 0.72 0.73 0.73 0.73 0.74 0.74 0.74 0.75 0.76 0.77 0.77 0.79 0.80

14,400

1800

1500 SPEED, RPM: 3600 7200

0.51 0.57 0.60 0.63 0.65 0.66 0.68 0.69 0.70 0.70 0.71 0.72 0.72 0.73 0.73 0.74 0.74 0.74 0.75 0.75 0.75 0.76 0.76 0.76 0.76 0.76 0.77 0.77 0.78 0.78 0.79 0.79

0.00 0.07 0.13 0.18 0.22 0.25 0.28 0.30 0.32 0.34 0.36 0.37 0.39 0.40 0.41 0.42 0.43 0.44 0.45 0.46 0.46 0.47 0.48 0.48 0.49 0.50 0.50 0.52 0.54 0.56 0.59 0.61

0.20 0.28 0.34 0.38 0.41 0.44 0.46 0.48 0.50 0.51 0.53 0.54 0.55 0.56 0.57 0.58 0.59 0.59 0.60 0.61 0.61 0.62 0.63 0.63 0.64 0.64 0.64 0.66 0.67 0.69 0.71 0.73

SPEED, RPM:

0.37 0.44 0.49 0.53 0.55 0.57 0.59 0.61 0.62 0.63 0.64 0.65 0.66 0.67 0.68 0.68 0.69 0.69 0.70 0.70 0.71 0.71 0.72 0.72 0.72 0.73 0.73 0.74 0.75 0.76 0.78 0.79

14,400 0.48 0.54 0.58 0.61 0.63 0.65 0.66 0.67 0.68 0.69 0.70 0.71 0.71 0.72 0.72 0.73 0.73 0.74 0.74 0.74 0.75 0.75 0.75 0.76 0.76 0.76 0.76 0.77 0.77 0.78 0.79 0.79

CENTRIFUGAL PUMP HYDRAULIC EFFICIENCY

The Pump Handbook Series

FLOW, GPM

SPEED, RPM:

TABLE 2

7200

1250 SPEED, RPM: 3600 7200

SPEED, RPM: SPEED, RPM:

HEAD PER STAGE, FT

Example 1. Head = 500 ft Flow = 300 gpm At 3600 rpm: Single stage efficiency = 69% Two stages (250 ft/stage) = 75% Or, at 7200 rpm: Single stage efficiency =

ic speed and, thereby, pump efficiency. Mechanical efficiencies generally decrease somewhat with increasing speed ratios. Regardless of which approach is taken, the efficiency of the drive train must be factored into the pump evaluation to ensure that maximum overall efficiency is attained.

74%

Specify either increased speed or staging to improve efficiency effectively.

Example 2. Head = 750 ft Flow = 100 gpm Single stage efficiency: At 3600 rpm = 48% At 7200 rpm = 59% At 14,400 rpm = 64% Note the dramatic increase in efficiency by choosing a higher speed pump for this typical single stage application. Example 3. Head = 1000 ft Flow = 175 gpm At 3600 rpm: Single stage efficiency = 52% Two stage efficiency = 55% Four stage efficiency = 70% Note the increase in efficiency available with a four stage pump. Alternatively, one can select a single stage pump at higher speed. At 14,400 rpm, efficiency = 68%

How to Use the Tables Use the tables to determine the potential for improving the efficiency of a pump for a given head and flowrate. Both the effects of staging and higher speed can be evaluated quickly. The following examples will illustrate: For head, flow, or speeds not given explicitly in the table, use linear interpolation. Accuracy is sufficient for most comparison purposes.

Concluding Remarks Today’s marketplace presents ample opportunity for pump users to conserve energy through the optimization of pump efficiency. Pertinent information regarding the use of speed and staging can assist with an educated pump selection. On the surface, it may appear that all increases in rotating speed or a higher number of stages will result in comparable increases in efficiency, as a result of higher specific speeds. Remember, however, that the optimum speed or number of stages for a given set of hydraulics may be a tradeoff between efficiency gains and losses as a result of specific speed and size effects. A disciplined use of the accompanying table will give the pump user an appreciation of these tradeoffs, particularly as they impact the search for minimum

The Pump Handbook Series

consumed power.■ Dave Carr is a Project Manager with Sundstrand Fluid Handling (Arvada, CO). A graduate of Purdue University with a B.S. degree in Mechanical Engineering Technology and M.S. in Management, he has held various positions within the company’s sales, marketing, product engineering and product development departments. Bill Mabe has been an Engineering Manager with Sundstrand for more than 20 years and is primarily responsible for new product development. A graduate of the University of Missouri at Rolla with a Mechanical Engi-neering degree, he is a member of the advisory committees for Pumps and Systems magazine and the International Pump Users Symposium.

REFERENCES Karrasik, I. et. al. (1985), Pump Handbook, pp. 2,13, McGraw Hill: New York Furst, R. B. et. al. (1973), Liquid Rocket Engine Centrifugal Flow Turbopumps, NASA Space Vehicle Design Criteria (Chemical Propulsion) SP-8109, pp. 15-17. Anderson, H. H. (1980), Prediction of Head, Quantity, and Efficiency in Pumps, Performance Prediction of Centrifugal Pumps and Compressors. ASME, pp. 201-211.

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CENTRIFUGAL PUMPS HANDBOOK

A Guide to ANSI Centrifugal Pump Design and Material Choices Here’s a pragmatic look at the commercially available design variations and material choices now being offered on pumps for corrosive/erosive and ultrapure service. By Dan Besic

ver the past decade a number of significant articles in the trade press have focused on the wide variety of available pump designs that affect flow characteristics, service life, maintenance, safety, product purity, operating costs and other factors. Too often these erudite articles – written by highly specialized hydraulic, mechanical or materials engineers – are more valuable to pump designers and manufacturers than to those who specify, purchase, operate or maintain the equipment. This observation is particularly valid in the area of corrosive, abrasive, hazardous or ultra-pure fluids. The adage that necessity is the mother of invention is a good explanation for the extensive design and material variations available. The whole field of exotic metal and non-metallic pumps has been and continues to be market driven. The standard metal pump market is commodity oriented. In the pumping of neutral fluids such as potable water or ordinary wastes, the key factors are initial costs, availability and service. Design standardization and available modifications are aimed at cost reduction and operatorfriendly aspects. Those responsible for price, delivery and dependable service are the main decision makers.

O

198

All of these purchasing influences are important when it comes to purchasing or specifying pumps for critical service applications. However, the handling of acids, caustics, solvents, halogens, salts, contaminated water, process fluids, ultrapure water or reagent grade chemicals and pharmaceuticals demands insight into design variations and material choices that are often specific to the fluid being handled or the operation involved. Pump design is still market driven but the speed with which new designs are translated into product availability at acceptable costs is controlled by the size of the market for the solution being offered. Plant engineering and maintenance personnel play critical roles. So do consulting and system engineers planning new installations or upgrading existing ones.

Historical Overview Centuries of pumping experience involving water and pH neutral or mildly corrosive liquids lie behind the standard design of horizontal centrifugal pumps. As new materials of construction were developed, pumps became more compact, more efficient and more resistant to mildly acidic or alkaline solutions, as well as to various atmospheric conditions and temperature variations. With the growth of the chemical processing The Pump Handbook Series

industries, pumps manufactured of stainless steel (types 304 and 316) became industry standards, with the molybdenum bearing type 316 holding sway because of its broader resistance to corrosion. As the market demanded even greater resistance, particularly to sulfuric acid, Alloy 20 became prominent. For even greater resistance to heat, corrosion or abrasion, higher nickel alloys and various exotic materials were made available. Costs were driven up as alloy content rose. Plastic pumps came into being about 50 years ago when heart/lung operations made it necessary to handle blood without contaminating it or destroying fragile red blood cells. The flexible liner, peristaltic type rotary pump developed for this lifesaving critical application utilized an acrylic material called Lucite for the casing or pump body and a pure gum rubber as the flexible component. These were the only two parts in contact with the fluid. The gentle “squeegee” action on the blood and the use of noncontaminating nonmetallic materials solved the potential problem of red blood cell destruction and helped save lives. This unique sealless pump design filled an unmet need for transferring corrosive/ abrasive fluids at low flows without corrosion, contamination or seal leakage. Coincidentally,

there was a major investment by DuPont and others to develop an extensive line of synthetic elastomers suitable for the liner. Industry researchers also sought to provide solid chemically resistant plastics such as Teflon, reinforced fluoropolymers, polyethylene and polypropylene for the casing. The mix and match potential for casings and liners (a dozen of these are now standard) has metamorphosed this “invention” for a particular problem into an almost universal design, a relatively low cost do-everything, sealless plastic pump. In another application, pump users handling bromine demanded improved service life and better worker environments than they were getting from stainless steel or nickel pumps. The answer found by pump designers relied on a new fluoropolymer, trade named Kynar, for all structural components. The market driven need for this corrosion and abrasion resistant material has so universalized its application that it has since become a standard recommendation for centrifugal pumps in an extended list of difficult-to-handle chemical-related processes, as well as those demanding high purity, such as pharmaceuticals or chemicals. More recently, the demand for non-metallic ANSI horizontal centrifugals dimensionally interchangeable with metal ANSI pumps has become sufficiently large to enable pumps made of polypropylene, polyvinyl chloride and to be costeffectively manufactured for a wide range of processing operations, as well as for water treatment and wastewater handling. For all these and similar market related reasons, non-metallic pumps in a variety of thermoplastic and thermoset formulations have emerged as serious alternatives to metal ones. The burden of reasonable choice, in many cases, is no longer linked to a particular innovative manufacturer of a specific pump in a specific material for a specific application. Users now have a choice among many metals and nonmetallics. The purpose of this article is to present an overview of commercially available design variations and mate-

rial choices now being offered to users and specifiers of centrifugal pumps for corrosive/erosive and ultrapure service. When in doubt, there is no substitute for direct interchange with the manufacturer and your own or their experience with the specific fluids and service conditions involved with an application.

Centrifugal Pump Design

reinforced plastics. The thermosets have physical properties closer to metals than the thermoplastics, and they combine this with many of the weight and chemical resistant advantages of non-metallics. Pressure ratings at ambient temperatures for flanges in thermoset pumps are equivalent to 150 lb. metal flanges of the same dimension. The pressuretemperature gradient is a linear factor and starts to degrade the material above 100ºF, so if fluid temperatures run higher, some pump manufacturers will add metal backup rings to the thermoset support head, or they have support heads which include a backup ring. Thermoplastic pumps designed so that the flanges and casings are completely supported and protected by heavy sectioned metal armor present no problem with respect to the pressure temperature gradient. Any of the available armored ANSI thermoplastic pumps will conform to the standards without modification.

Taking the ANSI end suction centrifugal process pump as the standard, let us compare the design characteristics of the critical components made with thermoset and thermoplastic pumps, assuming for the moment that all these material choices are satisfactory for the application. First, a word about plastic-lined metal pumps. Plastic-lined metal pipe is proving to be an economical approach to transferring corrosive fluids, but when it comes to pumps, linings are not as easy to apply or maintain. They are, however, on the market. Generally, pumps are produced with a corrosion resistant lining applied in the wetted areas. Most common designs are those with fluoropolymer or thermoset linings approximately 1/8-3/16” thick. These pumps may be subject to early failure from erosion, pin holing, peeling or physical damage from solid particles, as well as from temperature fluctuation. Resistance to corrosion and abrasion in a pump is a lot more complicated than it is in a pipeline. Since plastic lined pumps do not typically offer service life comparable to that achieved in pumps having solid molded plastic pump casings and impellers, the design variations of lined pumps will not be covered in this article.

As a general rule, the allowable nozzle loadings for thermoset pumps are lower than those indicated for metal pumps. Published data suggest that these pumps be exposed to nozzle loads 1/2 to 2/3 that of similarly sized stainless steel pumps. Armored thermoplastic pump designs conforming to ANSI B73.1 specifications have the discharge pipe and flanges reinforced (metal armored) so that they can accommodate the same nozzle loadings as the steel pumps they replace. Connections to metal pipe put the full load on the metal armor rather than on unsupported plastic as is the case with thermoset designs.

Casing

Shafts

Dimensionally, all pumps meeting ANSI B73.1 specification for process pumps are the same. The choice is yours. You can remove a metal pump and replace it with a plastic one, or vice versa, without changing the piping. As the need for a non-metallic pump to replace the stainless steel pumps in corrosive service became evident, it was natural for the manufacturers to offer designs made of various fiberglass

Most non-metallic pumps use stainless steel or other stress proof steel shafts. Regardless of which material is used, it is recommended that the shaft be completely sleeved in plastic to isolate it from the fluid in the wetted area. High strength carbon steel can be safely used when the shaft is so sleeved. The sleeve can be made of injection molded glass reinforced polyphelene sulfide (PPS) when the

The Pump Handbook Series

Nozzle Loadings

199

corrosive is relatively mild, but for best results shaft assembly should be made of the same material as the pump. Thermoplastic pumps should have sleeves of PVC (Polyvinylchloride), PP (Polypropylene), or PVDF (Polyvinylidene fluoride). The shaft sleeve in a horizontal centrifugal pump should be independent, not welded to the impeller. When so designed, a damaged shaft sleeve can be replaced by simply removing the impeller and then the sleeve. This approach is recommended because most users do not have plastic welding capability.

Mechanical Seal For maximum corrosion resistance it is important that mechanical seal configurations be completely non-metallic, or mounted so that the inboard or wetted non-metallic seal face is exposed to the fluid. Since seal changes represent a significant maintenance item, it is critical that the pump design permit easy inspection of the seal without disassembling or removing the impeller shaft. In standard centrifugal pump designs which are not made to ANSI specifications, or which do not permit back pullout, the available work area for seal inspection and maintenance is crucial. In some ANSI, as well as nonANSI pumps, a sliding bar design, which permits backing up the primary seal for inspection without affecting shaft alignment, keeps downtime at a minimum. Externally mounted seals facilitate maintenance of the pump/seal area by allowing personnel to inspect seal placement visually. This permits proper setting of the seals without relying, with fingers crossed, on shims or measurement. Pump stuffing box designs, which permit the widest choice of readily available single or double mechanical seals, provide a definite advantage.

Shaft Deflection Over the years the complicated formula for shaft deflection, an important component for comparing potential seal life from centrifugal pumps, has been abbreviated L3/D4. This relationship between the length of the shaft overhang from the front

200

or inboard bearing to the impeller and the diameter of the shaft at the seal face is generally satisfactory when comparing two metal pumps of similar design. However, it is useless and misleading when comparing a metal pump with a non-metallic one. The reason is simple. The full formula takes into consideration the diameter and the weight of the impeller. Impeller diameters of engineered plastic pumps are equal to or only slightly smaller than metal ones, but impeller weights are quite different. The lighter weight of the plastic impeller and the corresponding reduced downward thrust or force on the impeller end of the shaft are ignored by the L3/D4 abbreviation. These variables significantly affect the length/diameter ratio, making the common acceptable ratio (50) meaningless when applied to thermoset or thermoplastic pumps. Don’t get caught in this trap. Insist on knowing the actual vibration level at the bearings – the actual shaft deflection at the seal face, not the “shaft stiffness” factor that the simplified formula provides.

Cost Factors There was a time when plastic pumps were considered less expensive substitutions for metal ones. Those days are long gone. Nonmetallic pumps for corrosive/erosive/hazardous or other specialized services are carefully engineered products with initial costs competitive to similar pumps of type 316 stainless steel. Cost advantages, however, become appreciable for those applications involving the higher alloyed materials and for pumps requiring titanium, nickel and other exotic metals. Published cost comparisons indicate that plastic pumps tend to be 20% lower in price than those of Alloy 20 (a nickel-chromium-copper alloy originally developed for handling sulfuric acid), and half or less than the cost of Hastelloy C, the high nickel stainless alloy developed for severe corrosive service. Initial cost is but one factor, however. Service life is even more significant. Plastic pumps have been in service sufficiently long in highly corrosive/abrasive applications for The Pump Handbook Series

the purchaser or user to ask for data on anticipated service life based on previous installations. Laboratory tests on metal or plastic “coupons” prove to be poor guides to pump life because they usually involve immersion time under static conditions. Corrosion or wear rates for parts rotating at 1800 or 3600 revolutions per minute, or stationary parts handling fluid flows to 5000 gallons per minute, are not comparable to those based on laboratory tests. Ask the pump manufacturer for actual case history data to be sure. There are other cost factors to be considered when comparing metal versus plastic pumps, or one plastic pump versus another. Metal to plastic weight comparisons impact on costs involved in shipping, installation, disassembly/assembly, and time related problems such as galling or seizing of threaded components. Seal life, spare parts requirements and general maintenance requirements must be considered. In all of these areas, the non-metallics offer clear cut advantages because of their lighter weight (1/2 to 1/8 the weight of metal), lower component costs, non-corroding characteristics and longer seal life. The differential in cost between thermosets and thermoplastics is negligible if we compare polypropylene and fiberglass reinforced plastic (FRP). Comparative service life is more critical, but this has to be based on the particular thermoplastic and specific thermoset. Generally speaking, homogeneous thermoplastics have much broader chemical resistance than thermosets, and thermosets offer physical properties higher than those of thermoplastics. For example, if the superior corrosion and abrasion resistance of the fluoropolymers is required, the higher cost of fluoropolymer thermoplastics such as PVDF or ECTFE is often repaid many times over by the extended service life of and/or product purity advantages these fluoropolymers offer.

Maintenance Factors Standard, routine preventive maintenance programs are advisable for all pumps regardless of materials

of construction or design factors. Your own service experience is the best guide to proper scheduling. Where the installation is substantially different from what you have been involved with, it is wise to rely on the specific experience of the manufacturer. When shifting from metal pumps to plastic pumps, consider the following: A. If you are connecting thermoset horizontal centrifugal plastic pumps to metal pipes, you may need to use expansion joints to reduce nozzle strain, particularly if there is a significant difference in the modulus of elasticity of the two materials. This may not be necessary for amored. B. For vertical centrifugal pumps used in deep sumps, make sure the design accommodates the differential in axial thermal expansion between the long plastic support columns and the available impeller clearance. This, however, is not a factor for horizontal pumps. C. Although non-metallic flanges are dimensionally the same as metal ones, if they are unsupported you may need metal washers under the bolt heads and nuts. Casing bolting for non-metallics requires careful adherence to torque specified by the manufacturer. This is true for all pumps but not as critical with metal or metal armored plastic ones. D. Do not steam pressurize thermoplastic or thermoset pumps. This can cause components to be pressurized beyond ratings. E. Protect outside mechanical seal with a seal guard or cover.

Comparative Material Characteristics It is not the purpose of this article to concentrate on the long list of sophisticated materials of construction or erudite design factors over which the user has little or no control. These are basically the responsibility of pump manufacturers. Pump users need not be metallurgists or materials engineers. Their concern is and should be with understanding the relationship between available construction materials and the service conditions faced. Pump designers have an endless variety of

materials from which the critical wetted end components can be selected. In reality, however, users are limited to choosing from among a relatively short list of commercially available materials. Exotic metal or customized composites may be required for pumping problems that can’t be solved with materials that can currently be produced economically. Limited production potential often makes it difficult to produce pumps of these materials at reasonable cost and delivery schedules. The pump materials listed below are readily available for your consideration. Familiarity with the basic characteristics of these materials will help you select the most cost effective material for your applications.

Materials of Construction Metals Stainless Steel (type 316): A general purpose austenitic chromium (18%), nickel(8%), molybdenum (3%) alloy providing broad resistance to a long list of acids, caustics, solvents. Widely used for its atmospheric corrosion resistance and in products requiring sterilization. Not recommended for sea water, brine, bromine or strong oxidizing acids. Alloy 20: Originally developed to resist various concentrations of sulfuric acid for which type 316 stainless steel wasn’t suitable. This stainless alloy has higher nickel (28%) than chromium (20%) content, less molybdenum (2%), plus copper (3%). It offers good resistance to dilute and strong, as well as mixed acids, sulfate and sulfites, sulfurous and phosphoric acids, chlorides and brines. Hastelloy “C”: With an increase in the nickel content to approximately 50%, this nickel, chromium (22%), molybdenum (13%, tungsten (3%) alloy is recommended for uses in which the stainless steel and 20 alloys fail. It offers much greater resistance to oxidizing acids and acid mixtures, wet chlorine, chlorides, mineral acids, sea water and plating/pickling solutions.

Thermosets Fiberglass Reinforced Plastic (FRP): Although there is an endless The Pump Handbook Series

variety of thermoset formulations, many indicated by trade names, the most common standard formulations consist of vinyl ester and epoxy resin reinforced by glass fibers. Since the two components bring different values to the composite – vinyl for corrosion resistance, epoxy for solvent resistance – competitive formulations can be customized for a particular service. Thermoset materials generally offer broad resistance to most acids, caustics, bleaches, sea water and brine. Special formulations are available for mild abrasive service. Thermoset pumps of FRP have higher physical properties than those of thermoplastics and are available for flows to 5000 gpm, twice the flow of current thermoplastic offerings.

Thermoplastics Polyvinyl Chloride (PVC): Widely used in chemical processing, industrial plating, chilled water distribution, deionized water lines, chemical drainage and irrigation systems. Good physical properties and resistance to corrosion by acids, alkalies, salt solutions and many other chemicals. Not suitable for solvents such as ketones, chlorinated hydrocarbons and aromatics. Chlorinated Polyvinyl Chloride (CPVC): Its chemical resistance is similar to, but slightly better than, PVC. It is also stronger. Excellent for hot corrosive liquids, hot and cold water distribution and similar applications. Polypropylene (PP): A light weight polyolefin chemically resistant to organic solvents as well as acids and alkalies. Generally not recommended for contact with strong oxidizing acids, chlorinated hydrocarbons and aromatics. Widely used in water and wastewater applications and for laboratory wastes where mixtures of acids, bases and solvents may be involved. Polyvinylidene Fluoride (PVDF): Strong, tough and abrasion resistant fluoro-carbon material. Resists distortion and retains most of its strength at elevated temperatures. Recommended for ultrapure water and reagent chemicals. Resistant to most acids, bases and organic sol-

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Weight Loss/ 1000 Cycles

Polyvinyl chloride

140ºF

PP

Polypropylene

185ºF

PE

Polyethylene

200ºF

(UHMW)

5 mg.

CPVC

Chlorinated polyvinyl chloride

210ºF

PVDF

5-10

FRP

Fiberglass reinforced plastic

250ºF

PVC

12-20

PP

15-20

PVDF

Polyvinylidene fluoride

275ºF

CPVC

20

ECTFE

Ethylene chlorotrifluoroethylene

300ºF

Stainless Steel

50

PTFE

Polytetrafluoroethylene

500ºF

FRP

388-520

PTFE

500-1000

Table 1. Temperature

vents and equally suited for handling wet or dry chlorine, bromine and other halogens. Ethylene Chlorotrifluoroethylene (ECTFE): Resists an extremely broad range of acids, alkalies, organic solvents and combinations of them, as well as other corrosive and abrasive liquids. Also resistant to oxidizing acids and hydroxides. Ideal for ultrapure water applications. Polytetrafluoroethylene (PTFE): This crystalline polymer is the most inert compound known. It has useful mechanical properties at elevated temperatures. Impact resistance is high, but tensile strength, wear resistance, and creep resistance are low in comparison with other engineered plastics. Its coefficient of friction is lower than almost any other material.

Selection Criteria Temperature: Temperature Weight

Specific Gravity

PP

.91

PE

.92 - .94

PVC

1.30

CPVC

1.49

PVDF

1.75

ECTFE

1.75

PTFE

2.14 - 2.20

FRP

3.4 - 5.0

316 Stainless

7.9

Table 2. Weight differentials of construction materials

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Material

PVC

parameters are not critical in determining the choices among the various metals. Metallurgically, the differences may be great, but all of the metals considered for handling corrosive fluids are stable at most operating temperatures. When it comes to the plastics, however, anticipated temperature fluctuations are very critical. Table 1 provides upper temperature limits for the nonmetallic pump materials currently available. Weight: The specific gravity of the material can be a significant variable when comparing metal to plastic pumps because of weight-related costs such as shipping, installation, support structures and in-plant repositioning. It is also vital in the comparison of various plastics with each other. As anyone who has lifted a fluoroethylene polymer casing can tell you, not all plastics tip the scale in the same way. A PVDF casing may

Polyethylene

Table 3. Abrasion resistance (Tabor Abrasion Tester)

weigh twice as much as one molded in polypropylene. Of course, when comparing metal-armored plastic pumps or plastic-lined pumps with all metal ones, the apparent differences are minimized. Table 2 shows the comparative weight of the materials of construction. If the weight factor is significant, ask the manufacturer for total pump weights. You may need these to assist in installation and piping. Abrasion Resistance: Strange as it may seem, stainless steel pumps are relatively poor compared to plastics when it comes to resisting wear from abrasive fluid streams. A major reason for this is that the oxidized surface which protects passivated chromium-nickel stainless steel from corrosion is continuously removed by abrasive particles. The smooth,

Material

Tensile Strength (psi)

Hardness R = Rockwell D = Shore B = Barcal

PE PP PTFE PVDF ECTFE PVC CPVC FRP

3,500 - 5,600 4,000 - 5,000 2,000 - 5,000 5,500 - 8,250 6,500 - 7,500 6,000 - 7,500 7,500 - 11,000 10,000 - 13,000

R 35-40 R 80-110 D 50-55 D 80 D 75; R 93 R 113 R 121 B 35-40

Table 4. Strength of the non-metallics The Pump Handbook Series

Impact (IZOD)

1.5 - 12.0 0.5 - 2.2 3.0 3.6 - 4.0 No break at 73ºF 0.4 - 2.0 0.6 –

uniform interior surface of a molded thermoplastic casing is a significant factor in reducing friction and turbulence, both of which contribute to wear. The FRP materials rely on the epoxy/vinyl ester to contain the glass. If the surface is subject to fast flowing process streams containing solid particles, abrasion can be severe. The reinforcing fibers may be exposed causing degradation of the composite or a wicking/bleeding action that can contaminate the fluid. For these reasons, homogeneous thermoplastics offer much greater resistance to abrasion. Table 3 shows typical weight loss of the materials being considered. Strength of Non-Metallics: The differential in tensile strength and impact resistance between metal and non-metallic pumps may be of academic interest, but insofar as users are concerned, it is significant to keep in mind that unarmored non-metallic pumps need more care in handling than metal ones. Metal armored thermoplastic pumps can be treated as metal ones, but thermoset pumps require a bit more attention. Reinforced epoxy/vinyl composites can be brittle (like cast iron), so precautions should be taken to provide protection from falling overhead objects, fork lift trucks or careless handling. Tensile, hardness and impact strength of the various non-metallic materials are shown in Table 4. Corrosion/Chemical Resistance: When pump specifiers and users are asked why they are considering non-metallics, corrosion resistance is by far the number one reason given. The resistance of the basic construction material is significant in terms of service life, safety, maintenance and, in many cases, the purity of the product being pumped. This latter criterion is essential for those using deionized water, reagent grade chemicals or other ultrapure fluids in their manufacturing or processing operations. When metallic or other contaminants cannot be tolerated, this resistance is critical because it may seriously affect the quality or the value of the product being produced. Many handbooks and materials engineering articles give comparative

corrosion resistance data for an exhaustive list of metals and plastics. These are rated against page after page of different chemicals at varying concentrations and at various temperatures. The tables are helpful guidelines, but for the most part, they are based on corrosion or deterioration when the subject material is immersed in the fluid under static and unchanging conditions. Conditions of service in the real world are seldom that uniform or controllable. These lists and tables are no substitute for actual pumping experience – your own and that of your suppliers’. Described below are highlights of some interesting pumping operations in which material selection played a critical role. These examples may help you see your own pumping problems in a different perspective. If this article encourages you to review your company’s pump purchasing and maintenance picture with a view to how you might extend service life, reduce parts inventory, simplify maintenance procedures, improve product quality, meet current or anticipated environmental regulations, or even feel comfortable about your current operating equipment and procedures, it will have served its purpose. Examples: 1. Sulfuric Acid Waste Stream Containing 100 Micron Fines This corrosive/erosive mixture was destroying 316 stainless steel pumps. The short service life and costly downtime could not be tolerated. When the switch was made to polypropylene pumps to handle the 3 pH abrasive wastewater, the problem was solved. 2. Corrosive/Abrasive Mine Water Acid run off with a pH of 1-2 from coal mines in South Africa were severely corroding stainless steel pumps. The answer was found with a combination of non-metallics. Pump casings were supplied in polypropylene, but a PVDF impeller was specified because of the superior abrasion resistance of this fluoropolymer. To reduce problems with The Pump Handbook Series

early seal failure, a Teflon (PTFE) packed gland and product flush arrangement was used instead of the standard single seal product flush. The pump has been handling 16-18 million gallons of acidic fluid per week at 1800 gpm. No problems. 3. Etching Glass with Hydrofluoric Acid A major glass manufacturer experienced severe pump maintenance problems when it came to handling a slurry of hydrofluoric acid with a high content of proprietary gritty compound. Original equipment utilized a composite fiberglass pump with a stainless steel impeller. Service life averaged about 2 months. The plant manager tested a variety of non-metallics and finally decided on PVDF impellers for the corro-sive/abrasive service and polypropylene for the casing. To isolate the shaft from the fluid, he specified a thick sectioned PVDF sleeve and arranged to have the mechanical seal reverse mounted. The pumps are driven by 7.5 hp electric motors at 1750 rpm. The 10% production increase experienced is credited to keeping the production line operating with thermoplastic pumps. 4. Acid Mixing Plant in Alaska An oil field acidizing and well service company in Alaska ordered a modular plant built in Texas and shipped to their Prudhoe Bay site. Some of the pumps are polypropylene throughout. Others are fitted with ECTFE impellers to handle the hydrochloric acid and zylene. Although the pumps are located indoors, the fluid temperature could be minus 60°F because the chemicals are stored outdoors. Service reports indicate no problems. 5. Circulating Phosphoric Acid The use of lined metal pumps to handle phosphoric acid was causing extensive shutdown due to pinholes and tears that allowed the metal casings to contaminate the solution and downgrade it. Part of the problem was caused by heat generated in the pumps, which tended to destroy the thin linings. When solid PVDF thermoplastic pumps were substituted,

203

product purity was assured and overall maintenance substantially reduced. 6. Ultra Purity for Hydrogen Peroxide Production Specifications for these centrifugal pumps called for the use of pump casings, impellers and shaft sleeves to be made of virgin, unpigmented fluoropolymer totally resistant to 70% hydrogen peroxide being pumped at 50gpm against a total dynamic head of 80 feet. The material selected was ECTFE, ethylene chlorotrifluoroethylene, which is noted for its broad resistance to chemicals, pharmaceuticals and oxidizing acids. The seal rings were specified in the fluoroelastomer Viton and the casing jacket in Teflon PTFE. 7. Bleaching in White Paper Mills The most corrosive applications in white paper mills involve pumping chlorine bleach. Traditionally, this has been done with extremely expensive titanium pumps. When environmental concerns required a switch from chlorine bleach to chlorine dioxide, pulp mills were able to change from titanium to epoxy vinyl ester pumps at much lower initial cost, reduced spare parts inventory and ready availability of pumps and parts. 8. Sulfuric Acid and a Caustic Chaser A large can manufacturing division of one of the largest breweries in the world faced costly downtime due to excessive corrosion of the metal pumps transferring dilute sulfuric acid for the etch and strong caustic for the required neutralization. The decision was to change from stainless steel to PVC thermoplastic centrifugal pumps. All wetted end components were specified in the same homogeneous thermoplastic. The plastic pumps have served in this round the clock operation for more than a dozen years with only routine maintenance. 9. Hot Oil/Hydrofluoric Acid/ Caustic Mixture A refinery using hydrofluoric

204

acid in a process to produce high quality gasoline ran into difficult pumping problems due to the varying pH of the 150°F oil/HF/caustic mixture. They tried a variety of pumps from cast iron to 316 stainless steel, but a combination of corrosion plus impeller and seal damage from accumulated solids required pump replacement on a monthly schedule. Since installation of molded thermoplastic polypropylene pumps with double mechanical seals and pressurized water jackets using 8 to 10 gph of clear water to cool the seal jacket, the pumps have performed flawlessly. 10. Tall Oil Production in Kraft Mills Tall oil soap, a byproduct of the Kraft paper making process. Skimmed from the evaporated cooking liquor, the soap is acidulated with sulfuric acid in a reactor vessel to achieve a pH of 2.5 - 3.5 and allowed to settle out. The tall oil rises to the top, and the spend acid can be either fresh acid (98%) diluted with water to 30% or with spent acid from the bleaching operations. FRP pump manufacturers report that the thermosets show improved service life over pumps made of 316 stainless or 20 Alloy. 11. Pumping Deionized Water for Laboratory Use Medical laboratories utilize sealed diagnostic kits to analyze blood samples. To assure correct diagnosis, it is critical that the water used to prepare the various chemical solutions be free of contaminants. Pretreatment of the processed water requires a sand filter, deionization equipment and special fine pore filtration media to eliminate all particulates. Once the water is purified, it is protected by a fluid handling system composed entirely of chemically inert plastics. It is stored in polyethylene tanks and pumped through a closed loop system of rigid PVC pipe by a horizontal centrifugal pump with all wetted components of PVDF fluoropolymer. This system uses close coupled pumps with an integral pump/motor cantilevered shaft that enables the pump to run dry for The Pump Handbook Series

extended periods without damage. 12. Pump Maintenance Reduced by $25,000 The extreme corrosiveness of an ammonium chloride/zinc chloride solution at 140°F resulted in the failure of Alloy 20 pumps after a year of service. In addition to the high cost of this annual pump replacement, mechanical seal failure occurred on a monthly basis requiring a minimum of 2 hours of lost process time each month. When the metal pumps were replaced with those made of virgin thermoplastic polypropylene, the following results were reported: lower initial pump cost, improved resistance to both corrosion and cavitation, an average seal life of nine months, and better than two years of service before removal of the pumps for reconditioning. Annual maintenance costs decreased by $25,000. 13. Silicon Wafer Production Demands PVDF Pumps Production processes in the manufacture of high quality silicon wafers require the use of hydrochloric acid, hydrofluoric acid and various caustic solutions. The manufacturer had standardized on thermoset materials for the pumps, but the corrosiveness of the chemicals and mixed acid wastes proved too severe. Hazardous fluid leakage and reduced pump performance, particularly in the handling of the hydrofluoric acid waste streams, necessitated a changeover to polypropylene pumps with PVDF shaft sleeving in the wetted area. Plant management reports that leakage and capacity problems are a thing of the past.■ Dan Besic is Chief Engineer at Vanton Pump & Equipment Corp. (Hillside, NJ).

CENTRIFUGAL PUMPS HANDBOOK

Self-Priming Pumps: It’s in the System! Coordinate pump with piping system for optimum performance. By Ray Petersen

W

Centrifugal Pump Operation The basic theory of operation of a centrifugal pump – from which its name is derived – relates to centrifugal force. A rotating impeller imparts velocity energy to the fluid between the impeller vanes, and that velocity

is subsequently converted to pressure energy in the volute section of the pump casing. This increase in pressure energy (total dynamic head on a pump curve) is proportional to the centrifugal force applied to the fluid. Since centrifugal force is directly proportional to weight, the difference between the discharge head of a centrifugal pump filled with water and a centrifugal pump filled with air is in the order of 810 to 1! (This is the difference between the density of water and the density of air at atmospheric pressure.) A pump handling a mixture of air and liquid will exhibit discharge heads between these two limits.

impeller designs such as vertical and submersible pumps and converted horizontal pumps utilizing auxiliary vacuum producing equipment or suction priming tanks that allow a standard centrifugal pump to operate with a positive liquid head at all times. The most popular self-priming centrifugal pump used commercially is the “recirculation” or “peripheral priming” type. It is characterized by a liquid reservoir either attached to or integrally constructed with the (Photo courtesy of Fybroc Pump Div.)

ith the advent of increasingly stringent environmental regulations pertaining to storage tank connections below liquid levels, selfpriming centrifugal pumps are being utilized in a wide variety of applications. In the past, horizontal designs have been employed in these services due to the inherent advantages of centrifugal pumps. Installation was relatively simple, and if the suction and discharge requirements of a particular pump were met, satisfactory performance could be expected. Converting a centrifugal pump application that previously operated off of a positive suction head to one consisting of a negative, or combination negative and positive suction head, requires additional criteria to be met for a successful application. The addition of a compressible fluid (air) in the suction line imposes conditions on the piping system that must be overcome for a self-priming pump to remove air from the suction line and pump fluid as a centrifugal unit is designed to do. Piping systems composed of check valves, pipe loops and liquid traps should be reviewed with respect to the pump operation during the priming cycle to assure that all air can be evacuated from the suction line prior to the pump moving liquid only.

TDH ~ centrifugal force = wω2r g w = weight ω = angular velocity r = radius g = gravitational acceleration Density water at atmospheric pressure = 62.4#/ft3 Density air at atmospheric pressure = .077#/ft3 From this it is clear that large static heads cannot be imposed on the discharge side of a centrifugal pump if it handles air during some phase of its operation.

Self-Priming Centrifugal Pump Operation There are many types of selfpriming centrifugal pumps. Included in this category are submerged The Pump Handbook Series

The popular “recirculation” or “peripheral priming” type self-priming pump is characterized by a liquid reservoir either attached to or integrally constructed with the pump casing

pump casing (Photo 1). The suction connection is usually located above the impeller centerline in order to contain and trap a volume of liquid used in the priming cycle. Before the pump is initially started, the liquid reservoir must be filled manually. When the pump is shut down, a syphon breaker or internal suction check valve retains a quantity of the pumped fluid in the reser-

205

AIR

LIQUID AIR LIQUID MIXTURE

AIR

RECIRCULATION PORT

Figure 1. Internal construction of a self-priming centrifugal pump

voir for successive starts. The internal construction of a self-priming pump is similar to a conventional centrifugal pump except for the addition of a recirculating port in the volute passage that is connected to the fluid reservoir (Figure 1). As the impeller rotates during the priming cycle, the liquid in the impeller and volute passage is discharged out of the volute into an expansion or air separation area of the liquid reservoir. However, before the air separation area of the pump can be filled with liquid and the pump considered primed and capable of generating a pressure against the discharge piping system, air from the suction line enters the impeller eye, causing a drop in pressure due to the relative densities of air and liquid. As the impeller continues to rotate, the mixture of air and liquid moving at high velocity will draw additional liquid into the impeller

Discharge Piping Systems One important fact about the priming cycle is that the air evacuated from the suction line is at a very

PRESSURE AT PUMP DISCHARGE

FULL PRIME

LIQUID OUT

LIQUID REACHES HORIZONTAL RUN AT PUMP SUCTION

AIR OUT

and volute area through the recirculation port. This mixture is then discharged past the volute cutwater into the expansion area of the reservoir. In the expansion section, the liquid and air bubbles separate. The air, being lighter, vents upward out the pump discharge while the heavier liquid returns to the reservoir and continues to recirculate and entrain more air. This cycle continues until all the air in the suction line and impeller is evacuated, at which point the pump is primed (Figure 2). With the pump primed, the reservoir is at full pump discharge pressure. Flow through the recirculation port is minimal as pressures are nearly balanced, and in some designs the port functions as an auxiliary cutwater – thus reversing flow direction. Efficiencies on this type of pump approach efficiencies on standard centrifugal pumps for the same capacity and head range. Suction lifts up to 25ft are attainable with this design, depending on impeller diameter and speed. Close clearances between the impeller diameter and cutwater tip assure that the liquid/air mixture is discharged out of the volute area and not recirculated, which would cut down on suction lift capability. Some self-priming pumps are designed with replaceable or adjustable cutwater tips to compensate for wear. Thus, plant personnel can renew priming lift ability without replacing the entire pump casing.

AIR

AIR BLEED

(A)

(B)

(C)

(D)

UNSATISFACTORY

Figure 4. Discharge piping systems

low pressure at the pump discharge. As noted earlier, little pressure energy can be recovered as low density air is passed through the impeller and separated from the liquid. A typical plot of pressures observed at the pump discharge during the priming cycle is shown in Figure 3. Traditional applications for selfpriming pumps, such as dewatering an excavation or pumping out a flooded basement, normally don’t require discharge piping (Figure 4a). Air from the suction line is easily expelled at the pump discharge until the unit is primed. However, when the application requires the addition of a discharge piping system and incorporates a check valve to prevent backflow or to stop water hammer in high vertical runs of pipe when the pump is shut down (Figure 4b), a problem arises during the priming cycle. When the pump is started, the check valve in the discharge line prevents the air from being evacuated out of the suction line because it cannot develop enough pressure to overcome the head of liquid keeping the check valve closed. With no place for the air to vent, the pump will not prime, which can lead to pump or mechanical seal damage. Figures 4c and 4d illustrate options to prevent discharge piping Vd

Vs

PUMPING TIME PRIMING

PUMPING

Figure 2. The self-priming cycle

206

Figure 3. Typical pressures observed at pump discharge during priming cycle The Pump Handbook Series

Figure 5. Example of discharge piping system containing liquid loop or trap

system backflow yet allow the selfpriming pump to evacuate air from the suction line. The air bleed line in Figure 4c should not be installed below the liquid level or contain any liquid traps to impede air flow from the pump. The air release valve shown in Figure 4d allows the air to escape and seal once the pump is primed. Discharge piping systems containing liquid loops or traps similar to those shown in Figure 5 should be avoided. If the air occupying the pump and suction line volume (Vs) cannot be added to the air in the discharge volume (Vd) without exceeding the low pump discharge pressure during the priming cycle, provisions should be made to allow the suction air to vent as in Figures 4c and 4d.

Suction Piping Systems As applications for horizontal self-priming pumps have expanded, suction piping systems have progressed from a simple suction hose in a ditch to more complex piping arrangements designed for transferring liquids from tanks or rail cars to other storage facilities. The suction system shown in Figure 6 is a piping arrangement for tank car unloading operation. At startup, full discharge pressure is experienced as the pump is filled with liquid. The unit continues to pump the initial volume of suction liquid into the discharge system, and when the air in the suction line is encountered, the discharge pressure drops as the pump enters the priming mode. The location of the initial volume of pumped liquid in the discharge piping system should be

Figure 6. Suction piping system for tank car unloading

reviewed, as previously discussed, to be sure it does not present a restriction to the evacuation of the suction piping air. Again, a bleed line or air release valve may be necessary.

Vortexing Vortexing is the introduction of air into a pump due to insufficient depth of liquid above the pump suction line. High velocity flow patterns at the suction pipe entrance induce vortices or whirlpools in the liquid, and these open up a channel allowing air to enter the pump. When a tank is pumped-down to its lowest level, vortexing may occur. At this point in the pump operation, the discharge pipe system will be completely filled. To continue pumpdown below this level, the air must be able to pass through the pump at reduced discharge pressures. As the liquid flow through the pump is reduced from the air drawn into the suction, the vortexing will subside and the pump will reprime itself provided the air can pass through the pump. Upon re-prime, full flow will be realized and the vortexing will reappear. This alternate primere prime cycle should be avoided since it can lead to premature bearing or seal wear if the frequency is too high. This problem can exist in other self-priming centrifugal pumps. Submersible installations can experience difficulty if air enters the pump suction from vortexing and cannot exit because of the closing of a discharge check valve due to reduced pressure.

Suction Volume The volume of air to be evacuated from a suction piping system should be examined to prevent pump and/or mechanical seal problems. During the priming cycle heat is being added to the fluid in the pump reservoir from the recirculation of the priming liquid. In most cases the mechanical seal is being cooled by the pumped liquid, and in the priming cycle the heat being generated by the seal faces is also being absorbed by the priming liquid. During extremely long priming cycles and high suction lifts the priming liquid The Pump Handbook Series

can evaporate. This means the pump never reaches prime, a situation that can result in mechanical seal failure. Data on most commercial selfpriming pumps include priming time curves in addition to head-capacity characteristics. They show the time required to evacuate vertical suction lines as a function of the distance of the pump above the liquid level. These characteristics are commonly referred to as lift curves. Lift curve characteristics are based on a liquid with a specific gravity of 1.0. If different specific gravity liquids are to be pumped, an equivalent lift should be used to determine the priming time. equivalent suction lift = suction lift x specific gravity Increasing the suction pipe diameter to reduce friction losses, or including long runs of horizontal pipe to be evacuated, lengthens the time required for priming. Increasing the vertical suction pipe size results in an increase in priming time proportional to the square of the pipe diameters. Horizontal runs of suction pipe and varying pipe diameters require a numerical integration of the lift curve based on the location of the diameter changes and horizontal runs to determine the resultant priming time. If system priming times exceed the maximum times shown on the lift curve, the manufacturer should be consulted. In some instances with large suction volumes, an auxiliary line can be installed to add make-up liquid to the pump to prevent the priming fluid from boiling off.

NPSHA/Suction Lifts One important item to check on the suction piping arrangement of a self-priming centrifugal pump is the net positive suction head available (NPSHA). While NPSHA should be reviewed in all pumping applications, it is especially necessary on pumps operating with a suction lift. NPSHA is the absolute pressure available to push the liquid through the suction piping up to the pump suction.

207

NPSHA is defined as: NPSHA = PBAR - PVP - PFR+ PHGT Where PBAR = barometric pressure PVP = liquid vapor pressure PFR = suction line pressure drop due to friction PHGT= height (pressure) of liquid surface relative to the pump Note that all the terms making up NPSHA are pressure terms and can be expressed in feet of liquid. In the case of a suction lift, the last term (PHGT) would be negative as the pump is above the liquid surface. Noting that PBAR (barometric pressure) at sea level is 34ft of water; with high suction lifts, the NPSHA declines rapidly even before the vapor pressure and friction loss terms are deducted. The resultant NPSHA must always be greater (usually with a margin of 2-3 ft) than the NPSHR of the pump to prevent cavitation. NPSHR, the net positive suction head required, is a characteristic of the particular pump. It is determined by test and shown on pump performance curves as a function of flow rate. NPSHR can be viewed as a measure of the ease of moving liquid through the pump, and it increases with flow rate and speed.

Air Leakage While leakage is not normally considered in piping design, it is being mentioned here because it is a

208

primary source of problems in selfpriming pump performance. While a self-priming pump can produce a fairly high vacuum, it is not designed to handle large volumes of air. A small suction air leak at a high suction lift will, in most cases, prevent the pump from ever reaching prime. A self-priming pump with a 3” vertical suction line and a 10 ft suction lift that can prime in 30 seconds has an average air handling capacity of only approximately 1 cfm. The equivalent leakage area to flow 1 cfm at a 10 ft suction lift vacuum is only .002 in2! For this reason gaskets and seals should be in good condition to prevent air leakage during the priming cycle. Packing is not recommended for use in sealing self-priming pumps since it is prone to leakage under the negative operating pressures encountered during priming. If priming difficulties are experienced and air leakage is suspected, a common practice is to isolate the suction piping system from the pump and check the pump for blank vacuum. Blank vacuum gauge readings should be greater than the suction lift requirements. Typically, they run in the 2025 in.Hg vacuum range.

Summary Horizontal self-priming centrifugal pumps have all of the inherent advantages of standard centrifugal pumps such as: • low initial cost • high capacity

The Pump Handbook Series

• high suction lifts • relatively high discharge heads • ease of installation • ease of operation • ease of service • ability to handle dirty or solids laden liquids In addition, a self-priming pump can evacuate the air in the suction line prior to pumping liquid. However, while it can produce high lift vacuums, it cannot discharge the evacuated air against any back pressure. The key to a successful installation is to match the pump to the application, then match the piping system to the pump. Avoid the two common pitfalls of self-primer installations – namely regarding the pump as an air compressor and imposing back pressure on it during the priming cycle, and allowing air leakage in the suction line or pump seal area. Following these guidelines will result in relatively trouble free operation and allow the pump to systematically remove the air out from the suction line. This will result in efficient pumping of liquid, and upon shutdown, the pump will retain enough liquid to repeat the cycle.■ Ray Petersen is Manager of Engineering for the Fybroc Pump Division of Met-Pro Corporation (Telford, PA). He holds a master’s degree in mechanical engineering from Drexel University and has more than 25 years of experience in the centrifugal pump industry.

CENTRIFUGAL PUMPS HANDBOOK

Know the Inside Story of Your Mag Drive Pumps Accurate monitoring of temperature and pressure is key to reliable operation. By Harry Schommer grooves applied and t [°C] there is no defined flow 40 through these bearings PA’ TA Rotor back vanes provided in conven35 Tcs’ Pcs PS tional sleeve bearing design. Liquids such TE as propane, ethylene 30 oxide or methylene T’E chloride provide no lubrication capability. 25 he pump shafts in sealless p4 Similar to the situation magnetic driven pumps are p3 between the faces of held in place by sleeve bearTcs mechanical seals, only 20 ings. Therefore, these beara stable fluid film is ings must be located in the pumped TA required between the liquid. Today, the most common slide faces. The stabili- 15 bearing material is pure silicon carty of this film depends bide. Silicon carbide bearings work on temperature and without problems in low viscosity TE = TE’=NPSH - stable pressure. If tempera- 10 liquids such as chemicals, hydrocar10 20 30 40 50 ture rises inside the bons, solvents, acids, all kind of Minimum flow Qmin Q [m3/h] Abb. 1 pump above vapor temhydroxides and also in abrasive perature of the pumped Figure 1. Internal circulation, temperature behavior pumpage. With an additional dialiquid, vaporization mond-layer, the material provides internal circulation will lead to will break down this film. Under dry running capabilities. demagnetization of the coupling. these conditions, the bearings will The widely used term “process run dry and fail sooner or later. A lubricated bearings” is not quite corTemperature Rise in Single reliable temperature monitoring sysrect since there are no lubrication Stage Volute Casing Pumps tem is required to avoid with Magnetic Couplings this situation. Besides vaporization Internal Circulation of fluid inside the pump, Sealless pumps with magnetic dry running of an empty couplings and metallic containment pump is the worst operatshells generate eddy currents that ing condition. Because of lead to heat and cause temperature the starved suction, there rise of the pumped liquid in the conis no flow to any part of tainment shell. In order to prevent the pump. Although the this, heat must be dissipated through diamond-layer of the SiCan internal cooling flow. This cooling bearings will tolerate this flow – branched off as a partial flow situation because no hyfrom the main flow and led through draulic loads are acting, the gap between internal rotor and the built up heat which containment shell – is shown in cannot be dissipated Magnetic drive centrifugal pumps with metallic Figure 1. because of the starved Editors Note: Mr. Schommer, of Waldkraiburg, Germany prepared this article for Pumps and Systems. To assist our U.S. readers not entirely familiar with SI (metric) units, we have included an english/metric conversion chart at the end of this article.

∆T

T

containment shells

The Pump Handbook Series

209

Temperature Rise, Minimum Flow Conditions Figure 1 further displays the temperature behavior in a volute casing pump with a magnetic coupling. The pump size is 50/200, 2900 rpm; magnetic losses are 3,0 kW and the pumped liquid is water. When reviewing the temperature curve it must be considered that the flow leading through the magnet chamber is dependent only on the geometry of the rotor back vanes and on the speed. This means that, independent from capacity and differential head, a stable circulation flow exists that adopts the dissipated heat and leads it into the main flow. Since the magnetic losses of a given magnet coupling will not change during operation at constant speed, an almost stable temperature increase ∆T is produced in the range to the right of the minimum flow. However,

210

[°F] [°C] ▲Tcs, H2O 16

p[bar]

60

parameters Pv [kW]

15 58 56

Vapor pressure PD

▲T-values based on H2O O recommended Qmin H2O

14

Pcs Liquid

13 17kW

54

12

52

11

50

10

∆Pcs

8.3kW

48

9

PS

8 46

4.3kW

7 44 6

Thermal stable (▲T const.)

5 4

34

3 2.0kW

2 1 1 5

1.5

2.5 10

4 15 20

6

8 10

15

Tcs

Tzul.

TD T[°C]

Figure 3. Containment shell, vapor pressure curve

2.7kW

40

36

TE

Vapor

∆TProduct

5.6kW

42

38

∆TS ∆PD

The circulation flow is drawn from the discharge side behind the impeller, led into the chamber between the slide bearings and through the pump shaft via the rotor back vanes, and returned to the discharge side. This arrangement pressurizes the slide bearings and the containment shell with nearly the full discharge pressure, and helps to avoid flashing of the liquid in this area caused by heating up the product. Where the temperature increase is critical, the maximum pressure P4 prevails. It should be noted that there is no heated liquid flowing back to the suction side or impeller. Therefore, no negative influence on the NPSH required will occur. For this type of pump, handling of volatile liquids is not a problem. In pumps without rotor back vanes or rear impellers, the internal cooling flow is driven by the pressure gradient within the pump from discharge side to suction side, i.e. back to the impeller eye. In this case, problems may arise in pumps handling volatile liquids. Sufficient NPSH reserves must be available to accommodate the heat-conditioned rise of the pump’s NPSH value. Exact temperature measurements of the cooling flow after passing the magnet area are also impossible.

20 30

50

30 40 50 70 90 120 150 200

70 m3/h 300 U.S.G.P.M.

Figure 2. Temperature increase

if the minimum flow drops below, temperatures will rise remarkably. This is the reason why these pumps cannot be operated against closed discharge valve. Experience has shown that most of the slide bearing damages are a result of neglecting this fact. If process conditions dictate this operation, a bypass line must be installed from discharge to suction vessel. Numerous temperature measurements on magnetic drive pumps with different sizes, rotor diameters, containment shell materials and speed have proved a direct relation between the capacity Q, the magnetic losses Pv and the temperature ∆Tcs inside the containment shell. These relationships are displayed in Figure 2, based on water at 20ºC. Determination of the actual temperature increase inside the containment shell for other liquids depends on the product. ∆Tcs,product = ∆TH20 . spec.heatH20 . densityH20 [ºC] spec.heat product density product

Knowing the inlet temperature TE, the containment shell temperature Tcs, product is determined as follows: Tcs, product = TE + ∆Tcs,product [ºC]

The Pump Handbook Series

Maximum Allowable Containment Shell Temperature When handling volatile liquids or products with a vapor pressure that complies with the pump suction pressure Ps, the relation between containment shell temperature, containment shell pressure and boiling point of the liquid must be taken into consideration. Only when the operating conditions are not beyond the boiling point – that is, liquid is not flashing inside the containment shell – can safe operation be guaranteed (Figure 3). The condition point must always be in the liquid state. Basically, the pressure rise ∆Pcs inside the containment shell during operation must always be higher than the heat-conditioned rise of vapor pressure ∆PD of the product. To determine the maximum allowable containment shell temperature Tzul, the vapor pressure curve of the product must be available (Figure 3). The boiling temperature TD can be taken from the intersection point between containment shell pressure at duty point of the pump and the vapor pressure curve. By adding a certain safety margin ∆Ts, the maximum allowable containment shell temperature Tzul can be determined. However, it must always be higher than the calculated containment shell temperature Tcs in order to avoid vaporization. This means a high pressure rise ∆Pcs in the containment shell area provides a higher safety factor. Pumps with internal circulation from discharge to discharge side, through back vanes or rear impellers

Based on the vapor pressure curve for NH3 (Figure 4) and on the calculated containment shell pressure of 10,6 bar, a boiling temperature of +25ºC is given which must not be exceeded during operation. For final determination of allowable temperature Tzul, the actual expected containment shell temperature at the thermal stable minimum flow of 8 m3/h (Figure 2) must be calculated:

(Figure 1), are pressurized in the containment shell area by approximately 80% of the differential head plus the inlet pressure Ps. The containment shell pressure for this pump design can be calculated as follows: Pcs = Ps + ∆Pcs [bar] The pressure increase ∆Pcs depends on the rated differential head H and the density of the liquid: H . p . 0,8 [bar] ∆Pcs = 10,2

Tcs = TE + ∆Tproduct = 0+4 . 1 . 1 = 12,3ºC 0,492 0,66 Considering the boiling temperature of 25 ºC and the actual temperature of 12,3 ºC, the allowable containment shell temperature Tzul can be defined at 20 ºC.

H[mLC], ρ[kg/dm ] 3

Pump series with circulation from discharge side to suction side have lower ∆Pcs values. Exact values can be learned from the pump manufacturer or taken from the pump data sheet.

Temperature Monitoring

essential to monitor magnetic driven pumps with temperature probes to ensure an automatic switch off before serious damage occurs.

Temperature Probes For permanent control, resistance temperature probes are the preferred method of monitoring containment shell surface temperature, although this kind of protection is not activated under dry running conditions. Nor can containment shell rupture by the outer magnets resulting from worn antifriction bearings be avoided. Typically, the probes work with a measurement rheostat of platinum, showing an electrical resistance of 100Ohm at 0ºC. Temperature changes at the measuring point lead to a change of the resistance and, in turn, to a change of the voltage. If the preadjusted temperature limit is exceeded, the voltage change switches off the driver through a connected controller.

Practical Example

General

A standard chemical pump with metallic containment shell (Hastelloy C) and partial circulation as indicated in Figure 1 is required for handling ammonia (NH3). Service conditions are as follows:

Centrifugal pumps with magnet coupling are not considered electrical equipment. Level detectors or temperature control connection head adapter piece devices are not required to motor circuit Spring by government authorities for these pumps, Spring loaded Extension even when they are sealing installed with explosion Ceramic sand proof motors in haz- Cement ardous areas. However, PT 100-element Protection tube Shroud protection application experiences Bottom drive magnet have shown that the internal circulation main reason for operashroud tional troubles – besides driven magnet worn antifriction bearings – is the practice of exceeding the allowable Figure 5. Temperature probe (PT 100) containment shell temFigure 5 shows a standard probe perature. Consequences of excessive adaptable for containment shells. It temperature rise are has a flat bottom for sufficient concavitation in the area of tact to the containment shell surface. the containment shell The element is located directly on and dry running of the the bottom of the probe. Continuous slide bearings due to contact between the probe and conproduct flashing. Caustainment shell surface is guaranteed es for this can be operaby an integrated compression spring. tion below minimum Another type of temperature flow or against a closed probe, located in the internal recircudischarge valve, blocklation flow, measures the fluid teming of the inner rotor, perature after passing the magnet clogged circulation area. This system works sufficiently +40 +50 holes and/or solid partiif the pump is properly filled with cles between rotor and liquid. It protects against exceeding stationary shroud. the boiling point of the liquid in the Therefore, it is

• Liquid: NH3, TE = 0 ºC, Ps = 4,4 bar, Q = 0,66 kg/dm3, C = 0,492 g/cal/ºC • Differential head 120 mLC • Capacity 40 m3/h • Coupling losses Pv = 2,7 kW The expected containment shell pressure Pcs is to be determined first: Pcs = Ps + ∆Pcs Pcs = 4,4+120 . 0,66 . 0,8 = 10,6 bar 10,2 PD [bar] 30 20 Vapor pressure NH3 10,6 10 8 6 4,4

4

2 -30

-20

-10

0

+10

+20

+30 +25

Figure 4. Vapor pressure curve, NH3

The Pump Handbook Series

211

net area). This is only possible in designs with rotor back vanes or rear impellers. If the temperature probe is located in the precirculation flow (this is the case when circulating to the suction side), serious problems occur when pumping volatile liquids. With this setup, an increase in temperature is indicated only when the complete pump has already become hot – too late to prevent flashing in the magnet end.

T 1 , T 2 [°C ] 500

Coupling design: PNK : 21,5 kW / 2900min-1 PV : 2,0 kW

450 400

T1

350 300

T 1 = MAG SAFE T 2 = PT 100

250 200 150 100

T2

50 20 60

120

180

240 [sec]

Figure 6. Containment shell surface temperature

MAG-SAFE Temperature Monitoring

Figure 7 shows the area of the containment shell caused recently developed MAG-SAFE temby excessive temperature rise. perature monitoring system that is The problems of protecting able to read the temperature directly pumps against dry running through on the heat source. It records the temperature probes have already actual temperatures occurring been pointed out. Given an empty between the magnets inside the conpump, i.e., under dry running conditainment shell and converts them tions, it has been proven that the coninto a linear output of 4 to 20 mA. tainment shell surface temperature Therefore, it is possible to preadjust T1 in the center of the magnet couthrough a trip amplifier any shut-off pling (Figure 6) deviates remarkably temperature within the range of -5 to from the surface temperature on the 250ºC. measuring point T2. The reason for Compared with common temthis are the eddy currents occurring in the center of the magnets that rapidly 10 8 increase the containment shell tempera9 ture. The temperature 3 7 2 12 probe at measuring 5 11 6 4 1 point T2 cannot detect this temperature rise in time because of the bad thermal conductivity of the shell material (18.10.CrNi or Hastelloy). Consequently, these type of probes are not able to protect magnets against overheating of an empty pump. To obtain reliable 1. Containment shell 7. Connection piece measurements with 2. Outer magnets 8. Transmitter this type of monitor3. Thermocouple wire 9. Connection clamps ing, the pump must be 4. Thermocouple 10. Cable inlet vented and the probe 5. Connecting block 11. Protection device located at the reverse 6. Thread connection 12. Bearing bracket internal cooling flow (after passing the mag- Figure 7. MAG-SAFE temperature monitoring

212

The Pump Handbook Series

perature control (Figure 5), the MAGSAFE system features: • Extremely fast reaction time to all temperature rises (Figure 6) , and switches off under dry running conditions. • Since the containment shell is a heat source because of the eddy currents induced in it, a temperature rise can be detected before the liquid temperature in the containment shell is remarkably affected. Exceeding the boiling point can be prevented if the limit temperature is correctly set. SI Unit

kW m3/h ºC mLC

Approximate Conversion Factor

Resulting English Unit

1.34 4.4 9/5 ºC + 32 3.28

hp US gpm ºF feet(of head)

(meters liquid column)

• Excessive temperatures at containment shell surface within the Exarea can be prevented. • The direction of the internal circulation flow has no influence on the temperature indication. • Worn out ball bearings cause eccentric run of outer magnets and will lead, if not detected, to erosion at the protection device and containment shell rupture. With the MAGSAFE system, drive magnets cut the connection wire 3 if such a condition is not recorded in time, and the pump is switched off before serious damage can occur.■ Harry J. Schommer is Chief Engineer of Dickow Pumpen KG in Waldkraiburg, Germany, with responsibility for the design and development of the company’s sealless magnetic drive pumps. He has authored several papers on pumps and shaft-sealing systems, and conducted seminars for pump users in Germany, Europe and the Far East.

CENTRIFUGAL PUMPS HANDBOOK

Pump Rebuilding at Avon This refinery’s successful program meets strict emission standards and enhances pump reliability at the same time. By Stephen C. Rossi and John B. Cary ost states have adopted regulations that limit fugitive hydrocarbon emissions from mechanical seals in centrifugal pumps. In California, limits as low as 100 ppm have been imposed. Users, faced with few choices to meet these strict standards, have turned to dual seals and sealess pumps to comply. Many users, however, have found that they can meet current fugitive emission limits with single seals by paying careful attention to detailed retrofit and repair procedures. This article discusses the rigors required to rebuild, maintain and operate pumps successfully – in most cases, with a single seal. These techniques also enhance pump reliability and have been applied to more than 100 pumps in harsh refinery environments.

M

Background Pump History at Avon The Avon Refinery is more than 80 years old. In 1913 crude oil pipe stills were built by a group of San Joaquin Valley oil producers on the Carquinez Straits near Martinez, California. Shortly after building a wharf for receiving the crude oil, they commenced construction of what is now the Avon Refinery a few miles away. The oldest operating process plants date back to the late 1930s. Several facilities were built at the onset of World War II to support the defense effort. They were equipped with high speed centrifugal pumps almost exclusively. New plants have been added every decade since, but

the average age of a centrifugal pump for the whole complex is almost 20 years. Pumps of almost every type were installed over the years. Many were converted from packing to mechanical seals without regard to shaft flexibility and operating point. Over time poorly assembled and maintained auxiliary piping and flush systems were added, modified or abandoned. In addition, unclear or nonexistent operating procedures left operators to use their judgment. All of these factors contributed to poor pump/seal reliability. Evolution of BAAQMD Regulations for Pumps In the early 1970s the California Air Resources Board (CARB) mandated emissions standards for refineries. Two regional agencies were formed to monitor enforcement of the new standards – the Bay Area Air Quality Management District (BAAQMD), and the South Coast Air Quality Management District (SCAQMD). These agencies were also authorized to develop compliance standards for their jurisdictions. Historically, these YEAR

LEAK STD

standards have been much more stringent than federal emissions standards. The first hydrocarbon emissions standards for pumps limited release of Volatile Organic Compounds (VOCs) to 10,000 ppm (parts per million), expressed as methane. The BAAQMD also required no less frequent than quarterly monitoring of pump seals. The district also mandated detailed record keeping. Fifteen days were allowed to repair pumps found over the emissions limit. The equipment could not be returned to service until the emissions limit was met. These limits were attained in most cases by replacing mechanical packing with single pusher-type shaft mounted mechanical seals. As pumps were brought into compliance, the regulators began to “ratchet down” (their term!) on refiners. The regulation became increasingly stringent and, beginning in 1993, requirements went into effect regulating the percentage of equipment that could continue to operate and be put on a future or turnaround repair list (Table 1).

TIME TO REPAIR

1992 & prior 10,000 ppm

15 days

Jan. 1, 1993

minimization in 24 hours, repair in 7 days

1,000 ppm

%WAITING T/A REPAIR

REMARKS

no limit 10%

Spared equip. may not be on T/A list

July 1, 1993

same as above same as above

same as above

Nat Gas added

Jan. 1, 1995

same as above same as above

same as above

Methane

Jan. 1, 1997

500 ppm

1%

Nat Gas added

same as above

Table 1. Summary of BAAQMD regulations for pumps The Pump Handbook Series

213

As these emissions standards went into effect, it became more difficult – and in some cases impossible – to maintain emissions levels in older style pumps. Slender shafts and long spans between bearings created too much shaft deflection at the seal faces. Small seal chambers and inadequate clearances added to the difficulty of retrofitting pumps with modern mechanical seals. At the same time, seal technologies were evolving. Cartridge mounted seals and requirements for larger seal chambers drove Avon managers to embark on a major fugitive emissions reduction program.

Initial Approach In January 1993 the BAAQMD mandated that pump hydrocarbon emissions levels be reduced to 1000ppm. To meet this more rigorous standard, the refinery reliability group took the lead in identifying the scope of the program, with the intent of turning over the project to engineering for execution. The following approach was used for the initial scoping project: • data collection and analysis • seal performance testing and pump evaluation • vendor selection Data Collection and Analysis Data was collected and analyzed to develop a list of pumps requiring work. At this point, the work itself was undefined. Formal fugitive emissions testing was done every quarter. Three years of data on approximately 500 pump seals was analyzed to identify all pumps that had failed their emissions test more than once per year over that period. This information was compared with the refinery’s “bad actors” list and prioritized accordingly. Estimates and schedules were developed for the annual budgetary cycle. Other data sources were reviewed to gain a thorough understanding of the scope of the project. These included: • pump importance and its effect on the process • failure history, including mean time to repair • maintenance cost history (total

214

and per repair) • bad actors lists • hydraulic performance (NPSH and the difference between BEP and actual operating point) • future hydraulic requirements External information was also collected. A benchmarking study was undertaken in other West Coast refineries to determine what their experience had been with various seal designs and seal vendors. Seal Performance Testing and Pump Evaluation The initial data indicated that there were about 150 pumps requiring some degree of retrofitting. The majority of these pumps handled low viscosity products such as propane, butanes and light gas oils. Based on the scope of the project and the time allotted, the decision was made to rely on one seal manufacturer. This would save the time involved in competitively bidding each seal, and it would reduce the overall cost through volume purchases. Three identical pumps in the same service were chosen for initial seal testing. These pumps were selected because they were in light hydrocarbon service and moving liquid near its vapor pressure. In addition, the pumps were in a unit that was going into a maintenance turnaround, providing an opportunity to rebuild all three pumps to identical specifications. Three pump vendors were invited to participate in the project. They were asked to submit proposals to provide their “best available control technology” (BACT), and they were given the specifications for the test pumps. The objectives for the test project were to: • Establish a control technology for light hydrocarbon services to be used on pumps that must be retrofitted to meet future emissions requirements. • Provide single seals designed to meet stringent emissions standards, avoiding expensive and complicated dual seal installations. • Document this technology and The Pump Handbook Series

performance so it could be applied to retrofits as well as new installations. A technical support agreement for the test program was developed for the participating seal vendors, stipulating the obligations of each to provide technical support and assistance as a full partner in seal selection, retrofits and testing. Each vendor assigned a field service engineer to participate in the overhaul, installation, start-up and field monitoring of the test seal. The test pumps were meticulously overhauled and carefully installed. The seals were tested over a three month period, with the following data recorded daily: • emissions readings • suction and discharge pressure • pumped fluid temperature • vibration (radial and axial) • seal flush pressure • quench steam flow Vendor Selection At the end of the seal trial period, field test data were compared with results of the alliance selection process led by the plant’s purchasing group. Vendor performance was weighed against several pre-established performance criteria, such as degree of technical support, seal performance, response time, level of technical expertise and experience. Long term alliance agreements were then drafted, budgets prepared and proposed schedules developed.

Project Approach Defining the Project After establishing the breadth of the project, Avon officials initiated a formal program to implement the work. The program was divided into phases corresponding to annual budget cycles. A project core team was assembled, consisting of a project manager, project engineer, full time vendor technical representative, draftsperson, clerk, mechanical contractors, pump alliance partner and a buyer from the Purchasing Department. Various area maintenance and operating personnel became ad-hoc participants, depending on the location of the retrofit work.

The charter of the project team was as follows: • Develop reliable technology for pump emissions control to comply with the 1993 through 1997 emissions limits. • Implement this technology as proactively as possible to avoid penalties and fines for non-compliance. • Complete detailed designs required to achieve reliable 1993-1997 emissions compliance. • Provide project management support for all phases of equipment upgrades. • Develop and evaluate equipment hydrocarbon emissions data to prioritize and schedule optimal cost effective repairs and upgrades. • Integrate the activities of the project with that of the maintenance and operating departments to minimize disruption in the operation of the refinery. • Recommend and coordinate equipment upgrades meant to improve equipment reliability in conjunction with emissions compliance modifications. Candidate Pump Evaluation Emission levels were measured on all pumps identified as having potential VOC compliance problems. Qualitative data on all VOC program pumps were surveyed to assess the need for upgrades to the sealing system to meet 1997 regulations. Pumps found to be above the 500 ppm limit for 1997 were considered for possible upgrade. The company’s third party contractor for emissions compliance used a data base system to generate queries on all pumps with emissions levels at or above the limit. Detailed analysis of each pump system produced a breakdown of anticipated upgrades and replacements (Table 2). Proactive Approach to Problem Pumps A proactive approach was taken because an in-kind, reactive repair program would have increased cost to the company and created major repair backlogs when more stringent

Number Recommendation of Pumps Seal Upgrade only 96 Power End Retrofit 45 Complete Replacement 5 No Change 24 Total 170

“Proactive” Approach (continuation of modifications and thermal oxidizer installations)

$400,000

1995: 25 emissions repairs per year at $5000 each $125,000 1996: 15 emissions repairs per year at $5000 each $ 75,000 1997: 15 emissions repairs per year at $5000 each $ 75,000

$300,000

TOTAL COST

Table 2. Recommended modifications $500,000

$200,000 $100,000 $0

1992

1993

1994

Figure 1. Annual cost of emission

“Reactive” Approach (in-kind repairs on as needed basis)

1995: 25 emissions repairs per year at $5000 each $125,000 1996: 25 emissions repairs per year at $5000 each $125,000 1997: 40 emissions repairs per year at $5000 each $200,000 TOTAL COST

$450,000

emissions limits went into effect. The cost of emissions related pump repairs made in 1992, 1993 and 1994 is shown in Figure 1. In 1994, 40 pumps required emissions related repairs. This was down substantially from 92 in 1992 and 64 in 1993. The total cost for all repairs during this three year period was more than $1 million. The following projections illustrate the potential for savings with a proactive vs. reactive approach to emissions compliance. (Repairs were expected to remain constant through 1996, then increase by 50% in 1997 when the emission limit was reduced to 500 ppm). Savings over this three year period was estimated to be at least $175,000 as shown in Figure 2. A complimentary maintenance cost savings was also expected due to The Pump Handbook Series

$200,000 $180,000 $160,000 $140,000 $120,000 $100,000 $80,000 $60,000 $40,000 $20,000 $0

1995

$275,000

1996

1997

Reactive Program Costs Proactive Program Costs

Figure 2. Project pump repair costs

increased reliability. In conjunction with emissions history, mean time between repair (MTBR) and hydraulic performance requirements were evaluated for possible upgrade. Experience had shown that inkind repairs of equipment that failed emission monitoring would, in most cases, not comply with 1997 levels due to pre-existing problems such as pipe strain, unstable foundations and poor suction conditions. These conditions were corrected as part of the upgrade process. This approach reduced chronic mechanical seal failures and subsequent emissions, improved long term reliability and lowered total life cycle cost of the equipment. Tremendous effort was expended in identifying pumps to be included in the program. Data from a number of sources was scrutinized to justify each pump’s inclusion. This was an important exercise because it prioritized the upgrade sequence (shortest MTBR to longest MTBR), identified detailed scopes of work, and scheduled each retrofit to coin-

215

cide with other refinery activities. At the end of this process, a well defined list of pumps targeted for upgrade emerged. The plan was communicated to all affected refinery personnel and was used as the basis document for all project work.

Preparation of Specifications Specifications for Pump and Seal Purchase

The project team immediately set out to establish minimum requirements for pump and seal specifications. A series of meetings was held with both the seal and pump alliance partners, as well as with company machinery specialists, to create project specific specifications. The basis for the specifications was API 610 with clarifications in the areas of fits and tolerances. A draft version of API 682, (Centrifugal Pump Shaft Sealing Systems for Refinery Services; first draft 9/92)was also used for guidance. The intention was to create an environment for the seal that would allow it to function as intended. These specifications were later adopted company-wide. Examples of Critical Fits Areas of concern and additional requirements to API 610 are tabled below: Additional requirements: • seal chamber register (radial) to shaft within 0.001 T.I.R. • component match impeller to shaft fit to achieve a goal of 0.000 inches tight to 0.001 inches loose • component match of rolling element thrust bearings to achieve 0.004 inches maximum axial float • shaft sleeve bores equal to the maximum diameter of the shaft with a tolerance of +0.0010 inches to -0.0000 inches Modifications: • Squareness of seal chamber face register to shaft axis was reduced from 0.002 to 0.001 T.I.R. • Sleeves will have a relief centered axially and the minimum sleeve thickness can be 0.090

216

inches within the relieved area. It was agreed that these specifications would be the minimum quality level expected. Repair and Installation Specification The quality of pump installations – including foundation preparation, grout (or lack of it), piping strain, alignment and other key factors – varied considerably throughout the plant. No company repair and installation standard existed other than the original equipment manufacturer (OEM) guidelines and a few rules of thumb. Consequently, the project team created a repair and installation specification. This document covered setting of new pumps and drivers, rebuilding of pumps, installation of retrofit kits, and seal flush, vent and drain connections. The following are important focus areas addressed in the specification: • “As found” pipe strain effects were checked and recorded (Figure 3) prior to pump removal, using a laser alignment tool attached to the coupling. These were checked again when final piping connections were made. • Existing pump base was checked for voids and flatness (Figure 4). Pressure injection grouting and field machining of mounting pads were carried out when tolerances were exceeded. • On non-retrofit pumps, new 17-4PH pump shafts were fabricated and all fits reclaimed to tolerances in the new pump specification. • All pumps, including retrofits, were fitted with close clearance carbon throat bushings to maintain seal chamber pressure. • New dynamically balanced, multiple disk spacer couplings with register fits for the hubs and center section were provided on all upgrades.

DATE ___________________ MACHINIST_______________ APPROVED BY ____________ EQUIPMENT ID ___________ UNIT NUMBER ____________

DIS-ASSEMBLY

ZERO POSITION LASER

HOT ALIGNMENT ANGULAR PARALLEL 0° 90° 180° 270° 360°

COLD ALIGNMENT ANGULAR PARALLEL 0° 90° 180° 270° 360°

TORSIONAL DISPLACEMENT PIPE ALIGNMENT ANGULAR PARALLEL NO DSPLACEMENT L-R F-B

1/4 HOLE 1/2 HOLE 3/4 HOLE

Figure 3. “As found” pipe strain record form

MOTOR

PUMP

USE THE ABOVE TO MARK ANY AREAS OF CORROSION, VOIDS, OR PITS ON THE BASE PLATE INCLUDING GROUT. USE THE FOLLOWING V = VOID C = CORROSION P = PITS

Figure 4. “As found” baseplate condition

• Completed pump and seal assemblies were leak tested before field installation. This minimized the need for rework after the pump system was filled with process liquid. Support Systems Specification and Selection Existing piping and seal systems

The Pump Handbook Series

dure tag installed on nearby piping. The message addressed both seal and pump venting (Photo 2). Instrumentation

Photo 1. Piping reinforcement detail

Only essential – or remote and unattended – pump seal systems were instrumented. Typically, a level switch on a seal reservoir was the only signal back to a control room. Most installations did not warrant remote readout instrumentation since the operators were better informed of the general health of the pumps by observing them in person. Secondary Containment (VRS) A number of pump seal systems used existing vapor recovery systems (VRS) for 100 percent containment when it was available close to the pump installation. In many remote locations total containment was required, but VRS was not available. Alternative Technologies

Photo 2. Seal and pump venting procedure

were upgraded. Most of these pumps had screwed connections (potential emissions sources) throughout. As part of the upgrades for piping reliability, the following were addressed: • All screwed connections on the pump casing or process piping were replaced with Schedule 160 nipples; the nipples did not exceed 4” in length and were gusseted in two planes, seal welded and flanged – eliminating screwed connections if possible (Photo 1). • All tubing connections to the primary seal were 1/2” 316 stainless steel with a wall thickness of 0.065”. Smooth radius bent 3/4 inch tubing was used for thermosyphon cooling when required. To reduce seal flange distortion and make installation and removal of the seal easier, the staff made sure that the final tubing connections were not more than 18 inches long. Vent systems were provided on seal chambers in addition to the pump case to ensure filling of the seal chamber prior to startup. These venting systems included instructions in the form of a venting proce-

Where total containment was required and the only choice was a nitrogen pressurized dual seal arrangement, an alternative involving new emissions control technology was used. This technology utilized compressed air passing through a jet ejector to pull emissions from pump seals or barrier fluid reservoir vents through a flameless reactor that thermally oxidized the VOC’s to water vapor and carbon dioxide, as shown in Photo 3. Four units have been installed with good success; the sys-

tem is 99.99% effective in reduction of VOC’s. Each unit currently controls emissions from 10 pumps. Future expansion up to 20 pumps is possible. These systems are expected to avert as many as 15 emissions-related repairs per year. Many non-compliant pumps are scheduled to be connected to a thermal oxidizer within the next year, possibly eliminating the need for further modifications.

Part 2 S eal performance is affected by numerous internal and external forces. How a pump is sized for an application and how it is actually operated have a significant impact on seal life. In fact, both of these factors can shorten seal life in a pumping system. Mechanical problems such as misalignment, unbalance and flatness, as well as poor concentricity and perpendicularity are fairly well understood and relatively easy to control. Hydraulic forces, on the other hand, are generally not as well understood or recognized by personnel responsible for daily pump operation and maintenance. Furthermore, it is usually more difficult to remedy a hydraulic problem since it may often relate to the original design of the system. Shaft deflection and vibration caused by unbalanced hydraulic forces can be very destructive to a pump and severely diminish seal life. Before embarking on a project to improve seal performance, it is imperative that the pump’s hydraulic performance be verified. The closer a pump operates to its best efficiency point (BEP), the longer the seal will last. This has been demonstrated many times in the field and was recently proven analytically in a computer model specifically designed to predict seal life based on a pump’s proximity to BEP. A discussion of the process for evaluating pump hydraulics is included in the appendix to this article.

Developing Alliance Partnerships

Photo 3. Thermal oxidizer system The Pump Handbook Series

To be lasting, an alliance relationship must be profitable or beneficial for all parties. Partners need to

217

take mutual responsibility to ensure that the desired goals are achieved. One of the first steps the newly formed Avon project alliances undertook was to develop well defined objectives along with a mission statement.

• When a pump was added to the list, based on emissions survey, the pump vendor and company representatives interviewed operations personnel for possible insights on performance deficiencies and opera-

Alliance Team Mission Statement

PUMP EVALUATION SUMMARY

Adopt a fundamental philosophy of decreasing mechanical seal life cycle costs through increased equipment reliability. • Maximize equipment availability • Manage and document change accurately and completely • Improve data quality (new and existing) • Obtain accurate process information from the owner/user • Analyze the root causes of failure • Maintain honest communication about failure by owner • Strive for total buy-in by management and staff – down to the last person The alliance teams also developed matrices to assess the benefit of the arrangement and continue improving it:

tional problems. This information, along with service data, was then assembled into a file. • Process data – including head and flow requirements – and physical data was then collected on the fluid

____ PUMP TYPE___________ UNIT___________NO.______ SISTER PUMPS_________ TRI NO. PU__________________ MODEL______________________ DATE EVALUATED___/___/___

HYDRAULIC PERFORMANCE ORIGINAL

EXISTING

# OF NOTICES ________________ LAST NOTICE ________________ PPM ________________ DATE ________________ MAINTENANCE COST OVER LAST 7 YEARS _______________$ # OF SEAL FAILURES __________ PERIOD_________ MTBF________

%BEP ADEQUATE?

❑ YES

HEAD FLOW

❑ NO

NPSHA/R

YES

VAP. PRES. TEMP. S.G.

NO PRELIMINARY SEAL SELECTION ❑ SINGLE ❑ TANDEM

NO

IS THE SEAL CURRENTLY MEETING THE 1000 PPM LIMIT? ❑ YES ❑ NO YES

POSSIBLE REVAMP? ❑ YES ❑ NO

AXIAL

CASING DESIGN ❑ RADIAL ❑ AXIAL

YES

RADIAL

AFTER 1960 DESIGN? ❑ YES

❑ NO

AT LEAST 1 NO

CS or SS

PUMP CASING MATERIAL ❑ CS ❑ SS ❑ CI ❑ NI

DESIGNED FOR SEALS?

Dollars 1) cost of new seal purchases 2) cost of seal repairs 3) inventory reduction 4) market share

Reliability

❑ YES

❑ YES

❑ YES

❑ NO

SUFFICIENT STUFFING BOX AREA FOR TANDEM CARTRIDGE? ❑ YES

❑ NO

Contractor Performance

STEEL BEARING HOUSING

❑ YES

❑ YES

❑ NO

❑ NO

ALL 7 YES

The key is to involve alliance partners in all facets of project activity. Candidate Pump/Seal Evaluation Both the pump and seal alliance partners participated in all facets of the pump/seal evaluation.

REVAMP ❑ RETROFIT BACK PULLOUT ❑ OTHER

SEAL UPGRADE

1) appropriate paperwork 2) on-time payment 3) provides pump access

____/____/____

STUFFING BOX VENT?

BEARING HOUSING ACCEPTABLE?

Company Performance

RESCHEDULE

❑ NO

1) number of repairs 2) mean time between failures 1) plant-wide pump survey status 2) on-time delivery 3) failure analysis submittal

CI or NI

❑ NO

SHAFT DEFLECTION AT SEAL < 0.002 INCHES?

$___________

$___________

FIELD MODIFICATIONS

$___________

SEAL SYSTEM

FIELD MODIFICATIONS

$___________

PIPING

$___________

FIELD MODIFICATIONS

$___________

PIPING

$___________

INSTRUMENTATION

$___________

PIPING

$___________

REMOVE/INSTALL

$___________ $___________

INSTRUMENTATION

$___________

REMOVE/INSTALL

$___________

PUMP MODIFICATION (INCL CPLG)

$___________

PUMP MODIFICATION (INCL CPLG)

_______________________

$___________

_______________________

$___________

TOTAL

$___________

TOTAL

$___________

REMARKS

Figure 5. Pump evaluation summary

218

NEW PUMP

SEAL SYSTEM

SEAL SYSTEM

The Pump Handbook Series

$___________

INSTRUMENTATION

$___________

REMOVE/INSTALL

$___________

PUMP MODIFICATION (INCL CPLG)

$___________

_______________________

$___________

TOTAL

$___________

SEAL PROPOSAL ADDENDUM PUMP#/UNIT: PU.... / PUMP REPAIR TYPE: SEAL: PRODUCT: TEMPERATURE: SUCTION PRESSURE: VAPOR PRESSURE: DISCHARGE PRESSURE: SPECIFIC GRAVITY: API PIPING PLAN: SUCTION RETURN?: QUENCH?: PROPOSAL#: TOSCO CATALOG#: PUMP TYPE: NUMBER OF BOXES: COMMENTS:

FUGITIVE HYDROCARBON EMISSIONS PROJECT PUMP COMMISSIONING CHECK LIST

NEW SEAL ORDER CHECKLIST FOR OVERHAUL PUMPS

PUMP NO.:_______________

PUMP NUMBER: _______________

MACHINISTS:

F PSIG PSIA PSIG

SIGN-OFF: STEVE ROSSI_________ GIL TIGNO__________ Figure 6. Seal proposal addendum

being pumped. The information, which included vapor pressure and solids concentration, was then summarized on a Pump Evaluation Summary form (Figure 5) and used to determine the best fix based on pump type, emissions, maintenance history and performance data. Selection – The pump and seals alliance consultants submitted proposals for the agreed upon upgrades. Attached to each was a seal proposal addendum (Figure 6) that provided design details for construction. The seals consultant also prepared a new seal order checklist (Figure 7) to further define the construction details. When field measurements were required, the pump was taken out of service and checked to ensure that all the components fit precisely. Installation & Startup – After the installation was completed, a QA/QC evaluation was made, and the Pump Commissioning Check List (Figure 8) was signed by the project representative and the operator prior to startup. The seal vendor usually witnessed the startup and recorded initial emissions levels. Living Program Maintenance – As part

FIELD MEASUREMENTS

DATE:__________________

Name:____________________________________

1. COUPLING GUARD SECURE

_____

2. HOLD DOWN BOLTS INSTALLED

1) Physically verify: [ Shaft diameter___________ Bolt circle_____________ Stud size_________________ First Obstr____________ Gage Ring dist____________ Bolt Orientn__________

]

2) Suct Press___________ Disch Press ___________

[

]

3) Rotation from driver end - CW/CCW

[

]

4) Temperature ___________ 5) Make a diagram of the bearing web and the exis- [ ] ting seal piping. Where can new seal piping be located? 6) Note cage ring tap location(s). Can the seal box be [ vented through the cage ring taps?

]

7) Is the existing seal the same model as indicated by [ the files?

]

8) What is the O.D. of the current seal gland.

[

]

1) Verify the vapor pressure if possible. Make sure [ that the box pressure is sufficient to keep adequate vapor suppression.

]

EVALUATIONS

2) Design the seal flush piping system including [ ] orifice sizing and throat bushing clearance in order to get the required flow and vapor suppression. 3) Make sure there is adequate room for an O-Ring [ ] groove and multi-port injection between the box bore and the inside the stud holes. (especially important if box is to be bored) 4) Verify that the seal selected will fit. OK any box [ ] boring that will be required with Tosco and pump manufacturer.

Figure 7. New seal order checklist

PUMP _____ MOTOR _____

3. FLANGES PROPERLY MADE-UP

_____

4. JACKING BOLTS BACKED OFF

_____

5. SEAL DRIVE COLLAR BOLTS TIGHTENED

_____

TORQUE: ___________________ 6. SEAL SETTING PLATES ROTATED AWAY FROM COLLAR COMMISSIONING ENGINEER:

_____ Name:___________________

1. PROPER OIL LEVEL

PUMP MOTOR PRESSURE GAUGES INSTALLED/ORIENTED PIPE PLUGS INSTALLED GASKETS INSTALLED SMALL BORE PIPING SUPPORTED NOMALLY-CLOSED VALVES CLOSED ORIFICE PLATE INSTALLED WITH INSCRIBED TAB COOLING WATER FLOWING VENTING PROCEDURE SIGN POSTED AND VALVES TAGGED 10. AREA CLEANED UP 11. LOCKS/TAGS REMOVED 12. VENT PROCEDURE DELIVERED 13. SCREWED PIPING LEAK TESTED 2. 3. 4. 5. 6. 7. 8. 9.

_____ _____ _____ _____ _____ _____ _____ _____

_____ _____ _____ _____ _____

IF THERE ARE ANY PROBLEMS WITH THIS PUMP AFTER COMMISSIONING CALL STEVE ROSSI AT EXT.3263

Figure 8. Pump commissioning check list

Conclusions and Recommendations Accomplishments

of the long range compliance strategy, data is still being collected on all VOC equipment to determine future direction. Some areas where the alliances are now concentrating their efforts include: • Providing on-going training for maintenance and operations personnel featuring detailed information on seal installation and operation. It is expected that this training will greatly increase the MTBR and reduce life cycle costs of pumping systems throughout the refinery. • Continuing to develop records on seal life, failure analysis and life cycle costs, with focus on solutions to bad actor pump/seal systems. • Incorporating seal and pump parts and repair services under a single manufacturer for each.

The Pump Handbook Series

• All of the 102 pumps modified in the project beginning in 1991 met 1993 emission limits. • More than 60% of these pumps had initial emissions levels of 1000 ppm or more, and the MTBR initially averaged 8 months; after retrofitting the MTBR has increased to an average of 16 months. • The alliances established criteria for cost effective procurement of pumps and seals, and they provided accessibility to the most current technology resources. • An engineering standard for further VOC pump upgrades and a repair and installation standard for pumps and seals was put in place. A key contributor to this success was the ability of those involved to

219

view the solution as an overall system of modifications. Pre-existing conditions such as pipe strain, unstable foundations and misalignment were corrected to eliminate vibration, stresses and distortions. As an added benefit, the reliability and safety of the equipment improved, thus lowering equipment life cycle costs. Correlation of Bad Actors to Emissions Compliance From 1990 to 1992 numerous inkind maintenance repairs were made on equipment in response to emissions violations. Most of these repairs lasted only 3 to 6 months before another violation notice was received. The upgrades undertaken have in many cases doubled or tripled the time between emissions failures and eliminated chronic reliability problems. Lessons Learned There are three primary causes for premature seal failures and/or excessive vapor emissions from upgraded pumps: • installation errors • changes in the chemical composition of the pumpage • operational and hydraulic problems (such as dry running and cavitation) The first problem is the most controllable. The others are more challenging and require continuous education and training. In addition to initial equipment installation, an improved focus on equipment reliability through troubleshooting to resolve premature failures is needed. This is expected to take the form of additional training for both maintenance and operating personnel, revision of operating procedures, and continuous measurement of MTBF and life cycle costs. A skilled team of dedicated experts can rebuild a pump perfectly and still fail to achieve the final objective if the system is not started up and operated properly. A number of details must be attended to in order to achieve success: • Include process operators in the installation process. Have the pro-

220

duction department assign responsible operators to the start-up team. Communicate their responsibilities for a proper start-up and continued operation. Then conduct training on any special requirements of the seal system. Use seal and pump partners to develop materials and provide training. • Develop pre-startup checklists that include the following procedures: a. Steam, flush and purge the pump casing prior to introducing product. (Minimize the time spent doing this to avoid contamination and overheating in the seal chamber.) b. Prepare for hot alignment checks (P.T. > 300ºF and steam turbine driven pumps). c. Review existing pump start-up instructions. • Prior to starting a pump, gather responsible core team members together, including alliance partners. Review the start-up procedure and the duties of each team member. Develop a start-up checklist that incorporates the following information: a. Pump start-up procedures (including venting of all air and vapors from the seal chamber prior to and during start-up.) b. Expected normal, minimum and maximum operating parameters (flow, temperature, pressure, viscosity, cooling, etc.) c. Performance parameters, including suction and discharge pressures, flow temperatures, suction strainer differential pressure and so on. d. Program for continuous monitoring after start-up. e. Troubleshooting guidelines for operators and mechanics. • Pump and seal alliance partners should be full participants with users in the successful commissioning and operation of retrofit pumps. As stated, they conditionally guarantee their equipment if all repair, installaThe Pump Handbook Series

tion and start-up conditions are met. For this project, the seal alliance partner guaranteed that fugitive emissions levels would not exceed BAAQMD limits for three years of continuous operation. • The ability to exercise a warranty is dependent on good documentation. Post start-up documentation requirements must be agreed to with alliance partners as part of the initial parameters of the arrangement. As a minimum, the following data should be collected: Fugitive emissions levels – Initially, the project team collected this data monthly until levels stabilized, at which time the monitoring was turned over to the contractor responsible for collecting quarterly compliance data. Vibration data – This information is also taken more frequently in the beginning, to catch infant mortality-type failures. When readings stabilize, then routine (documented) monitoring can resume. The importance of documentation can not be overstated. Proper documentation [API Standard 682, Shaft Sealing Systems for Centrifugal and Rotary Pumps, First Edition, October 1994] is required throughout the entire process from start to finish. The minimum requirements are listed below: Seals 1. Completed API Standard 682 data sheets 2. Cross sectional drawing of all seals (modified typical) 3. Schematic of any auxiliary system (or systems) including utility requirements 4. Electrical and instrumentation schematics and arrangement/connections 5. Seal manufacturer qualification test results, if specified 6. Detailed cross sectional drawings of all seals (specific, not typical) 7. Detailed drawing of barrier/buffer fluid reservoir (if included) 8. Detailed bill of materials on all seals and auxiliaries

On-going Program Performance

alliance partnerships cannot be overemphasized. These relationships require constant nurturing and attention. Resistance to using the alliance will be an ongoing issue for the core team members. The alliance must constantly review the performance of the partnership itself and compliance with stated goals and objectives. A formal and periodic review process should be formulated. Additionally, long term issues such as how to provide continuous improvement (CI) to the alliance relationship are important. There are many facets to CI, but it typically involves empowering employees to pursue improvements actively. It also means providing technical support at the front line, rigorous root cause failure analysis and use of advanced analytical techniques. Along with CI is the need to maintain the new way of doing business. This includes purchasing quality spare parts, maintaining quality control and standardization, doing meticulous pump and seal overhauls, developing consistent, detailed documentation, and retaining a highly skilled and motivated work force.■

In closing, the importance of continuing to work within the

Stephen Rossi is Principal Engineer for Rotating Machinery with Tosco

9. Material safety data sheets on all paints, preservatives, chemicals and special barrier/buffer fluids 10. Installation, operation and maintenance manuals 11. Pre and post start-up checklists 12. Routine performance monitoring data sheets Pumps 1. Completed API Standard 610 data sheets 2. As-found and as-built specifications (rebuilt pumps) 3. Pump manufacturer performance test results, if specified 4. Detailed cross sectional drawings of all pumps (specific, not typical) 5. Detailed bill of materials on all pumps and auxiliaries. 6. Material safety data sheets on all paints, preservatives and chemicals 7. Installation, operation and maintenance manuals 8. Installation checklists 9. Pre and post start-up checklists 10. Routine performance monitoring data sheets

The Pump Handbook Series

Refining Company, a Division of Tosco Corporation. Steve re-joined Tosco to direct the engineering effort for the company’s Fugitive Hydrocarbon Emissions Project in 1992. Previously he spent eight years at Chevron as a mechanical equipment consultant on pumps, turbines, engines and compressors. John B. Cary consults in the area of reliability improvement and is currently a Manager in the Maintenance Services unit for ERIN Engineering and Research in Walnut Creek, CA. Mr. Cary has more than 20 years of experience in the hydrocarbon processing and petrochemical industry. Prior to joining ERIN, he was a Reliability Superintendent for Tosco. Editor’s Note: This article has been reproduced with permission of the Turbomachinery Laboratory. Edited from Proceedings of the Fourteenth International Pump Users Symposium, Turbomachinery Laboratory, Texas A&M University, College Station, TX, pp. 2534, Copyright 1997. Acknowledgment: The authors would like to thank Gil Tigno of the Tosco Refining Company for his contributions to the success of this program.

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CENTRIFUGAL PUMPS HANDBOOK

Evaluating Sealless Centrifugal Pump Design & Performance Caught in the sealed versus sealless crossfire? This article sheds light on the appraisal process so you can determine the best fit for your service. By Dave Carr

ealless pumps have been available for more than 50 years, yet it has been only during the last decade and a half that they have gained popularity in North America. Sealless pumps are used much more often in the European and Japanese markets. It is not coincidental that North America’s interest in this category of pumps coincides with legislation aimed at protecting the work place and the natural environment through increased safety and minimization of fugitive emissions. Many reviews have previously been published regarding the two drive technologies that dominate this class of sealless pump – canned motor and magnetic couplings. Paramount to both designs is the fact that no dynamic shaft seal is required to contain the pumped fluid. Rather, a stationary containment shroud is used to isolate the pumpage from the ambient environment. This is possible because power is transmitted across the shroud through magnetic lines of flux that induce rotation of the impeller shaft. The significance of this design feature is the fact that sealless pumps can negate leakage that is a fundamental byproduct of the typical mechanical face seal. The corresponding benefits to personnel and the environment are obvious. The chemical processing industry (CPI) is a leader in the adoption of

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sealless centrifugal process pumps and a model for their use in North America. The CPI is currently experiencing a heightened awareness of volatile organic compound (VOC) and volatile hazardous air pollutant (VHAP) emissions as a result of growing regulatory requirements. Today’s environmental regulations also have related monitoring and reporting costs in association with the 1990 Clean Air Act. Sealless pumps fit within the broadest exemption from emission instrument inspections due to their compliance with the without an externally actuated shaft penetrating the pump housing definition [Ref.1]. These costs vary from plant to plant but, in aggregate, can account for a substantial budgetary allowance with equipment that is pumping volatile liquids. Likewise, they represent an input that should be included within an assessment of “total evaluated costs” (TEC) – i.e., all costs accrued from inquiry through the life of the pump. Numerous studies show that shaft seals are the principal cause of failures in chemical process pumps. One such study [Ref. 2], of more than 1250 operating pumps (and more than 2500 installed pumps), indicates that approximately 50% of the recorded primary failure causes were the result of leaking seals. Additionally, officials for a major chemical manufacturing plant [Ref. 3] have The Pump Handbook Series

concluded that their pump repairs are 30 to 60% less costly/frequent with canned motor and magnetic driven sealless pumps than with chemical duty pumps equipped with multiple mechanical seals. It is, therefore, important to recognize the role that maintenance/operating availability should play in the pump evaluation process. When factored into the TEC equation, it has often been cited as the component that sways a decision toward a sealless offering. Consequently, the need for a succinct “sealed versus sealless pump” decision process has surfaced. The following discussion reviews a logic stream that can be used to assist in this exercise. It is reasonable to expect that the outcome of any analysis model can not be considered absolute since customer-specific application and experience-sensitive inputs are often required to make an informed equipment selection. The basic tactic, however, can be used to determine when a sealless design is a viable choice.

Sealless Hydraulic Capabilities Performance capabilities to 2000 gpm and 750 feet TDH (though not necessarily simultaneous) are representative of widespread sealless pump experience and coincide with

single suction, single stage, centrifugal pump capabilities. Production designs are available to further extend this region, but field experience has been typically limited to head and flow combinations that fit within 200 horsepower. Requirements outside of this range will likely mandate the use of larger single stage, multistage or high speed centrifugal pumps or positive displacement designs. The vast majority of sealless manufacturers’ equipment is built around 150 and 300 pound ANSI flange ratings, which are consistent with the needs of a typical chemical processing plant. Once it has been determined that your service conditions roughly fit into the range of today’s sealless centrifugal pumps, an appraisal process to determine the best fit can begin. Pump users are particularly challenged when operating with liquids that are characterized by one or more of the volatile, toxic, flammable or corrosive definitions. In the CPI, this represents a broad range of fluids that include acids, alkalis, salts, esters, hydrocarbons, monomers/ polymers, alcohols, ethers, halogenides, nitrogen/sulfur compounds and even some extreme water conditions. Some characteristics common to these broad definitions are a propensity for the fluids to be unstable, poisonous, noxious, dangerous, destructive or chemically reactive with air. They may also precipitate components, solidify easily, or need to be handled at extreme temperatures. Sealless pumps are well suited for these applications, but some basic application criteria need to be established to ensure a troublefree installation.

Look at the Liquid A qualification of the pumped liquid is the first major decision point when considering the use of a sealless pump. Early applications were relegated to services addressing a need to meet emissions limits for volatile fluids (VHAP/VOC) or where mechanical seals had proven ineffective. Often those applications were associated with situations in which fluid volatility was a significant application parameter. It should immedi-

SEALLESS CENTRIFUGAL OPERATING CHECK LIST Liquid Types: hazardous, or volatile vapors, toxic, flammable, corrosive heat transfer fluid Flows: <2000 gpm <750 ft TDH Field experience typically limited to head/ flow combinations that fit within about 200 hp Temperatures: <750°F (without auxiliary cooling) <850°F (with auxiliary cooling) >-140°F (specialized designs to -260°F) Viscosity: >0.15 cP <200 cP (Note: Lower viscosities may require bearing design changes to support shaft) Solids Content: <150 Micron 3-5 wt % ately be recognized, however, that a simple change from sealed to sealless options is not appropriate without a corresponding analysis of the strengths and suitability of each pump. By design, a mechanical seal face requires a liquid film to act as a cooling and lubricating medium between the stationary and rotating members. Corresponding leakage across those faces, be it ever so slight, is a necessary consequence to promote acceptable seal life. This situation is complicated with volatile fluids having vapor pressure characteristics that result in a change from liquid to vapor states at atmospheric pressure. Consequently, leakage can cause liquid mechanical seals to be a source of fugitive emissions even when operating properly. A hybrid category worth specifically noting is heat transfer fluids (HTF), which can exhibit one or more of the above mentioned attributes. Many users believe that HTF applications represent the single most prevalent application for sealless pumps in the CPI today since those services transcend safety, environmental and maintenance issues. Any of the three criteria by itself can be used to justify a sealless pump purchase decision and, collectively, they represent a prime opportunity to exploit the leak-free nature of sealless pumps. Only sealless and multiple mechanical seal (either liquid or gas configurations) pump designs should be considered to meet demands of liquids in the volatile, toxic, flammaThe Pump Handbook Series

ble, corrosive or HTF categories. Applications that fall outside of those parameters can reasonably be controlled with single mechanically sealed pumps as an alternative to more sophisticated designs.

Frigid or Fervid? Extreme temperature conditions represent good opportunities for the use of sealless pumps. This belief is attributable to the fact that such conditions present a difficult situation with regard to maintaining the necessary seal face integrity that complements low leakage. High values are commonly associated with reference to “extreme” temperature conditions. It is generally accepted that a “high temperature” definition refers to operating levels greater than approximately 250ºF (121ºC), and that temperatures above this threshold act as a handicap to the operational ease, simplicity and effectiveness of liquid sealing. The development of high strength magnets and insulation materials has yielded sealless designs capable of withstanding temperatures as high as 750ºF (399ºC) without the use of auxiliary cooling. Current technology limits cooled versions to as high as an 850ºF (454ºC) rating, but these limits are continually being challenged by new product designers intent on extending application limits. Cryogenic extremes, however, are often disregarded when considering the use of sealless pumps. Most sealless pump manufacturers rate their standard equipment for levels

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to -140ºF (-96ºC), and specialized designs are available for temperatures approaching -260ºF (-162ºC). A side benefit to a canned motor pump’s operation with cold liquids is the fact that increased cooling capacity can increase the power transmission capability of a particular frame size. The basic advantages that attract pump users to sealless designs for high temperature services are equally pertinent for low temperatures. The pumped liquid’s sensitivity to temperature changes must also be considered to complement its discreet temperature evaluation. This includes liquids that have peculiar melting and/or freezing points, e.g. MDI, TDI and many acids, and ones that experience polymerization or crystallization with accompanying temperature changes, such as formaldehyde. Some liquids (caustics, for example) experience crystallization when they come in contact with air. This can occur as the liquid leaks across a conventional seal face or when air is drawn into the pump – e.g., in a system where the pump’s suction is less than atmospheric pressure. These circumstances are troublesome to mechanical seals since the crystal residue can be abrasive to the sealing faces and inhibits face adjustments with accumulation in the secondary area. Also, in vacuum applications process disruptions can occur with the introduction of air. Sealless pumps address these problems by nature of the pumped liquid’s absolute isolation from the ambient environment and the ability to add or subtract heat from the system. Jackets are available for drive sections, bearing housings and pump cases to meet minimum or maximum temperature constraints of a particular liquid. For the successful application of pumps in such services, however, the pump user must disclose a liquid’s temperature sensitivity characteristics. It is imperative as well to recognize that unobstructed flow paths, i.e., without the accumulation of solids, crystals or polymers, are essential to ensure lubrication and heat dissipation within the drive section. If a heat medium is not available for liquids that have the propen-

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sity to polymerize, crystallize or freeze, an externally flushed sealless or a multiple mechanical sealed pump should be considered.

Thick or Thin? Centrifugal pumps are generally applied to relatively low viscosity liquids, and hydraulic-end viscosity corrections are the same for a sealed or sealless design. When high viscosity liquids are passed through the drive section, however, increased parasitic losses occur. The Hydraulic Institute specifies that an external flush fluid should be used when the viscosity exceeds 200 centipoise [Ref. 4]. Any simplified evaluation of viscosity’s impact toward parasitic losses must recognize the fact that smaller/slower speed designs are much less affected by viscosity changes. The correction factor corresponding to the drive section’s coolant/lubricant flow is an exponential step function, rather than a linear relationship, therefore parasitic losses must be well understood. Special designs are available to extend a sealless pump’s viscous handling capabilities, but the earlier comment on the value of specific experience, be it the manufacturer’s or user’s, is again appropriate. On the opposite extreme, many liquids described in the “Look at the Liquid” section of this article have viscosities that are less than that of water and can benefit from the use of sealless pumps. Included within this group are liquids such as HF acid, ammonia, fluorocarbon refrigerants, hot water, heat transfer fluids and hydrocarbons, all of which are regularly used in many CPI processes. It should be recognized that there are long-standing debates about the acceptability of such liquids with product lubricated bearings. Experience has shown, however, that they can conservatively be used, with well designed bearing systems, down to a 0.15 centipoise level, and some manufacturers have exhibited experience at even lower values. Properly designed product lubricated bearing systems result in legitimate sealless pump alternatives to multiple mechanical seals for low viscosity services. A liquid’s contamination with The Pump Handbook Series

solids is another critical concern in the appraisal of a sealless pump’s viability. Various manufacturers’ literature specifies a range of 2 to 6% by weight. A 3 to 5% guideline is a slight compromise from the maximum published range, but it accurately addresses the capabilities of most manufacturers and offers a reasonable rule of thumb. Clarification with regard to the size of the solids, however, is required since sealless pumps have small passages within the drive section circuit. A 150 micron judgment limit allows approximately a 4:1 safety factor to the clearance between the rotor and containment shroud with worst case designs. Invariably, one of the smallest clearance locations is at the shaft (sleeve) to journal bearing. This clearance typically ranges between 0.001 and 0.0035 inches (radial) for many manufacturers. Therefore, further analysis is required to predict the impact of the concentration, size and abrasiveness of solids with regard to the materials used for the bearings and other components of the drive section. Pump applications that fail to pass the critical questions discussed within the “liquid type” sections may be more economically serviced by a single mechanically sealed design. Extreme temperature and pressure conditions often prove to be an exception to the single seal decision and present a competitive opportunity for the sealless pump. There are times when the leak-free nature of sealless pumps is desirable, but the temperature sensitivity, viscosity and solids content characteristics of the pumped liquid converge to demand an examination of a barrier liquid’s viability. The reasonableness of using a barrier fluid to provide the drive section with a higher quality coolant/ lubricant is analogous to clean liquids that would support mechanical seals (API-32/52/53 seal support plans). Disparities are generally the result of supply and consumption differences between the sealed and sealless configurations. Sealless pump manufacturers have become extremely innovative with flush designs for sealless drive sections,

Figure 1.

which effectively isolate the pumpage from that area. If a barrier fluid analysis finds that the sealless approach is not practical, the choice defaults to multiple mechanical seals, which will still require an ancillary support system.

The Internal Flow Circuit Many of today’s sealless pump applications are with liquids exhibiting good thermal stability. When that is the case, the basic design understanding of all manufacturers will ensure stable pump operations. Volatile liquids, however, are an application field in which additional scrutiny is required for operating success. We have previously discussed the fact that a sealless pump’s internal circulation system is used to supply the bearings with lubricant and to dissipate the heat generated in the drive section. The corresponding flow paths are demonstrated in Figure 1 for a typical magnetic driven pump. Inadequate flow and/or pressure will ultimately result in distress to the bearings and, if unchecked, lead to ultimate failures. Active participation in the equipment selection and operation processes will minimize these problems. The sealless decision must, therefore, include a thorough engineering analysis for such services so as not to trade a seal problem for one with a sealless pump’s drive section. Two critical liquid parameters that must be understood are vapor pressure and specific heat. The pump designer requires vapor pressure information at the rated temperature and at some level greater than that to understand the fluid’s state as a

result of pressure and temperature changes of the drive section’s flow. Normally, vapor pressures at rated temperature plus 10ºF (5.6ºC) and 20ºF (11.1ºC) from suction will give the pump manufacturer sufficient information to make an informed analysis. A fluid’s specific heat value is critical since most manufacturers’ temperature rise analyses are derived from testing with water, which can absorb considerably more heat than many of the liquids in today’s critical processes. A thorough understanding of coolant flow, pressure and temperature must be demonstrated to support troublefree pump operation in the field. Pumps operating on volatile liquids typically exhibit minimum flow constraints that are dictated by thermal, rather than mechanical, limits. One of Murphy’s laws is that pumps never operate at their design point. Therefore, the pump supplier must also conduct his analysis at offdesign points to establish a recommended operating “window” – i.e., a minimum to maximum flow regime – to give good operating flexibility. There are many successful sealless pump applications with steep vapor pressure liquids attesting to the fact that reputable manufacturers possess the ability to meet these demanding services. Ammonia, isobutane and propane are examples. The effort expended to conduct this flow circuit analysis will ensure that the drive section is continually supplied with a liquid and not a gas. This situation is analogous to the topic of dry running a pump. There are a number of design means that can be used to minimize the effects of a loss of pumped liquid, including impellers with special or no wear rings, impregnated or coated bearing materials and nonmetallic shrouds. As with any pump, whether sealed or sealless, the goal should be to prevent or at least to minimize the occurrence of a pump being operated without liquid. The successful application of The Pump Handbook Series

sealless pumps on numerous loading and unloading services is one example of the viability of the design under dry running circumstances. Success is conditional, however, upon the minimization of the dry run duration and having sufficient instrumentation to protect the machine as configured. With a sealless pump, it must be recognized that the drive and pump sections may run dry and equal care should be exercised to prevent either occurrence. Recent technological advances have been made with non-invasive electrical devices that can sense the fluid condition – liquid, gas or two phase state – within the sealless pump’s drive section, thereby ensuring reliable pump operation.

Instrumentation Too often sealless manufacturers are confronted with installations in which no precautions have been made, supplied instrumentation has not been installed or excessive instruments (which yield unnecessary complications for field personnel) have been employed. A balanced instrumentation plan is recommended to preclude the possibility of these occurrences. Flow measurement is the surest way to guard against a loss of pump flow, but it is typically not available in most chemical plant installations. A pseudo measurement of flow is a change in the driver’s power requirement. For electric motors greater than approximately 10 horsepower, a current sensing device works well, but smaller motors may be better protected with watt meters. This protection should not, however, be used as a minimum flow protection device! Power or current sensing does not have the sensitivity to act in that manner but can readily indicate the onset of a “no-flow” condition. Pump flow and power instruments will not indicate a vaporization state in the drive section that has been caused by blocked flow passages or insufficient cooling. A temperature detection device, such as an RTD or thermocouple,is commonly used to indicate that condition and can be positioned to sense the temperature of the containment shroud.

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It is generally accepted that a nominal 20ºF (11.1ºC) setting, greater than the calculated temperature rise, will protect against a dry run occurrence. Further, insufficient cooling will cause a CMP to experience a rapid increase in the stator winding temperature, a fact that highlights the need to always wire its thermostats. To meet electrical safety standards, Europeans have used the science of ultrasonics to ensure a “wet” condition within the drive section of their canned motor pumps. Innovative adaptations to this technology are finding their way into sealless pump designs in North America to guard against overheating a motor or magnet assembly. Similarly, instruments are now available to indicate a canned motor pump’s direction of rotation without physically seeing the shaft turn. This advancement addresses a long-standing complaint by users resulting from instructions to observe a dry run bump start to guarantee proper motor wiring. A number of manufacturers now offer bearing wear detectors for use with carbon bearings. These instruments can be either electrical or mechanical and do a good job of alerting the user to an impending minimum material condition. Consequently, routine maintenance can be scheduled in a manner consistent with production needs. Growing competition in today’s sealless pump

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marketplace has prompted manufacturers to take a more active role in the application of pumps, and their counsel should be sought when developing a sealless pump instrumentation plan for the first time. This article has attempted to demonstrate that sealless centrifugal pumps are a valuable asset in the arsenal of today’s pump users. The fact that they completely eliminate fugitive emissions complements increasingly strict environmental and safety objectives. To capitalize on a sealless pump’s cost effectiveness and inherent reliability, however, applications should be well understood. Wisdom in this area results from a combination of complete customer input, a manufacturer’s thorough technical analysis/application expertise and adequate protective instrumentation. Worldwide experience shows that these pumps should not only be considered for new installations. Many conversion opportunities exist for the retrofit of existing, poorly performing pump installations. Manufacturers’ standardization with ANSI dimensions facilitates this process and makes retrofitting a cost effective alternative to living with leaking pumps. There will continue to be a need for mechanically sealed pumps that operate outside of the hydraulic capabilities of sealless equipment. Movements toward a continuous improvement philosophy, increasing recogni-

The Pump Handbook Series

tion of sealless pump maintenance advantages, continued developments in materials and significant strides toward pricing parity can be expected to further encourage the worldwide utilization of sealless technologies.■

References: 1. Guidelines for Meeting Emission Regulations for Rotating Machinery with Mechanical Seals, STLE Special Publication SP-30 (1994) 2. Bloch, H.P. Practical Machinery Management for Process Plants, Improving Machinery Reliability, Gulf Publishing Company (1982) 3. Fischer, K.B., Seifert, W., and Vollmuller, H. Canned Motor and Magnetically Coupled Applications, Operations and Maintenance in a Chemical Plant, Proceedings from the Tenth International Pump Users Symposium, Houston, Texas (1993) 4. American National Standard for Sealless Centrifugal Pumps ANSI/HI 5.1-5.6, Hydraulic Institute (1994) David M. Carr is a Project Manager with Sundstrand Fluid Handling in Arvada, Colorado and a frequent contributor to Pumps and Systems. He has been with the company since 1980 and has held various positions within the sales, marketing, product engineering and product development departments. He is a graduate of Purdue University with a B.S. degree in mechanical engineering technology and an M.S. degree in management.

CENTRIFUGAL PUMPS HANDBOOK

Vertical Motor-Under Pumps Expand Their Range By Pumps and Systems Staff with Gaylan Dow of Hayward Tyler, Inc., and Joe Campanelli of Air Products & Chemicals, Inc.

f you asked two individuals, one an experienced utility machinery engineer and the other an experienced process machinery engineer, to describe an integral motor sealless pump (think canned motor pump for now) you probably will get two very different descriptions. Like the proverbial elephant and the blind men trying to understand it through touch, sealless pump technology is perceived differently at different ends of the marketplace. This article highlights the features of a type of sealless pump utilized extensively in the utility industry. It also explains how many of these features are being used with benefits in process environments. Most specifying engineers and pump users are familiar with the sealless pump/motor combination design of a canned motor/centrifugal pump of end suction configuration with a single primary radial impeller that does the work. It is mounted horizontally on a bracket or baseplate, and, short of being compact in design and without a coupling, it would look similar to most process pumps. If you pressed harder, you might learn that it uses a mechanism known as “hydraulic balance” to locate the impeller pretty much in the center of the volute. Another class of sealless integral motor pumps has been used for decades in the utility industry. In addition to boiler feed pumps found in all conventional power plant boilers, many boilers are designed to use pumps that force the water to circulate instead of relying on natural convection circulation. With thousands

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of units installed worldwide, the vertical motor-under pump is one of the unsung success stories in the pump industry.

Proven Design for Utility Service Boiler water circulation is arduous duty, with pressures around 3000 psig, temperatures of about 650 ºF and horsepower to 1600. The pump that has evolved to suit this duty is fabricated into and suspended by the plant pipe work (Figure 1). In this setup the motor is slung beneath the pump and is, in turn, supported by a hot neck, which thermally isolates the pump and motor. The whole assembly, not tied to any baseplate or foundation, is free to move with thermal expansion of the plant. In many instances the pumps will move more than one foot from cold start-up to hot running condition. This movement represents relief of piping strain which would normally have been imposed on the pump through the nozzles. Water at boiler pressure is circulated through the motor and heat exchanger by a secondary (auxiliary) impeller that is designed into the thrust bearing. Because of the large horsepower involved, these units are designed to maximize the efficiency of the motor. However, though they are highly efficient for fluid-filled motors, even a small fraction of a large horsepower yields a significant amount of waste heat that needs to be disposed of. This precludes the use of a canned type approach with a liner that separates the winding from the The Pump Handbook Series

Figure 1. The glandless recirculation pump is fabricated into and suspended by the plant pipe work.

fluid in the bore of the motor. The heat generated would create unacceptably high temperatures in the winding cavity. Consequently, a design approach similar to that used in many types of bore hole submersible pumps is used. The windings are coated with a high dielectric strength polymer and immersed in the water itself. This is termed a WSU (wet stator unit) (Photo 1) as opposed to a DSU (dry stator unit/canned motor), and it ensures that a large surface area is cooled by the internal water flow. Additionally, with the absence of an internal liner (or can), the rotor to stator gap is increased, and a failure point (damage to the can) is removed.

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fore extremely quiet. There have been reported incidents of boiler circulator pumps being left on while the boilers have been completely drained. Although this is not recommended, having the motor (and therefore bearings) surrounded by fluid minimizes the potential for damage during dry pump running. • The design utilizes a hydrodynamic thrust bearing to locate the rotor axially. Because of the rotor’s massive weight on boiler circulators, generating and closely controlling the large thrusts necessary to accomplish axial hydraulic balance is impractical. Therefore, the pump is designed to thrust positively in one direction, and it incorporates thrust bearings capable of accepting thrust loads in all cases. This provides the added benefit that the pump has a wide latitude of operation over the curve because there is no hydraulic balance to upset.

Photo 1. Wet stator unit. A secondary impeller incorporated in the motor rotating element circulates water through the motor cavity to the top of the heat exchanger and back to the lower part of the motor. This recirculation dissipates the heat generated by motor electrical losses.

• Radial bearing loads are reduced by the weight of the rotor. Additionally, a multi-vaned diffuser in a concentric casing is sometimes used to further minimize radial loads. • Suction conditions are optimized. Rather than mounting an elbow on the pump suction as is typical of an end suction pump, a longer

A special power feed-through is utilized. It forms both the terminal for connection to the power supply (460V to 6.6 KV) and part of the pressure vessel of the motor.

run of straight pipe can come down directly into the pump suction. • Design and construction costs are minimized for hot systems. The piping goes in, the piping goes out. Much like a motor operated valve, there is no need for long runs of pipe, numerous elbows or expansion joints to relieve the piping strain on the pumps. This greatly simplifies the analysis and design of the piping system. There is no foundation to design or pour to be done and no baseplate to design and fabricate. Also, there is no grouting or hot coupling alignments. • Heat tends to travel sideways or up. Locating the motor below the hot pump end minimizes the heat exchanger load and lifetime operating costs. • In addition to boiler circulation duty, this configuration is also used in the utility industries in attemperator spray pumps as well as in nuclear applications such as reactor internal pumps and reactor water clean up pumps. To minimize the piping involved with reactor internal pumps, the pump ends are actually located directly in the reactor vessel.

Process Versatility Because the pump is continuously self venting, it eliminates an opera-

Practical Advantages Advantages of the motor-under design include the following:

• The pump remains full until intentionally drained. These are large pumps that are full of fluid and there-

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(photo courtesy of Hayward Tyler, Inc.)

• The unit is continuously self venting. Most conventional horizontal canned motor pumps with hot necks and heat exchangers have the latter located horizontally above the motor. On hot standby, this arrangement promotes thermal siphoning and keeps the motor insulation system within temperature limits. The setup will often require a venting operation from the high point (the heat exchanger) when initially filled. This operation is not required with the motor mounted beneath the pump.

Photo 2. A 6000 psig/300ºF canned motor-under pump in service at a chemical process plant. Large quantities of gas are injected and consumed as part of the process. This gas becomes entrained and the primary concern is loss of liquid in the pump end. The process has upset conditions with rapid pressure spikes to 6000 psig . The Pump Handbook Series

(photo courtesy of Hayward Tyler, Inc.)

Photo 3. This canned motor-under pump at a chemical processing facility – an end suction configuration – has to contend with entrained gas and thermal shock.

tor function such as venting on startup – always a good idea. Additionally, some process systems (such as heat transfer) generate non-condensable vapors. In other applications, gas that is injected purposefully as part of the process is consumed, and more and more systems have gas injected at seals and as blankets. This gas tends to collect through centrifugal force about the pump shaft and can make its way back into the motor cavity. On a canned pump with a top mounted heat exchanger, this usually requires the use of a level detection device at the heat exchanger, as well as operator intervention to vent the pump during operation. Even on canned or other sealless pump applications not requiring heat exchangers, the gas can centrifuge out around the rotor and cause possible dry running of the bearings. In addition, because the pump remains full of liquid until intentionally drained, it is protected against dry running, and the motor is impervious to large quantities of entrained gas in the pump end. Typical applications which this benefits include tank car unloaders and batch processes in which tanks are completely evacuated on purpose. The sealless integral motor pump utilizes a hydrodynamic thrust

bearing to locate the rotor axially. A canned motor pump relies on a set of clearances about the impeller to balance the rotor assembly hydraulically and locate it axially. Theoretically, this could be accomplished in a vertical motor-under design by designing in just enough thrust to lift the rotor and locate it axially. In practice, however, using a real thrust bearing has its advantages. Positioning it in the bottom of the unit protects the pump and motor assembly from system upsets. Dry running, or the collection of entrained gas behind the impeller, negates the lift generated by the impeller. Thermal shock dramatically changes the internal pump wear ring clearances, which temporarily and drastically alters the dynamics of a thrust balance design. Also, history has shown that despite efforts to control water chemistry, many boilers are loaded with oxides, which have a tendency to wear clearances and block balance holes, upsetting the thrust characteristics of the impeller. In various process applications rust, scale, catalysts or product solids can be present, and these can create similar circumstances. Any of these conditions would cause a hydraulically balanced unit to lose lift or thrust upward enough to generate considerable wear and damage to the pump and eventual failure. Radial bearing loads are reduced with the motor-under pump. This can be a real advantage for processes in which fluid viscosity is low. And for service in hot systems, these pumps are cost effective to install. The same design constraints of trying to bolt a massive hot piping system to grade through the pump are present in process plants as well. The piping wants to move, and it is better to let the pump move with it. Vertical motor-under canned motor pumps can also be attached directly to a chemical reactor. This setup further reduces piping (and thus lowers costs) as well as minimizes the heat tracing requirements.

Conclusion A vertical motor-under canned pump can be useful in applications requiring zero leakage, including situations in which large quantities of The Pump Handbook Series

gas are present or the pump runs dry. It can also provide reliable service for hot systems in which significant thermal strain can be imparted to the pump flanges, or on systems which run both hot and cold and alignment is an issue. If other service requirements can be met, the motor under design can also lower plant construction costs and be an advantage when floor space is at a premium – such as on oil production platforms or in multilevel chemical processes. Last, it is also useful in situations in which the minimization of piping is attractive, such as those requiring heat tracing, or for lethal services.■ Gaylan Dow is Sales & Marketing Manager for Hayward Tyler, Inc.

Upgrading Boiler Water Circulation Process By Joe Campanelli, Air Products and Chemicals Inc. rocess waste heat boiler circulating pumps have traditionally been horizontal pumps of either end-suction, overhung impeller design (for very low flowrates) or double-suction, impeller-between-bearing, single-stage units for higher flow applications. Boiler water circulation is a very demanding service, with the pump taking its suction directly off the boiler steam drum. Due to the elevated suction pressure and temperature duty, mechanical seal problems are frequent, as are other maladies commonly associated with high pressure/temperature pumps, in-

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cluding: • Pump casing deflection. This is caused by thermally induced pipe strain resulting in rotor rubs, seal leakage/failures, bearing failures, casing cracks, casing splitline leakage and piping cracks. • Thermal stratification in hot standby units. This results in rotor bows, seal leakage, internal rubs on startup and rotor lockup. • Seal leakage and seal maintenance problems. These are constant issues on boiler circulating pumps – often requiring replacements and additional spares. And catastrophic seal failures, while rare, do occur and can expose operating and maintenance personnel to significant releases of hot, high pressure water, which immediately flashes into a plume of steam. The recent introduction of vertical canned motor pumps into this application has enabled our process pump users to capitalize on the extensive experience gained by Hayward-Tyler (HTI) in the utility industry. The HTI vertical canned motor pump offers a proven, robust mechanical design with optimum materials of construction. The application of these

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pumps to process waste heat boiler circulation service has solved the problems associated with conventional horizontal pumps and has provided significant benefits in pump reliability, maintenance and spare parts cost reductions. Two recent waste heat boiler applications that have utilized the HTI pumps with good results are detailed in the chart. The pumps in both applications have performed extremely well to date and will be chosen for

this type of application on future projects. ■ Joe Campanelli is a Lead Mechanical Engineer in the Production and Delivery Organization of Air Products and Chemicals Inc. He has 19 years of experience on a wide variety of rotating equipment in the process industry and is currently responsible for the safe, reliable and efficient operation of the machinery in 28 process plants.

Facility

LaPorte, TX

Pasadena, TX

Pump manufacturer Pump type Pump model Capacity - gpm TDH - ft. Inlet Pressure - psig Discharge Pressure - psig Inlet Temperature - °F Speed - rpm Power - hp Pumps installed On stream date

HTI Dry stator 3 x 4-8/DSU 656 170 745 802 510 3490 72 (4) 2 trains, 2x 100% Jan ‘96

HTI Wet stator 12 x 14-14/200 WSU 8052 76 620 646 489 1780 200 (2) 2 x 100% Oct ‘96

The Pump Handbook Series

CENTRIFUGAL PUMPS HANDBOOK

Centrifugal Pump Suction Pump problems often begin on the suction side. Here is a guide for checking existing operations as well as information on proper designs for new suction systems in the planning stage. By John H. Horwath, Ampco Pumps Co. ump not working properly? Chances are something is wrong in the suction sector. To begin with, there are a multitude of circumstances that can cause suction problems. To cover this subject effectively and systematically, you are being provided a format for checking existing operations as well as information on proper designs for new suction systems in the planning stage. The system’s suction arrangement must provide the energy to move the liquid to the eye of the rotating pump impeller. Until the liquid reaches the leading edge of the impeller vane, the pump cannot impart its energy to move the liquid onward. One must establish a suction requirement for a pump that meets or exceeds the system’s minimum suction availability. It is also the user’s responsibility to see that the suction system provides and maintains the stated condition at the required flow and that the pump is maintained in good operating condition.

P

Existing Operation When a previously successful pump-suction system fails, there are three primary areas that need to be looked at for probable cause. They are the pump, the suction and the liquid being pumped. The most common causes of failure in each sector are: Pumps • Leaky seal • Reduced speed • Pump modification • Wrong direction of rotation • Plugged impeller • Worn wearing rings System Line • Leaky suction line

• Plugged line (including pump inlet) • Sticky valve • Suction height increased • System line modification • Corrosion or product buildup in suction line Liquid • Increase in temperature • Drop in liquid level • Aeration due to vortex or processing change • Viscosity change • Specific gravity change • Other physical property changes Once the problem is identified, it can be corrected. In the case of a processing change, it provides awareness of potential problem areas to check new requirements against the capabilities of the existing pump to see if the unit remains adequate. The eye-balling method of measuring cannot be relied upon to define operating conditions effectively. Use proper instrumentation to measure pressure, flow, suction lifts, speeds and the temperature rise of motors. When motor speeds are incorrect, check connections and measure voltage at the motor terminals. Remember, too, that flow and pressure readings should be taken in areas where stable accurate readings can be obtained. This usually requires 5 to 10 pipe diameters of straight piping after going through an unsettling section such as one that includes elbows, valves and reducers. In addition to measurement, be aware of sounds different than those emanating during normal operations, as well as vibrations and surges.

Planning Stage It is during this phase that most The Pump Handbook Series

potential problem areas can best be dealt with. Begin with careful planning and installation of the pump suction line. Design the line to provide the shortest practical direct route utilizing large radius elbows with an absolute minimum number of fittings and valves.

Suction Piping Air pockets or high spots in a pump suction line invariably cause trouble. Piping must be laid out so it provides a continual rise or at least a perfect horizontal run without high spots from source of supply to the pump. For the same reason, an eccentric reducer instead of a straight taper should be used in a horizontal suction line. Another way to remove vapor trapped in a suction line because of a high point is to vent the sector back to the vapor space in the supply vessel. If an air pocket is left in the suction pipe when the pump is primed, it will often pump properly for a time and then lose its prime or have its capacity greatly reduced. The small pocket of air under a partial vacuum condition will expand and greatly reduce the effective flow cross-section of the pipe, thus starving the pump. Or air will be drawn into the pump, resulting in loss of prime. The suction pipe should be submerged to a depth of at least 3 feet when the water is at its lowest level. Frequently, foot valves are installed on the end of suction pipe for convenience in priming. They should be sized to provide a flow area 50% greater than the pipe area. To protect the foot valve, pump and other equipment from being fouled by refuse such as sticks, rags, light plastics and miscellaneous somewhat buoyant solids, a screen strainer with an area at least three times that of the pipe area should be installed ahead

231

of the foot valve. After planning the required piping system, you should determine the NPSH available of the system. (This is a desirable procedure to follow where your height plus dynamic losses exceed 15 feet or handling a hot liquid or operation in a closed system.) Keep in mind that a pump’s location can be of utmost importance in establishing its suction requirements. Often you will find it more feasible to relocate the pump than to call for a special low NPSH pump. The cost of these changes must be weighed against the cost of a usually larger or special pump required for a lower available NPSH condition.

Net Positive Suction Head NPSH (Net Positive Suction Head) can be defined as the absolute pressure at a datum line (normally the centerline of the impeller eye at the suction nozzle) minus the vapor pressure of the liquid (at pumping temperature) being pumped.

tank or atmospheric pressure in open tank ps = gage pressure reading in the pump section centerline (steady flow conditions should exist at the gage tap; five to ten diameters of straight pipe of unvarying cross-section are necessary immediately ahead of tap). The corrected gage reading is a minus term in the equation if it is below atmospheric pressure. v2s 2g= velocity head at point of measuring ps (based on actual internal diameter of pipe) at point of pressure taps hvp= absolute vapor pressure of the pumped liquid at the pumping temperature

Hp

Proposed System A. Proposed System (units in feet of water) NSPH Available= hp ± (*) hz – hf – hvp where: hp = absolute pressure acting on suction liquid surface hz = liquid suction height above or below the pump impeller centerline hf = total head losses in the suction including exit, entrance, fitting and pipe losses at the intended flow rate hvp = vapor pressure of the pumped liquid at the pumping temperature (*) for liquid level above centerline + applies;for liquid level below centerline – applies

+Hz

Hf

-Hz

Figure 1. Proposed system design Pg

Ps

Existing System B. Existing System (units in feet of water) NSPH Available = pg ± ps + v2s – hvp where: 2g pg = gas pressure in closed

232

Figure 2. Existing system design The Pump Handbook Series

Pumped Liquid Characteristics The liquid being pumped can have a profound effect on the pump’s suction capability. Keep in mind that a centrifugal pump is not fully capable of handling gases or vapors and that it is always necessary for the absolute pressure in any sector of the pump to be higher than the vapor pressure of the liquid being pumped. Entrained air also has an adverse effect on pump performance. As little as 1% entrained air by volume can reduce pump head and capacity substantially. Under no circumstances should a standard centrifugal pump unit be expected to handle more than 3% air by volume as measured under pump suction conditions. Substantial effort should be made to keep air out of the liquid entering the pump. This commonly occurs if a vortex develops at the suction pipe inlet and adequate submergence or effective baffling can help prevent this condition. A leak in the suction or pump stuffing box operated under a vacuum may also introduce air into the pump’s suction. A well-designed entry coupled with good maintenance practices will alleviate most entrained air situations. Dissolved gas or gas evolving from a chemical reaction can also be troublesome. Common practice calls for the use of a large diameter impeller (not necessarily a larger inlet) to meet the hydraulic performances based on handling cold clear water alone. Where the percent of air (or gas) exceeds the recommended limit for standard pumps (which may vary for different designs), it may be appropriate to consider less efficient pumps designed specifically for handling two-phase flow. The boiling of a liquid that can occur at reduced pressures is dependent on the liquid properties – pressure, temperature, latent heat of vaporization and specific heat. The stated suction characteristics of pumps based on cold water is standard throughout the pump industry. Determination of other liquids must usually be made through testing, although the Hydraulic Institute’s Standard 14th Edition (Figure 70 in their publication) does provide

correction for some liquids at temperatures up to 400ºF. The same operating conditions as those with cold water are usually maintained in the absence of data. Other contributing factors are liquid surface tension, specific gravity and viscosity. Some adjustment can be made for certain liquids, however, to avoid any chance of cavitation. The same operating conditions as with water are usually maintained. Always remember that the pressure at any point within the pump must remain higher than the vapor pressure of the liquid if cavitation is to be avoided.

Cavitation When the absolute pressure becomes equal to or less than the vapor pressure of the liquid being pumped, bubbles consisting of dissolved gases begin to form. The bubbles are carried by the liquid flow into an area beyond the leading edges of the impeller vanes where higher pressure being developed causes the bubbles to condense and collapse, creating severe mechanical shocks. The term used to indicate this process is “implosion.” The bursting bubbles begin to damage the pump’s interior surfaces in the immediate vicinity, and this damage can be quite destructive over a period of time, depending on the pressures developed, their collapsing rate and the base material being subjected to attack. Four common symptoms of cavitation are: Noise – caused by the collapse of vapor bubbles as they enter the high pressure area. This is typically identified as a light hissing and cracking sound at the onset of cavitation and a rotating noise when fully developed. Vibration – caused by the impacting of the bursting bubbles on the impeller surface, this condition can also result in a premature bearing and shaft seal failure. Drop in Efficiency – indicates the onset of a cavitating condition. The degree of efficiency drop-off increases precipitously as cavitation increases. Erratic Flow – commonly occurs. The severity of the fluctuating flow is determined by the degree of cavita-

tion and pump design.

The Pump The suction system must provide the energy to move the liquid into the eye of the impeller. The determination of a centrifugal pump’s NSPH requirement is established empirically through a series of performance tests run under specific conditions. Of prime importance in the suction sector is the pump’s inlet vane angle. This is usually in the range between 10 and 27 degrees independent of the pump’s design. An often used flow angle is 17 degrees – a compromise between efficiency and a low NSPH requirement. For best efficiency the vane inlet angle would be nearer 27 degrees while for a low NSPH requirement one would go closer to 10 degrees. Four to six vanes are commonly used in most designs of the smaller units available from several pump vendors. Manufacturers have developed various pump design innovations to improve suction capability. Extremely low NSPH availability may require the inlet blade angles to be reduced even further. Some commercial units provide a separate axial impeller (inducer) in the suction entry just ahead of the main impeller to further induce flow into the standard impeller’s eye. When a low angle inlet is introduced it may be necessary to reduce the number of vanes to decrease the blockage effect. Some manufacturers reduce the length of every second inlet blade tip into the impeller passage, speed (rpm) and profile centroid radius. Pump manufacturers today incorporate design features that provide efficient conversion of velocity energy to pressure energy, smooth cast surfaces for lower friction losses and sharp vane tips to reduce entry shock losses in the impeller eye. Smooth surfaces can also defer the onset of cavitation because they delay the formation of microscopic bubbles that form prematurely on rough surfaces prior to the inception of a cavitating condition. Accurate determination of the start of cavitation requires very careThe Pump Handbook Series

ful control of all factors that influence pump operation. A number of test points bracketing the point of change must be taken and the data plotted. Because of the difficulty in determining just when change starts, a drop in head of 3% is usually accepted as evidence that cavitation is present.

Summary An important aid in diagnosing a pump or application problem is a maintenance file card or folder kept on the pump that lists its history of operational problems as well as parts replacement over the years. In cases where severe attacks occurred in the impeller suction area caused by erosion or cavitation, photographs of the attacked area should be kept as well as commentary describing operation at that point – be it noise, vibration, erratic flow and/or surges. Become familiar with the terminology and procedures provided in this article since they can aid in identifying and resolving many of the problems encountered on the suction side of the pump. Understanding basic pump suction concepts will enable you to more clearly explain your situation to your pump representative when a persistent problem can’t seem to be corrected. When purchasing a new or replacement pump for a sensitive suction service, go beyond the price, pictorial and geographical presentations and ask the following questions: 1. Were the pump’s calibration tests conducted in conformance with the Hydraulic Institute’s standards? 2. Can the vendor provide performance data specific to your anticipated suction requirements? 3. Can the vendor furnish a typical pump impeller and casing set for you to examine?■ John H. Horwath is a Senior Technical Consultant with Ampco Pumps Company in Milwaukee, WI.

233

CENTRIFUGAL PUMPS HANDBOOK

Pumping Options for Low Flow/High Head Applications Familiarity with the major designs will help you match the right one to your application.

here is a trend toward low flow/high head pumps in the chemical processing industry. This article will focus on several types of such pumps, describing the advantages of each and telling when they should be used. We will address the more popular low flow/high head pumps, giving basic terms and discussing one specific design in greater detail.

T

Low Flow/High Head Many industrial processes require approximately 15 gpm with a pressure of 200 to 1000+ ft TDH. Many companies use standard end suction pumps in an attempt to meet these needs. Most end suction pumps, however, are not designed for this type of application – a fact that leads to premature failure. As an ANSI pump is operated to the left of its Best Efficiency Point (BEP), several things happen. First, there is suction recirculation. Second, with even less flow, impellers, bearings and mechanical seals wear out faster. Last, in shutoff conditions, there is no flow and significant rise in temperature. All of these result in shorter pump life, less reliability, shorter seal life, and shorter MTBPM. However, several options are available to apply the right pump to the right application.

Option 1 – Special ANSI Design Impellers Due to the large installed population of ANSI pumps in the world, many manufactures have designed

234

special type of impellers for low flow applications. There are basically two configurations. The first is the “Barske” design. Basically, this is a semi-open impeller with balance holes and it requires a special concentric casing. The advantage of this design is stable low flow/high head applications, and only the impeller and casing have to be replaced on an existing ANSI pump for a conversion. The second low flow impeller design is an enclosed type that provides stable operation at low flows but has added benefits: the impeller is the only part that needs to be replaced, and there are no wear rings. Thus, an existing ANSI pump can be changed out with minimal cost and time through a simple replacement of the impeller. Figure 1 illustrates this type of design. Side Note: One company offers a heavy duty vertical enclosed-design chemical sump pump for low flow/high head applications that offers an additional benefit. It has, as an option, balance holes in the back cover that allow the pump to run against a closed valve (dead headed) for extended periods of time with no adverse effect on pump life. This is ideal for applications in which tanks are being used for a batch process that only requires intermediate amounts of product. With the pump’s ability to run continuously, the customer can let the unit run and operate a valve to meet requirements. Also, the pump has ANSI impellers and casings as a standard, which offers increased interchangeThe Pump Handbook Series

(courtesy of Sterling Fluid Systems)

By Patrick Rienks, Sterling Fluid Systems (USA), Inc.

Figure 1. Enclosed design for ANSI pumps

ability of parts and less inventory in the field. Both basic impeller designs have been available for some time from various pump companies. In addition, these designs allow for field upgrades while keeping cost and down time minimal. Furthermore, if these are purchased as new units, all other parts would be interchangeable with other installed ANSI pumps. These designs will handle a small amount of entrained air and/or solids that may be present. The NPSHR is also less than standard ANSI pumps running at the same service. These pumps are available with all the ANSI options –e.g., jacketing and large bore stuffing boxes and a wide range of materials.

Option 2 Pitot Tube Pumps In a unique single stage centrifugal pump known as the Pitot tube

pump, the Pitot tube is stationary, and the inner casing rotates (See Figure 2). The liquid enters the pump through the suction line, passing the mechanical seal, then enters the rotor (Part 2) where it is brought up to rotor speed. The liquid near the largest rotor diameter has a pressure that obeys the basic mechanical laws of centrifugal force. A stationary wing-shaped Pitottube (Part 1), is placed inside the rotor (Part 2) which has a circular opening near the largest rotor diameter. The Pitot-tube has a double function. First, the liquid is forced to enter the tube due to lower pressure in the tube. Second, when the rotating liquid hits the specially shaped stationary tube, the liquid speed is transformed into pressure energy. 17

31

2

1

Option 3 Side Channel Pumps

26 29 36

E

D

I

F G

H

J

C

B

A

Figure 4. Cross section of CEH side channel impeller stage

Photo 1. CEH side channel pump (courtesy of Sterling Fluid Systems)

This operational principle enables the pump to generate pulsation-free flow and a stable NPSHR curve. The Pitot tube design offers higher pressures – up to 6200 feet of TDH – whereas most ANSI pumps are limited to about 900 feet of TDH. The pressure can be significantly higher in Pitot-tube pumps because the mechanical seal (Part 29) is subjected to suction pressure only, and thus the seal pressure is low, where in an ANSI design the seal is subjected to a combination of suction and discharge pressure. Also, the Pitot tube design has lower NPSH requirements than the ANSI style, and it can handle some solids. Pitot-tube pumps offer greater efficiency, 64% versus 35% for ANSI styles. A wide range of alloys is available, and the design provides stable flow in a wide range of applications.

(courtesy of Sterling Fluid Systems)

Figure 2. Combitube – Pitot-tube style design

handling capabilities and self priming. Also, unlike ANSI designs, side channel pumps will not vapor lock. ANSI will handle approximately 7% entrained gas vs. 50% for the side channel design. For example, this pump can handle volatile liquid up to 150 gpm at 1200 feet TDH with an NPSHR of less than 1 foot. No other design can accomplish this. Additionally, it can achieve suction lifts of more than 25 feet, has “floating” impellers that eliminate end thrust on bearings, and can operate at low speed and under 25” Hg vacuum or greater for long periods. The side channel pump is a compact design available in a wide range of materials. Figure 4 illustrates the design principle of lateral channel stages in turbine pumps. This design can be made in one to eight stages. From the diagram it can be seen that liquid or

(courtesy of Sterling Fluid Systems)

The side channel design is used more in other parts of the world than in the USA, but it is starting to draw domestic interest. Photo 1 shows a typical side channel design. Figure 3 shows a cross-sectional view. This proven end suction multistage design brings the liquid into the eye of the impeller. The inlet stage is centrifugal, and the impeller is enclosed, optimizing suction requirements and providing superior NPSHR values. Additional performance is achieved by the open impeller stages behind the inlet stage, each of which generates pressures many times greater than standard centrifugal pumps running at the same speed. The special configuration of the stages gives the pump excellent gas

(courtesy of Sterling Fluid Systems)

4

And there is another advantage. To increase flow and head in most situations, one needs to change only the Pitot-tube and/or the speed. This provides a quick and inexpensive solution versus changing entire sizes of ANSI designs. The Pitot tube design has been in industry for more than 35 years. Thus, for special applications it is a proven problem solver. Typical applications include cleaning, descaling, injection, boiler feed, process handling and hydraulic and spraying systems in chemical, plastic, paper, steel and other industries.

Figure 3. Cross section of CEH side channel impeller stage The Pump Handbook Series

vapor or mixtures enter the stage through the inlet port (A) in the suction intermediate plate (B). Although not shown, it should be noted that the internal face of this plate is flat, not channeled as in conventional turbine pumps. The mixture, once it encounters the rotating impeller (C), makes several regenerative passes through the unique side channel design shown (D), located in the discharge intermediate plate (E). Due to centrifugal action, the liq-

235

uid, being the heavier component in a mixture, is forced toward the periphery of the chamber, whereas the lighter vapor or air collects near the center at the base of the impeller blades. Most of the liquid exits through the discharge port (F), with the remainder then guided along the mini channel (G), which eventually dead ends at point (H). There the liquid is forced to turn toward the impeller hub, thus compressing any vapor or entrained air at the base of the impeller blades. This compressed vapor or air is now forced through the secondary discharge point (I), after which it rejoins the major portion of the liquid that was discharged through discharge port (F). Thus, the problem of continual air or vapor build-up within the chamber has been overcome, and the liquid vapor mixture continues on to the subsequent stage and is eventually discharged from the pump. Where multiple stages are utilized,

236

they are staggered radially to bring about balance and minimize shaft deflection. The impeller in each stage, although keyed for radial drive, is allowed to “float” axially, thus assuming a running position in the equilibrium brought about by balance holes (J), which are appropriately positioned in the impeller hub. The floating action of the impeller also eliminates axial thrust on the pump’s external ball bearings. This design, also available in a mag drive version, can be applied to the same services as Pitot-tube and ANSI style pumps, but it can be additionally applied to self priming applications. Another option is a barrel design enclosing the entire pump for added safety. However, due to the lateral gap between the vane wheel impeller and the guide disk in this design, these pumps should not be used where solids may be present. The barrel pump is also available in many materials, including non-

The Pump Handbook Series

metallic PAEK.

Summary There are many possibilities for low flow/high head applications. We have explained some basic guidelines. However, if you have a specific application, its best to ask the applications group of a pump company to review the data in order to select the right pump for your needs. In today’s changing industry, there are many solutions to help you keep your process running, lower your costs and greatly increase your MTBPM by selecting the right pump for the job.■ Patrick Rienks is a Product Manager for the Industrial Division of Sterling Fluid Systems (LaBour, Peerless, SIHI, and SPP Companies) in Indianapolis, IN. He earned his B.S. in Chemical Engineering from Tri-State University and is currently enrolled at Rose-Hulman Institute of Technology for his Master’s of Engineering.

CENTRIFUGAL PUMPS HANDBOOK

Anti-Friction Bearings in Centrifugal Pumps A detailed overview of bearing design, along with application and maintenance tips. By William E. (Ed) Nelson, Turbomachinery Consultant Introduction earings used in centrifugal pumps are categorized according to the direction of the forces they absorb, either radial or axial thrust. Most small centrifugal pump bearings use anti-friction bearings, ball or roller types, because they can be designed to handle a combination of both radial and axial thrust loads. Anti-friction bearings use balls or rollers instead of a hydrodynamic fluid film to support a shaft load with minimal wear and friction (Figure 1). Cleanliness, accuracy and care are required when installing ball bearings. The ball bearing is a piece of

B

1. Oil Formation

A–Bearing-static condition.

B–Partial operating speed.

C–Operating at full speed. Figure 1. Theory of ball bearings

precision equipP ment manufactured to extremely close tolerances. 0.004" R1 To obtain the 0.002" 0.002" 0.002" maximum service from it, the shaft and housing must also be machined to rigid tolerances. For example, Mech. Bearing Bearing R2 seal locating shoulders radial axial must be at right angles to the shaft centerline so the Impeller 10"a 10"s bearing will be squared with the Figure 2. Loads on radial and thrust bearings shaft, and the between-bearings design pumps, or housing bores must be in almost perthe closest one to the coupling on a fect alignment to ensure that the back pull-out design, is fixed axially. bearing will not be forced to operate The inboard bearing is free to slide in a twisted position. Maintenance of within the housing bore to accommoball bearings is simple. Protect the date thermal expansion and contracbearing from contaminants and tion of the shaft as shown in Figure moisture, and provide proper lubri2. Since the outboard bearing is fixed cation. in the housing, it must carry both Bearing Loads axial and radial thrust. The axial thrust is considered to be acting Assuming that the above condialong the centerline of the shaft and tions are satisfied, the life of a ball therefore is the same at the outboard bearing depends on the load it must bearing as it is at the impeller. The carry and the speed of operation. The radial and axial loads combine to creloads on pump bearings are imposed ate an angular load at the outboard by the radial and axial hydraulic bearing. forces acting on the impeller. Radial thrust acting on the In any two-bearing system, one impeller creates radial loading on of the bearings must be fixed axially both bearings. The magnitude of the while the other is free to slide. This load at each bearing can be deterarrangement allows the shaft to mined by the use of Figure 3 and expand or contract without imposing these equations: axial loads on the bearings. At the Pxa same time, the arrangement locates R1 = s one end of the shaft relative to the stationary parts of the pump. GenerR2 = P(a+s) ally, the outboard bearing of s The Pump Handbook Series

237

Where: R = Radial load on bearings 1 and 2 (pounds) P = Radial thrust on impeller (pounds) a = Distance from the centerline of the impeller to the center line of the inboard radial bearing (inches) s = Distance from the centerline of the inboard radial bearing to the centerline of the out board bearing (inches) Radial loads come from other sources as well. The weight of the rotating assembly (shaft, sleeve and impeller) gives one load. Unbalance and external misalignment of the shaft give still another. The weight of the overhung coupling also creates a bearing load. Pump designs should limit the shaft deflections at the seal face to under 0.002 inch at the worst conditions. For single stage horizontal pumps, this will be with the maximum impeller diameter at “shut off” conditions – i.e., closed discharge. For larger double suction pumps, this load might well occur at the far end of the performance curve. Attention paid to cutwater clearances, Gap “B,” and “overfiling” of impeller vanes can reduce some of the hydraulic loads.

Ball Bearings Types A ball bearing normally consists of two hardened steel rings and several hardened balls utilizing a separator to space the rolling elements and reduce friction as shown in Figure 3. The many ball bearing designs used OUTER RING SEAL OR SHIELD NOTCH

OUTSIDE DIAMETER

BORE

INNER RING FACE

OUTER RING FACE

Figure 4. Conrad bearing design

in industry are classified according to the type of loading they receive: radial, thrust and combined. Sizes and classes of precession of bearings are governed by the Anti-Friction Bearing Manufacturing Association (AFBMA) and by the Annular Bearing Engineers Committee (ABEC). There are five ABEC Classes, 1, 3, 5, 7 and 9. Class 1 is standard, and Class 9 is high precision. Pump bearings generally are Class 3, loose fit. ABEC Class 9 is factory order only and has no longer bearing life or higher speed rating than ABEC 1. Three types of bearings are generally used in centrifugal pumps: Conrad Type – The Conrad type – identified also by its design features as the deep-groove or non-loading groove type – is the most widely used (Figure 4). A general purpose bearing, it is used on electric motors or wherever slight axial movement of the shaft is permissible. The deepgroove rings enable this bearing to carry not only radial loading, for which it is primarily designed, but about 75% of that amount of thrust load in either direction, in combination with the radial load. API 610 Standard “Centrifugal Pumps for PetroWIDTH leum, Heavy Duty Chemical and Gas CORNER RADIUS Industry Services” pumps require that SHOULDERS single row or double CORNER RADIUS row radial bearings be Conrad type with INNER RING Class 3 or loose fit. BALL RACE This permits enough flexibility to let the SEPARATOR shaft correct for any OUTER RING misalignment betBALL RACE ween the housing and the shaft.

Figure 3. Elements of a ball bearing

238

The Pump Handbook Series

Maximum Capacity Type–Another radial bearing is the maximum capacity or filling-slot type. It is provided with a filling notch that extends through the ring or ring shoulders to the ring way, permitting a larger number of balls to be placed between the rings than can be done in the same size Conrad bearing. The supposed advantage of these bearings (larger load carrying capability) has vanished with the availability of better steels and lubricants. Filling slots often are not precision machined and can enter the ball contact area of the rings. This will result in early bearing failure. Several bearing manufacturers consider a filling slot bearing to be unreliable and discourage their use. The 1981 and later editions of API 610 prohibit their use although they are still permitted by ANSI specifications.

Figure 5. Angular contact bearing design

Figure 6. Back-to-back duplex mounting of angular contact bearings

Angular Contact Type – This design allows the carrying of high radial loads in combination with thrust loads from 150 to 300 percent of the imposed radial load (Figure 5). Unlike the Conrad design, the contact angle is not perpendicular to the bearing axis. Several different angles are available, offering a variety of radial and thrust loadings. Filling slots are not used in this design, and

Contact angles for 7000 series duplex bearings Contact angle Bearing manufacturer 20° 25° 30° 35° 40° A B C D

X

X

X

X X

X X

UNSTAMPED FACES OF OUTER RINGS TOGETHER

CLEARANCE BETWEEN OUTER RING FACES THESE INNER AND OUTER RING FACES NOT FLUSH

INNER AND OUTER RING FACES CLAMPED TOGETHER FACES FLUSH ON BOTH SIDES

Figure 7. Contact angles of various bearing manufacturers

Before Mounting STAMPED FACES OF OUTER RINGS TOGETHER

Mounted

Figure 9. Face-to-face (DF) mounting moment arm

Not all ball bearing companies manuINNER RING FACES CLEARANCE facture the 40 degree CLAMPED TOGETHER BETWEEN INNER RING angle angular contact THESE INNER AND FACES OUTER RING FACES bearing. If the local ARE FLUSH bearing supply house is not on its toes, pump repairs could Before Mounting Mounted be made with nonspecification bearings. Figure 8. Back-to-back (DB) mounting moment arm The 40 degree contact angle gives an 18 to 40 percent thrust loads can be imposed from increase in capacity over the 30 one direction only. degree angle depending on bearing Duplex Type – These are identical size. This differential is a very imporangular contact bearings placed side tant piece of information, and unforby side. The contacting surfaces tunately, it is communicated or must be ground to generate a specidesignated in industry by a confufied preload (Figures 6). This special sion of suffixes. The numerical code grinding allows the two bearings to used in bearing identification is mostshare loads equally. Without it, one ly standard among the various bearbearing in the pair would be overing manufacturers. However, the loaded, the other underloaded. alphabetical prefixes and suffixes are API pump specifications require not. When identifying bearings from duplex 40 degree contact angle thrust codes for the purpose of inter bearings mounted back- to-back with changing bearings, care should be a light (100 pounds or 45 kg) preload exercised that the meaning of all as the best choice. While the requirenumbers and letters is determined so ment of a 7000 series, 40 degree conan exact substitution can be made. tact angle, light preload bearing Most manufacturers supply crossshould be a fairly tight specification, reference tables for identifying equiit is not. First, there are three 7000 valent bearings. A very good source series bearing designs – light, mediis the bearing manual of the AFBMA. um and heavy. There is about a 50% change in capacity from one design Mountings to another. Second, some manufacThere are five types of duplex turers use more than one contact angular contact bearing mountings, angle. The contact angle is the source although only two are commonly of considerable confusion as shown used. Rigidity of the shaft and bearin Figure 7. There are no standard ing assembly depends in part on the designations to identify the 40 degree moment arm between ball-contact angle. The Pump Handbook Series

angles of duplex bearings. Within reasonable limits, the longer the moment arm, the greater its resistance to misalignment. API 610 calls for the DB or backto-back mounting of angular contact bearings. They are placed so that the stamped backs of the outer rings are together. In this position the ball-contact angles diverge outwardly, away from the bearing axis. With DB bearings the space between the diverging contact angles is extended (Figure 8). Shaft rigidity and resistance to misalignment are correspondingly increased. DF bearings are intended only for face-to-face mounting. They are placed so that the faces (or low shoulders) of the outer rings are together (Figure 9). Ball-contact angles thus converge inwardly, toward the bearing axis. With DF bearings the space between the converging contact angles is short. Bearing-shaft rigidity is relatively low. However, this arrangement permits a greater degree of shaft misalignment than other mounting methods. Some older multistage pumps use this mounting arrangement.

Special Designs Centrifugal force in the unloaded thrust bearing causes its balls to move out of their intended track and operate on a skewed axis. The balls begin to slide rather than roll during rotation. The increased friction that results reduces the viscosity of the oil film, accelerates wear of

239

Bearing Misalignment Capability

Figure 11. Double row bearings

B. Pump Type: Overhung single suction pump and high speeds, in excess of 1750 cpm, with large thrust loads coupled with radial load Bearing Style: 40 and 15 degree, DB mounting

Figure 10. Special design duplex angular contact bearings

the raceways and leads to early failure. Some recent work in bearing selection indicates that adherence to the API specification of 40 degree bearings mounted DB may not be optimal. For single stage pumps in which the thrust action is steady and in one direction at all flows, the use of a 40 degree angular contact bearing to absorb the primary thrust and a 15 degree angular bearing for any reverse thrust has extended the life of pump bearings. The 15 degree bearing decreases the tendency for ball sliding and increased friction. The bearings also have machined bronze retainers to reduce internal friction further. The bearings are packaged in pairs and are marked so that when they are mounted they will accommodate the primary thrust load (Figure 10). This arrangement is better in some but not all applications. Recommendations for bearing use are summarized as follows: A. Pump Type: Overhung single suction pump and low speeds, 1750 cpm or below, with any load conditions Bearing Style: 40 degree duplex, DB mounting, or 40 and 15 degree

C. Pump Type: Double suction between bearings and high speeds, in excess of 1750 cpm, mostly radial loads with low thrust loads Bearing Style: Duplex 40 degree, DB mounting

The ability to tolerate misalignment between the bearing housing and the shaft is dictated by ball and ring geometry. Table 1 is a chart showing the relative capabilities of three bearing types. Knowledge of these relative capacities can help a maintenance engineer or supervisor make substitutions to get out of many bearing problems. Note the Conrad bearing’s rated radial capacity and speed limit are taken as unity for comparison purposes. The angular contact bearing can carry almost double the Conrad’s radial rating in thrust. The Conrad can only carry 75% of its radial rating in thrust, and the self-aligning ball bearing can carry only 20%. This means that thrust load on a self-aligning ball

Double-Row Type Essentially, the double-row bearing is an integral duplex pair of angular contact bearings with built-in preload (Figure 11). It resists radial loads, thrust loads or combined loads from any direction. Two basic types are available, corresponding to the face-to-face and the back-to-back mounting of conventional duplex bearings. Avoid using two doublerow bearings on the same shaft because this makes the mounting too rigid. The bearings on each end of the shaft will tend to impose loads on each other. While double-row bearings can be constructed with angular contact, such designs require a filling slot for assembly of at least one row. API 610 prohibits the use of the design because of its greater vulnerability to failure in reverse thrust applications.

Figure 12. Load zones and retainers of a ball bearing

bearing is prohibited. Note the angular misalignment capability of the various bearings. The Conrad can withstand 15 minutes; the self-aligning ball can take 16 times as much or 4 degrees. Values in this chart are for comparison purposes only. Actual catalog values for load ratings and limiting speed should be used.

Average relative ratings Type

Radial Thrust

Conrad type Angular contact 40° Self-aligning

1.00 1.00 0.70

0.75 1.90 0.20

Limiting Misalignment speed 1.00 1.00 1.00

+ - 0 deg 15’ + - 0 deg 2’ + - 4 deg

Table 1. Capacity for various bearing designs

240

The Pump Handbook Series

Other Bearing Problems There are a number of problems associated with anti-friction bearings utilization that impact pump reliability.

Retainers A retainer ring or cage is used to make all the balls of a bearing go through the load zone (Figure 12). The most common retainer material is low carbon steel (1010 analysis) attached by fingers, rivets or spot welding. Riveted or spot welded steel strip retainers are more subject to fatigue failures. When a bearing ring or cage is misaligned, the balls are driven up against the ring shoulder, the top ball to the left and the bottom ball to the right. The center balls on each side, at this particular point, tend to stay in the center of the ring because balls in this position relative to the misalignment are not thrust loaded. The net effect of this action is to flex the ring in plane bending. As the inner ring turns, a cyclic retainer bending stress occurs. The load on the retainer pocket is also cyclic. At the high thrust positions, the retainer exerts the maximum force in maintaining the ball space. Since the retainer and ball are in rubbing contact, the thermal load is at its highest on one side (no thrust load point). In this manner, the retainer is subjected to both a flexing and thermal cyclic load that can lead to fatigue cracking at retainer stress

Some manufactures use pressed brass, machined bronze, machined phenolic and molded plastics in an effort to reduce the heat generation. The relative desirability of bearing separator types is as follows: 1. phenolic 2. machined bronze 3. pressed brass strips 4. pressed steel strips 5. riveted steel strips

Bearing Carriers In order to use a larger radial bearing and still be able to remove the impeller or mechanical seal from the shaft, some manufacturers utilize a bearing carrier similar to the one shown in Figure 13 on between-bearing pumps. The carrier is a shouldered sleeve with a small clearance between it and the shaft. The radial bearing is then shrunk onto the outside diameter of the carrier. The problem with this design is that if the bearing begins to heat up, due to lack of lubrication or some other reason, the carrier also heats up – expanding until it comes loose. At this point, even though the bearing has not failed yet, the carrier may be free to spin on the shaft, a condition that will eventually cause the shaft to

Suction Stuffing Box

bend or fail. Current API 610 specifications require that the bearings be mounted directly on the shaft. There are a lot of carriers still in service.

Snap Rings Snap rings are flat, split washerlike devices used by some manufacturers to position components axially – e.g.,ball bearings and seal sleeves on shafts. As with bearing carriers, there are two major problems in using snap rings. First, removal requires the use of a tool that is not normally found in the pump machinist’s tool box. When used in a shaft, they must be positioned in a groove. The addition of a radial groove in the shaft effectively reduces the diameter and may weaken the shaft. When used in a bearing mounting, the rings permit considerable end float of the bearing. Current API 610 specification prohibits the use of snap ring mounted bearings.

Bearing Arrangements Different arrangements of antifriction bearings can handle various loadings imposed on the pump. Pump design is crucial in determining possible bearing arrangements.

Impeller Suction Stuffing box

Figure 13. Typical bearing carrier design

points. Notches and rivet holes may form. Due to the rubbing contact between the retainer and balls, the lubrication requirements here are more critical than for the rolling contact between the balls and rings. Shock loading of the bearings also causes retainer failure at the pockets.

Radial Bearing

Thrust bearing Discharge

Figure 14. Typical pump bearing arrangements – between bearing pump The Pump Handbook Series

241

Horizontal Pumps Overhung impeller pumps usually employ ball bearings only. In a typical bearing housing arrangement (Figure 2), the radial ball bearing is located adjacent to the impeller or inboard position. It is arranged to take only radial loads. The thrust bearing is located closest to the coupling and usually consists of a duplex pair of angular contact bearings. The bearings are mounted back-to-back so that axial thrust load can be carried in either direction. This duplex bearing pair carries both the unbalanced axial thrust loading as well as radial load. In between-bearing pumps, the ball radial bearing and the ball thrust bearing combination have individual bearing housings (Figure 14). The radial bearing is normally located at the coupling end of the pump. The ball thrust bearing is located at the outboard pump end. The thrust bearing must be secured axially on the shaft to transmit the axial thrust load to the bearing housing through the bearing. The bearing is usually located against a shoulder on the shaft and locked in place by a bearing nut. This means that the shaft diameter under the thrust bearing is less than the shaft diameter under the radial bearing. Thus, by mounting the radial bearing on the inboard (or coupling) end of the pump shaft, a larger shaft diameter is available to transmit pump torque from the coupling to the impeller. The thrust bearing, on the other hand, is locked axially in the thrust bearing housing, the radial bearing is axially loose in its housing to allow for axial thermal growth. A popular combination for between-bearing double suction pumps consists of journal type radial bearings and a ball thrust bearing. In such an arrangement, all radial pump loads are handled by the journal radial bearing. The ball thrust bearing is mounted in the thrust bearing housing such that only axial loads are carried by the thrust bearing. The housing around the ball thrust bearing is radially loose. A metallic strap is employed on the outer rings of the thrust bearing. This strap locks into the bearing housing

242

to prevent rotation of the outer rings. Such an bearing arrangement is useful in higher horsepower and higher speed applications where ball radial bearings would be impractical due to speed, load and lubrication limitations. Because the ball thrust bearing is located on the outboard end of the shaft, the shaft diameter under the ball thrust bearing can be relatively small since no torque is transmitted from this end of the shaft.

THRUST BEARING

RADIAL BEARING

Vertical Pumps Most vertical pumps differ from horizontal pumps in that the entire axial thrust, consisting of axial hydraulic forces as well as the static Figure 15. Vertical pump motor – solid shaft weight of the pump and the driver rotor, is supported by the driported at the top of the motor shaft. ver thrust bearing. Therefore, the sizClearance is provided between the ing of that bearing becomes a joint outside diameter of the pump shaft effort between the pump manufacand the bore of the motor shaft. On turer, the driver manufacturer and solid shaft units, a solid coupling is the end user. furnished by the pump manufacMany end users require that this turer to provide rigid attachment bearing be rated to handle at least between the pump shaft and the twice the maximum thrust load, up motor shaft extension. Thus, the or down, developed by the pump in a pump shaft is retained radially by worn condition with two times the the lower motor bearing. This is internal clearances it had when new. generally considered to be a better This requirement came into effect arrangement for most vertical after users experienced problems in pumps since the shaft runout will the field that result from the followbe less and thus the seal or packing ing facts or principles. life will be longer. Further, larger 1. The calculation of pump thrust is not highly accurate. 2. Pump thrust increases as internal clearances increase. 3. The thrust load varies with the vertical position of the impellers with the casing(s). 4. The thrust load varies with flow. (In some cases it may even reverse direction.) A reasonable margin should be provided between the driver thrust bearing rating and the maximum calculated pump thrust.

Motors for vertical pumps are available as solid shaft or hollow shaft units. On hollow shaft units, the pump shaft extends upward through the motor shaft and is supThe Pump Handbook Series

diameter pump shafts can be coupled to solid shaft motors than can pass through the bore of hollow shaft motors, and this also provides increased shaft rigidity. Because of potential field problems, the thrust bearing should be mounted in the driver top bearing housing, farthest from the solid coupling and the pump (Figure 15) . In the event that a thrust bearing fails, any subsequent drop in the driver/pump shafts could result in a mechanical seal failure that could release hydrocarbons to the atmosphere. Also, these could be ignited by a hot bearing.

Ball Bearing Fits Unfortunately, many pump manufacturers do not indicate the proper bearing fits for shaft and housings to guide shop repairs. The original dimensions of both the housing and the shaft will change with time due to oxidation, fretting, damage from locked bearings and other causes. Every bearing handbook has tables to aid you in selecting fits. The vibration effect of looseness on the bearing fits is different for the housing and the shaft. Housing Fits – Ball bearing fits in the bearing housing are, of necessity, slightly loose for assembly. If this looseness becomes excessive, vibration at rotational speed and multiple frequencies will result. Do not install bearings with O.D.’s outside of the given tolerance band since this might result in either excessive or inadequate outer race looseness. Table 2 shows rules of thumb.

outside of the given tolerance band since this might result in either excessive or inadequate shaft tightness. Table 3 gives rules of thumb.

Detection of Anti-Friction Bearing Defects

Anti-friction bearing defects are difficult to detect in the early stages of a failure because the resulting vibration is very low and the frequency is very high. If monitoring is performed with simple instrumentation, these low levels will not be detected, and unexpected failures will occur. The vibration frequencies transmit well to the bearing housing because the bearings are stiff. Detection of defects is best done using accelerometers or shock pulse meters. There are guidelines for evaluation of bearing deterioration. For example, a ball passes over defects on the inner race more often than those on the outer because the linear distance around 1. Bearing OD to housing clearance - About the diameter is shorter. 0.00075 inch loose with 0.0015 inch maxiThere are four dimensions mum. of a ball bearing that can 2. Bearing housing out of round tolerance is be used to establish some 0.001 inch maximum. feel for its condition: 3. Bearing housing shoulder tolerance for a 1. A defect on outer thrust bearing is 0 to 0.0005 race (ball pass frequency inch per inch of diameter off square up to outer) occurs at about 40% a maximum of 0.002 inch. of the number of balls

Table 2. Rules of thumb for housing fits

Shaft Fits – A loose fit of the shaft to the bearing bore will give the effect of an eccentric shaft, at a one times running frequency vibration pattern. The objective of the shaft fit is to obtain a slight interference of the anti-friction bearing inner ring when mounted on the shaft. The bearing bore should be measured to verify inner race bore dimensions. Do not install bearings with an I.D.

times running speed. 2. A defect on inner race (ball pass frequency inner) causes a frequency of about 60% of the number of balls multiplied by running speed. 3. Ball defects (ball spin frequency) are variable with lubrication, temperature and other factors. 4. Fundamental train frequency (retainer defect) occurs at lower than running speed values.

A simple check for verification of poor bearing condition is 1. Fit of bearing inner race bore to shaft is made by shutting off the 0.0005 inch tight for small pump and observing that sizes: 0.00075 inch tight for large sizes. the high bearing frequency remains as the pump speed 2. Shaft shoulder tolerance for a thrust reduces. This high frequenbearing is 0 to 0.0005 inch per inch of cy signal will normally diameter off square up to a maximum of remain until the pump 0.001 inch. stops. The frequency indication is normally from 5 to Table 3. Rules of thumb for shaft fits The Pump Handbook Series

50 times the running speed of the machine. In next month’s conclusion to this article we will examine various lubrication methods for anti-friction bearings, as well as the role of bearing protection devices, labyrinths and magnetic seals.

PART II This second half of our series examines ways to keep your anti-friction bearings in top operating condition. Grease, oil flood, ring-oil and mist lubrication systems are all detailed, as well as some advantages and disadvantages of three bearing housing protection devices. Anti-friction pump bearings can be either grease or oil lubricated. Failure from lack of effective lubrication, either in type or quantity, constitutes a major source of bearing difficulties. The primary purpose of oil, or the oil constituent of grease, is to establish an elastohydrodynamic film between the bearing’s moving parts as shown in Figure 1. This film results from a wedging action of the oil between the roller elements and raceways. The formation of the film is, to a major degree, a function of the bearing operating speed and, to a lesser degree, the magnitude of the applied load. Lubrication for ball or roller type bearings can be developed in three ways: 1. full elastohydrodynamic oil film — no metal-to-metal contact 2. no separating oil film — metalto-metal contact all the time 3. mixture of the above methods (boundary lubrication) — occasional metal-to-metal contact with an oil film present only part of the time While the surfaces of bearings are highly finished, there are, nevertheless, small surface imperfections. Use of correct viscosity lubricant ensures development of a full oil film between rotating parts. In boundary lubrication, metal-to-metal contact occurs, and friction wear develops. If the bearings are operated with the correct viscosity lubricant for the speeds and loads involved, a full elastohydrodynamic film will develop between the rotating parts (Figure 1).

243

Under these conditions the oil film is thick enough to separate the bearings completely, despite the unevenness of their surfaces. Since there is no metal-to-metal contact with full film lubrication, there is no wear on the bearing parts. The only time metal-to-metal contact occurs is on startup, or when the bearing is brought to rest. A lubricant with a viscosity too low for the operating loads and speeds permits the moving parts to penetrate the oil film, which results in their making direct contact. In boundary lubrication, this metal-to-metal friction causes wear of the surfaces to increase rapidly, as the film is frequently ruptured. Viscosity requirements for both ball and roller type bearings are expressed in terms of DN value, a factor used to compare the speed effects of different sized bearings. The DN value is obtained by multiplying the bearing bore size in millimeters by the actual rotating speeds in revolutions per minute. Once determined, the DN value is checked against standard tables to determine which viscosity oil to use. It is important to remember, however, that the advantages of proper viscosity can be offset by high speeds if too much lubricant is placed in the bearing and housing cavity. At high speeds, excessive amounts of lubricant will churn the oil and increase the friction and operating temperatures of the bearing. Lubricants also help protect highly finished bearing surfaces from corrosion and, in the case of grease, aid in the expulsion of foreign contaminants from the bearing chamber through periodic regreasing. Increased operating temperature reduces oil viscosity and film thickness and accelerates deterioration of the lubricant. Petroleum based lubricants operated beyond the 200°F range will experience a 50% reduction in life for every 18 degrees above this level. At high temperatures, the more volatile components of oil and grease begin to evaporate and can carbonize or harden within the bearing cavity.

sump, or oil reservoir. This sump maintains a level of oil at or near the centerline of the lowest ball of the bearings, kept constant by means of a constant level oiler. There are two problems associated with this type of lubrication. First, if the level is too high, frothing and foaming may occur, generating heat within the reservoir. Second, there is a very small range between the proper level and the bottom of the balls, below which the bearings are dry.

Figure 16. Grease lubricated pump bearings

Grease Grease lubrication is normally limited to small, low horsepower, noncritical pumps which operate at relatively low speeds and temperatures (Figure 16). The grease can be located in the bearing housing surrounding the bearing or packed in the bearing and then sealed. Oil Flooded A more common lubrication system for centrifugal pumps is the oil flood (Figure 17). In such an arrangement, the bearing housing provides a

CONSTANT LEVEL OILER

LINE BEARING HOUSING

LINE BEARING

BEARING HOUSING COVER VENT

OIL DRAIN GROOVE SHAFT

OIL LEVEL

WATER SHIELD

CASING

OIL DRAIN PLUG

Figure 17. Oil flood lubrication

244

Ring-oiled Ring-oiled systems are often used to lubricate anti-friction bearings in larger horizontal pumps where, because of speed or loads, a simple flood system is not adequate. Rings of a diameter larger than the shaft ride on top of the shaft and dip into the oil reservoir below (Figure 18). These rings are located axially between, but adjacent to, the bearings. The rotation of the shaft causes the rings to rotate and carry oil from the reservoir up to the shaft. The oil is then thrown from the shaft by oil flingers, located next to each oil ring, directly into each bearing to assure complete lubrication. As the oil is circulated through the bearings, it is returned to the oil sump. For proper system function, the oil level in the reservoir must be maintained so that at least 1/4 to 3/8 inch of the ring bore is immersed.

The Pump Handbook Series

oil aerosol into the bearing housing. The equipment necessary for such a system is shown in Figure 20. There are two types of inlet fittings or reclassifiers of the oil particles. They differ in the degree of coalescence from essentially none for the pure mist to partial for the purge mist as shown in Figures 21 and 22.

SHAFT

OIL RING

CAP

SIGHT GLASS (FRONT)

STEM MARK HERE

OIL LEVEL: • HIGH • LOW

0.375" (9.5mm)

0.250" (6.3mm)

Figure 18. Ring oiling

OUTER SLEEVE LOCK SCREWS

ADJUSTMENT RANGE

OIL LEVEL

Figure 19. Constant level oiler

The constant level oiler (Figure 19) is a device used with both flooded and ring-oiled lubrication systems and acts as a small reservoir for extra

oil while maintaining a predetermined level in the bearing housing. The constant level oiler uses a liquid seal to keep the oil in the bearing reservoir constant. When the oil level in the bearing recedes due to consumption, the liquid seal on the spout is temporarily broken. This lets air from the air intake vents enter the oiler reservoir, which releases oil until the liquid seal and proper level are re-established. Unfortunately, each oiler installation is slightly different, so some thought must go into setting this position. The proper level is usually indicated (by either a name plate, casting mark or stamp) on the side of the reservoir. Oil Mist Lubrication The basic concept of the oil mist lubrication system is dispersion of an

MIST HEADER STD. GALV. PIPE

REDUCER

SNAP DRAIN

INSTRUMENT AIR SUPPLY

VERTICAL PUMP

PUMP MIST MONITOR RECLASSIFIER

AIR FILTER

PUMP

2” GALV.

MIST DISTRIBUTION MANIFOLD

S.S. TUBE RELIEF VALVE

BULK OIL SUPPLY

S.S. TUBE-MALE PIPE CONNECTOR SNAP DRAIN VALVE

OIL FILTER

PUMP TURBINE

PUMP

ELECTRIC MOTOR

Figure 20. Basic oil mist system The Pump Handbook Series

Pure Mist In pure mist lubrication (Figure 21), an oil/air mist is fed under pressure directly into the bearing housing. There is no reservoir of oil in the housing, and oil rings are not employed; the pressurized mist flows through the bearings. The moving components of the ball bearings produce internal turbulence, causing impingement and collection of oil upon the surfaces of the ball bearing. Vents are located on the backside of each bearing to assure that the mist travels through the bearings, and a drain in the bottom of the bearing housing prevents the buildup of condensed oil as shown in Figure 23. The advantage of pure mist is that it creates an uncontaminated environment in which the bearings can operate, and protects them from adverse environmental conditions while effectively eliminating heat buildup. The oil mist system will follow the path of least resistance. The back-to-back mounting of the angular contact thrust bearing will have the most windage, so most of the flow of a single inlet fitting will go through the radial bearing, which has less windage and hence less resistance to flow. As a result, heavily loaded bearings may require two spray inlets. The mist must flow from the inlet fitting through the bearing, then to the vent. For duplex angular contact bearings, the flow should be in the same direction as the thrust (Figure 23). The position of the vent and the center spray can be interchanged. Vent area should be about twice the reclassifier bore area. This will create a slight back pressure in the bearing housing, which keeps dirt out. All vents should carry approximately an equal portion of air in multi-vent installations. Different sized reclassifier orifices are needed

245

MIST INLET

CONSTANT LEVEL OILER

PURE MIST

CL 3/10 DIA. HOLE

REMOVE OIL RING

BEARING HOUSING

OPERATING OIL LEVEL

1/4 IN. VENT VENT

Figure 24. Vented oil sight glass bottle

OIL SIGHT BOTTLE

REPLACE CONSTANT LEVEL OILER WITH PLUGS

Figure 21. Oil mist lubrication - pure mist MIST INLET

VENT

housing with oil up to the shaft level. The shaft should be turned 3 or 4 revolutions so that the bearing is coated and the oil drained out of the housing. 2. Connected to the oil mist system and operated 8 to 12 hours to “plate out” an oil film on the bearing.

PURGE MIST FITTING

Purge Mist Another version of oil mist lubrication is called purge mist (Figure 22). This system is employed in conjunction with a conventional oil flood or oil ring OIL SIGHT SEE CONSTANT LEVEL lubrication system. It BOTTLE OILER DETAIL combines the advanFigure 22. Oil mist lubrication - purge mist tages of the positive oil circulation created by SPRAY INLETS the oil rings or oil flood system with the pressurized uncontaminated oil VENT mist system. When this combination is employed, a constant level oiler DIRECTION OF THRUST with overflow feature is used to prevent buildup and flooding of the bearings, which can result in excessive heat buildup (Figure 24). Figure 23. Flow pattern of oil mist

according to bearing size. During the mounting process, bearings must be heated to about 250°F to go on the interference fits of the shaft. Most of the oil on the bearing will flow off. To replace this oil, the pump bearing should be either: 1. Reoiled by filling the bearing

246

Bearing Housing Protection Devices There is a close relationship between the life of rolling element bearings and mechanical seals in pumps. Liquid leakage from a mechanical seal may cause the bearings to fail, while a rolling element bearing in poor condition can reduce The Pump Handbook Series

seal life. Only about 10% of rolling element bearings achieve their three to five year design life. Rain, product leakage, debris and wash-down water entering the bearing housing contaminate the bearing lubricant and have a catastrophic effect on bearing life. A contamination level of only 0.002% water in the lubricating oil can reduce bearing life by as much as 48%. A level of 0.10% water will reduce bearing life by as much as 90%. To improve the conditions inside a bearing housing, various types of end seals are used. In almost every case, the normal operating life and quality of the end seal is not nearly as good as that of the rolling element bearings. Improving the quality of the end seals will increase the life of rolling element bearings. Lip seals The advantages of the frequently used lip seals are low initial cost, availability, and an easily understood technology. New lip seals provide protection in both static and dynamic modes. Their major disadvantage is short protection life due to wear of the elastomer. Life expectancy of a common single lip seal can be as low as 3,000 hours, or three to four months. Thus, while a bearing is designed to last from three to five years of continuous operation, the lip seal will provide protection for only a few months. The temperature limits of lip seals are - 40°F to 400°F (-42°C to 203°C) for Viton. Labyrinths Labyrinths are devices that con-

SUMMARY OF BEARING HOUSING PROTECTION DEVICES Lip Seal Labyrinth Magnetic Seal Acceptable Wet Sump Acceptable Possible Short Life Preferred Purge Oil Mist Acceptable Possible Short Life Acceptable Pure Oil Mist Acceptable Possible Short Life Not Recommended Grease Acceptable Acceptable Vertical Shaft Positions

Acceptable

Top Position Only

Top Position Only

High Humidity or Steam with Temperature Cycling

Possible Short Life

Acceptable

Preferred

Direct Water Impingement

Possible Short Life

Special Design

Preferred

Dirt and Dust Atmosphere

Very Short Life

Acceptable

Acceptable

cause a premature failure, not just the bearings. However, anti-friction bearings — their installation and the environment they operate in — are a major factor in pump life expectancy. Well cared for bearings can extend mean time between failures (MTBF).■

REFERENCES 1. Dufour, J.W., Nelson, W.E., Centrifugal Pump Sourcebook, McGrawHill, 1992. 2. SKF Staff, Bearings in Centrifugal Pumps, An Application Handbook, SKF USA Inc., 1993.

Table 4. Summary of bearing housing protection devices

tain a tortuous path, making it difficult for contaminants to enter the bearing housing. Unfortunately, there are both well designed and poorly designed labyrinth seals. The advantages of labyrinths are their non-wearing and self-venting features. With no contacting parts to wear out, a labyrinth can be reused for a number of equipment rebuilds. Because the labyrinth provides an open, however difficult, path to the atmosphere, the bearing housing vent can be removed and the tapped hole plugged with a temperature gauge. The disadvantages of labyrinths include a higher initial cost than lip seals and the existence of an open path to the atmosphere, which can enable contamination of the lubricant by atmospheric condensate as the housing chamber “breathes” during temperature fluctuations in humid environments. Also, they do not work as well in a static mode as in a dynamic, rotating mode. The temperature limits of labyrinths are determined by the elastomers driving the rotor and holding the stator in place, the same as for the lip seal. Magnetic seals Magnetic seals use a two-piece end face mechanical seal with optically flat seal faces held together by magnetic attraction. They have a

design life equivalent to mechanical seals and rolling element bearings and can be repaired. The major advantage of magnetic seals is the hermetic seal they provide for the bearing housing. Because of the positive seal, other arrangements must be made to allow for the “breathing” that results from expansion and contraction of the air pocket above the lubricant during normal temperature changes. Disadvantages of magnetic seals include higher initial cost and a shorter life than the almost infinite life of a labyrinth. Magnetic seals are generally not recommended with dry sump, oil mist lubricated bearing housings or grease-lubricated bearings. The upper operating temperature limit of magnetic seals is lower than that of labyrinth seals, in the range of 250°F (121°C). Table 4 summarizes suggestions for the application of bearing housing end sealing devices with various types of lubrication systems and environments. Bearing life can be extended by improving the environment of the bearing housing, and this can be accomplished simply by improving the end seals.

Conclusions The effective working life of a pumping system is influenced by many factors that are not necessarily apparent to the facility engineer. Look at all possibilities that can The Pump Handbook Series

247

CENTRIFUGAL PUMPS HANDBOOK

Centrifugal Pumps Operating in Parallel When it comes to pumps and flow, one plus one doesn’t necessarily equal two. by Uno Wahren, Consultant ften it happens that a user buys a pump for a given system. Later its capacity proves to be inadequate, and operations people request that capacity be doubled. A common mistake in this situation is to purchase another pump of equal capacity to add to the system while using the same discharge piping configuration. After the new pump has been installed, it becomes apparent to everyone’s chagrin that the flow has not doubled. The problem is that the operating point has shifted because of increased friction losses in the discharge piping system due to higher fluid velocity. The tools available to analyze a pump system are the head-capacity or pump performance curves and the system curve. Pump suppliers provide head-capacity curves with their pumps as shown in Figure 1. On the y-axis they plot the head, or pressure. The x-axis shows the flow. Pump efficiency curves and often the NPSH required are also shown. The buyer plots the system curve on an x-y axis as in Figure 2. This curve represents the required discharge head and the friction losses in the discharge piping. The x-axis shows the flow in gpm and the y-axis the friction losses in feet. The desired discharge head is a straight, horizontal line, since it remains constant from minimum to maximum flow. When selecting a pump, the buyer specifies the differential head and the capacity. This is the operat-

O

248

ing point of the pump. There are many reasons to have pumps operating in parallel. The most common is flexibility. If only one pump is used for a service and that pump breaks down, the pumping system or process shuts down. A spare pump is therefore often mandatory in critical systems. The spare pump can be identical to the main pump. Project specifications often demand 50% sparing; one or two pumps in parallel supply the total flow, with an additional pump as a spare. This gives extra flexibility for maintenance and in case the process flow varies. Buying pumps for a new system

is fairly straightforward. The first step is to determine the desired flow, the required head, the desired pipe diameter and length, and the valves and fittings included in the system. Next, plot a system curve. The y-axis shows head losses, the x-axis flow. Let’s say a particular application requires 900 gpm at a constant head of 82 feet. Suction is flooded. This constant head may be the pressure of a pipeline from which the liquid is to be pumped, or the elevation of a reservoir, tank or pressure vessel. The optimum discharge pipe diameter is determined to be 8”. The last step is to plot the friction losses against flow. NPSH

14

HEAD IN

12

FT

8

100

6

170

4

160

2 0

150 140

66%

21" DIA

74% 80%

130 120

19" DIA.

HEAD CAPACITY CURVE 84%

110 100

86% 17" DIA

84%

90

80%

80

74%

70 60

66%

OPERATING POINT

NPSH

50 40

60 50

30

40

20

30

10

1 2

3 4

5

6 7 8

9

10

12

14

FLOW IN 100 USGPM

Figure 1. Typical head-capacity curve The Pump Handbook Series

16

18

20

22

24

26

recirculation can become a problem. The shaft and bearings may not be designed for the increased torque and loads of the higher speed and horsepower requirements. A gear train means added maintenance problems. A larger driver may be required. Why get into a situation like that? Assuming that the installed impellers have an outside diameter of 6”, the maximum size impeller for that pump is 6.3” O.D. The affinity law for a constant speed pump is:

HEAD IN FT 170

150

130

SYSTEM CURVE

110

90

70

50

D1/D2 = Q1/Q2 = √H1/√H2 STATIC HEAD

30

The flow and head for a 6.3 diamter impeller are:

10 1

2

3

4

5

6

7

8

9

10

11

12

FLOW IN 100 USGPM

Figure 2. System curve

Figure 3 shows a straight line parallel to the y-axis drawn at 900 gpm and intersecting the system curve at 82 ft. This particular system requires a pump that delivers 900 gpm with a discharge head of 82 ft. Two pumps each delivering 450 gpm, or three pumps each delivering 300 gpm, will achieve the same result. In this case, the decision is to use two pumps running in parallel, each with a capacity of 450 gpm. Each pump operating alone delivers approximately 520 gpm at only 68 ft. This is adequate because the end of the line or static head is 60 ft. If the calculated system head curve is above actual, it will be necessary to throttle the pump flow, using either the discharge isolation valve or an eventual control valve on the discharge line. Suppose a system has one pump delivering 450 gpm against 68 ft. The pump discharge line is 6” new steel pipe. The fluid is water with a specific gravity of 1.0. There is a request to double the flow. In a case like this it is very common to request another identical pump — one that will deliver 450 gpm at 68 ft. That is a mistake. With the addition of the second 450 gpm pump, the combined (1 and 2) head-capacity curve will bisect the system curve (Point B) at 790 gpm. The total head required at that flow is 78 ft as shown in Figure 4.

If the new pump is already bought and installed, flow can be increased by changing the discharge piping from 6” to 8”. This is usually not an economical solution. The discharge line may be long. Replacing valves and fittings may be expensive. Another solution is to buy a third pump and run the three pumps in parallel. This can also be expensive and adds to maintenance costs. Changing the impellers to a larger diameter that will still fit into the pump casing may also solve the problem. The characteristics of a centrifugal pump are such that with constant speed (rpm) and a specific impeller diameter, the curve will not change, regardless of the properties of the liquid being pumped. However, the curve will change if either the speed or the impeller’s diameter change. The performance curve shown in Figure 1 will change if the pump’s rpm changes from 3560 to 4200 through a speed increasing gear as follows: Q1 = 450 gpm; H1 = 160 ft Q2 = 450 x 4200/3560 = 530 gpm H2 = 160 x (4200/3560)2 = 223 ft Increasing pump speed is uncommon. Pump characteristics change when speeds are higher than design specifications. As NPSH requirements increase, internal The Pump Handbook Series

Q2 = (450 x 6.3)/6 = 472 gpm H2 = (√68 x 6.3/6)2 = 75 ft (Note: Some impeller designs do not precisely agree with the affinity laws for impeller diameter changes. Always discuss this with the pump manufacturer.) As shown in Figure 5, the two pumps with the larger impellers operating in parallel still do not deliver the full 900 gpm. In this case, the total flow might be enough for the particular process. On the other hand, it may not. Increasing the impeller size will often not solve the problem. To buy a pump that will double the flow requires plotting the system curve against the head-capacity curves. The plot shows that at the increased flow the system will require a total discharge head of 82 ft. When the first pump (pump 1) runs alone on the system curve, the flow is 450 gpm and the head is 68 ft. Next select a pump (pump 2) which, running in parallel with pump 1, will deliver the required flow at the required total discharge head (TDH). For this example, an adequate pump is one with a flow of 645 gpm with a TDH of 72 ft. The pumps are operating in parallel on the same system curve. The pumps are sized so that pump 1’s head-capacity curve intersects the system curve before point A. Therefore, pump 2 is the commanding pump. If pump 1 starts first (before pump 2) it will back off and the system will deliver only about 650 gpm against 72 ft of head. Figure 6 shows how the two dis-

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similar pumps can operate correctly only if the head-capacity curve intersects the system curve on the AD portion of the head-capacity curve. The pumps may be throttled, but not further back on their curve than 525 gpm (Point A), for then pump 1 will back off, letting pump 2 deliver only its full capacity. For the system to operate, pump 2 has to start first. After pump 2 has reached full flow, pump 1 can be started. (Figure 6). When selecting an additional pump, bear in mind that no pump will operate satisfactorily from zero flow to the end of curve flow. Pump manufacturers show minimum allowable constant flow on their pump curves. This means that below that flow, the pump is unstable. The

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minimum flow requirements for low head, low capacity pumps range from 20 to 25% of flow at Best Efficiency Point (BEP). It is not uncommon for large multi-staged pumps to have minimum flow requirements as high as 65% of BEP capacity. Always pick a pump where the required flow is less than the flow at BEP. It is good practice to specify that a pump will operate satisfactorily at 120% of the rated flow, since flow requirement may increase. The higher the flow, the higher the NPSH requirement. At some point beyond the BEP, the pump curve will collapse. Operations beyond that point will cause cavitation and severe pump damage.■

The Pump Handbook Series

Uno Wahren is a Registered Professional Engineer in the State of Texas with over 30 years of experience as a project and rotating equipment engineer in the oil and gas industry. He has a B.S. degree in Industrial Engineering from the University of Houston, and a B.A. in Foreign Trade from the Thunderbird Graduate School of International Management, Glendale, AZ.

CENTRIFUGAL PUMPS HANDBOOK

Fire Pump SystemsDesign and Specification Think your fire pumps are just like the rest of your fluid-movers? Think again. Rigorous standards and certifications make sure these life-savers are up to snuff. by George Lingenfelder and Paul Shank, Precision Powered Products oint of view is everything when discussing fire pump systems. For the engineeringcontractor, they are relatively uncommon systems referencing specifications unheard of in conventional process units. To the purchaser, they are mandated systems that take time, space and capital away from money-making units. For users, fire pump systems are (or should be) a once a week test requirement. But in spite of the time, space and money constraints, fire pump systems must work as required. Period. Hundreds of lives and millions of dollars in hardware and production costs rely on the performance of these systems. Fortunately, they almost always work. Because of the critical nature of this service, one might think that industry standards could simply be invoked to insure reliable design and specification. Of course they can. They just never are. At least not without the addition of supplemental proprietary specifications that can conflict with industry standards, government regulations and sometimes with the basic system design. Seemingly minor requirements can result in the loss of a listing agency label and render a perfectly functional system unacceptable to insurers or governmental agencies.

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Pumps The basic fire pump system includes a UL (Underwriters

Laboratory) or FM (Factory Mutual) listed pump, driver and controller. Pumps are rated in discrete increments starting at 25 gpm and extending through 5000 gpm. Each capacity designation is tested for compliance with NFPA 20 (National Fire Protection Association) requirements for performance and by independent testing institutions such as UL or FM for design, reliability and safety. A single pump can be certified for more than one operating point, but it must meet all performance and design criteria at both points. In the area of basic hydraulic performance, a pump must deliver at least 65% of rated discharge pressure at 150% of rated flow to achieve NFPA 20 acceptance. So a 1500 gpm pump rated at 150 psig must deliver 2250 gpm at a discharge pressure of not less than 97.5 psig. This operating range then establishes the design parameters for the piping system and fire fighting equipment. Another requirement establishes that the maximum shut-off head must not exceed 140% of design head. Both horizontal and vertical centrifugal pumps are available as listed pumps. (For brevity, we will use the term "listed" to refer to any equipment certified by UL, FM, the Canadian Standards Association (CSA) or other agency as suitable for fire pump system application.) While there are some obvious design differences between the requirements for these two styles, The Pump Handbook Series

Photo 1. Fire water pump with diesel engine and air start with a nitrogen back-up system

the performance prerequisites remain un-changed. It is important to note that listed pumps are not designed or manufactured to API 610 standards. Materials of construction vary with the type of system and the fire fighting fluid. The standard cast iron case with bronze impellers and wear surfaces is the most common in landbased applications. This material is generally suitable for fresh water and sometimes brackish or even salt water services, depending on the length of service and anticipated life of the unit. Even in fresh water services, it is critical to consider casing corrosion rates when choosing materials. For more aggressive fluids, many manufacturers offer various grades of bronze, including zinc-free and nickel-aluminum bronze. Higher alloys of 300 series stainless steel and Alloy 20 may also be available for very corrosive services or situations in which the expected project life is extremely long.

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Figure 2. Plan view of horizontal fire water pump system with diesel engine drive, discharge piping main relief valve and discharge cone Figure 1. NFPA 20 vertical fire pump system with right angle gear, diesel engine drive and elevated fuel tank with containment reservoir

Vertical pumps are available in a much wider range of metallurgies because of their applications in corrosive offshore environments. Various grades of bronze are common, as are higher alloys. Carbon steel is very rare. Material options for horizontal units include bronze casing materials, moving up to carbon steel and then austenitic steels for both case and internals. As with vertical pumps, higher alloys are also available for special situations. From a commercial standpoint, cast iron-bronze fitted (CIBF) pumps are the basic choice. Standard aluminum bronze materials such as ASTM B148 Alloy 952 (for vertical units) generally command price premiums in the 4.5-5 multiplier range. Nickel-aluminum bronze (B148 Alloy 955) will increase this pricing by an additional 10-15%. Carbon steel (for horizontal units) will raise standard CIBF costs by 3-4 times. Stainless steel, typically 316SS, increases base costs by a factor of 5-8 depending on size. Higher alloys can increase CIBF costs by up to 8 times. Auxiliary equipment and accessories offer many possible options for the pump. While these will be discussed in greater detail later, it is important to note that most fire water pump manufacturers only

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offer a limited number of materials and options for listed pumps. Pumps (like drivers and other equipment) can be listed for fire pump service only when they are manufactured in the same materials and configuration as the design tested by the certifying agency. Changes in casing materials, even when poured by the same foundry using the same patterns, can prevent a manufacturer from labeling the pump. Likewise, changing accessories, especially on engines, can result in the loss of the label. In practice, this is usually a semantic issue and rarely becomes a significant problem, provided that the original design or unit was labeled and that the changes are clearly upgrades in materials, control and/or reliability. The importance of labeling also varies with the location and type of fire pump service, as well as responsible government agencies and insurance carriers. Most manufacturers offer labeled equipment in the most elemental fashion for application in the widest range of markets. This makes it possible for the specifying engineer and user to work together with the supplier to develop the best system for their application. Even The Pump Handbook Series

though the end user might customize the system and add accessories that void the label, the original equipment was labeled, and this shows the intent of the user to install labeled equipment.

Drivers While the pump is the mechanical heart of a fire water system, the driver is a critical component, frequently requiring more attention in specification and design. Diesel engines power the vast majority of fire water pumps, creating another difficulty for the specifying engineer. Most of us have limited experience with this category of equipment. They are rare in general plant design and frequently used only because the power required exceeds the available UPS (Uninterruptible Power Source). Unlike the situation with motors, the specification and selection of diesel engines cannot be foisted off on another discipline. The basic, listed fire pump engine, offered by many manufacturers, is more than adequate in terms of reliability and service life. After all, this system will only operate once a week for thirty minutes until it is called on to run in a real emergency. Then it will operate for eight hours (the standard fuel tank sizing) or until it is destroyed in the conflagration. A great deal of time and energy is expended on the starting system,

Photo 2. Custom fire water pump system designed for Class 1, Groups C & D and Division II. The controller features all NEMA 7 switches mounted in a stainless steel cabinet. The battery charger and battery case are purged and constructed of 316SS. The system has a primary electric start with a back-up hydraulic system.

Photo 3. Vertical fire water pump for offshore installation. Compressed air start with back-up nitrogen system and expansion receiver

which often includes back-up start systems and sometimes even multiple starting systems. Basic systems include dual sets of batteries. Also commonly available are pneumatic systems—compressed air with bottled nitrogen back-up and hydraulic starting systems. Many of the larger engines, above 200 hp, offer two ports for starters, so that two types of starting systems can be used. This having been said, most experienced engine users agree that if the unit does not start within the first three to five seconds, the likelihood of starting at all is just about nil. For this reason, it is very important to pay close attention to the starting cycle of the unit during the weekly exercise sessions. Remedy any difficulties or malfunctions at once. Although relatively uncommon, some systems also use motors, espe-

cially when the firewater pumps are also used for water lift or washdown services. Whether they are more reliable remains an open discussion. The simple fact of their predominance as drivers has made them more acceptable. Like other major components of fire pump systems, motors are available in UL/FM and CSA listed varieties from a wide range of manufacturers. While still not widely accepted, the most recent revision of NFPA 20 (January 1998) now requires fire pump motors to be UL listed.

Controllers Controllers were the last major component of fire pump systems to receive certification by various listing agencies. Unlike other components, however, they come with a wide range of options and even provide remote system contacts for the other pump, driver and other options that may need to be incorporated into the control scheme. A few areas in the design and specification of the controller should be reviewed carefully. The first is area classification. Unusual as it may seem to designers and specifiers who are unaccustomed to it, standard drivers and controllers are not rated for any National Electrical Code (NEC) area classification. Depending on the manufacturer, they can add this as an option or build a custom controller that meets a specific area’s classification. Meeting a Division I or II classification usually involves adding a Zpurge system, which can be expensive relative to the cost of the controller (typically increasing cost by up to 50%), but this is currently the most cost-effective method to meet these requirements. Custom controllers, a very expensive alternative, will generally incorporate hermetically sealed or NEMA 7 enclosures for all potentially arcing devices. In addition, diesel engine control sensors and battery start systems are not intended for a classified area. Electric motors must also be specified for the area. Pneumatic engine controllers with The Pump Handbook Series

air start systems are an obvious alternative when area classification becomes an issue. Although not commercially available as UL/FM listed units, it is generally agreed that they are inherently explosion-proof because they have no electrical components. Custom built electric controllers, however, can meet the area classification requirements of any facility. Controllers incorporating PLC logic are also available as standard commercially available, but not as UL/FM listed units, or pneumatic controls as custom built units. Custom manufactured controllers are becoming more common, especially in the HPI/CPI markets to meet NEC area classifications. It is important to note, however, that while these units may consist entirely of UL (or other) listed components, this does not convey a UL listing to the controller. Thus, the controller does not meet the UL218 requirements. This may appear to be a subtle differentiation, but local codes and regulations, as well as insurance requirements, may require sacrificing an area classification for a UL/FM listing.

Instrumentation Fire pump systems are not process systems. Thus, much of the advanced instrumentation applied to process systems is relatively uncommon in the world of fire pumping. However, some systems include transmitters and other “smart” devices that provide additional information to the user. Carefully review the purpose and function of these devices to determine their value under the anticipated operating conditions and to insure that they provide valuable information without contributing to needless complication. Unlike a process system, where almost every eventuality can be understood, evaluated and prepared for, the full array of conditions under which a fire pump system will be used is almost impossible to imagine. Controls should be as simple and straightforward as possible, lest they fall victim to the law of unintended consequences.

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Photo 4. A 570 hp V-12 diesel engine driver for all nickel aluminum bronze fire water pump producing 3500 gpm at 180 psi.

Industry Standards Two major organizations are the primary promulgators of standards for fire pump systems, and their roles, far from being conflicting, are complementary. National Fire Protection Association The National Fire Protection Association, through its NFPA 20 Standard for the Installation of Centrifugal Fire Pumps, provides the most complete set of provisions for the design and installation of various components as well as the overall system. The NFPA 20 sets standards for design and construction of all the major components in the fire pump system, including minimum pipe sizing tables, electric motor characteristics, performance testing, and periodic testing and system design. NFPA 20 further attempts to develop a safety standard or level of performance for centrifugal fire pump systems to provide a reasonable degree of protection for life and property. Under these provisions, alternate arrangements and new technologies are permitted—and in fact encouraged. The National Fire Protection Association does not, however, certify or evaluate compliance with its various specifications and guidelines. NFPA 20, while the source of some performance requirements, is best known for its detailed design provisions for all components commonly found in fire pump systems.

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Unlike many industry standards that address only the design, manufacture and testing of specific components, NFPA 20 is a veritable 'how-to' manual for the engineer or user who needs to develop guidelines and specifications for fire water systems. Taken as a complete document with referenced texts, this standard alone will assure the purchase and installation of a reliable fire pump system. Take care in developing additional specifications, whether stand-alone or ancillary, to avoid conflicts with NFPA 20. Because of the detailed nature of this specification, add-on requirements can frequently have the effect of actually diminishing the standards. Here again it is important to keep in mind the function of the equipment in a fire pump system. Contrary to the design considerations in process units, fire pump systems should be designed to operate reliably under the most adverse conditions for as long as required, but typically this is only a matter of hours. In a process unit, systems are designed to protect themselves with various shutdowns and monitors. Fire pump systems are designed to protect the plant and personnel even if that means operating to destruction. In other words, three hours into a fire fighting event, the vibration level of the diesel engine driver is of no significance so long as the pump continues to deliver adequate flow. NFPA 20 includes pump design guidelines for vertical shaft turbine pumps, horizontal (both end-suction and split-case) and vertical in-line pumps. It also includes standards for motors, both horizontal and vertical, right angle gears and diesel engine drives. Additionally, the guidelines are an excellent resource for auxiliary and ancillary equipment found in most fire water pump systems. For example, the diesel engine specifications include not only requirements for the engine itself, but also the fuel supply, exhaust system and control system operation. The Pump Handbook Series

Underwriters Laboratories Underwriters Laboratories (UL) reviews equipment and systems from a performance orientation. The organization provides an independent, third party evaluation of manufacturers for a variety of components, but for fire pump systems the primary specifications for our consideration are UL448 for pumps, UL1247 for diesel engines and UL218 for controllers. In certifying equipment to these standards, UL reviews construction design, materials of construction and overall performance. These reviews assess the ability of the equipment to perform the task for which it is to be rated. Following initial certification, UL maintains an ongoing surveillance program to insure continued adherence to design and performance criteria. UL field representatives make unannounced visits to all manufacturers displaying the UL label. If team members encounter problems or questions, they may schedule more frequent visits. UL standards are among the most stringent in the industry, and equipment certification is necessarily very specific. So what happens when a listed product is modified to meet customer specifications or site specific requirements? If the change is clearly an upgrade to the existing listed design and not a major change in the product, UL will frequently certify the modification without re-testing. Major modifications, of course, must go through the complete certification process. The manufacturer is responsible for initiating changes to the listing. UL448 sets the standards for certification of centrifugal pumps for use in fire water systems. As previously noted, UL448 is based on the ability of the unit to perform under the conditions of service required and according to the manufacturer's specifications with regard to total differential head (TDH), capacity, and efficiency or power requirement. Materials of construction are reviewed for strength and corrosion resistance. Each unit or size to be

certified is given an operational performance test and is also hydrotested to twice the manufacturer's published Maximum Allowable Working Pressure. This is a substantial increase over pump industry standards of 150% hydrotest pressures. After certification, UL448 requires that each unit bearing the UL mark be tested successfully for hydrostatic integrity and hydraulic performance. Diesel engines fall under the scrutiny of UL1247. Again, the emphasis is on performance. Certification requirements include extended testing on dynamometers as well as speed control. Because of the importance of driver speed in centrifugal pump applications, speed control and overspeed shutdown operation are critical areas. Units are also tested for their ability to start under a wide range of conditions, both hot and cold. After certification, each production unit shipped must be subjected to a dynamometer test including performance checks of the speed control and overspeed shutdown systems. Engine and motor controller specifications are covered by UL218, which is written and administered by the Industrial Controls section of Underwriters Laboratories. While this section is closely aligned with UL's primary purpose to review equipment for fire and shock hazards, the critical nature of fire pump controllers makes performance an important concern of this standard. These units are reviewed for safety, of course, but the importance of starting a fire water pump under emergency conditions warrants a different philosophy in certifying the controllers. Controllers are inspected with special attention to their ability to signal—that is, to notify remote personnel of any abnormal conditions in the controller system—as well as to be able to accept remote instructions for starting. Diesel engine controllers, besides providing a starting signal to the engine, must also provide charging current for the batter-

ies, perform a weekly starting and run test of the engine and driven equipment train, as well as provide visual and audible indication of various engine failures. These include failure of the engine to start, shutdown from overspeed, battery failure, battery charge failure and other abnormal conditions. Diesel engine controllers also include pressure recorders to sense pressure in the fire protection system and confirm unit performance on demand or during weekly tests. Electric motor control requirements also focus on the ability of the unit to start the motor drive. This overriding concern is demonstrated in unit design, for fire system controllers are different from standard motor controllers. This is perhaps best demonstrated by their use of either a listed fire pump circuit breaker or a non-thermal instantaneous trip circuit breaker with a separate motor overcurrent protective device for protection against overcurrents and short circuits. Another difference between these and standard controllers is that once a fire pump controller is under emergency conditions, it is prevented from shutdown except when it is under a condition more threatening than the fire or until conditions return to normal standby. Surprisingly, UL lists no fire pump controllers for installation in hazardous (classified) locations. To meet Division I or II requirements, controllers must be custom designed with a suitable hazardous location protection method. They must either have a purge system or be explosion proof. Custom designs, of course, will incorporate either NEMA 7 components or enclose elements in explosion-proof boxes. The X and Z-purge systems are the most common. For international installations, the problems are compounded because no harmonized IEC (International Electrotechnical Commission) standards exist that address fire pump controllers This situation, however, is being addressed by a joint working committee of the NFPA and UL that is developing an IEC standard that

The Pump Handbook Series

will incorporate NFPA 20 and UL218 requirements. At present, the committee has developed a draft specification and submitted it to the IEC for inclusion as a new work item proposal (NWIP). Following acceptance as a NWIP, the IEC will convene a group to review and comment on the standard for development as an IEC publication. Given the importance of international markets, this move will be an appreciated step for U.S. based manufacturers.

Conclusion Fire pump systems are designed to protect life and property. In this area, specifiers, owners and operators need to change their perspective on equipment, controls and operation. Performance and reliability must be the priority in design, purchase and maintenance. While site specifications and purchaser requirements certainly must be considered, simpler is usually better. The ability of the system to perform reliably under the most severe conditions, for the protection of life and property, must be the driving consideration. n

Special thanks to Kerry Bell and Dave Styrcula at Underwriters Laboratories for their assistance. George Lingenfelder is a principal and founder of Precision Powered Products, a systems design and fabrication firm serving the oil and gas industry the petrochemical/refining markets. He worked for major engineering and construction companies as a Systems Engineer for ten years before starting Precision Powered Products in 1984. Mr. Lingenfelder received his B.S. degree in mechanical engineering from the University of Houston Paul Shank is with Precision Powered Products in Houston, Texas. He has more than 20 years of sales and engineering experience in the specialty process equipment field. He has authored technical articles and coauthored several software programs for the selection and pricing of pumps, compressors and steam turbines. Mr. Shank received a B.S. degree from the University of Houston.

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CENTRIFUGAL PUMPS HANDBOOK

Well Pump Applications for Mine Dewatering Choosing the right pumps means knowing drainage requirements, dewatering schedules and well construction as well as system and fluid conditions. by Mark List, Miller Sales and Engineering

nce the economics of a mineral deposit have been determined as favorable and a decision made to proceed with mine development, significant financial commitment is placed at risk in expectation of a calculated return on investment. Many aspects of mining carry relatively high levels of uncertainty that contribute to the overall degree of risk. One important consideration is groundwater control, should it be a factor, during mining operations. Where mining must take place below the water table in highly permeable geologic material, operations would not be possible without effective control of groundwater. Several mines in the western United States have developed large well fields, up to 70,000 gpm capacity, to intercept inflow and divert groundwater from the workings.

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Mine Dewatering Objectives There are two general objectives to most mine dewatering programs: Keeping the working conditions relatively dry and maintaining the stability of the excavation or opening. Operating costs and production rates are directly influenced by working conditions. Wet floors and working faces create poor ground conditions for heavy loading and hauling equipment. They also increase tire wear, reduce cycle times and impact tonnage factors. Safety becomes a direct issue if elec-

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tric powered machinery is used. In the worst case, a submerged working level becomes inaccessible and production is halted. The stability of open pit walls and underground openings is of vital concern for safety and economic reasons. Adequate drainage must occur in order to keep pore pressure at acceptable levels based on geotechnical stability analysis. General Types of Mines and Groundwater Control Methods Mines are either open pit excavations or underground excavations, or both. In some cases large scale open pit mines have succeeded prior episodes of underground mining in the same area. In other situations, underground mines are developed adjacent to or from within existing open pit mines. Driven by metal prices and advancing technology, companies have continued to explore deeper and/or more challenging geologic territory for mineable orebodies. Mine dewatering methods have evolved out of necessity in response to the increasing groundwater control requirements of contemporary mining. Where conditions permit, open pumping from collection sumps is a standard practice. This method is commonly used in open pit mines to control surface water drainage and in both open pit and underground mines where groundwater inflow rates are small enough to be The Pump Handbook Series

Photo 1. In-pit well in service with loading operation on left, blasted material on right, and high wall in background

managed in this manner. Booster stations might be required employing positive displacement pumps or horizontal centrifugal pumps designed for high head dirty water, abrasive solids or slurry service. Depending on the hydrogeologic setting, however, some mines cannot be effectively or safely dewatered using this method alone. Well Field Systems for Dewatering Several mines in northern

units are applied. Submersibles are also installed in series using tandem 1000 hp (2000 hp) units, tandem 1170 hp (2340 hp) units and combination 1500 hp/800 hp (2300 hp) units with the lower setting at depths exceeding 2000 feet. Pit perimeter wells typically discharge to gathering mains at relatively low well head pressures. Inpit well pumps can be staged for the additional head required to discharge to a surface location outside the pit, or to a booster station in the pit. Vertical turbine can pumps and horizontal centrifugal pumps are in service for this application.

Understanding the Application Photo 2. Twin 800 hp, 5000 gpm vertical turbine can boosters with 42” discharge main in background

Nevada require the use of wells to intercept and lower groundwater levels around open pit and underground excavations. The orebodies associated with these mines are hosted in fractured bedrock formations along mountain ranges that contact alluvial basins high in groundwater storage. Mining companies drill large diameter deep wells in bedrock fracture zones tested for favorable production yield. Other wells are completed so as to intercept shallow recharge or promote drainage of less permeable zones. Wells are typically located outside the open pit, but drilling sites in the pit are often unavoidable due to local hydraulic compartmentalization. A well field can be comprised of 30 or more individual wells with completion depths up to 2000 feet or greater and production casing diameters up to 24 inches. Well-specific capacities can exceed 60 gpm per foot of drawdown. Vertical turbine line shaft pumps (400 hp) are in service at setting depths of 1020 feet, and 1500 hp vertical turbine line shaft pumps are in service at setting depths of 800 feet. Single 2200 hp submersible units work at setting depths of more than 1800 feet, and at more than 2000 feet deep single 1500 hp submersible

The uncertainties involved in developing an efficient mine dewatering program become much better understood as operations progress. Groundwater flow information available at the onset of large scale dewatering can be very complete and supported by sophisticated model simulations, but such information is usually based on field test data that cannot be conducted at a scale proportional to what will actually be undertaken. Granted, pump applications engineers are most comfortable when customers assume all risk by specifying the necessary conditions for pump equipment selection. The outcome is likely to be better for all parties involved if knowledge is shared prior to establishing the conditions for equipment selection. Getting the Right Concept Well field dewatering involves lowering the water pressure level in the mine area to permit safe, efficient excavation. Over the life of the mine, the change in pumping lift from the initial static water level to some future pumping water level can be large—on the order of 1000 feet or more. Individual well production capacities can decrease dramatically depending upon aquifer system characteristics. From a pump applications perspective, this means selecting equipment with initial operating points that best match starting conditions and which can be made to fit, if possible, conditions that are expected to occur as The Pump Handbook Series

formation dewatering progresses. Dewatering Schedule and Pumping Rates The rate at which a mine is expected to be deepened below the static water level is an important planning factor. It is used to establish a schedule for lowering groundwater heads before excavation begins, and it is a major consideration in predicting the required overall pumping rate. The change in pumping lift over time, indicated by the dewatering schedule, is the variable component required to evaluate intermediate and final TDH conditions for pump selection. Pump capacity range can be estimated assuming that sufficient test or operating data are available to be confident in doing so. The most reliable values for individual well production capacity and efficiency (well drawdown) are not available until the well is constructed and test pumping has been completed. However, the overall mine development schedule might not allow for the long delivery times that may be needed for special pump engineering or construction. If this potential problem is not addressed during project planning, pump equipment orders can be placed with results that are not costeffective over the long term. Well Construction Well dimensions limit the size and type of pump equipment that can be installed. Although very costly to construct, large diameter deep wells can accommodate the installation of the large diameter four-pole (1800 rpm) 1500 hp and 2200 hp submersible electric motors that are required for high volume deep set applications beyond the practical setting depth limitations of line shaft pumps. Similar deep set applications with smaller casing diameters can require the use of two-pole (3600 rpm) submersible motors with an overall length of 100 feet or longer. In both situations a well must be drilled deep enough to achieve the design pump intake elevation and to accommodate motor equipment length and standard clearances.

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In addition to well diameter, well alignment is of critical importance for deep set line shaft pumps. Even the closest attention to construction alignment standards, however, cannot prevent ground movement from adversely affecting well alignment as dewatering progresses. Depending on the severity of ground movement and resulting deflection, shaft vibration can result in shaft and motor bearing failure.

Photo 3. 400 hp vertical turbine line shaft (1020' setting) with mineral processing plant in background

System Considerations System conditions vary in response to production well field changes and discharge method modifications. A well head pressure condition can usually be determined for use in staging the well pump, but a conservative approach is often taken to ensure that the desired pumping capacity can be maintained. If required, throttling is used to impose pressure temporarily until system conditions are within pump operating conditions.

ment materials and coatings. Unfortunately, geologic formation water conditions can and do change during dewatering. Partial aeration can occur with the rapid displacement of groundwater, and this can lead to unanticipated corrosive damage. Bronze and bronze alloys should be considered if conservatism is justifiable. If standard materials are selected, the first pump tear-down will reveal what doesn’t work.

Fluid Conditions Water temperature and corrosivity are major factors influencing the selection of dewatering well pump equipment. Water temperature can influence the type of construction and materials used in a line shaft pump, but elevated water temperature adds significantly to the cost of submersible electrical equipment and thus can be a limiting factor in selection. At one particular dewatering operation, line shaft pumps are not an option, and submersible motors rated at up to 2200 hp are operating in water temperatures of 140°F. These are oil filled motors of specialized construction sometimes fitted with heat exchangers. Because the motors are located in the lower reaches of the wells, below the pump intake, shrouds designed for adequate water flow past the motors are usually required for cooling purposes. Unforeseen corrosion damage to pump cases, impellers and column pipe joints can ruin the best efforts in hydraulic applications engineering. Corrosion potential can sometimes be estimated up front by water quality analysis, and should be taken into account, if possible, in the specification of pump equip-

258

Equipment Selection Making a reasonable attempt to understand the factors that dictate initial conditions and influence future conditions is key to selecting dewatering well pumping equipment that will remain effective under actual operating conditions. Nevertheless, there are limitations, and well yields can eventually decline to the point that pumps must either be operated intermittently or replaced with lower capacity units. Electric Submersible vs. Line Shaft Pumps Vertical turbine oil lubricated line shaft pumps are operating successfully at setting depths of more than 1000 feet. The slower pump and driver speeds (1800 rpm or less) of these units are favored by many operators. Pump damage from abrasive particles or partially aerated formation water is significantly less intense at slower speeds. Electrical problems are much simpler to troubleshoot and correct because motors are located at the surface above the discharge head. On the downside, slower speeds require larger bowls and well casing diameters. Line The Pump Handbook Series

Photo 4. Pump rig installing deep set line shaft pump

shaft equipment is mechanically complex and requires special engineering and manufacturing for bowl tolerances to accommodate the effects of relative shaft elongation under high thrust loading. Well alignment problems can adversely affect shaft and bearing life or even preclude the use of line shaft equipment. High capacity submersible equipment is available in both 1800 rpm and 3500 rpm classes. Small diameter high yield wells or setting depths greater than 1000 feet generally restrict pump equipment to the submersible type. The limitation here is motor power output. Four-pole 20inch diameter motors (1800 rpm) are available to 2200 hp for 24-inch casing applications. Slim line two-pole motors (3500 rpm) can be coupled in tandem to produce more than 1000 hp. These two-pole motor assemblies are more than 100 feet long, requiring additional well depth. Because the motors are installed below the pumps, electrical faults that occur in the down hole power cable or motor system require retrieval of the equipment string from the well for testing and repair to take place.

The Pump Curve For a desired initial performance and estimated final performance, there is a simple rule of thumb for dewatering pump selection: start on the right side of the H-Q curve, run back to the left through the Best Efficiency Point, and plan to refit the pump end with additional stages if necessary to conform with estimated future conditions. Another method is to throttle the pump during initial operations if the range of expected conditions indicates that this will eliminate the need to pull and refit the pump end. Throttling is most common in deep set submersible applications involving relative certainty in the drawdown rate and final conditions. Pump Mechanical Considerations Line shaft applications involving deep settings and high thrust require special consideration for relative shaft stretch and bowl endplay requirements to establish adequate lateral impeller clearance under running conditions. Enclosing tube tension design as well as manufacturing tolerances for tube, shaft and column pipe lengths also need to be considered. Surface Equipment Power transformers and switching equipment are available in modular form on skid mounted platforms. Also, individual components can be custom assembled on a common skid

or placed on individual pads if preferred. Mine power distribution systems often are plagued with swing loads and transients depending on the variety of electric machinery in service. Power factor correction, system protection and motor control requirements vary with the application. Vertical hollow shaft type motors used with line shaft pumps are usually 460 volt or 4160 volt. Submersible motors are typically 460 volt, 2400 volt or 4160 volt. Installation Considerations Line shaft pump installation can be more mechanically involved than submersible pump installation. The oil tube and shaft are usually shipped assembled in lengths of 20 feet and must be individually placed in each piece of column pipe before installation in the well. Because the column, tube and shaft assembly are run in the well casing, three threaded connections must be properly made at each 20-foot interval. The projection dimensions of the tube and shaft, which start at the pump discharge case, must be maintained over the length of the column assembly for proper fit at the discharge head and motor coupling. Depending on the manufacturer of the submersible pump equipment, motor system and pump assembly during installation can be more or less complicated, generally requiring manufacturer’s field ser-

The Pump Handbook Series

vice in addition to the installation crew and equipment. However, once the pump and motor equipment are assembled in the well and tested for continuity, column installation is a straightforward process of making one threaded joint per pipe length and securing the power cable to the column. This goes relatively quickly, especially if the pump rig can handle pipe lengths of 40 feet. Operating Considerations It is important to follow up on the performance of a pump after it has been placed in service. The operator will no doubt inform someone associated with the sale of the equipment if a failure has occurred or performance is not as represented; conversely, the operator will be concerned with other matters if equipment performance is acceptable. Either situation involves information that can assist the applications engineer in selecting proper equipment and recommending the most effective modifications. n Mark List has more than 25 years of experience in the construction and mining industries, with 10 years of practice in groundwater investigation and mine dewatering equipment design, construction and operation as Chief Dewatering Engineer for a western gold mine. He is a registered engineer and state water rights surveyor in Nevada.

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CENTRIFUGAL PUMPS HANDBOOK

Design Practices for Safe Handling of Hazardous Fluids in API Applications Safety begins at the specification stage when dealing with dangerous pumpage. By Ron Forsberg, Sundyne Corporation he chemical and hydrocarbon processing industries have typically specified mechanically sealed and sealless centrifugal pumps for hazardous applications. Sealless pumps, both canned motor and magnetic drive, have been used in the chemical processing industry (CPI) for more than 40 years, but their use in the petroleum processing industry (API) was extremely limited until the early 1990s. Passage of the 1990 Clean Air Act and implementation of new environmental regulations related to the monitoring and reporting of volatile organic compound (VOC) and volatile hazardous air pollutant (VHAP) emissions have made sealless pumps an economically attractive alternative to conventional mechanically sealed pumps. This has led to the development of vertical-in-line API canned motor pumps. The significant feature of the canned motor pump and the magnetic coupling designs is the absence of a dynamic mechanical shaft seal to contain the pumped fluid. With sealless pumps, a containment shroud or liner isolates the process fluid from the environment. In both sealless designs the shroud contains the main rotating elements (impeller, inducer and rotor shaft) and magnetic lines of flux transmit rotational power through the shroud to the shaft.

T

toxic, flammable or corrosive services where leakage to the environment is regulated or exposure is hazardous or lethal. These can include methanol production, caustic circulation and acid alkylation processes. The alkylation process is a typical API hazardous application that bears more detailed discussion.

Description of a Typical Alkylation Process Alkylation is a process used in petroleum refineries to produce gasoline blend stocks with octanes between 95 and 100. The feed stocks are light olefins and iso-butane. The olefins (unsaturated hydrocarbons— mostly butene with some propene) react with iso-butane to form isooctane and other similar high-octane components used in gasoline.

Typical Hazardous Application The API market is beginning to use more sealless pumps in volatile,

260

Figure 1. The typical alkyation process The Pump Handbook Series

There are two major process variations: 1) the HF alkylation process that uses anhydrous hydrofluoric acid as the catalyst and 2) the sulfuric acid alkylation process that uses sulfuric acid as the catalyst. While both processes are hazardous, the sulfuric acid process uses 15-19 lb. of acid per barrel of alkylate, whereas the HF acid process uses only 0.1 lb. (Ref. 1). Less acid, of course, reduces the hazardous content of the process. Figure 1 depicts a typical HF alkylation process showing representative pump locations.

Hazards There are several hazards inherent in the alkylation processes. Physical contact with either HF or sulfuric acid can cause severe burning. HF acid can be particularly haz-

Figure 2. Centrifugal pump with mechanical seals for API services (Sundyne LMV-311)

ardous. If HF acid leaks to the atmosphere, it vaporizes and forms a cloud. Contact with the cloud can cause burning, and inhalation can be lethal. For these reasons, it is critical that potential leakage points such as gasketing and mechanical seals be eliminated or kept to a minimum. Unique Pumping and Piping Requirements Because of the hazardous nature of the process, pump and piping requirements are unique and focus on eliminating leakage. In HF alkylation, the feedstock streams (olefins and iso-butane) do not contain HF and standard cast steel construction is satisfactory. Streams containing HF require special metallurgy for all wetted parts. Internal parts such as impellers, inducers, diffusers, shafts and shaft sleeves are some grade of Monel. Bolting is K-500 Monel. The user may specify carbon steel casings and seal housings, but all contact and sealing surfaces are Monel clad.Flange gaskets are typically spiral wound Monel with Teflon‚ fill. Oring seals are Litharge Cured Viton.

Centrifugal Pump Selection and Design The first factor in the selection of the centrifugal pump is the required duty point. Conventional mechanically sealed pumps can be used without much regard for head and flow limitations. Sealless pumps, on the other hand, are typically single suction, single stage and operate at synchronous speeds, which usually limit the application to 2000 gpm and 750 feet TDH. (Higher heads can be achieved with variable frequency drive technology.) The second factor is the process fluid. In addition to metallurgical material compatibility, careful consideration must be given to basic fluid properties such as density, temperature, viscosity, specific heat and vapor pressure. Technology Alternatives Pump shaft sealing technology generally falls into three distinctive categories: double mechanical seals, magnetic drives and canned motors. Double Seals For many years users have speciThe Pump Handbook Series

fied conventional centrifugal pumps with mechanical seals in API services (Figure 2). However, since mechanical seals require the passage of some fluid across the seal face for heat transfer and lubrication, a double seal configuration with a pressurized buffer system is necessary. With proper design, instrumentation and operation, this system can be quite satisfactory. However, loss of buffer fluid or pressure, mechanical seal failure, or loss of pump shaft control results in leakage to the environment. Recent developments in gas seal technology have led to their increased use for volatile fluids. Gas seals are being applied in double configurations with a safe gas buffer such as nitrogen, or as a backup to a liquid seal where the vaporized fluid vents to a flare or to a vapor recovery unit. Because of its externally actuated shaft, this type of pump configuration is subject to VOC and VHAP emissions monitoring and reporting. Proper application must give special consideration to seal material selection. Face seals with alpha sintered silicon carbide stationary faces and tungsten carbide rotating faces with nickel binders have been very successful. Magnetic Drives Magnetic coupling drive pumps are available in end suction, top discharge horizontal configurations and can provide positive single containment of hazardous fluids. Figure 3 shows a typical magnetic drive for API applications. The manufacturer can design the casing to withstand API nozzle forces and moments, and the baseplate can meet API requirements. Slightly higher than synchronous speed heads are achievable with variable frequency drive technology, however, the magnetic coupling torque rating decreases proportionally with speed increase. The potential for magnetic de-coupling significantly increases because pump power requirements rise faster than motor and coupling output do. Canned Motors Canned motor pumps with casings designed per API specifications provide the most reliable method of preventing hazardous escape to the environment. The process fluid circulates through the canned motor

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Single Enclosed Casing Gasket

Mechanically Retained Outer Magnet Assembly

Fully Enclosed Impeller

Non-Sparking Bump Strip

High Capacity Radial and Thrust Bearings

Impeller/Casing Wear Rings Front & Back

Optional Flanged Casing Drain

Provision for Optional Secondary Containment Systems Large Recirculation Path

Rugged One-Piece Containment Shell

Oil Lubricated Power Frame

FULLY CENTERLINE SUPPORTED

Figure 3. Magnetic coupling drive pumps (Kontro GSA-Frame 1 & 2)

Figure 4. Vertical in-line canned motor pump (Sundyne VIP-801)

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The Pump Handbook Series

to provide lubrication for the bearings and cooling for the motor. Hermetically sealed liners (generally of Hastelloy® construction) isolate the rotor core and stator windings from the process fluid. The manufacturer designs the outer stator band and electric terminal connections, which provide true secondary containment per the ASME Section VIII, Division 2, Pressure Vessel Code (including API corrosion allowance). Figure 4 shows a crosssection of an in-line API canned motor pump. As stated in previous articles in Pumps and Systems Magazine, the bearing material of choice for HF service in canned motor pumps is 100% alpha sintered silicon carbide. The bearing support system is as important as the bearings. The silicon carbide bearings are extremely hard and operate on a hydrodynamic film. A radial load applied to the rotor at the impeller tilts the rotor. If the bearing housings secure the silicon carbide sleeve bearings firmly, the tilt of the rotor will cause point loading of the bearing and loss of the hydrodynamic film. Extreme heating will occur, leading to bearing failure. One patented bearing support system enables the bearing to articulate as unbalanced hydraulic forces apply an impeller radial load. Since the direction of the radial load varies with the flowrate of the pump, the housing must allow the bearing to articulate around the full circumference of the bearing. This articulation enables the bearings to align themselves to the rotor’s axis of rotation. Similarly, with a vertical configuration, the entire weight of the rotor, in addition to hydraulic axial forces, must be supported by a thrust bearing system. Again, 100% alpha sintered silicon carbide is the material of choice. As with the radial bearings, the thrust bearing must be able to align itself with the rotor thrust runner. In one vertical in-line design, the thrust bearing rests on a metallic support plate, which in turn rests on a tilt washer, providing a gimbaling effect. Canned motor pumps offer the added benefit of reduced noise levels. This is primarily due to the elimination of motor fan and gear related noise. Figure 5 shows the sound pressure levels (dB) of two sizes of vertical-in-line canned motor pumps.

Canned Motor Pump Design for Unique Applications In some applications it is desirable to keep the process fluid out of the canned motor. This was true for a recent application in the Phillips Petroleum Company’s Reduced Volatility Alkylation Process (ReVAP), where the pumpage of HF acid, acid soluble oils (ASO) and additive could have clogged the close tolerance fluid passages within the motor. The Phillips Petroleum Company’s licensed proprietary ReVAP process uses a modified hydrofluoric (HF) acid catalyst system that, when used in existing HF alkylation units, makes it possible to reduce the airborne HF by as much as 60-90% in the event of an accidental release. The three principal advantages of the ReVAP process are its use with existing HF alkylation units, its environmental friendliness and its relatively low cost. Phillips and Mobil Oil Corp. jointly developed the process and installed it in late 1997 at the Phillips Woods Cross, Utah, facility and Mobil’s Torrance, California refinery. The Woods Cross installation is a skid mounted packaged unit where space is at a premium. The customer selected the Sundyne‚ VIP-801 vertical in-line canned motor pump with the special HF compatible metallurgy defined above, and with design modifications to isolate the canned motor from the pump. More information about the Phillips Petroleum Company’s ReVAP process can be found on their website at www.phillips66.com (Ref. 2). Several modifications to a standard canned motor were required for this particular hazardous application. These modifications included the addition of a mechanical barrier seal and replacement of hydrodynamic bearings with hybrid rolling element bearings. Mechanical Barrier Seals Options for the isolation of the canned motor from the pump included either a close clearance throttle bushing or a mechanical face seal. Throttle bushings require a relatively large amount of flush fluid to effectively isolate the process fluid from the canned motor. In contrast, mechanical barrier seals provide excellent isolation with very little buffer fluid leakage. In this case the user selected the

Figure 5. Vertical in-line canned motor pump sound levels

Figure 6. Vertical in-line canned motor pump with slurry circulation system The Pump Handbook Series

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mechanical barrier seal as the most desirable configuration because it posed the least possible contamination of the process fluid. In addition, the manufacturer developed a new seal with a balance of hydraulic and spring forces that resulted in the ability to seal 150 psi differential in either direction. The primary seal face leakage control from the motor to the pump with secondary seal leakage control from the pump to the motor. The mechanical seal mounts in the diffuser cover, and a rolling element bearing system accommodates it without lengthening the motor shaft. A pair of angular contact ball bearings is face-to-face in place of the lower sleeve bearing. A single deep groove ball bearing replaces the upper sleeve bearing. These rolling element bearings provide the required axial and radial load support and eliminate the need for more complex bearing housings. Since the bearings are fully submerged, hybrid bearings consisting of standard tool steel races, composite cages and silicon nitride balls were selected. The hybrid bearings provide extended life, tolerance to marginal buffer fluid conditions and reduced heat generation. Buffer System An API Plan 53 pressurized buffer system performs the dual function of cooling the motor and lubricating the bearings. A low viscosity, product compatible mineral oil circulated by an auxiliary impeller within the motor buffers the canned motor and seal from the hazardous process fluid. The API Plan 53 buffer system materials must be compatible with the hazardous fluid. For this application, the cooling coil, interconnecting tubing, valve fittings and wetted instrumentation components are Monel. A hand pump provides pressurized buffer fluid replenishment while the unit is operating. Pressure and liquid level alarms and a pressure regulator for plant nitrogen are important auxiliary components typically specified for this type of critical application. Containment The canned motor configuration provides the benefit of dual containment. In the event of catastrophic seal failure, loss of buffer fluid or buffer pressure, HF laden fluid will be contained in the buffer system

264

and won’t escape to the environment. If the motor liner fails, the outer stator band provides the secondary containment.

Application Guidelines When applying vertical in-line canned motor pumps in hazardous services, there are several critical areas that both the user and manufacturer must consider. (Ref. 3) • Performance envelope - Will the required service duty fit within the performance range of a canned motor pump? Consider head, flow, suction pressure, temperature and power requirements. • Circulation system - Will the fluid’s properties permit its use in the canned motor? Consider cleanliness, viscosity, specific heat and the vapor pressure curve. • Materials - Does the hazardous fluid require special metallurgy and gasketing requirements? Consider not only the pump components but also the motor components. Reaction to the materials, as well as processes such as welded joints and heat-affected zones at elevated temperatures, must be evaluated. • Costs - The user should evaluate initial purchase price, predicted maintenance costs and regulatory compliance costs such as VOC and VHAP emissions monitoring and reporting.

Field Experience The vertical in-line canned motor pump is currently in service around the world in a variety of hazardous, lethal and environmentally sensitive applications. A partial list includes: HF Alkylation Sulfuric Acid Alkylation Benzene Naphtha Sour Water Butane Iso-butane Methanol Butylene Toluene Caustic Methyl Mercaptan Propylene Alkylate Kerosene Kerosene Disulfide

The Pump Handbook Series

Concluding Remarks Successful, safe handling of hazardous fluids presents some unique requirements for both the user and the pump supplier. The most important requirement is a complete understanding of the pumping system requirements and fluid properties. Once the decision has been made to apply and specify a sealless pump, the type of pumping system must be selected, i.e., mechanically sealed, canned motor or magnetic coupling drive. As with any type of pump, you must make a careful evaluation and selection of product compatible materials throughout the entire operating range of pressures and temperatures. The process fluid must be evaluated for its suitability for motor or magnetic coupling cooling as well as mechanical seal and bearing lubrication. The thermodynamic properties of the fluid must be evaluated against the expected thermal and pressure gradients within the sealless pump and motor or magnetic coupling throughout the pump’s entire operating range. Finally, the proper fluid circulation system must be selected to assure material compatibility, fluid thermodynamic stability and system reliability. The user and the supplier must work together in the application and selection of the pump, materials and circulation system to ensure the safe handling of the hazardous fluid. "

References 1. Hydrocarbon Processing, Vol. 75, No. 11, November 1996, pp. 91-93. 2. Phillips Petroleum Company news releases, http://www.phillips66. c o m / n ew s r o o m / r e l 1 7 0 . h t m l , 04/24/98 and http://www.phillips66. com/newsroom/rel1201.html, 11/23/98. 3. Carr, Dave. “Evaluating Sealless Centrifugal Pump Design & Performance.” Pumps and Systems Magazine, August 1997, p. 14. Ron Forsberg is a Senior Project Engineer with Sundyne Corporation in Arvada, Colorado. He has been with the company since 1973 and has held various positions within the quality assurance, maintenance, product engineering and product development departments. He is a graduate of the U.S. Merchant Marine Academy at Kings Point, NY with a B.S. degree in marine engineering.

CENTRIFUGAL PUMPS HANDBOOK

Bearing Reliability in Centrifugal Pumps Bearing handling, installation, monitoring and care require a significant degree of maintenance expertise and warrant a high priority by management. By Dave Mikalonis, SKF USA Inc.

earings play a crucial role in the longevity of centrifugal pumps. Choose the wrong bearings, install them improperly, or fail to maintain them, and the consequences could be costly. Unscheduled downtime costs a lot more than planned maintenance (generally a factor of 10X), and can lead to huge costs if a production run is interrupted. The most common causes of bearing failure include misalignment, poor lubrication, contamination, excessive loading, cavitation and skidding in unloaded bearings. If properly cared for from the time of purchase to its installation and use in an application, a bearing will often outlast the expected design life.

B

Ball Bearing Types Optimizing bearing life begins with the selection of the proper bearing for a given application. A variety of ball bearing types are available for centrifugal pumps. The most commonly used are: deep groove ball bearings, double row angular contact ball bearings and single row angular contact ball bearings. Deep Groove Ball Bearings Single row deep groove ball bearings, which have greater than Normal (C3) internal clearance, are generally used in centrifugal pumps for support of radial loads (see A in Figure 1). The greater initial clearance allows for radial expansion of the bearing inner ring in pumps that move high temperature fluids, or

when bearings are mounted on stainless steel shafts and a temperature differential exists between ambient and operating. Some stainless steels have coefficients of thermal expansion that are 40-50% higher than most carbon and alloy steels. This difference can reduce the internal clearance in the Normal bearing, leading to excessive internal bearing friction. The use of the greater C3 clearance enables the use of the standard ISO k5 shaft tolerance—mean interference of 0.0125 mm (0.0005 in.)—regardless of the material. Double Row Angular Contact Ball Bearings Double row angular contact ball bearings are used to support a combination of radial and axial loads in chemical process, double suction and submersible pumps. They con-

trol the position of the pump shaft, impeller and mechanical seal. Both Conrad (uninterrupted raceway shoulders) and Filling Slot configurations are available. The Conrad type is recommended for centrifugal pump applications because experience has shown that they tend to operate at a lower temperature. Determine internal axial clearance for this type of bearing using the following guidelines: 1. Use industry-designated “Normal” internal axial clearance unless the pump shaft speed is more than 75% of the bearing speed rating in oil, or the shaft temperature is higher than 80ºC (176ºF) due to heat conducted from the pump. Contact the manufacturer or consult the manufacturer’s catalog for speed ratings.

Figure 1. Deep groove ball bearings are used to support radial loads (FrA). The Pump Handbook Series

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2. Use the ISO k5 shaft and H6 housing tolerances. In general, the use of Normal clearance rather than C3 with this type of bearing helps control the motion of the unloaded ball set and reduces internal wear. Open bearings (i.e., those without shields) are recommended as long as you can re-lubricate them and protect them from solid particle contamination. Bearings with a single shield can be used for protection from contamination. In grease-lubricated applications, the shield is oriented in the housing so that grease is supplied between the bearing shield and the shaft seal. In centrifugal pump applications with high axial loads, tests have shown that bearings with a contact angle of 30º operate at lower temperatures than bearings with a lower contact angle. Single Row Angular Contact Ball Bearings Single row angular contact ball bearings feature high radial and axial load capacity as well as a high speed rating. They can support pure thrust load or combined radial and axial loads when paired in back-to-back or face-to-face arrangements. Back-toback mounting provides superior shaft rigidity and is most common. Face-to-face mounting arrangements are sometimes used when operational shaft deflection is inherent. They typically operate with a small axial clearance or a light preload for good positioning accuracy. Preload can increase the fatigue life of a bearing by improving internal distribution of the applied external loads. Too large a preload, however, can reduce bearing life (Figure 2). In centrifugal pumps, preload is principally used in balanced thrust load applications to avoid light load skidding due to axial deflection in angular contact ball bearings and to provide shaft positioning accuracy necessary for mechanical seals. Preloaded bearings are more sensitive to misalignment and incorrect mounting than bearings with clearance. For this reason, single row angular contact ball bearings with a small axial clearance are recommended for centrifugal pump applications where specific operating conditions are

266

only generally known and other clearance/preload options cannot be satisfactorily evaluated. Universally matchable single row angular contact ball bearings can be arranged as pairs to support loading in either of the axial directions. One manufacturer has introduced a 40º contact angle ball bearing paired with a 15º contact angle ball bearing together in a set called PumPac®. The 40º bearing is mounted to support the applied axial load; the 15º bearing supports momentary reversing loads and radial load. Tests show the set operates at cooler temperatures than a same-size pair of 40º contact angle ball bearings because of reduced skidding in the inactive 15º bearing. This in turn lowers the temperature of the lubricant, enabling operation at higher speeds or longer bearing service life. The PumPac® set should be used when the pump axial load acts predominately in one axial direction such as in single-stage, end suction centrifugal pumps. Several cage material options are available for single and double row bearings, including polyamide, steel and machined brass. The polyamide cage provides quieter operation, but it will not last as long if the lubrication is poorly maintained or temperatures are excessive 120°C (248°F). The quiet operation, which continues even if there is a bearing problem, means that users might have insufficient warning of impending bearing failure. Metallic cages can help ensure the highest reliability and are recommended for centrifugal pump applications.

Inspection While it is important to select the proper type of bearing, it is equally essential to choose a bearing supplier that will provide a quality product both in design and manufacture. Most major bearing manufacturers do so and ensure this by noise-testing and measuring 100% of their bearings during the manufacturing process. Most users, in turn, establish a vendor qualification program to verify that the manufacturers supply high quality products. Once the quality practices of the bearing supplier are verified, there is no need to inspect each incoming bearing. In fact, such a step can have an adverse effect on reliability The Pump Handbook Series

LIFE

PRELOAD

CLEARANCE

Figure 2. Preloading can increase bearing fatigue life by improving internal distribution of the applied loads.

because it introduces the possibility of contamination or damage. Instead, the most crucial step for incoming product is to verify that the right part number has been received. A wide variety of bearing types exist for different applications, so smart users will make sure that they get the right one.

If the bearings are not going to be used immediately, they should be stored in a vibration-free area where humidity and temperature are reasonably constant. Bearings not stored in their original packaging should be protected against dirt and corrosion.

Installation Improper installation is much more likely to cause problems than poor product quality. Two basic methods are available for installing bearings: cold mounting and hot mounting. Cold mounting is recommended only for units with an outer diameter of 4” or less. Use a press whenever possible. Lubricate the shaft and place a mounting sleeve between the bearing and the press, resting it on the ring with the interference fit. Make sure that the end faces are flat, parallel and free of burrs. A hydraulic press can be safer and more efficient than a mechanical one. Hammer blows to a sleeve can be used to mount a small bearing, but avoid soft-headed hammers—these can leave fragments behind. Never directly strike a bearing with a hammer. In hot mounting, the bearing is heated to at least 79ºC (175ºF) more than the seating so that the differen-

Figure 3. Condition monitoring provides early warning of impending bearing failure.

tial expansion is sufficient to reduce the force required to mount the bearing. This method is recommended for larger bearings since the force required to mount a bearing increases rapidly with size. The bearing should never be heated to more than 121ºC (250ºF), however, as this can cause dimensional changes in the bearing. Monitor the temperature carefully and wear clean, protective gloves when handling a hot bearing. If a bearing requires an interference fit in the housing, a smaller temperature increase, about 21-49ºC (70120ºF), is usually enough. Optional tools for heating a bearing include hot oil baths, heating cabinets, heating rings and induction heaters. The preferred method is induction heating, which creates heat by inducing electric currents. Although it requires expensive equipment, induction heating is the fastest and cleanest of the alternatives. Most plants today use the induction heating method, but a note of caution is necessary. Induction heating requires that the bearing be magnetized. After the bearing has been heated, it must be demagnetized. The newer induction heaters demagnetize bearings automatically; older equipment does not, and users must make sure to take the extra step of demagnetizing. Oil baths provide even heating

and make it possible to maintain a bearing at a certain temperature, but they introduce more opportunity for contamination. Cleanliness is also crucial during the installation procedure. Keep bearings in their protective packages until you are ready to mount them. Cover all equipment with plastic, waxed paper or a clean dry cloth. These precautions will protect exposed components during the process. Hands and tools that will contact the bearings must also be clean. If possible, move the machine (or the part of the machine to be fitted with new bearings) to the workshop or an area where it will not be exposed to airborne debris such as metal particles, sawdust, sand, cement or corrosive substances. (Bearing tolerances are in tens of microns or 1⁄5 the thickness of a human hair, even smoke particles can be detrimental.) Shafts, housings and other bearing arrangement parts need to be thoroughly cleaned and dried. The bore and outside surfaces of the new bearings should be cleaned with a lint-free cloth. Proper alignment is also crucial— one study found that more than 50% of rotating machinery failures can be traced to shaft misalignment. Modern flexible couplings enable more misalignment in connecting shafts, but sloppy alignment still The Pump Handbook Series

reduces bearing life and should not be tolerated. Various shaft alignment equipment is available from some bearing manufacturers, ranging from easyto-use electromechanical systems to advanced laser alignment systems. An increasing number of companies are opting for the laser systems as they strive to realize longer life from their equipment. As part of the alignment process, the installer should inspect for soft foot and pipe strain. Soft foot occurs when a machine is not resting equally on all supports. A dial indicator applied to the top of the foot will let you know if there is a problem. Loosen the bolt and check the movement on the indicator—if it is greater than 0.051 mm (0.0020”), you’ve got a problem. Pipe strain in the suction and discharge piping occurs when force is used to connect two flanges that are several inches offset from each other. The resulting torque will introduce stresses into the machinery. A visual inspection will tell whether the piping needs to be refabricated. Another key step before installing new bearings is checking assembly drawings for tolerances, fits and internal clearances for the shaft and housing. If the shaft and housing are not to tolerance, the bearing will not operate as designed and can fail prematurely. (The bearing will either ride too loosely on the shaft or the fit will be so tight that the intended clearances will disappear and the bearings can run hot and quickly fail.) If the dimensions are not correct, the shaft will have to be replaced or modified, or the housing will have to be bored out and a sleeve added to achieve the right dimensions.

Housing Seals A bearing housing seal can significantly influence pump friction and operating temperature. Tests of a horizontal centrifugal pump revealed that shaft temperature was reduced 38ºC (100ºF) when labyrinth shaft seals were used instead of the original radial lip seals. This improved the operation of the mechanical seal, increased the useful life of the lubricant and provided long term effective shaft sealing that excluded

267

Figure 4. Lubrication requirements depend on the bearing size, operating temperature and speed.

contamination. Labyrinth shaft seals, OEM made or commercial types, offer frictionless sealing and effective exclusion of solid and liquid contamination.

Condition Monitoring Once the bearings are installed, a condition monitoring program will help you get the most out of them (Figure 3). The purpose of condition monitoring is to provide better information for better decisionmaking and to provide early warning of an impending problem. This can help in two areas: optimized timing of scheduled maintenance and improved planning for predictive maintenance. Scheduled Maintenance A pump manufacturer may say that a pump should be overhauled every 18 months, but that amount of time includes a margin of safety. By monitoring the condition of the

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pump, you might be able to keep it going several months longer, perhaps as long as 27 months, so that in 54 months the pump will be overhauled two times instead of three. At an average of $5,000 to $8,000 per pump rebuild, the savings are significant. Predictive Maintenance For predictive maintenance, the savings can be even more dramatic. Condition monitoring can tell you if a failure is likely to occur before scheduled maintenance. One rule of thumb illustrates the possible savings—if a machine has to be fixed after a sudden failure, it will cost ten times more than fixing it as part of a scheduled maintenance. The extra cost comes from factors such as lost production time, overtime pay, emergency ordering of required parts and other expenses that could be avoided if the failure was anticipated and prevented. If the failure stops a production run at a critical The Pump Handbook Series

juncture, it can cause a loss of workin-process, and the expense will be even greater. The main indicator for condition monitoring is vibration. A machine will exhibit a particular vibration signature depending on its condition. The frequency of the vibration indicates the type of problem; the amplitude tells how bad the problem is. If the machinery is out of balance, the vibration signals will be at the rate of the shaft speed. Misalignment generates a signal that is twice the shaft speed. Looseness leads to signals corresponding to the harmonics of the shaft speed. By following the trends of the various signals, you can predict when a failure is likely to occur and take steps to avoid it. One of the key vibration indicators is the acoustic noise generated by metal-to-metal contact. Such contact occurs when the lubricant film in a bearing assembly breaks down, which can happen for a variety of reasons, including damage to the bearing raceways. Contamination, both solid and liquid, in the lubricant can also lead to metal-to-metal contact. These signals can be easily detected with one of the newer, more effective condition monitoring techniques called Spectral Emitted Energy (SEE) measuring. A SEE device takes high-frequency acoustic measurements and combines them with enveloping techniques to diagnose bearing condition. In enveloping, the vibration signal is filtered to leave only the high frequencies, which takes out most of the vibrations caused by structural vibration, misalignment and other factors. The remaining defect signals are reduced in energy content, but still occur at the same intervals. High frequency vibrations are generated each time a bearing defect is overrolled. SEE is replacing shock pulse testing at most sites. Shock pulse testing measures overall amplitude of vibrations, but it does not provide enough information to analyze the causes in detail. Condition monitoring may be carried out periodically with portable equipment or continuously with hardwired equipment dedicated to a particular machine. The second method is recommended for continuous running machinery in critical processes, and it often includes the

Lubrication Proper lubrication plays a key role in prolonging bearing life. Oil or grease are typically used to separate the rolling elements and raceway contact surfaces, lubricate the sliding surfaces within the bearings, and provide corrosion protection and cooling. Figure 4 shows the recommended minimum viscosity at operating temperature for lubricants used in centrifugal pumps. Mineral oils, synthetics and, less commonly, grease, are used for lubrication. The viscosity of synthetic oils is less sensitive to temperature changes, so synthetics are most often used in applications where temperatures vary widely. In general, avoid synthetics that contain solid additives. If temperatures will exceed 100ºC (212ºF), synthetics should be used because mineral oils have a short life and will quickly begin to oxidize at those temperatures. Mineral oils used at elevated temperatures should be changed every three months. Longer intervals between replacements are possible with lower operating temperatures. Frequent oil changes contribute to long bearing service life; however, the recommended frequency depends on the specific oil conditions. In any case, lubricant should be changed when the following levels are exceeded: • solid contaminants: 0.2% • water: 0.002 to 0.2% • acidity: maximum 1 unit increase (in mineral oils) If κ > 4, use κ = 4 curve As the value of ηc (Pu/P) tends to zero, aSKF tends to 0.1 for all values of κ Figure 5. Adjustment factor for predicting bearing load life

ability to automatically shut down the process when a particular level of vibration is exceeded. Portable equipment is available that gathers vibration and temperature data, provides alerts when conditions exceed programmed limits and shares data with other programs. Temperature can also be monitored as part of a program, but vibration analysis, when done properly, is much more efficient and can give a lead time of three to four months in predicting problems. By the time a temperature signal indi-

cates a problem, it might be too late. The condition of the lubricating oil can also be monitored. This usually involves sending an oil sample to a laboratory that analyzes water and contamination levels. It is generally recommended that water levels be kept below 200 ppm. In high humidity areas or for processes that involve a lot of moisture, water filters are recommended. The type of contaminants found, and their shape and size, are also indicators of bearing condition.

The Pump Handbook Series

The method of lubrication also impacts bearing life. Options include: oil bath, oil-ring, oil mist, and grease. Oil mist is becoming the most commonly recommended method. In oil mist lubrication, the oil is atomized into tiny droplets that are conveyed by compressed air. The mist is produced by a mist generator and is pressurized to 1 or 2 psi above the ambient pressure. Oil mist brings a constant flow of clean, cool lubricant to the bearings, and the slight pressurization serves to keep contaminants out of the bearing housing. The mist, which represents a very small amount of oil, can be left on when the pump is idle. This provides maximum protection from contamination and condensation.

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A small vent opposite the point where the mist enters the housing enables the mist to flow freely. Directed oil mist is recommended if the bearing ndm value is greater than 300,000 and if the bearing supports a high axial load. Synthetic or special de-waxed oils are often used for oil-mist lubrication because paraffins in standard oils can clog the oil mist fittings. Oil mist can be combined with an oil bath in a configuration known as purge oil mist. But pure oil mist, without the bath, is preferred. It has been shown to significantly improve bearing life. The system, however, must be safeguarded with alarms to avoid bearing failure in the event that the mist system fails. To ensure adequate initial lubrication, the bearings should be prelubricated with oil or connected to the mist for an extended period of time before the pump is started. The explanation for the effectiveness of oil mist lubrication lies in the nature of the forces that generate friction in a bearing. In pump applications, the most dominant friction is from the load-independent friction due to the high speed motion of the balls moving through the lubricant. The more balls and lubricant in the bearing, the greater the friction. Oil-mist, with its extremely low quantities of oil, provides the lowest friction and the coolest operating temperatures.

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Oil mist lubrication requires a larger up front cost, so in operations with hundreds of pumps, companies are applying it gradually—as pumps are rebuilt the oil mist system is added. Sealing is also crucial to bearing life, and in fact the bearing isolator, introduced several years ago, facilitates the broader application of oil mist lubrication. The bearing isolator is an elaborate labyrinth seal that enables the pressurizing of the housing and prevents leakage of the oil mist, even when the shaft is idle.

Bearing Rating Life Many of the factors described above that impact bearing life can be accounted for in a new method for estimating bearing life being considered by ISO. The earlier ISO method was based on the Bearing Dynamic Load Rating and the magnitude of the applied load and the operating speed. A multiplier was added to account for the operating conditions of the bearing lubricant. The predictions made with the old method, however, did not always correlate with experience, especially for lightly loaded bearings and those operating in clean conditions. In those cases, service lives were significantly longer than predicted. The SKF Life Theory (adapted and now being reviewed by ISO) adjusts the calculation to account for bearings operating under low loads in clean environments with satisfactory

The Pump Handbook Series

lubrication and manufactured to accurate tolerances. It adds an adjustment factor based on a fatigue load limit, a contamination factor and a lubrication condition. The graph in Figure 5 shows the SKF factors for radial ball bearings.

Application at Your Facility Bearing handling, installation, monitoring and care require a significant degree of maintenance expertise and warrant a high priority by management. Comprehensive bearing maintenance training seminars, conducted at user locations, are available from some bearing manufacturers. The seminars cover all the topics mentioned in this article and more. Properly maintained bearings yield reliable operation and will generally outlive the pumps in which they are installed. Think of the time needed to train technicians in bearing maintenance as a critical investment in plant operating efficiency. " David R. Mikalonis is the Global Engineering Manager-Fluid Machinery Segment for SKF USA Inc., Kulpsville, PA. Prior to this he worked as an SKF Applications Engineer providing application recommendations bearing selection, and failure analysis to customers in the petrochemical and electric motor industries. Mr. Mikalonis received a B.S.M.E. degree from Drexel University in 1989.

CENTRIFUGAL PUMPS HANDBOOK

Magnetic Drive Pumps Without the Hype The attraction of these pumps is due to more than the powerful pull of their rare earth magnets. By Craig J. Bailey, ITT Industries/Goulds Pumps

T

Magnetic Drive Pump Construction The primary components that induce flow through a centrifugal pump, the impeller and casing, are generally not affected by the use of sealless drive devices. The main impact on the pump package is the added inefficiency of the drive mechanism itself. This is manifested in two ways. The first is an increased horsepower requirement that adds amperage draw and can require an increase in motor size. The second is that the inefficiency has to go somewhere and, guess what, the process fluid acts to

remove the generated heat. Fluid temperatures will typically be higher at the discharge of these pumps than for comparable sealed pumps. This should be considered, since in many cases where sealless/zero leakage is desired, the fluids or processes are sensitive to temperature increases. Magnetic drive devices eliminate the use of rotary type sealing including packing and mechanical seals. The following construction details should illustrate how this works. Magnetic Coupling Before I get into the specific discussions of construction of the magnetic coupling, I would like to offer some general comments on the magnets used in these types of pumps. They are generally one of the following three types: aluminum nickel cobalt (AlNiCo), neodymium iron boron (NeFeBo) or samarium cobalt (SmCo). AlNiCo is the weakest of the three in magnetic strength, which is measured in gausses (G). The NeFeBo and SmCo types, also known as “rare earth” magnets, are much stronger and therefore a system will require either fewer of them or less total volume to achieve the same horsepower ratings. Figure 1 shows relative strength, by volume, of the magnet materials. The one advantage of the AlNiCo magnet has been its resistance to the degradation that is often caused by exposure to high temperatures. This, combined with its long-time use in eddy current couplings, has made it The Pump Handbook Series

attractive in hot oil services, which have temperatures as high as 800°F. The magnetic strength of the rare earth materials has led to everincreasing power in magnetic coupling designs. Several manufacturers offer pumps in the hundreds of horsepower range. Now, on to how these magnets are used. The magnetic coupling is just that, a coupling. It enables transmission of power from the driver— whether an electric motor, a steam turbine or an engine—to the impeller. It is not the motive power. So we take the drive magnet and connect it to a driver. This can be accomplished by directly attaching it to the motor or using a conven-

Today’s Magnets 1 0.9 0.8 0.7 Relative Strength

o know what magnetic drive pumps can mean to you and your process needs, it is important to understand their method of operation. As you read this article, note the references to design criteria. Mag drive pumps can handle many applications, even though you may be restricted in which particular design you can use. These pumps can offer reliable, safe, zero leakage operation. Magnetic drive designs are classified as centrifugal process pumps with magnetic couplings that eliminate the sealing requirement associated with a rotating drive shaft. The category of true “sealless” pumps includes many more designs. Most of these are positive displacement and include diaphragm (flat or tubular; single or double), peristaltic (“hose”), gear and screw designs.

0.6 0.5 0.4 0.3 0.2 0.1 0

AlNiCo

SmCo

NeFeBo

Magnet Materials

Figure 1. Magnet Comparison

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tional shaft coupling. The drive magnet (sometimes referred to as the outer magnet) is a cylinder holding the magnets on its inside surface parallel to the rotating axis. The magnets are attached to the cylinder either with epoxies or a combination of epoxy and mechanical bonding. This cylinder rotates around the same axis as the impeller. Of particular importance is the balance of these sometimes large, heavy components. Proper balancing of the coupling minimizes movement that could jeopardize clearance from the adjoining stationary containment shell. Similarly, the support bearings for the drive magnet, whether they are motor or power frame bearings, must be properly designed to provide for long, trouble-free operation. Of added significance is the fact that the coupling is outside the process liquid and within the frame of the pump body. This enables the use of less costly alloys in its manufacturing. The driven or inner magnet is constructed much the same way (Figure 2). It too is cylindrical with a smaller diameter that enables it to fit inside the drive magnet. The driven magnets are positioned on the outer circumference and matched to a drive magnet by polarity. This generates a magnetic flux field of attraction between the two, which causes the driven magnet to follow the drive magnet in a synchronous fashion. The two rotate at the same speed and in the same direction. Using stronger magnets or just more of the “weaker” ones can alter the strength of the magnetic field, and therefore the coupling. Increasing the number of magnets can be accomplished with proportional increases in cylinder diameter, enabling the placement of more magnets around the circumference. Another method is to Containment Shell Drive Magnet

Shaft

Driven Magnet

Figure 2. Driven magnet construction details

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lengthen the cylinder and add rows or “banks” of magnets. In each case, physical dimensions and mechanical forces become limiting factors. The largest thief of power is the distance between the magnets, followed closely by the conductivity of any material that is passed between them. Magnetic flux is greatest when the magnet bars are touching. As they are pulled apart, the magnetic field strength reduces proportionally to the distance. This demonstrates the importance of optimum design of the drive and driven magnet components, the support structure of each, and the containment shell thickness and material of construction. The driven magnet acts directly upon the impeller either through a shaft (typically) or through a bearing carrier. It is immersed in the process liquid and therefore must be constructed similarly to the rest of the wetted components. The driven magnet core is encapsulated in a thin sheath that offers protection while minimizing the negative impact of magnet separation. The sheath is usually stainless steel as a minimum, but it should always offer chemical resistance equal to or greater than the casing and impeller. Note that because this sheath rotates with the magnet coupling there is no associated disruption of the magnetic flux field. Containment Shell Now we have power going to a drive magnet that is surrounding a driven magnet that will follow its movement. However, remember that we are constructing a “sealless” zero leakage pump; the design must keep the liquid surrounding the driven magnet isolated from the environment. This is accomplished using a containment shell or isolation can. This “can” is best described as a dome shaped object with a cylinder diameter somewhere between that of the drive and driven magnets. It acts with the casing to hold the process liquid, but it permits close proximity of the magnets. To hold the process liquid, the containment shell must be compatible with the process, chemically resistant, and have adequate mechanical integrity to withstand the full range of pump pressures. While offering these features, it must also be as invisible as possible to the magnet The Pump Handbook Series

coupling. This enables maximum transmission of power between the drive and driven magnets. To review, we want thick for pressure containment but thin for allowing closeness of magnets and minimum disruption of magnetic flux. We want chemical resistance with minimum cost, but cannot allow for corrosion that would add extra magnetic flux disturbance. Add to these the concerns of resistance to direct impact or rubbing, which would breach the containment, and you can see why there is such diversity among containment shell materials. You should also note the significance of changes in these materials and constructions as the classification of pump changes (e.g., ANSI/ISO, API, light duty, general service, etc.). It is not unusual to see non-metallic containment shells on light duty and general service pumps. API pumps will have a metallic, robust design for higher pressure and temperature ratings. Some examples of material used in containment shells are stainless steel, Alloy 20, Hastelloys®‚ (B & C), ceramics (silicon carbide or zirconium oxide), and non-metallics (PEEK, PAEK (polyaryletherketone), TEFZEL/fiber reinforced vinyl ester resin, PTFE/carbon fiber reinforced plastic, and others). Keep in mind that these containment shells can be formed in as many ways as there are materials to construct them. Key to material choice and construction design is the severity of the duty for which the pump is intended. Two additional design elements of the containment shell are support for a rear bearing for the inner rotating assembly and “secondary containment.” With regard to the first element, bearing support, both nonmetallic and metal containment shells can be designed so that the inner rear bearing is positioned in a holder as part of the rear flange or dome of the containment shell. These designs are mentioned because they add design criteria to consider. If secondary containment is desired, some manufacturers offer dual or double containment shells. This provides a backup barrier in case the primary one leaks. In some setups, placing a monitoring device between these two shells can give

One Piece Brg. Holder Drive Magnet

Impeller

Shaft Driven Magnet Containment Shell

Figure 3. Inner rotor construction details

Impeller

Front Brg. Holder Drive Magnet Rear Brg. Holder

Shaft Driven Magnet Containment Shell

Figure 4. This arrangement uses both forward and rear bearing holders

early warning of a leak. There are some concerns with this configuration. You are still limited as to the total depth of these two shells, but they must provide minimum resistance to the magnetic flux. Because these shells are relatively thin, you must consider what the true protection of each is independently. If one is breached, will the other hold? Secondary containment can be offered in other manners on magnetic drive pumps. One such configuration has the power frame gasketed and a sealing mechanism added to the drive magnet shaft. This can be anything from a simple lip seal to a magnetic, non-contacting seal. Notice the loss of the “sealless” designation here. (The general consensus is that canned motor pumps offer a superior product if secondary containment is important.) Inner Rotor The purpose of a pump is to move liquid. To do this in centrifugal pumps, you must have an impeller— and it must rotate. This is accomplished by coupling it directly to the driven magnet either through a conventional shaft or a rotating bearing carrier. If a shaft is used, the rotor must be supported by bearings anchored to the pump casing/frame.

There are two designs using impeller, shaft and driven/inner magnet rotor support. The more common is a rigid single bearing carrier (Figure 3). This carrier is secured between the pump casing and the power frame/adapter frame. It usually contains one or two bearings. The general consensus is that these bearings can be more accurately aligned. And once aligned, the alignment is easy to maintain. The second practice is to use a forward bearing carrier to hold a single bearing and then support a second bearing aft of the driven/ inner magnet (Figure 4). This is done using a bearing support attached to the inside face of the rear portion of the containment shell. The thought here is that the often-substantial weight of the driven magnet more evenly distributes the bearing loads when it straddles the support. Two additional concerns of this design are the added load criteria to the critical containment shell and the alignment of the bearings (necessary to maximize their performance). Because the bearings are in two separate components, assembly procedures can be complicated. If a bearing carrier drives the impeller, the shaft or spindle is usually held stationary by the casing/ frame and the containment shell. This configuration is more generally used in light to medium duty applications with limited horsepower. In either of these situations, the interface of the rotating and stationary components, or the bearing surfaces, becomes the critical, dynamic state. Support Bearings In the vast majority of setups, this interface/bearing surface is handled with a hydrodynamic bearing design. The process liquid is used as a fluid film to support the sleeve bearings. This is important because it makes the process liquid properties critical to the selection of these pumps. Secondly, these properties can impact the choice of bearing materials. In general, bearing loads are so low that most liquids can generate sufficient film thickness to support them within the pump’s normal operating ranges. The most common bearing material is silicon carbide. It has excellent chemical resistance and its surfaces can be finished very smooth, which The Pump Handbook Series

reduces drag. The material easily supports the loads associated with the increasing sizes of magnet couplings and the pump hydraulics to which they are applied. An additional feature of the material is its hardness, which minimizes abrasive wear when handling unclear liquids. The second most popular bearing material is carbon. It has many of the same advantages as silicon carbide, but it is softer and will “wear” more rapidly during normal use. Any solids present in the liquid will accelerate this beyond carbon’s practical consideration. While normal wear can be anticipated and planned maintenance can preclude any damage, both these options necessitate more frequent shutdown and entry into processes which, as the use of a sealless pump suggests, are hazardous.

Critical Design Issues Magnetic drive pumps generally have three critical design issues: containment shell, bearings and lubrication paths. Containment Shell/Can What we have discussed are the required features of pressure containment, resistance to chemical attack by pumped fluid and minimum disruption of magnetic field. As you might expect, these can conflict with one another. One example would be hot oil services where the chemical resistance of carbon steel is more than adequate but the induced eddy currents generate enormous hysteresis losses, rendering magnetic drive pumps impractical. In another case, a non-metallic containment shell could offer excellent chemical resistance and superior eddy current resistance but not be able to handle expected pressure containment. Generally, if pressure containment becomes the most critical issue, as with secondary containment, you might want to consider a canned motor pump. By virtue of their design, they often become the preferred selection for these situations. Keep in mind that the can is supported by the considerable mass of the stator and housing. The method of constructing the containment shell in magnetic drive pumps is an often-debated issue. The two generally accepted choices

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Bearing

Shaft

Front Bearing

Rear Bearing

Figures 5 and 6. Misalignment can play a key role in the performance of bearings

are the welded tube section with an end cap or the hydro-formed onepiece dome. Central to the concerns are integrity of the weldments and uniformity of the cross section. (Important here is that you understand the design limits of each choice and that your supplier provides consistent high quality products.) I will leave the remainder of the discussion to your favorite mag drive pump manufacturer. Bearings Bearings are probably the most critical component of these pumps. If the bearings are operating properly (centering the rotating assembly in the brackets and supporting the normal ranges of dynamic loads) the pumps will be quiet, exhibit low vibration and, as experience shows, run for many years. However, if they wear out, are improperly aligned or have inadequate lubrication, the resultant failure can be severe. Wear associated with sleeve (journal) bearings is affected by the load acting on the bearing, the differences in hardness of the mating rotating/stationary components, the lubricity of the fluid film and the abrasive nature of the fluid. In general, loads on these bearings are well within the range of acceptable limits, even at extreme off-design performance. The more severe condition is run-out or high flow operation. When this happens, the loads are relatively high, but, of more significance, the pressure differential across the bearings is at its lowest.

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This jeopardizes the production of a film thick enough to support these higher loads. However, if properly selected, these pumps should rarely see inoperable bearing loads. Because there is such a close relationship between the hardness of mating bearing components and the lubricity of the process fluid, I will combine their discussion. Suffice it to say that the lubricity of the pumped fluid is the single biggest factor contributing to successful operation of these pumps. With adequate lubrication, the sleeve bearings will have a very low coefficient of friction. Heat generation due to friction will be minimal, thereby reducing any negative impact on the fluid properties that are critical to bearing performance. Associated wear will be immeasurable. When there is a great difference in the component’s hardness, any loads on the surfaces, whether lubricated or not, will cause faster wear of the softer material. This is demonstrated in carbon vs. chrome alloy journals. The carbon bearings are often replaced three to five times more frequently than the corresponding journals. Alignment, or rather misalignment, plays a key role in the successful performance of these types of bearings (Figures 5 and 6). The design anticipates a uniform film thickness supporting an evenly distributed load. Therefore, any angularity that affects the film profile is harmful. This includes bearing-tojournal fits not being parallel and bearing-to-bearing alignment not being concentric. In either of these situations the result will be point contact of bearing and journal. With materials not designed to contact, rapid and substantial bearing failure is inevitable. Some methods used to ensure bearing alignment are rabbet fits on flanges of mating component surfaces, precisely machined bearing cartridges, single-piece concentric bore bearing carriers, and self aligning bearings and thrust washers. Lubrication Paths Now that we have these bearings properly aligned and made from a material compatible with the dynamic loads and process fluids, we must provide an environment conducive to long-term operation. Because, as stated, these bearings are internal to the containment shell, we The Pump Handbook Series

will generally rely on the process fluid for this environment. I call it an “environment” in that this fluid flow or lubrication path serves three primary purposes: providing liquid for the hydrodynamic bearing film; providing flush for any solids either entrained in the process or captured from pump components (e.g. bearing wear); and removing the heat associated with the sealless configuration (hysteresis) and internal bearing system. In short, it carries the load and washes and cools the drive. The lubricating fluid, most commonly the process liquid, is drawn from the pump casing, typically through the bearing carrier front flange or the casing back plate (Figure 7). In most cases it originates at a high pressure point and flows through internal bores/channels to the close clearances of the bearing area. From there it moves to the far reaches of the rotor assembly, and then to the high velocity areas of the driven/inner magnet. The lubricant finally reaches the containment shell gap, where it rejoins the process fluid at some lower pressure point, most often at the suction. In these areas it supports the rotating assembly through all of its practical operating conditions. It also circulates through all open areas internal to the containment, continually removing debris and solids buoyed by the flush. The fluid also removes both the heat of hysteresis loss generated by the disruption in the magnetic or electromagnetic flux fields passing through the containment shell and, to a lesser degree, the friction heat of the bearings. This is probably one of the most critical aspects of sealless pump operation. With various manufacturers’ designs, the combination of internal flow circulation rates (ranging from less than 1% to as much as 5% of pump flow) and hysteresis

Figure 7. Discharge to suction circulation path

Figure 8. Flow originating at backside of impeller splits into dual flow path for improved circulation, optimum cooling and lubrication

losses (from 0% to as much as 15% of drive ratings) dictate various heat removal requirements. Temperature rises of this fluid through its path can range from 1-15°F when handling water. When the fluids more typically associated with sealless requirements are handled, these temperature fluctuations change inversely proportional to the specific heat and density of the liquid. When the temperature of the process fluid lubricating the bearings rises above its vapor pressure and flashing of liquid occurs, the fluid support of the bearing is interrupted. The phenomenon to watch for is often referred to as “bearing cavitation.” The bearings do not perform as hydrodynamic bearings and contact occurs similar to misalignment. Factors that can cause this condition are the design path of internal circulation, handling liquids too close to their vapor pressure or operating at extreme flow conditions (particularly high flow). Since there are often choices in internal circulation flow paths, care should be taken to understand your options. The most common is “discharge to suction.” This, as it suggests, takes process fluid from discharge pressure (actually about 80-85% of differential above suction pressure) and routes it through the

drive components to the eye of the impeller (suction pressure). In many such designs the pump’s Net Positive Suction Head required (NPSHr) would be increased. While this choice generally gives higher internal flow rates, it exposes the heated fluid to lowest possible pressure. When operating close to vapor pressures, the preferred method is “discharge to discharge.” This really means discharge, as stated above, back to a pressure maybe 40-50% of differential above suction. This area is always outside of impeller-to-casing running fits. The typical result is a lower circulation rate, unless an internal impeller or repeller is provided in the containment shell area to “boost” the circulation pressure. With either the “discharge to suction” or the “discharge to discharge” method, the route of the flow can be a factor. In some setups the flow goes immediately to the bearings and then on to the containment shell area to remove heat. This affords the bearings the most pressure at the lowest temperature. In other designs the flow is around the containment shell, then across the bearings (Figure 8). Each and every design can work properly. It is important to understand your own application constraints. Then consult your sealless pump supplier for choices and The Pump Handbook Series

advice about what is best for you. Since this internal circulation is directly affected by the pressure differential across the pump, it is important to carefully consider “runout,” or high flow conditions. They offer high bearing loads, use larger magnet couplings with corresponding larger losses and provide lower developed head for driving the liquid through the pump internals, regardless of the flow path selected. All of this could have an adverse effect on bearing performance by producing bearing cavitation. Magnetic drive pump manufacturers have become more flexible in providing “external” flushing arrangements. This change addresses a multitude of process conditions previously known to be detrimental to sealless pump operation. The newest Sealless Centrifugal Pump Standards from the Hydraulic Institute (HI 5.1 - 5.6) show twelve such conditions. They cover everything that has been associated with seal flushes for years. My experiences strongly support these as viable choices in extending the performance of sealless pumps when confronted with marginal process conditions. Consult an experienced supplier for details and restrictions on the use of these plans.

Conclusion Now that you better understand the design of these pumps, please allow me to offer some interesting cases of customer satisfaction with improvements in their process as a result of using this technology. The first that comes to mind is a large chemical producer handling hot (250-300°F) acids. The company was experiencing repeated failures of sealed pumps, and even some sealless designs. Complicating the problem was a relatively high suction pressure. After years of having sealed pumps fail as frequently as 23 times a month and lesser design sealless pumps not survive a 60-day guarantee, we provided a Teflon®lined, heavy duty magnetic drive pump. This pump was offered with a reinforcement collar on the casing to provide for higher working pressures at the elevated temperature. The containment shell was Teflon®lined with a carbon fiber reinforced polymer pressure shell. This non-

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metallic design reduced any internal temperature increases to that generated by bearings. In magnetic drive pumps, the bearing loads (axial) associated with various suction conditions are pretty well balanced. The resulting magnetic drive performance has been an uninterrupted run of 22 months. Think of all the repair costs, environmental concerns of acid leaks and general process interruptions that were eliminated! The maintenance guys that were required to wear slicker suits in this area are most appreciative. In some situations it is not the operation of the pump that causes problems but the procedures relating to the pump process. One case of this is a recent experience where after a brief period of successful performance, the pump was shut down. During the next start-up the pump failed when the inner magnet became locked up in the containment shell. Upon dismantling the pump it was discovered that the inner liner of the containment shell had collapsed around the driven magnet. A review of the shutdown

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and start-up procedures showed nothing. Upon further review, it was found that between runs the pump was evacuated using a vacuum system. The force was significant enough to suck the liner over the driven magnet. The remedy was to convert the dual non-metallic containment shell to a “bonded” shell designed for operation in vacuum service. This enabled the use of the same pump and did not require alteration of the customer’s normal procedures. It is important to use the above occurrence to reiterate the significance of close communications between customer and salesperson to evaluate all aspects of the service for which magnetic drive pumps, or any pumps for that matter, are intended. The data sheets will not always give the whole picture of forces acting on pumps. As you would expect, the more you know about the application, the greater the success rate. And that will benefit all involved! "

The Pump Handbook Series

Acknowledgments I would like to acknowledge the assistance of the following individuals in contributing various levels of assistance with this article: Denny Fegan, President, Teikoku USA, Houston, TX Doug Twyford, Marketing Manager, Sundstrand Fluid Handling, Arvada, CO Venetta Diesel, Marketing Assistant, ITT Industries, Cincinnati, OH Cliff Dodge, Industrial Advertising, Goulds Pumps, Seneca Falls, NY Craig J. Bailey is a 1971 graduate of Virginia Polytechnic Institute with a B.S. in Mechanical Engineering and Nuclear Engineering. Mr. Bailey’s focus on sealless products, which started in 1984, has included work as Sealless Product Manager for ITT A-C and Richter. He is now Strategic Account Manager for Goulds Pumps.

Reliability Tips for Operating Magnetic Drive Pumps By Nick Valente, ITT Industries/Goulds Pumps • Do not operate mag drive pumps under no flow conditions.

the manufacturer’s recommended operating envelope.

• Do not operate mag drive pumps against a closed discharge valve.

• Consider temperature controlling devices such as heat jackets or steam tracing for pumps that are subject to process fluids whose characteristics (viscosity, specific gravity, crystallization, coagulation, solidification or vaporization) change with variable process temperatures.

• Do not operate mag drive pumps with solids that exceed the manufacturer’s maximum limits in particle size or concentration. • Before operating mag drive pumps, confirm the chemical compatibility of the process liquid with all wetted pump components. This will reduce corrosion, permeation and erosion. • Do not operate mag drive pumps with process liquids that may exceed the maximum temperature limits or fall below the minimum temperature limits defined by the pump manufacturer. • Do not operate mag drive pumps outside of the manufacturer’s recommended operating range. Otherwise, recirculate adequate flow through bypass lines when operating near or below the manufacturer’s recommended mechanical and thermal operating flow. This will prevent excessive temperature rise or recirculation cavitation. • Consider the temperature limits and recoverable flux density losses to increased temperatures of the inner and outer magnet materials. • Do not operate a mag drive pump without considering the process liquid’s vapor pressure characteristics over the temperature range of the application. Adequate NPSHa as well as vapor pressure is mandatory to prevent cavitation or vaporization in localized low pressure regions in the pump.

• Use temperature monitoring devices such as thermocouples, RTDs, temperature controllers or thermometers when the process fluid is susceptible to critical variations in temperature. • Use leak detectors such as fiber optic sensing devices or pressure monitoring devices when the fluid cannot enter the atmosphere. • Keep the process liquid in liquid form, that is, prevent the liquid from flashing. • Drive losses must be added to hydraulic power when selecting motor power, magnet materials and magnet sizes for metallic mag drive pumps. • Flush the pump sleeve bearings with clean flushing liquids from an external source or from a filtered internal recirculation piping system. • Do not exceed the manufacturer’s maximum viscosity limits. Internal fluid circulation velocities won’t be able to properly cool and lubricate the sleeve bearings. • Know how much temperature rise is acceptable to the process and proper protection of the pump. Use this formula:

∆T=

5.09 x hp Q x S.H. x S.G.

Where: T = Temperature rise in °F hp = total power required in hp Q = flow rate in gpm S.H.= specific heat in btu (lb/°F) S.G.= specific gravity or density relative to water • For pumps with internal lubrication circuits, the thermal rise of the liquid must be considered to obtain the total rise in the pump. Follow the manufacturer’s guidelines to determine temperature rise in the recirculation circuit. • Do not exceed the pressure and temperature limits of the pump as defined by the manufacturer. Consider the worst case scenario at shut off pressure when operating at maximum temperature and suction pressure. If your pumps will be modified with larger impellers for future capacity expansion, consider the maximum operating conditions for the largest impeller diameter you might install in the units. • Persons with surgical, metallic or electronic health devices should not perform maintenance on or handle mag drive pumps without consulting their physician. These pumps contain highly magnetic materials. • Remove electronic badges, credit cards, digital watches and similar devices from the magnetic field of pump parts to prevent damage. • Use non-metallic tools and work surfaces. Follow manufacturer’s precautions when disassembling, assembling or repairing mag-drive pumps.

• Use power monitoring devices when operating near or outside The Pump Handbook Series

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CENTRIFUGAL PUMPS HANDBOOK

Canned Motor Pumps: Back to Basics A straightforward look at what makes these sealless centrifugal designs unique. By Craig Bailey, ITT Industries/Goulds Pumps

anned motor” is a very descriptive term in a discussion of pumps. As the words suggest, the motor (electric) has been encapsulated in a can isolating the rotor and process liquid from the stator (Figure 1). The rotor directly connects to the impeller. In this configuration, the motor and pump are singular and allinclusive. The result is a compact package.

C “

Stator and Rotor The stator for this pump is very much identical to a conventional electric motor stator. It receives electric impulses from leads and induces a revolving electromagnetic field. All the concerns of these phenomena—including heat dissipation and electrical conductivity—are present. Similar to the stator, the inner

Figure 1. Canned motor construction

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rotor is typical of an electric motor. It consists of a metallic core that responds to the electromagnetic field of the stator. The rotor will be drawn along the rotating flux induced by the electric current in the stator. This rotation becomes the motive power for the impeller. The rotor is surrounded by the process liquid and therefore needs protection. Whatever chemical resistance is required for the pump casing and impeller should be considered. Since it rotates with the electromagnetic field, it has no hysteresis impact. Can/Containment Shell The can, as in canned motor pump, is a thin tubular caisson precisely machined and fitted inside the internal diameter of the stator. It extends from the front adapter plate that meets the pump casing to the bearing bracket/end cap supporting the rear bearing. It will totally isolate the process fluid from the motor stator windings. Since the can is in contact with the process fluid, it must resist corrosion and erosion. Additionally, since it is interfering with the natural flow of the electromagnetic flux field, it should be constructed so as to minimize this energy draw. In canned motor pumps, can thickness The Pump Handbook Series

is less critical since it is constructed in close contact with the stator, which offers mechanical support for pressure containment. This close fit is also important and facilitates maximum heat dissipation to the process liquid. Shaft The shaft in canned motor pumps, as in the more typical magnetic drive pumps, transmits power/torque from the rotor to the impeller. This rotating assembly needs the same bearing support as magnetic drive pumps. Typically, this is accomplished with bearing supports between the impeller and the rotor, and at the extreme end of the shaft, aft of the rotor. Notice that this is similar to the “straddle” mount design of some magnetic drive pumps. Here again, alignment of these bearing holders is critical during assembly and maintenance. Support Bearings Everything regarding magnetic drive pumps support bearings is relevant to canned motor pumps. Purpose and operation is identical. The evolution of bearing designs and materials has taken different paths with these two sealless technologies, but we are using similar designs with identical safety concerns.

Containment Shell/Can What we have discussed are the required features of pressure containment, resistance to chemical attack by pumped fluid and minimum disruption of magnetic field. As you might expect, these goals can conflict with one another. One example would be hot oil services where the chemical resistance of carbon steel is more than adequate but the induced eddy currents would generate enormous hysteresis losses, rendering the magnetic drive or canned motor pumps impractical. In another case, a non-metallic containment shell could offer excellent chemical resistance and superior eddy current resistance but not handle expected pressure containment. Generally, if pressure containment becomes the most critical issue, as with secondary containment, canned motor pumps by virtue of their design become the preferred selection. Remember that their can is supported by the considerable mass of the stator and housing. The method of constructing the containment shell in magnetic drive pumps is an often-debated issue. The two generally accepted choices are the welded tube section with end cap or the hydro-formed onepiece dome. Central concerns are integrity of the weldments and uniformity of the cross section. Important here is that design limits of any choice be understood and that your supplier provides consistently high quality products. Bearings Bearings are probably the most critical component of these types of sealless pumps. If the bearings are operating properly (centering the rotating assembly in the brackets and supporting the normal ranges of dynamic loads) the pumps will be quiet, exhibit low vibration and, as experience shows, run for many years. However, if they wear out, are improperly aligned or have inadequate lubrication, the resultant failure can be severe.

Wear associated with sleeve (journal) bearings is affected by the load acting on the bearing, the differences in hardness of the mating rotating/stationary components, the lubricity of the fluid film and the abrasive nature of the fluid. In general, loads on these bearings are well within acceptable limits even at extreme off-design performance. The more severe condition is run-out or high flow operation. At this point the loads are relatively high, but, of more significance, the pressure differential across the bearings is at its lowest. This jeopardizes the generation of an adequate film thickness to support these higher loads. However, if properly selected, these pumps rarely should see inoperable bearing loads. With such a close relationship between the hardness of mating bearing components and the lubricity of the process fluid, I will combine discussion of these features. Lubricity of the pumped fluid is the single biggest factor contributing to successful operation of these sealless pumps. With adequate lubricating qualities of the liquid, the sleeve bearings will have a very low coefficient of friction. Heat generation due to friction will be minimal, thereby reducing any negative impact on the fluid properties critical to bearing performance. Associated wear will be negligible. With greater difference in the component’s hardness, any loads on the surfaces, whether lubricated or not, will cause faster wear of the softer material. This is demonstrated in carbon versus chrome alloy journals where the carbon bearings may need to be replaced three to five times more frequently than the corresponding journals. Alignment, or rather misalignment, plays a key role in the successful performance of these types of bearings (Figures 2 and 3). Since the design anticipates a uniform film thickness supporting an evenly distributed load, any angularity that affects the film profile is harmful. This can be bearing-to-journal fits The Pump Handbook Series

Bearing

Shaft

Front Bearing

Rear Bearing

Figures 2 and 3. Misalignment can play a key role in the performance of bearings.

not being parallel or bearing-to-bearing alignment not being concentric. In either of these cases, the result will be point contact of bearing and journal. With materials not designed to contact, rapid and substantial bearing failure is inevitable. Some methods used to ensure bearing alignment are rabbet fits on flanges of mating component surfaces, precisely machined bearing cartridges, self aligning bearings and thrust washers and single-piece, concentric-bore bearing carriers. Lubrication Paths Now that we have these bearings properly aligned and of a material compatible with the dynamic loads and process fluids, we must provide an environment conducive to long term operation. Since, as stated, these bearings are internal to the containment shell or can, we will generally rely on the process fluid for this environment. I call it an “environment” in that this fluid flow or lubrication path serves three primary purposes: providing liquid (key is that this is a liquid) for the hydrodynamic bearing film; providing flush for any solids either entrained in the process or captured from pump components (e.g., particles from bearing wear); and removing the heat associated with the sealless configuration (hysteresis) and internal bearing system. In short, it carries the load and washes and cools the sealless drive.

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Figure 4. Discharge-to-suction circulation path

The lubricating fluid, most commonly the process liquid, is drawn from the pump casing, typically through the bearing carrier front flange or the casing back plate (Figure 4). In most cases, it originates at a high pressure point and flows through internal bores/channels to the close clearances of the bearing area, to the far reaches of the rotor assembly, and to the high velocity areas of the driven/inner magnet or canned motor rotor to containment shell/can gap. Here it rejoins the process fluid at some lower pressure point, most often suction. In these areas it supports the rotating assembly through all of its practical operating conditions. It circulates through all open areas internal to the containment, continually removing debris and solids buoyed by the flush. Of particular importance, the fluid removes both the heat of hysteresis loss generated by the disruption in the magnetic or electromagnetic flux fields passing through the containment shell or can—and, to a lesser degree, the friction heat of the bearings. This is probably one of the most critical aspects of the operating sealless pump. With various manufacturers’ designs, the combination of internal flow circulation rates (ranging from less than 1% to as much as 5% of pump flow) and hysteresis losses (from 0% to as much as 15% of drive ratings) dictates various heat removal requirements. Temperature rises of this fluid through its path can range from 1-15°F when handling water. When the temperature of the process fluid lubricating the bearings

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rises above its vapor pressure and flashing of liquid occurs, the fluid support of the bearing is broken. The phenomenon to watch for is often referred to as “bearing cavitation.” The bearings then do not perform as hydrodynamic bearings, and contact similar to misalignment occurs. Factors contributing to this condition are the design path of internal circulation, handling liquids too close to their vapor pressure or operating at extreme flow conditions (particularly high flow). Since there are often choices in internal circulation flow paths, care should be taken to understand those proposed. The most common is “discharge to suction.” This, as it suggests, takes process fluid from discharge pressure (actually about 80-85% of differential above suction pressure) and routes it through the sealless drive components to the eye of the impeller (suction pressure). In many such designs the pump’s Net Positive Suction Head Required (NPSHR) will be increased. While this choice generally gives higher internal flow rates, it exposes heated fluid to lowest possible pressure. This principle applies to both magnetic

drive and canned motor pumps. When operating close to vapor pressures, the preferred method is “discharge to discharge.” This really means discharge, as stated above, back to a pressure perhaps 40-50% of differential above suction. This area is always outside of impeller-to-casing running fits. The situation will typically result in lower circulation rates unless an internal impeller or repeller is provided in the containment shell/can area to “boost” the circulation pressure. In either the “discharge to suction” or the “discharge to discharge” situations the route of the flow can be a factor. In some cases, the flow goes immediately to the bearings, then to the containment shell area to remove heat. This affords the bearings the most pressure at the least temperature. In other cases, the flow is around the containment shell, then across the bearings (Figure 5). Each and every design can work properly. It is important to understand your own application constraints. Then consult your sealless pump supplier for choices and advice on what is best for you. Since this internal circulation is directly affected by the pressure dif-

Figure 5. Flow orginating at backside of impeller splits into dual path for improved circulation, optimum cooling and lubrication The Pump Handbook Series

ferential across the pump, it is important to carefully consider “run-out,” or high flow conditions. They offer high bearing loads, use larger magnet couplings with corresponding larger losses and provide lower developed head for driving the liquid through the pump internals, regardless of the flow path selected. All of this could adversely impact bearing performance by producing bearing cavitation.

Conclusion Both magnetic drive and canned

motor pump manufacturers have become more flexible in providing “external” flushing arrangements to address a multitude of process conditions previously known to be detrimental to sealless pump operation. The newest Hydraulic Institute’s Sealless Centrifugal Pump Standards, HI 5.1 5.6, shows 12 such conditions. They cover everything that has been associated with seal flushes for years. My experiences strongly support these as viable choices in extending the performance of sealless pumps when confronted with marginal process conditions.

The Pump Handbook Series

Consult an experienced supplier for details and restrictions on the use of these plans. " Craig J. Bailey is a 1971 graduate of Virginia Polytechnic Institute with a B.S. degree in mechanical engineering and nuclear engineering. Mr. Bailey’s focus on sealless products, which started in 1984, has included work as Sealless Product Manager for ITT A-C and Richter. He is now Strategic Account Manager for Goulds Pumps.

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Selecting the Right Thermoplastic Sump Pump For Corrosive, Abrasive and Ultrapure Services Understand the hydraulic properties of these pumps and systems before choosing the design for your application. By Erich A. Meyer, Vanton Pump & Equipment Corporation he assumption that sump pump selection requires little more than the utilization of the basic hydraulic formulas involving the horizontal and vertical distance traveled from the pump discharge connection, the viscosity of the fluid and the friction loss of the piping, is a dangerous simplification that can lead to serious consequences, especially if the sump depth is a significant percentage of the static pump head requirement. What is unique about all sump pump applications is the often ignored fact that the system static head loss undergoes continuous change. It is constant only at a given instant of time. As the liquid level in the sump changes, the system static head changes with it. In truth, the piping system head loss curve is actually a moving target whenever the pump is functioning. One purpose of this article is to clarify the significance of this concept. In addition to this basic hydraulic problem, the handling of corrosive, abrasive, hazardous, and other troublesome fluids requires special attention to

T

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construction materials. Much has been written about the chromium-nickel family of stainless steels, the special high alloys, and various exotic metals and alloys. Despite the fact that thermoplastic pumps have been in transfer, processing and effluent handling for half a century, information on their use in specific applications is surprisingly limited. There are notable exceptions to this shortcoming. One is the substantial editorial attention being paid to the use of thermoplastics when pumping semiconductor chemicals, reagent grade chemicals, pharmaceuticals and other liquids demanding complete freedom from metallic contamination. Another is the growing library of technical article reprints which form a unique application database. These are available from major publications, from pump manufacturers and on the Internet. I will attempt in this article to highlight key material selection considerations presented in these articles—or that I have learned from actual experience— that affect vertical pump selection. Another topic that will be addressed is design specifics that The Pump Handbook Series

differentiate thermoplastic vertical pumps from metal pumps. I will try to present this data in a way that will highlight the limitations as well as the advantages of the thermoplastics, so that specifiers will feel confident in the selection they make. When dealing with aggressive fluids, pump selection is a combination of textbook data and field experience. This holds true whether we are dealing with metals or plastics. There is no substitute for the experience of the user, the raw material producer or the pump manufacturer.

Look at the Hydraulics System Characteristics It is clear that before the cost effective pump size can be determined it is necessary to evaluate the piping system hydraulic losses as a function of the flow through the piping system. The system losses are the pressure (head) losses in the pump discharge piping from the point where the liquid leaves the pump casing to the discharge point of the piping system.

The two components that make up the system losses consist of static and dynamic head losses. Static head loss is the energy required to lift liquid from one elevation to a higher one. The dynamic head losses are primarily friction losses occurring due to the flow of liquid in the piping system. The dynamic losses will therefore vary with the magnitude of the flow through the system, pipe size and pipe type. A typical sump pump application is shown in Figure 1, where liquid needs to be transferred from an inground sump to a vessel above ground. The pressure of the liquid at the sump liquid surface is atmospheric (14.7 psia). The pressure at the exit of the piping P2, at the tank, is also atmospheric. The static head that the pump must supply is the energy required to lift and transfer the sump liquid from the air/liquid interface in the sump to the discharge point. The dynamic head losses that must be overcome by the pump consist of all the energy losses associated with the movement of liquid through the piping system from P1 to P2. A typical system loss curve is shown in Figure 2. When the sump liquid level is at its maximum, the static head loss is represented by the horizontal line at a loss value of H, ST, MIN. When the sump liquid level is at its minimum, the static head loss is represented by the horizontal line at a loss value of H, ST, MAX. The dynamic head losses are added to the constant static head loss to determine the total piping system loss as a function of flow. As explained previously, the system static head is constant only at a given instant of time. As the sump liquid level drops due to pump operation, the system static head loss constantly increases and reaches a maximum value of H, ST, MAX at the minimum sump level. The piping system head loss curve is actually a moving target in sump pump applications. More scientifically speaking, the piping system head loss curve is a function of both time and capacity.

Another sump pump application different from that in Figure 1 is shown in Figure 3. Here the transfer vessel is fed the liquid from the sump pump near the bottom of the tank instead of dumping into the top of the tank. The sump pump must therefore overcome a changing pressure P2 at the exit of the pump discharge pipe.

P2 increases as the transfer vessel fills with fluid and h,2 increases. It can be shown that the static head loss that the sump pump must overcome is h,3, the vertical distance between the liquid air interfaces in the two containment vessels. As is the case for the Figure 1 application, the system static head loss is a function of time as

Figure 1. A typical sump pump application where liquid needs to be transferred from an in-ground sump to a vessel above ground

Figure 2. A typical system loss curve The Pump Handbook Series

283

the sump pump operates. As the pump empties the sump, the distance h,3 increases not only because the sump level decreases, as in Figure 1, but also because the transfer vessel height, h,2, is increasing.

Sump Pump Head The Hydraulic Institute defines the total head as “the measure of work increase per unit mass of liquid, imparted to the liquid by the pump, and is the algebraic difference

between the total discharge head and the total suction head.” Total head can also be thought of as the difference between the discharge pressure, Pd, and the suction pressure, Ps, referred to the datum and expressed in feet of liquid pumped, plus the velocity head. Referring to Figure 4, the pump discharge pressure referred to the pump datum and expressed in feet of liquid is Pd = [P + γ(ZD)]/γ Where γ = liquid specific weight P= discharge pressure gauge reading The pump suction pressure is the pressure of the sump liquid at the datum given by Ps = γ (ZW) Neglecting the velocity head, the total pump head is then H = [Pd + γ(ZD)]/γ-[γ(ZW)]/γ Simplifying,

Figure 3. In this sump pump application, the transfer vessel is fed the liquid from the pump near the bottom of the tank instead of dumping into the top of the tank.

H = Pd/γ +[(ZD-ZW)] The quantity ZD-ZW is the distance between the centerline of the discharge gauge and the sump liquid level. The pump head is a constant for a fixed capacity. Therefore, although the discharge pressure of an operating sump pump will increase when the level of the liquid in the sump rises, the pump head remains constant because the distance ZD-ZW decreases with rising liquid level. Similarly, as the liquid level falls, the discharge pressure gauge reading of an operating pump will drop, but the pump head remains constant because the distance ZD-ZW increases.

Pump Selection and Operating Point To determine the flow at which a particular pump will operate when

Figure 4.

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The Pump Handbook Series

installed in a system, the pump headcapacity curve needs to be superimposed on the system head loss curve as illustrated in Figure 2. The pump operating flow is fixed by the intersection of the system head loss curve with the pump head-capacity curve. Referring to Figure 2, if the sump level remained fixed at the minimum sump level (as would be the case if the sump volume were much larger than the transfer vessel volume), then the pump would operate at flow Q, MIN. The pump size must be selected such that the operating flow Q, MIN is sufficient to fill the receiver vessel in a time frame that meets the project objectives. In many sump pump applications it is the intent to empty the sump in a reasonable amount of time, say 20 to 60 minutes, and the sump level does change considerably. In that case it is important to realize that if no flow control valves are employed in the pump discharge, then the operating

flow of the pump will vary from Q, MAX to Q, MIN as the sump is emptied. The initial operating flow rate is determined by intersection of the pump curve and the system loss curve when the liquid level is at its maximum. The pump flow then gradually diminishes to Q, MIN when the sump is at its minimum level. It is imperative that the selected pump performance is such that the minimum flow Q, MIN is sufficient to meet project requirements. This is also the worse case condition where the selected pump has the largest head requirement. The shape of the pump head-capacity curve must be investigated from shut off to Q, MIN to prevent the pump from operating at zero flow. If the head-capacity curve is relatively flat from shut-off to capacity Q, MIN, any slight change in the process or errors in the calculation of system losses could cause the pump to run at no flow or randomly oscillate in flow as the sump level drops.

Vertical Pump Design Considerations

Photo 1. Typical vertical thermoplastic pump with welded support ribs to provide rigidity to column

Assuming you are considering thermoplastic pumps for a given application, the basic design factor should be related to how the manufacturer assures there will be no metal in contact with the aggressive fluid. Regardless of pump length, pump shafts are furnished in 18-8 stainless steel or in high strength carbon steel, depending on pump length. A critical aspect is the precaution taken to assure that the portion of the metal shaft immersed in the fluid, or exposed to fumes emanating from it, will be totally isolated. One solution is to encase the shaft in a thick sectioned sheath of a thermoplastic material inert to the liquids or gases handled. For vertical pumps with shaft lengths under 20 feet, this can readily be accomplished with standard sleeving/sheathing practices, but for pumps in deeper sumps, the solution is difficult. In fact, until the development of the patented segmented shaft with its sliding coupling arrangement, The Pump Handbook Series

total isolation of shafts for pumps from 20 to 50 feet was impossible, or certainly impractical. Metal shaft length is no longer a deterrent since shorter shafts coupled to suit any sump depth can now be specified. This makes it feasible to install and economically maintain these very tall pumps in plant areas having limited headroom. It also simplifies and reduces shipping and handling costs. The spans for both the ball bearings and wet bearings need to be determined over the length of the pump shaft to assure proper rigidity and maximize service life. The inner wet bearings should be specified in a ceramic such as ultrapure (99.5%) aluminum oxide, while the outer bearings should be made of a material specifically selected for its compatibility with the liquids to be handled. This can be reinforced Teflon®, Vanite or ceramic. Lubrication or cooling of these bearings can be accomplished by the pumped liquid or clean water, with the type dependent on the characteristics of the fluid, including its solids or abrasive content, and the service duty cycle (continuous vs. intermittent). Recent developments eliminating immersed bearings in vertical pumps with shaft lengths to 5 feet have been attracting great deal of attention. Using heavy duty, larger diameter steel shafts, these fully cantilevered pumps are intended for services that preclude the use of wet bearings due to damaging solids content in the fluid being pumped or the inability to dilute the product with a water bearing flush. The bearingless cantilevered design also provides for extended dry running without pump or motor damage. The shaft is completely isolated from the fluid with thermoplastic sleeves. Prior to this development, lower cost thermoplastic vertical sump pumps were and still are available with a modified cantilever design. This design eliminates the use of wet sleeve bearings and incorporates a shaft stabilizing bushing in the casing to increase rigidity.

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Grease lubricated ball bearings above the cover plate should be protected by a V-seal designed to retain corrosive fumes within the sump and avoid corrosive damage to metal parts above the cover plate as well as to the motor. The external ball bearings in

Photos 2 & 3. Full cantilever thermoplastic pump designed to run dry for extended periods without damage to pump or motor. There are no bearings immersed in the fluid being pumped. Note the rugged stainless steel shaft completely isolated from the fluid with a thick-sectioned thermoplastic sleeve.

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the motor mounting bracket should permit easy, accurate adjustment of impeller clearance without removing the pump from the installation. Specifications frequently call for these bearings to be regreasable camlock designs to simplify positioning and locking of the pump shaft in position. The vertical column sections containing the pumped fluid are generally furnished in standard rigid, heavy walled thermoplastic pipe or tubing, with the material selected in accordance with its inertness to the fluids in which it will be immersed. To assure rigidity, it is suggested that all sump pumps with column sections longer than 96 inches be reinforced by stabilizing welded ribs of the same material as the column. The individual sections should be fitted with flanges at the ends to facilitate the bolting together of sections. All bolts used for this function or for mounting the pump casing, or for any other purpose within the fluid area, should be specified in a compatible thermoplastic material to minimize maintenance and avoid contamination of the pumped fluid. For maximum performance and low maintenance, it is recommended that pump casings and casing covers, as well as the impellers, be precision molded of virgin, homogeneous thermoplastics selected for the anticipated service conditions. The impeller should be molded around a dynamically balanced stainless steel reinforcing insert that has been bored and keyed for mounting on the pump shaft. In addition, the impeller should be designed to allow for movement within the pump casing to accommodate changes due to fluctuating temperatures. This design aspect, referred to as TFM—Thermal Fluctuation Modification, allows for axial movement of the impeller without interrupting operations. For the handling of fluids, sludges, slurries and solids-laden fluids, including those with stringy materials, vertical pumps can be specified with recessed impellers that prevent The Pump Handbook Series

clogging and impeller binding. These vortex designs with clog-free impeller positioning are available in the full range of vertical thermoplastic pumps for sump depths to 50 feet.

Material Selection The materials choices for fluid contact components in vertical nonmetallic pumps are similar to those for horizontal centrifugal or rotary pumps, and the determination depends on the same factors. The critical considerations are fluid temperatures, the corrosive nature of the chemicals, the abrasive

Photo 4. Close up of sump pump head showing heavy-duty solid thermoplastic components— column, casing, piping, strainer and hardware.

Photo 5. Close up of vertical thermoplastic pump shaft/impeller assembly showing precision molded PVDF impeller, thick-sectioned PVDF shaft sleeve and ceramic bearing.

Table 1.

Table 2.

characteristics of the liquid or slurry, the degree of product contamination that can be tolerated, and the current as well as the potential applications for which the pumps may be used. Although textbooks on available plastics list endless material compositions, the significant choices faced by plant and materials engineers, as well as system consultants, realistically narrow down to just these four: Polyvinyl chloride (PVC) Chlorinated polyvinyl chloride (CPVC) Polypropylene (PP) Polyvinylidene fluoride (PVDF) Horizontal centrifugal pumps can also be furnished in ethylene chlorotrifluorethylene (ECTFE), but at present this material is not available for the columns in vertical pumps, although it can be furnished for the casings and impellers. In Europe, vertical pumps are also frequently manufactured in polyethylene (PE), but the ready availability of polypropylene—and its superior resistance to strong oxidizing acids and broad range of waste streams—has made it an industry standard. Since this article deals with pumps made of

homogeneous, virgin thermoplastics, we have not included design aspects that must be considered in selecting fiberglass reinforced plastics as the basic construction material. The primary consideration in terms of material selection lies with the anticipated fluid temperature. Listed in Table I are the maximum service temperatures for all of the materials mentioned. These figures refer to sustained temperatures, not occasional, short term degrees. We have also included polytetrafluoroethylene (PTFE) for comparative purposes, although heavy duty vertical pumps are not available in this material. Vertical centrifugal pumps are widely used for the handling of corrosive waste streams with widely varying pH values, and frequently with multi-shaped and sized abrasive debris. The broad chemical resistance of polypropylene, coupled with its excellent abrasion resistance, relatively high temperature tolerance, ready availability and low cost, has made it the preferred material for process fluids as well as industrial and municipal effluents, both liquid wastes and noxious or hazardous fumes. For handling process fluids that are extremely corrosive or abrasive, or those that require tolerance of higher sustained temperatures, polyvinylidene fluoride is most commonly specified. Table II offers direct comparisons of the abrasion resistance of these materials as measured in weight loss when laboratory tested on Taber Abrasion Testing machines. In recent years there has been an increasing demand for pumping equipment that avoids all types of product contamination, not only those caused by contact with metal components. This is particularly significant in the transfer of ultrapure water, semiconductor chemicals, pharmaceuThe Pump Handbook Series

ticals and reagent grade chemicals. Thermoplastic vertical pumps with all wetted components made of (or encapsulated in) chemically inert nonmetallic materials, take care of the metallic contamination problem. However, in some extremely critical applications in which nonmetallic additives may be present in the molded or extruded plastics, care must be taken to specify natural or unpigmented thermoplastics to avoid product contamination.

Summary There is no substitute for experience in the transfer or metering of corrosive, abrasive, hazardous or ultrapure fluids. Your own experience coordinated with that of the pump manufacturer provides the best assurance of dependable, cost effective fluid handling. The consistent upgrading of pump designs and material improvements suggest the wisdom of reviewing your choices with your suppliers whenever you are expanding existing facilities, planning new construction or even replacing a troublesome pump. Recent plant efficiency studies show increased concern about life cycle pump costs, which are often ignored when direct replacements or costly repairs are authorized because the cost can be assigned to existing maintenance and repair budgets. Automatic replacement or repairs of the same pump without evaluation of its performance and related up-keep costs may prove “penny wise and pound foolish.” " Erich Meyer is an engineering specialist with extensive experience in hydraulics. Prior to joining Vanton Pump & Equipment as Senior Design Engineer, he served with Worthington Pump Company and Ingersoll-Dresser as a hydraulic specialist and research/ development engineer. Mr. Meyer has a B.S. in mechanical engineering and an M.S. in fluid dynamics and is the author of many articles on hydraulics and pump performance.

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CENTRIFUGAL PUMPS HANDBOOK

Successful Submersible Operation Part I: Pump Installation and Start-Up Getting these pumps set up and working right requires careful planning, execution, testing and follow-up. By Submersible Wastewater Pump Association areful preparation and planning is needed to ensure proper installation of all equipment in submersible sewage pumping stations. There must be effective coordination between the supervising engineer, the mechanical and electrical contractors and all suppliers. The first step is to make certain that the required station equipment is ordered completely and accurately according to specifications, and that—considering lead times—all of it will arrive prior to the scheduled installation. This includes not only the pump and electrical control panel, but the station itself (if prefabricated) and all accessories, such as piping, valves, and access covers.

C

Receipt and Inspection All equipment should be examined upon receipt for any signs of apparent damage. If damage is indicated, a claim should be filed immediately with the carrier and the supplier should be notified. All parts shipped loose or separately should be checked for loss or damage. If the equipment won’t be installed immediately, it is best to store it in a clean, dry location where it will be protected from possible damage. When storing equip-

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ment, follow the manufacturer’s recommendations and protect it from low temperatures or freezing. Check the equipment prior to storage to anticipate any possible problems at the time of installation. By checking sizes, design features called for on the plans and specifications, and all interface components, installation problems can be minimized.

The end of the power cable must not be submerged, as water may wick through the cable into the motor. After any period of storage, the pump should be inspected, tested and/or reconditioned in accordance with the manufacturer’s recommendations before it is put into operation.

Pump Station Installation Handling Pumps A submersible pump should be transported and stored in accordance with the pump manufacturer’s recommendations. Make sure that it cannot roll or fall over. Always lift the pump by its carrying handle, never by the motor cable. In the absence of specific manufacturer recommendations about storage, pumps must be protected against excessive moisture and heat. This precaution is necessary to prevent the possibility of moisture damage to internal components and the power cord, etc. The impeller should be rotated by hand occasionally (for example, every other month) to ensure free movement of the rotating elements. Make sure the cable entry seal conforms to the outside diameter of the cable to prevent leakage into the pump. The Pump Handbook Series

When beginning the installation, make certain that all equipment is on site and that all manufacturer instructional literature is available and has been reviewed. In addition to the pump(s), the following items are typically required: • discharge elbow/support base and/or sealing arrangement for connecting the pump to the discharge line • guide rail(s) consisting of specified material and size • upper guide rail brackets for attaching the rail(s) to the access cover frame or top of the station • intermediate guide rail supports as required • specified discharge piping and fittings • the proper check valves and shutoff valves

• level sensor or other control equipment • cable holder for level sensors or other sensor brackets • junction box or conduit box (if required) • access frame (with covers) • control panel

Access Cover Installation There are two common installation methods for access covers. One is to cast the unit into the poured concrete slab well cover. The other method is to frame flange the unit and drop it into the steel well cover. Both types of installation must align properly with the base plate. Prior to casting the door unit into concrete, the cover should be closed and checked to ensure that it rests on the frame all around. Shim the frame as necessary to ensure proper door closure. Remember that aluminum frames must be protected from the wet concrete by a bituminous coating. Installation procedures when casting the access cover in concrete are as follows: • Spring-loaded units—Caution: The cover is spring loaded. Do not remove the safety shipping bolt until the unit is ready to be installed in a normal horizontal operating position. • Non-spring loaded units—shipping bolt or banding should not be removed until the unit is ready to be installed and is in normal horizontal operating position. • On angle frame doors, be sure the concrete anchors are in the proper position when setting the frame. Then remove the banding or safety shipping bolt and plate and the outside latch release handle. • Place the access frame and cover on the opening so that the lift handle is in the desired location and the slide rail connections, junction box or conduit box provisions, and chain hooks (if all are furnished) are in the proper relationship to the pump baseplate and the discharge piping connections.

• Before pouring concrete or anchoring in place, open and close the door(s) and check to see that the door rests on the frame all around. If not, shim under the frame as needed at the corners and recheck for proper door closure. • With door(s) closed, place the access frame in position and install the anchor bolt. Pour the concrete flush with the top of the frame. Use care to prevent the concrete from getting into the frame or around the hinges. Be sure to support the frame to prevent sagging. Do not permit the weight of the concrete to push the frame inward and reduce the clearance between the door(s) and the frame. There is little or no maintenance required on access covers. Painted doors should be cleaned and repainted when necessary. Spare parts— such as springs, hinges, latches and arms—seldom require replacing. However, should there be a need to replace any of these items, they are usually readily available from suppliers. To ensure that replacement parts are ordered correctly, the broken part should be returned to be matched. If this is not possible, the model number, size, type of metal, year of purchase and shop order number should be provided.

Internal Assembly Now that the access cover is in the proper position, you can begin to install the internal station components. Place the pump discharge connection in position. Temporarily secure the guide rail(s) in the upper mounting brackets and the discharge bosses at the bottom. Install the intermediate support brackets, if required. Make sure the rails are in a true vertical position, so the pump will clear the access opening and will slide freely down the rails into place in the discharge connection. Once the rails are in proper alignment, bolt the discharge connection The Pump Handbook Series

into the floor of the station. Connect the discharge pipe to the discharge connection and proceed to install the check valve, shutoff valve, and fittings according to the plans and specifications. All level-sensing devices must be properly secured to the access cover frame or the wet well. Install them at the levels indicated on the plans and specifications. Control components, other than level-sensing devices, should not be installed in the wet well. Lifting equipment is normally required for handling pumps. It should be able to hoist a pump straight up and down in the station, preferably without having to reset the lifting hook. The lifting equipment must be securely anchored. Keep all personnel out from under suspended loads. Before lowering a pump into place, it must be checked for correct rotation, using one of the following methods: 1. Prior to installation in the wet well, lay the pump on its side and momentarily run (jog) and check it for rotation. When running the pump outside the wet well, stay clear of the impeller and provide a safe, temporary connection of the motor leads. The best procedure is to lift the pump with a hoist or tilt it slightly and observe the rotation. An arrow on the pump casting, or instructions in the Pump Instruction and Operation Manual (IOM), will show the proper direction of rotation. This is the most accurate method to ensure proper rotation; however, other methods are usually provided as a secondary check or for circumstances when the pump cannot be lifted from the station. 2. After installation, if the check valve has an external operating arm, observation will indicate which rotation of the motor

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opens the check valve widest. This indicates the higher flows and is the correct rotation for the pump. 3. If pressure taps are available, rotation can be ascertained by reading the shut-off head on a pressure gauge. The proper rotation will produce the higher pressure. 4. The least desirable procedure is to position the pump in the Hand mode, using the HandOff-Automatic (HOA) switch, lowering it to the bottom of the wet well, and checking the drop in water level in the well. The rotation resulting in the faster drawdown is the proper rotation. One of the above procedures must be used to check and ensure proper rotation of each pump before startup of the station. Keep in mind that visual inspection of the rotating elements is the most accurate method. The direction of rotation on three phase pumps may be changed by interchanging any two motor leads at their control panel connections. For single phase pumps, if improper rotation is observed, consult the IOM or contact the pump manufacturer. After the proper rotation is verified, lower the pump along the guide rail(s). Upon reaching its bottom position, it should automatically connect to the discharge. Adjust as required, following the manufacturer’s recommendations. Fasten the lifting chain on the access frame eyebolt and fasten the cables on the cable holder. Cable supports are required for deep installations. Run the pump and level control cables up to the electric control panel or into the junction box, if used. The pump and level control cables should be long enough to reach the control panel without splicing. If a junction box, conduit box or quick connector plugs are utilized, they should be located outside the wet well—or if located in the wet

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well, be of at least NEMA Type 4 construction. It is highly recommended that “conduit seal offs” be located between the wet well and the control panel to prevent gas migration into the control panel enclosure. Arrange for a cable, if needed, between the sump and the electrical control panel. Make sure the cables are not sharply bent or pinched and that all connections are sealed and watertight to prevent leakage from ground water.

Control Panels Before making a new installation, a qualified electrician or factory service technician should verify the horsepower, voltage rating and full load amperage of each pump. This information should be used to ensure that the control panel is of the correct horsepower, and that the heater coils furnished in each motor starter overload unit are sized or set correctly to match the motor’s operating current as given on the motor nameplate. The service voltage and frequency should be checked to ensure that they are the same as the motor rating. This information should be used to crosscheck the circuit breakers, fuse or disconnect ratings. All electrical work must comply with national and local codes and regulations. At the time of installation, the panel or equipment should be checked for missing or loose components—including a correct wiring diagram. All wire terminations should be checked for tightness. Care must be taken in handling the control panel and equipment during installation to avoid damage to the enclosure or any of its components. Location and provisions for mounting the panel should be shown on the plans for the job. Make certain the enclosure is of the type specified for that particular location. Adequate racks, mounting brackets and fastening hardware The Pump Handbook Series

must be used to securely mount the panels and devices. On flat surfaces, panels should have a slight clearance between the back of the panel and the mounting surface to allow for air circulation. While clearance is usually provided by the enclosure design, it may need to be furnished by mounting hardware. This assures heat dissipation and prevents moisture accumulation. If the panel is mounted on a rack, it must be constructed and braced to provide a flat, rigid surface that will not distort the panel and cause possible door alignment problems. Mounting bolts and hardware must be of sufficient size to provide stable positioning and of materials suitable for the operating environment.

Field Wiring All conduits and wires must be installed as required on the plans for the specific job. All cable entrances into the enclosure must be in accordance with the NEC and shall maintain the integrity of the NEMA enclosure to prevent intrusion of moisture, dust and gas vapors. When possible, attach conduits from the bottom of the enclosure to facilitate weatherproofing. Use weatherproof hubs and sealed fittings. The size of conduits and wires must be adequate for the specific requirements of the incoming service, pump motor leads and remote control devices. The NEC must be used as a minimum guideline. After the installation of the conduit and wiring, make certain that all terminations in the control panel conform to the panel manufacturer’s diagrams and instructions. The incoming service voltage must be correct for the panel. Typically, three phase systems are designed for a 208, 240, 480, or 575 volt, four wire service voltage. Single-phase systems are most often 208 or 240 volt, two or three wire, but are sometimes 120 volt, two wire service voltage. As a safety precaution, prior to connecting the motor leads to the

control panel or applying power to the pump, megohm readings should be taken and recorded with a 500-volt megger. Connect all motor leads together and check the combination to ground. The readings should be above 20 megohms at all points. A motor should not be run if any reading is below 10 megohms. If this is the case, find the source of the low readings and correct it. If the readings are between 10 and 20 megohms, the pump should be run for short periods and the readings rechecked. Only after tests are complete should the motor leads be connected permanently to the control panel. The pump leads must now be connected to the control panel following the panel diagram, using the pump manufacturer’s wiring diagram to distinguish and verify the connection and color coding of the leads. Internal pump/motor safety controls should be identified following the pump manufacturer’s wiring diagram and connected to the panel as specified by the panel manufacturer. Make certain the pump is correctly grounded. If automatic control of the pumps is provided by liquid level sensors in the wet well, all connections must be made to the proper terminal points in the control panel. Identify each sensor and its specific function. Proper connections ensure sequential and automatic operation of the pumps. Wires must be marked to provide future identification should re-connection be required during maintenance or troubleshooting. A record should be made in the permanent file for the control panel regarding field wire connections, wire sizes and types, cable lengths, and any other related information. This file can be valuable during normal maintenance, for emergency situations and as a reference for future installations. A qualified electrician should be present when incoming service voltage is available and the system is

complete and ready to be put into service. If the service voltage is 240 volt, three phase, four wire from Delta-connected transformers, the “high line” must be identified; it has a higher than 120 volt (usually approximately 210 volt) reading phase to neutral. The control panel must then be checked to make certain no control circuits or 120 volt operating devices are connected to this incoming high line. Some local codes require that the high line be on a particular phase connection, usually Phase B or sometimes Phase C. Most control panels are designed to use Phase A for all 120 volt phase to neutral circuits when no control transformer is used.

Start-Up and Testing Many pump and panel manufacturers have special forms that can be used during the start-up of the station (Figure 1). Some manufacturers require that these forms be filled out and returned to ensure warranty on the pumps, control panels and station components. This type of form provides a detailed description of the procedures and tests to be performed. The following summarizes these procedures. The panel is ready to be put into service after the incoming service voltage has been checked on the line side of the main circuit breaker or disconnect, and it is verified that all phases and neutral (if used) are present and at acceptable levels. All circuit breakers and selector switches must be in the off position. The main circuit breaker or disconnect can then be turned on. One at a time, each of the other circuit breakers should be turned on, to check the load side of each for correct voltage. This includes the control circuit. With sufficient water in the well, turn each pump operating selector switch Hand-Off-Automatic (HOA) to the Hand position to run each pump. Amperage readings should be taken and recorded on each motor lead with a clamp-on ammeter. The The Pump Handbook Series

phase-to-phase voltage must be checked and recorded at this time. If there is excessive amperage draw on one leg, start troubleshooting by checking the manufacturer’s recommendations. Consult with the power company only after all other checks have been made. The initial readings for amperage, voltage, and ohmmeter resistance, plus megohm readings, should be the start of a permanent maintenance file. Monthly, quarterly or annual readings should be taken as part of a good preventive maintenance program. They are the basis for scheduled checks, which can indicate the trend of the motor current draw and can help prevent major outages and costly motor rewind jobs.

Operational Checks When start-up and testing is completed, the system is ready to be checked for automatic operation. Both or all of the pump operating selector switches should be turned to the auto position, and the alarms turned on or reset as required for normal operation. The best way to check and correctly set the on and off levels is to provide an external water supply source to fill the wet well to various predetermined levels. This water can be from a nearby hydrant or other source. Be careful when discharging water into the wet well not to affect floats or other controls. By controlling the fill and then observing or setting the operating points, the automatic cycle or cycles of operation can be checked, no matter what type of level controls are employed. To achieve second and/or subsequent pump start levels, the circuit breaker or disconnect can be turned off for the lead pump. This allows the wet well to fill, without running the lead pump, to the lag pump on level for checkout. All pumps may be turned off to enable checking and testing of alarm

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levels. The pumps should be permitted to pump, to allow checking of the desired off level. If a low alarm level is included below the normal off level, a pump may have to be run in the hand mode to check it. If guaranteed submergence with redundant low water cutoff is specified, it should be checked to be sure the wiring has eliminated the possibility of hand operation. The control system should be cycled more than once to check the proper and automatic alternation sequence of the pumps. If the panel is supplied with a manual alternator switch, it should be operated and the pumps cycled again to check for sequencing. Each run cycle, after alternation, should be checked to run both lead, lag and any subsequent pumps to ensure complete, correct alternation of the duty cycle.

Simulation Testing The following steps should be taken only in those extreme cases where sufficient water is not available and the pump(s) cannot be run in the well. Because the pump(s) are not run and the actual well water level is not used to energize the level sensing devices, this procedure is not recommended. If automatic controls are used with level sensor switches, they can be removed from the wet well and the controls manually actuated. Mark each sensor. With all controls in the off position, a low wet well level is simulated. By actuating the controls (either manually or by lowering them into a large container of water) one at a time, a rising wet well level is simulated. It is very important that the correct sequence is achieved when simulating a rising level and also as the floats are turned back to a vertical position or raised out of the container of water. When simulating automatic operation, enough water must be maintained in the wet well to enable the pump(s) to run full cycles without damage; otherwise, they should be turned off at their circuit breaker or disconnect switch. When the pumps are turned off, the starters for each pump and their

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run indicator pilot lights can be observed. If each pump control circuit is isolated by turning off its circuit breaker or disconnect, the pump leads may have to be disconnected, and the circuit breaker or disconnect left on to observe starter operation while testing. Purged air bubbler control systems can be checked through simulation by someone who fully understands their operation. The bubbler line to the wet well must be disconnected and a needle type cut-off valve installed on the line from the controls. Some units may have factoryinstalled shut-off and bleed valves. With the test valve completely open, the level gauge should read zero to simulate a low wet well level. As the test valve is slowly closed, creating a back pressure on the controls, the gauge will indicate the simulated level and the pumps should start as the required levels are reached. When the test valve is slowly opened, lowering the back pressure on the controls, the pumps should turn off. This procedure may be repeated as required to adjust operating levels and check alternation with each duty cycle. During testing, the actual wet well level should be checked to ensure the pumps do not pump so far down as to damage them.

Other Checks During the check-out of control systems, the pump run pilot light indicators, elapsed time meters and any other associated controls should be reviewed for proper operation. The alarm levels, when activated, should operate all devices in the system. These may include an alarm light, pilot light, audible horn or bell and a relay to provide remote signal contacts or telemetering contacts. All the alarm devices should be checked for proper operation, including audible alarm silencing circuits or switches and any reset push-button, if used. If any special control or alarm features are included in the system, they The Pump Handbook Series

should be checked by simulation if actual conditions cannot be achieved. These may include seal failure indicators, motor thermal sensors, or telemetered incoming or outgoing signals. Another device often used and not always checked is the phase monitor. Most have three functions—loss or low voltage on a single phase, loss or low voltage on all three phases and phase rotation reversal. By turning off the service voltage to the panel—disconnecting one of the sensing leads to the phase monitor and then insulating it when the power is turned back on to the panel—the pumps should not operate in either the hand or automatic mode. Reversing the above procedure, the lead should then be reconnected. This will check the loss of a single-phase feature. On some units, if the service nominal voltage is low enough, the set point on the monitor can be raised sufficiently above nominal voltage to cause the running pump or pumps to drop out. This will check the three phase low voltage setting of the phase monitor. By turning off the service voltage and reversing two of the sensing leads on the phase monitor (to induce improper rotation) and then restoring the service voltage, the pumps should not operate in either the hand or automatic mode. Reversing the above procedure, the leads should then be reconnected. This will check the phase reversal feature of the phase monitor. These checks are important, since they offer the protection needed during normal operation of the system. The last is important if the service is ever disconnected and reconnected, to prevent the pumps from being run in reverse. If a station has a power plug for the use of a portable standby generator, this reverse phase feature is essential to ensure that the generator phase rotation is matched with the normal service phase rotation.

SAMPLE MANUFACTURER’S START-UP REPORT FORM This report is designed to ensure the customer that customer service and a quality product are the number one priority. Please answer the following questions completely and as accurately as possible. Please mail this form to: (Manufacturer’s Name and Address)

1)

Pump Owner’s Name ________________________________________________

6)

Liquid Level Controls: Model __________________________________________

Address ____________________________________________________________

Is Control Installed Away from Turbulence? ___ Yes ___ No

Location of Installation ________________________________________________

Operation Check:

Person in Charge______________ Phone ________________________________

Tip lowest float (stop float), all pumps should remain off.

Purchased From______________________________________________________

Tip second float (and stop float), one pump comes on. Tip third float (and stop float), both pumps on (alarm on simplex).

2)

Model_________________Serial No. ____________________________________

Tip fourth float (and stop float), high level alarm on (omit on simplex).

Voltage Phase __________ Hertz ___________Horsepower ________________

If not our level controls, describe type of controls ________________________

Rotation: Direction of Impeller Rotation

Does liquid level ever drop below volute top? ___ Yes ___No

(Use C/W for clockwise, CC/W for counterclockwise) ____________________ 7)

Method Used to Check Rotation

Does Impeller Turn Freely by Hand ________ Yes

3)

Control Panel Model No.

____________________________________________

Number of Pumps Operated by Control Panel __________________________

(viewed from bottom) ________________________________________________ ________ No

NOTE: At no time should hole be made in top of control panel, unless proper sealing devices are utilized.

Condition of Equipment ______ Good ______ Fair ______ Poor

Control Panel Manufactured By Others: ________________________________

Condition of Cable Jacket ______ Good ______ Fair ______ Poor

Company Name ____________________________________________________

Resistance of Cable and Pump Motor (measured at pump control)

Model No __________________________________________________________

Red-Black ______ Ohms Red-White ______Ohms White-Black ______ Ohms

Short Circuit Protection_________Type __________________________________

Resistance of Ground Circuit Between Control Panel and

Number and Size of Short Circuit Device(s) ______________________________

Outside of Pump______Ohms

Amp Rating ________________________________________________________

MEG Ohm Check of Insulation:

Overload Type _____________________

Red to Ground________ White to Ground ________

Amp Rating _________________________________________________________

Black to Ground________

Do Protective Devices Comply With Pump Motor Amp Rating?

Size __________________________

___ Yes ___ No 4)

Condition of Equipment at Start-Up:

Are All Connections Tight? ___ Yes ___ No

Dry_______ Wet_______ Muddy________

Is the Interior of the Panel Dry? ___ Yes ___ No.

Was Equipment Stored: ___ Yes ___ No.

lf “No,” correct the moisture problem.

If YES, length of Storage:

___________________________________________

Describe Station Layout ______________________________________________

8)

Electrical Readings: Single Phase:

5)

Liquid Being Pumped ________________________________________________

Voltage Supply at Panel Line Connection, Pump Off,

Debris in Bottom of Station? ___ Yes ___ No

Ll, L2 __________________________

Was Debris Removed in Your Presence? ___ Yes ___ No

Voltage Supply at Panel Line Connection, Pump On,

Are Guide Rails exactly Vertical? ___ Yes ___ No

Ll, L2__________________________

Is Base Elbow Installed Level? ___ Yes ___ No

Amperage: Load Connection, Pump On, Ll__________________________ L2__________________________

Figure 1. Sample Manufacturer’s Start-up Report Form

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SAMPLE MANUFACTURER’S START-UP REPORT FORM (Page 2) 11) Manuals:

Three Phase: Voltage Supply at Panel Line Connection, Pump Off,

Has Operator Received Pump Instruction and Parts Manual?

Ll-L2 _______________ L2-L3 _______________

___ Yes ___ No

L3-Ll _______________

Has Operator Received Electrical Control Panel Diagram?

Voltage Supply at Panel Line Connection, Pump On, Ll-L2_______________

L2-L3 _______________ L3-Ll _______________

___ Yes ___ No

Amperage, Load Connection, Pump On,

Has Operator Been Briefed On Warranty? ___ Yes ___ No

Ll _______________

Name/Address of Local Representative/Distributor ______________________

L2 _______________ L3 _______________

__________________________________________________________________ 9)

Final Check: 12) I Have Received the Above Information. ___ Yes ___ No

Is Pump Seated on Discharge Properly? ___ Yes ___ No Was Pump Checked for Leaks? ___ Yes ___No

Name of Operator __________________________________________________

Do Check Valves Operate Properly? ___ Yes ___ No

Name of Company ________________________________Date______________

Flow: Does Station Appear to Operate at Proper Rate? ___ Yes ___ No

I Certify This Report To Be Accurate.

Noise Level: ___ Acceptable ___ Unacceptable

Signed By: (Start-Up Person) __________________________________________

Comments:

Employed By: _____________________________ Date: ___________________

________________________________________________________

Date and Time of Start-Up: ____________________________________________ 10) Describe Any Equipment Difficulties During Start-Up:

________________________________

Present at Start-Up:

________________________________________________________________________

(

) Engineer’s Name ______________________________________________

________________________________________________________________________

(

) Operator’s Name ______________________________________________

________________________________________________________________________

(

) Contractor’s Name ____________________________________________

________________________________________________________________________

(

) Others ______________________________________________________

Figure 1. continued

Final Tests After the above tests have been completed, the panel should be thoroughly checked to ensure that all wires have been reconnected properly, all switches or jumpers used for simulations have been removed, and all circuits restored to normal operation.

With the service voltage turned off, all wire terminations should be rechecked for tightness and the panel cleaned to ensure a good, maintainable environment. The panel should now be ready for continuous automatic operation. The run time on elapsed time meters should be read and recorded for each pump. These readings should be the start of a permanent maintenance file. Periodic readings should be taken and recorded as part of a good preventive maintenance program. With these readings, the alternating duty cycles can be checked to

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ensure operation of the alternator. Many pump and motor maintenance schedules are based on hours of operation. Actual automatic cycles are monitored to be sure they are not excessive. For submersible motors, a maximum of 10 to 15 starts per hour are acceptable in terms of equipment life. Over time, this record may show system demand trends and indicate the need for larger pumps. By calculating the pumping rate of each pump and the hours of run time, you can estimate the gallons of water pumped for a given period. This might indicate problems with increased water infiltration into the system due to cracked or broken lines, based on an increase in gallons pumped with no other substantial changes within the system. Make periodic checks during the first few days of operation of a new The Pump Handbook Series

system. This can uncover unforeseen problems and help the operator become familiar with duty cycles and other characteristics of the pumping station. This special attention can help avoid future problems. ■ This article was adapted from the Submersible Sewage Pumping Systems (SWPA) Handbook, published by the Submersible Wastewater Pump Association, 1866 Sheridan Road, Suite 210, Highland Park, IL 60035. Phone: (847) 681-1868, Fax: (847) 681-1869, E-Mail: [email protected]. SWPA is a national trade association representing and serving manufacturers of submersible pumps for municipal and industrial wastewater applications. Founded in 1976, the association’s primary focus is on industry guidelines, education and promotion.

CENTRIFUGAL PUMPS HANDBOOK

Successful Submersible Operation Part II: Inspection And Maintenance Upkeep and regular attention keep these workhorses of the wastewater industry running smoothly. By Submersible Wastewater Pump Association

egular inspection and preventive maintenance ensure continued, reliable operation of the entire submersible pumping system. All stations, pumps and operating equipment should be inspected at least once a year—more frequently under severe operating conditions. One of the major advantages of a submersible station is the ability of the service technician to handle most maintenance and service onsite, without entering the wet well. All equipment in the station should be backed by manufacturers’ service manuals. This material should be carefully read, filed and consulted whenever servicing is required.

Safety Precautions

• Be aware of the risk of electrical accidents. • Check the explosion risk before welding or using electric hand tools in or near the station. Never weld or use electrical tools in the wet well after it has been in operation. • Make sure that all lifting equipment, when used, is in good condition and of adequate capacity. • Provide a suitable barrier around the work area—for example, a guard rail. • Make sure that all personnel have a clear path of retreat. • Use safety helmets, safety goggles and protective shoes or boots. • All personnel working with sewage systems must be vaccinated against any diseases that can occur.

CAUTION: Note and read all safety precautions before performing any operation or maintenance procedure. To minimize the risk of accidents in connection with service work, the following rules—as well as all applicable laws, regulations and manufacturers’ recommendations—must be followed: • Be aware of health hazards. Observe strict cleanliness.

Because sewage pumps are designed for use in liquids that can be hazardous to one’s health, make sure that all equipment has been thoroughly cleaned. To prevent injury to the eyes and skin, observe the following rules: 1. Always wear goggles and rubber gloves. 2. Wash and rinse pump thoroughly

R

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with clean water before starting work. 3. Wash and rinse any components in clean water after disassembly and then dry thoroughly. If you get hazardous chemicals in your eyes, rinse them immediately with running water for 15 minutes, and hold your eyelids apart with your fingers. Contact a doctor immediately. If you get hazardous chemicals on your skin, remove contaminated clothes, wash your skin with soap and water; seek medical attention immediately.

Recommended Inspections CAUTION: Before starting work on any pump, make sure it is isolated from the power supply and cannot be energized. This applies to the control circuit as well. One method is to tag and lock the control panel to let other personnel know that you are working on the station. Keep in mind that some systems have an override switch at the treatment plant or other buildings. Make sure that this switch is also off

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Problem 1. Pump will not start.

2.Pump will not start and overload heaters trip.

3. Pump runs but will not shut off.

4. Pump does not deliver proper capacity.

5. Motor stops and then restarts after short period, but overload heaters in starter do not trip.

Possible Cause

Remedy

No power to motor.

Check for blown fuse or open circuit breaker.

Selector switch may be in off position.

Turn to on position.

Control circuit breaker may be tripped.

Reset the circuit breaker.

Overload heater in starter may be tripped.

Push to reset.

Overload heater in starter may be burned out.

Replace the heater.

Unit may be improperly grounded.

Turn power off and check motor leads with megger ohmeter.

Motor windings may not be balanced.

Check resistance of motor windings. If it is three-phase, all phases should show the same reading.

Impeller may be clogged, blocked or damaged.

If no grounds exist and the motor windings check out satisfactorily, remove the pump from the well and check for impeller blockage.

Pump may be air-locked.

Turn pump off for several minutes, then restart.

Lower level switch may be locked in closed position.

Check to be certain the level control is free.

Selector switch may be in the "hand" position.

Switch selector to "auto" position.

Discharge gate valve may be partially clogged.

Open and unclog valve.

Check valve may be partially clogged.

Valve must be cleared—if there is an outside lever, move it up and down.

Discharge line may be clogged.

Use a sewer cleaner or high-pressure hose to clear the obstruction.

Pump may be running in the wrong direction.

Low speed pumps can operate in reverse with little noise or vibration. Check your manual for methods of establishing and correcting rotation.

Discharge head may be too high.

Check total head with a gauge when pump is operating. Compare against original design and precious operating records. If pump has been in service for some time and capacity falls off, remove the pump and check for clogged impeller.

Heat sensors in the motor may trip due to excessive heat.

Impeller may be partially clogged, resulting in the sustained overload, though not high enough to trip the overload heater switch.

Motor may be operating out of liquid due to failed level control.

Check location and operation of level controls.

Pump may be operating on a short cycle.

The wet well may be too small or water may be repeatedly returning to the well due to a leaking check valve. Both must be checked.

Table 1. Troubleshooting checklist for pumps (These are general guidelines only. Consult the specific manufacturer’s manual for detailed instructions.)

and tagged at the other building before you start working on the station. After the pump(s) have been isolated from the power supply and pulled to the top of the station, the appropriate manufacturer’s service manuals should be consulted and the following inspection guidelines adhered to.

Visible Parts on Pump and in Station 1. Check for vandalism or other station damage.

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2. Make certain the access cover works properly. Check the holdopen device to ensure that it is engaged. 3. Make certain that the guide rails are completely vertical. 4. Check the condition of the lifting eye, chains, hooks and wire ropes. 5. Make certain that all screws, bolts and nuts are tight. 6. Replace or repair worn or damaged parts.

The Pump Handbook Series

Pump Casing and Impeller 1. If clearance between the impeller skirt and the pump casing or wear ring exceeds manufacturer’s recommendations, it may be necessary to adjust the impeller or replace the wear rings. 2. Wear on the outlet flange from the pump casing usually causes corresponding wear on the discharge connection. 3. Follow the manufacturer’s instructions for disassembly, inspection, and reassembly of the impeller and

Problem

Possible Cause

Remedy

Service voltage is not on to the panel.

Turn on and check for proper voltage.

Main or control circuit breakers tripped or turned off. Main or control circuit fuses blown.

Turn on or reset and turn on all circuit breakers. Check and replace blown fuses.

Motor heat sensor connections not made properly.

Check motor heat connections and correct to panel.

2. Pumps 1 and 2 will not run in "hand" position. Run lights are on.

Motors not wired properly.

Check motor heat connections and correct to panel.

Voltage starter coils are the wrong ones.

Check and correct starter coils voltage rating to match control circuit voltage.

3. Pump 1 or 2 will not run in "hand" position. Run light not on. One pump operates in "hand" position.

Pump circuit breaker tripped or turned off.

Turn on or reset breaker.

Pump circuit fuse is blown.

Check and replace any blown fuses.

Motor starter overload tripped.

Reset overload after checking motor.

Motor heat sensor circuit open or not properly connected.

Check continuity of motor heat sensor. Correct connections.

Level in wet well not high enough to turn on pumps.

Fill or allow wet well to fill to required levels.

Level float switches may be incorrectly connected or have failed.

Check and connect each float correctly; replace if required.

Air bubbler supply may be off or failed.

Check and ensure air supply is on, bubbler line is working in wet well and has no leaks.

Pressure switches or sensors may not be adjusted or sequenced properly.

Check and adjust pressure switches to correct levels and sequence.

Relay or other control device failed.

Check and replace any control relay, alternator or other device with defective coil or contacts.

No probable panel problem.

None.

Possible system problem, for example, clogged discharge line.

Check and clear check valve or line of obstruction.

Temporary high level condition after power failure or influent surge.

Monitor station operation until "high level" is reduced.

6. Alarm light and/or audible alarm turns on, with one or both pumps not running.

Test hand operation of pump that is not running and refer to Problem 3.

See remedies for Problem 3.

Refer to Problem 4.

See remedies for Problem 4.

7. Circuit breaker tripped for motor power.

Motor not wired properly.

Check and correct connection to panel.

Short in pump cable, wiring or motor.

Disconnect motor and check wiring. Check motor for shorts or grounds.

Size of breaker too small and/or ambient heat problem.

Check and correct breaker size for motor and/or provide ventilation or compensation for ambient heat.

8. Blown fuse for motor power.

Same as Problem 7.

See remedies for Problem 7.

9. Pumps do not alternate. Same as Problem 7.

Alternator relay may be defective.

Check and replace alternator.

Improper sequencing of float switches or pressure sensors.

Check and correct sequence of controls to ensure off, lead, lag sequence.

Pilot light bulb has failed.

Replace lamp with correct voltage replacement.

1. Pumps 1 and 2 will not run in "hand" or "auto" positions.

4. Pumps 1 and 2 will not run in "auto" position.

5. Alarm light and/or audible alarm turns on with both pumps running.

Table 2. Troubleshooting checklist for control panels (These are general guidelines only. Consult the specific manufacturer’s manual for detailed instructions.)

volute. When it is disassembled, check the motor shaft, impeller and volute bore for wear or damage. 4. Follow the manufacturer’s instructions for disassembly, inspection

and reassembly of the shaft seal. It must be clean and properly seated before reassembly. 5. Always replace worn or damaged parts. The Pump Handbook Series

Electrical Insulation Perform a megger (insulation resistance) test between the pump motor leads and the pump casing. A low— 20 megohms or less—reading indicates

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moisture entry into the motor chamber or power cord, or other deterioration of the insulation system. Such problems should be corrected before a major breakdown occurs. Oil Quantity and Condition A. Oil-Filled Motors CAUTION: If there has been any leakage, the motor housing may be under pressure. Hold a rag over the inspection plug to prevent splatter when loosening the plug. Check the oil in both the motor housing and the seal cavity through the oil inspection plugs. The oil level may be low or emulsified (creamlike), which indicates that water has entered the cavity and a leak is present. One possible cause is an inspection plug that is not sufficiently tight. Check the sealing surface of the motor housing, the cable entry and the condition of the shaft seal. Whatever the problem, correct it and make certain that the oil is refilled to the proper level with the motor manufacturer’s recommended oil. B. Non Oil-Filled Motors If there is any liquid in the motor housing, a leak is present and all sealing faces should be checked as listed above under “Oil-Filled Motors.” Cable Entry 1. Make certain that the cable connection is tight. 2. If the cable entry leaks, it may be necessary to replace the cable seal. See the manufacturer’s manual for instructions. 3. When refitting a cable that has been used before, even when the jacket has not been damaged, always cut off a short piece of the cable so that the cable entry seal does not close around it at the same point. 4. If the outer jacket of the cable is damaged, replace the cable. Make sure the cable has no sharp bends and is not pinched.

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Controls 1. Check liquid level sensors throughout their entire range of operation. Clean, adjust, replace and repair damaged equipment. Follow the manufacturer’s instructions. 2. The same procedure should be used for checking the balance of the control system. In particular, check signals and the tripping function, and make sure that the relays, lamps, fuses and connections are intact. Replace all inoperative equipment. Piping and Valves Repair all flaws and replace inoperative equipment. Fault Tracing A voltmeter, ohmmeter, and ammeter—together with the job wiring diagram—are required to test, measure and carry out fault tracing on electrical equipment. Fault tracing must always be performed with the power supply disconnected and locked off, except for those checks which can be performed only with power. Electrical work must be performed by qualified electricians and all applicable local, state and national safety regulations followed. Observe the recommended safety precautions previously mentioned in this article.

Major Servicing Submersible sewage pumps can be serviced in the field at qualified facilities. If the pump is still under warranty, it should be serviced by an authorized shop. Manufacturers’ service manuals provide detailed instructions for the replacement of impellers, stators, seals and bearings. To facilitate field maintenance and service, many manufacturers provide a list of authorized service facilities, recommended spare parts and the maintenance equipment required. If the motor carries an approval label such as Underwriters Laboratories (UL) or Factory Mutual (FM), in order to The Pump Handbook Series

retain this approval label, major motor repair must be performed by an authorized UL or FM repair facility. ■ This article was adapted from Submersible Sewage Pumping Systems (SWPA) Handbook, published by the Submersible Wastewater Pump Association, 1866 Sheridan Road, Suite 210, Highland Park, IL 60035. Phone: (847) 681-1868, Fax: (847) 681-1869, E-Mail: [email protected]. SWPA is a national trade association representing and serving the manufacturers of submersible pumps for municipal and industrial wastewater applications. Founded in 1976, the association’s primary focus is on industry guidelines, education and promotion.

CENTRIFUGAL PUMPS HANDBOOK

Precision Solutions for Low-Flow Handling When you need less than 100 gallons per hour in increments as small as milliliters, metering pumps are the answer. By Ken Gibson, ProMinent Fluid Controls, Inc. he world of low-flow applications is very different from that of high flow counterparts. It is a very precise field, and, over the years, has experienced many changes and developments. Low-flow handling can be a very critical step in such applications as water treatment and chemical injection. To ensure that it is accomplished well, users must be aware of the process involved, and they must choose the correct tool(s) for the job. A low-flow application normally involves the injection of a small amount of fluid into a process. Since the term “lowflow” can mean different things to people, this article will concentrate on applications that are less than 100 gallons per hour. The injection of process chemicals into applications takes an accurate and precise metering pump. Besides the pump itself, there are certain accessories that should be considered to enhance its performance. This article will explore the area of low-flow applications using metering pumps and how they can be done simply and correctly.

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Low-Flow Pump Designs There are many low-flow pump types to choose from; they come in all shapes and sizes. Some popular versions are the reciprocating

1. Plunger 2. Packing 3. Flushing 4. Piston shaft 5. Valves 6. Flushing connector

Figure 1. Packed plunger metering pump

diaphragm, the packed-plunger/piston (Figure 1) and the peristaltic. The reciprocating diaphragm pump can be actuated by one of the following mechanisms: • solenoid • mechanical gear • hydraulic fluid • air Solenoid-driven pumps are some of the most widely used in low-flow situations (Figure 2). Some models can accurately meter flows well below one gallon per hour (gph), while other models can exceed 20 gph. The mechanical gear pump is a higher flow pump and can easily hit flows higher The Pump Handbook Series

than 600 gph. The hydraulicallyactuated pump is capable of high pressures and low flows. Air-driven pumps can be useful, since many applications have pressurized air available. Metering pumps have come a long way in the past ten years, especially the solenoid versions. The better models are microprocessor-based (kind of like having a mini-computer in your pump) and are extremely accurate and very dependable. They can operate manually or by an external source such as a pulse contact or a 4-20 mA analog signal. Fault relays also are common. They notify the

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Cutaway view of ProMinent Gamma/L solenoid-driven metering pump 4

5

6

8

9

10

12 2 11 3

7

1

1. 2. 3. 4. 5. 6. 7. 8. 9. 10.

Housing Liquid end Diaphragm Backplate Solenoid Solenoid coil Solenoid axle Armature Cover Stroke adjustment screw 11. Stroke adjustment axle 12. Stroke adjustment knob

Figure 2. Solenoid-driven metering pump

operator when the pump has experienced a problem such as loss of power or low chemical level. Many models have digital displays that show stroking frequency and stroke length. One manufacturer just released a model that displays flow in gallons/hour or liters/hour and gives a totalizing count of the amount of chemical pumped over a period of time. All of these features enable the operator to handle low-flow chemical feed with fewer hassles and more confidence. Applications and Pump Accessories Examples of low flow applications requiring a metering pump include industrial/municipal water and wastewater treatment, cooling towers, boilers, reverse osmosis, swimming pools, car washes, chemical process and laboratories. Industrial customers include food and beverage, pulp and paper, semi-conductor, chemical manufacturers, pharmaceutical and metal finishing/plating. Municipal customers include your local waterworks authority, where drinking water and subsequent wastewater are treated. If you visited these facilities and spent some time walking around, you might be surprised at the number of metering pumps clicking away.

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Depending on the application and space allowance, pumps can be mounted on chemical drums, prepackaged and mounted onto plastic or steel skids, or tucked away in areas of the plant where they are forgotten until a flow rate changes or a failure occurs (Figure 3). Regardless of location, the installation should consider flow rates, backpressure and the position of the pump in relation to the material being fed. This is one of the reasons accessories are essential to ensure proper feed and prohibit costly downtime (Figure 4). Foot Valves If you are pulling liquid up from a drum or tank, a foot valve is recommended. A foot valve is a check valve that enables the pump to pull liquid up into the suction line. At the same time, it prevents fluid flow back into the tank once the pump is stopped. This helps keep the suction line primed and makes it easier on the pump when it is turned back on. Injection Valve An injection valve is used on the discharge side of the pump to provide a little backpressure and to inject liquid into a water stream or tank. The injection valve can function like a quill and be secured into a pipe. This enables the liquid to be injected The Pump Handbook Series

directly into a water stream and not dribble down the sides of the pipe, which could cause corrosion. Backpressure Valves In any diaphragm-design positive displacement metering pump, a constant backpressure is recommended. This enables the pump to constantly work against the same pressure and not be subjected to the changing pressures that typically occur in processes. A backpressure or loading valve will do the trick, enhancing metering pump accuracy and repeatability. These valves are normally spring-loaded diaphragm valves that are adjustable up to about 150 psi. The backpressure valve normally is located close to the injection point of the liquid.

Figure 3. Many different arrangements for metering pumps are available. This particular pump is mounted on the tank it is drawing from.

Pressure Relief Valves The pump must be capable of overcoming both the valve and the system pressure. Pressure relief valves also are used in many applications. Similar to the backpressure valve in design, the pressure relief valve has an extra port that enables the pressurized liquid to be diverted in the event that the pump experiences overpressure. This valve protects the user’s investment against the metering pump

Metering monitor

Cable

Foot valve

Multifunction valve

Injection valve

Figure 4. Some important metering pump accessories

operating outside of its designed pressure rating, such as closing a valve on the discharge side of the pump, also known as a “dead head.” Pulsation Dampeners Another very important accessory for low-flow pumps is a pulsation dampener. This device actually dampens the pulsating flow of the pump, resulting in a more laminar-like flow. It acts much the same way as a surge tank does in a water system, preventing waterhammer, which will cause vibrations that result in damage to equipment, with untimely and avoidable expense. Pulsating flow of a reciprocating metering pump can be equated to waterhammer. A pulsation dampener also is recommended when the discharge line of the pump is extremely long—25 feet or more. Since the pump has to push against all of the fluid contained in the long pipe length, the dampener will hold a certain amount of fluid and then push the fluid through the pipe. This takes a lot of the load off the metering pump by reducing the frictional losses associated with pushing the column of fluid through the discharge piping. For installation purposes, the pulsation dampener should be close to the discharge of the pump, followed by the pressure relief valve.

Multi-Function Valves Another helpful accessory is the multi-function valve. It is a combination valve that has the properties of backpressure, antisiphon and pressure relief. The valves are normally installed directly on the discharge valve of the pump.

Metering Monitors A metering monitor is an accessory that enables the operator to see if there is flow going through the pump. The monitor runs off the pump’s electrical power and detects each pump stroke. The monitor can be programmed to shut the pump off if there is no fluid being transferred. Pressure Gauges A simple device, often overlooked but very important in the information that it provides, is a pressure gauge installed on the discharge side of a metering pump. Pressure gauges are standard operating tools in hydraulic processes under pressure. A metering pump application is a hydraulic system under pressure. The gauge will assist the operator in adjusting to, and determining the loading set points of both the pressure relief and backpressure valves. It will also give an indication as to whether the metering pump is operating within its designed maximum pressure specification.

Photo 1. These ProMinent Fluid Controls metering pumps are injecting phosphate at a water treatment facility. The Pump Handbook Series

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3-way valve (for self-fill calibration column) Isolation valve (for direction flow) Backpressure valve SCR drive Pressure gauge Pulsation dampener Calibration column Pressure relief valve

Motorized diaphragm metering pump Y-strainer

Figure 5. Pump manufacturers now have pre-packaged pump systems with a backup pump.

3-way valve (for self-fill calibration column) Backpressure valve

Pulsation dampener Pressure gauge Calibration column Pressure relief valve Solenoid diaphragm metering pump

Figure 6. A complete metering pump system package

Calibration Column One additional tool, which will assist the operator in providing precise outputs from metering equipment in low-flow applications, is the calibration column. This graduated column, scaled in milliliters or gallons, allows

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for determination of metering pump output under operating conditions. It also enables precise adjustments to pump output in a safe and reliable manner. Some or all of these accessories may be needed to ensure precision and The Pump Handbook Series

accuracy in low-flow situations. It is best to check with the pump manufacturer to select the proper accessories for your metering application. There are also software programs available that can help you select the proper accessories.

Maintenance and Operation In order to keep your low-flow metering application running smoothly, you should set up a maintenance schedule. Most diaphragm metering pumps are low maintenance to begin with. They are normally durable and consistent if they are treated properly and maintained. For solenoid pumps, there should be a routine check on the suction and discharge valves, as well as the diaphragm. If the valves are worked hard, they may need to be replaced due to the constant pounding of the valve balls on the ball seats. Also, O-rings become worn and some are subject to chemical attack of the liquid being metered. A good durable diaphragm will have a steel core center encased in Teflon® and an elastomer backing. This type of diaphragm can last for years if the pump itself is properly maintained. If the pump needs to be repaired, the cost is usually minimal. In a situation where a pump goes down, it is a good practice to have a spare one ready to go. If a spare is not available, some manufacturers have a “loaner program” where they will loan the operator a pump while the broken pump is being fixed at the factory. Still, other end-users prefer to throw the broken pump away and purchase a new one. This practice is a waste of time and effort that could be prevented if a better-quality pump were being used. Most metering pumps have a standard one to two year warranty that covers most parts and labor—as long as the pump is not abused. Normally the pump is sent back to the manufacturer for evaluation. Before shipping a used pump, it is extremely important to flush the liquid end thoroughly for safety reasons. After a thorough check, it then can be determined if the problem was due to manufacturer’s defect, chemical attack or abuse. The best scenario is to have a backup pump or system. Some pump manufacturers now offer pre-packaged pump systems with a backup pump (Figure 5). If the

main pump goes down for any reason, the backup pump will automatically start up and continue where the other one left off. This is a wise choice when metering a chemical into a critical process such as sodium hypochlorite for disinfection, or hydrofluosilicic acid for fluoridation in drinking water. Fluoridation is an excellent example of the need for a precision metering pump in a low-flow application. In high concentrations, fluoride ingestion is toxic to one’s health. In low concentrations in drinking water (11.5 mg/liter), fluoride is effective in minimizing tooth decay. The packaged systems can be simple—such as a pump and calibration column, or with most of the accessories mentioned previously. A complete metering pump system package will offer the greatest opportunity for providing safety, accuracy, and reliability (Figure 6).

With low-flow situations increasing (better control—lower costs), we now have a better way to handle them and keep all parties happy. ■ Kenneth J. Gibson is the Industrial Market Manager for ProMinent Fluid Controls, Inc. He has been with the company five years and also conducts product training for ProMinent authorized representatives. Prior to this position, Ken was an Equipment Product Manager for Calgon Corporation, where he worked for 20 years. Ken holds a B.S. degree in Chemistry from the University of Pittsburgh.

Metering Goes High-Tech Technology has certainly helped the metering pump industry and subsequently, low-flow applications. Solenoid pumps now come equipped with microprocessors that control all electrical functions in the pump. Control options such as analog (4-20 mA typically), fault relays, timers, flow monitoring and batch counters enable the operator to set up an entire feed application with just a keystroke. New technology has redefined the features and benefits of the metering pump, saving time and providing vital information and precision pumping. Motor-driven pumps have also jumped on the technology train. Some models come with built-in microprocessors, enabling some of the same functions found in the solenoid models such as analog control and fault relays. Pumps are now smarter and better than they were in past years. While some end-users still prefer the old technology, there are new ways of handling low-flow applications—specifically with metering pumps and innovative technology. The Pump Handbook Series

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CENTRIFUGAL PUMPS HANDBOOK

Key Centrifugal Pump Parameters and How They Impact Your Applications – Part 1 Pumps and Systems: They Go Together By Doug Kriebel, PE, Kriebel Engineered Equipment

he purpose of this article is to familiarize the reader with centrifugal pump manufacturers’ performance curves, show how a pump system head curve is constructed and demonstrate how to analyze the effect of changes in the system. A pumping system operates where the pump curve and the system resistance curve intersect. This important fact is one of the most overlooked concepts when designing or troubleshooting a pumping system. The pump curve has a head/capacity portion, which indicates the change in flow with respect to head at the pump. It is generated by tests performed by the pump manufacturer.

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Figure 1. Performance curve data

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We will review the key facts that this curve offers. The system resistance curve is the change in flow with respect to head of the system. It must be plotted by the user based upon the conditions of service. These include physical layout, process conditions and fluid characteristics. Anyone operating a pumping system must know where the pump is operating. Otherwise, it is like driving a car and not knowing how fast it’s going.

The Pump Curve

Differential Head (H is expressed in feet of pumped liquid). This is the difference between the Total Discharge Head (Hd) at the pump discharge and the Total Suction Head (Hs) at the pump inlet at any given flow rate (Q), expressed in gpm (gallons per minute). H = Hd-Hs (Eq. 1) All pressure is in feet of liquid. At this point, it’s important to revisit some pressure units:

The pump curve consists of a head/capacity curve (curve A of Figure 1) that shows the Total

Figure 2. Pressure equivalent The Pump Handbook Series

psi = pounds/in2 psia = pound/in2 absolute

specific gravity must be considered. BHP = Q x H x S.G./(3960 x % eff) (Eq. 2) The NPSHr is read from the bottom curve from flows and the NPSHr in feet is read on the lower right. The following examples apply to Figure 3. Example 1 For an impeller diameter of 15-15/16” (maximum), what are the head, efficiency, NPSHr and BHP at 900 gpm? (Answer: 110’; 76.5%, 5’ NPSHr and 32 BHP) Example 2 What are the head, efficiency, NPSHr and BHP at 1600 gpm? (Answer: 85’, 78%, 7.5’ NPSHr and 44 BHP)

Figure 3. Manufacturer’s performance curve

psig = pound/in2 gage psia = psig + barometric pressure The conversion from psi to feet of liquid is to divide the pressure, in psi, by its density: (lbs/in2)/(lbs/ft3) x (144in2/ft2) = feet of pumped fluid For water with a specific gravity (S.G.) of 1, you substitute its density (62.4 lbs/ft3) into the above equation and the result is: psi x 2.31/S.G. = feet of liquid. (S.G. is the ratio of the density of a liquid to water at 60°F.) See Figure 2 for illustration of feet, psia/psig/S.G. The pump curve also includes the pump efficiency (in percent–curve B), BHP (brake horsepower–curve C) and NPSHr (Net Positive Suction Head required–curve D) at any flow rate. This curve is plotted for a constant speed (rpm) and a given impeller diameter (or series of diameters). Figure 1 is a single line curve; Figure 3 is an actual manufacturer’s

performance curve showing the same information, but expressed as a series of curves for different impeller diameters (from 12-5/8” to In a pumping system, the capac15-15/16”). The pump impeller can ity of the pump moves along the be trimmed to any dimension pump curve based upon the differbetween the maximum and miniential head across the pump. For mum diameters. instance, for Figure 3, with the The H, NPSHr, efficiency and maximum impeller diameter, if the BHP all vary with flow rate, Q. The flow is 900 gpm and you want to best efficiency point (BEP) is the increase it, you reduce the head to point on the curve where the effi85’ and the flow will increase to ciency is at a maximum. For Figure 3, it’s 82% at 1300 gpm and 100’ TDH. BHP can be read from the curve at any flow rate; however, it is usually calculated from the efficiency, head and flow using Eq. 2. Pump curves are based on a specific gravity of 1.0. Other liquids’ Figure 4. Head vs. capacity vs. Ns The Pump Handbook Series

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1600 gpm. It’s by controlling the pump differential head that you control the flow of the pump. If you wanted 1300 gpm, you would have to throttle the head back to 100’. The slopes of pump curves vary depending on the pump specific speed (Ns). How they vary depends on Ns and is shown in Figure 4. The Ns is a dimensionless number calculated as: Ns = rpm x (Q1/2)/ H3/4 all at BEP (Eq. 3) The Ns for Figure 3 is: 1311 NPSHr is Net Positive Suction Head required by the pump—based on test. NPSHa is Net Positive Suction Head available by the system—must be calculated. NPSHa = P1 + hs- hvp - hfs (Eq. 4) P1 = pressure of suction vessel (atmospheric in open tanks) hvp = vapor pressure of liquid at its pumping temperature, t hs = static suction head, the vertical distance between the eye of the first stage impeller centerline and the suction liquid level. (This may be a positive or negative number, depending on whether the suction liquid is above the pump suction eye or below it, as on a “lift.”) hsf = friction and entrance losses on the suction side Pump curves are generated by manufacturers and show a range of impeller diameters from minimum to maximum size for the casing size selected. These are based on a single speed (rpm). Head/capacity relationship changes with speed or impeller diameter per the affinity laws. Affinity laws are discussed in all the references. Other information often found is suction specific speed (Nss), which is a dimensionless number calculated at the BEP as:

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Nss = rpm x (gpmeye)1/2/ NPSHr3/4 (Eq. 5) (Note: The gpm is per impeller eye, so for a double suction impeller, the flow is half the flow of single suction impeller.) The Nss in Figure 3 is 10,800. Nss is a function of NPSHr. To reduce NPSHr, the impeller eye is modified. If the eye is modified too much, it causes a reduction in the low flow capabilities of the pump and limits the low end point of operation.

System Head Curves (Figure 5) Step one in constructing a system/head curve is to consider the fluid energy requirement. At any point in a pumping system, there are four components making up its total energy (head in feet): • pressure • velocity • elevation • heat (ignored, since pumping is usually at a constant temperature across the pump)

of the pipe (velocity), roughness of the pipe’s internals, length of travel, and losses at fittings, control valves, and the entrance and exits. H system = Hd-Hs = (P2-P1) + (hd-hs) + (Vd2-Vs2)/2gc + (hfs + hfd)

(Eq. 6) Where: P1 and P2 = the entrance and exit pressures hd and hs = the elevations from the pump impeller eye to the entrance and exit liquid levels Vs and Vd = the entrance and exit velocities (velocities are given in ft/sec in this article) hfs and hfd = the frictional losses in the suction and discharge pipe system, expressed in feet To simplify: Hf = hfs + hfd is the total friction loss (at each flow rate) through the system (Eq. 7) h = hd- hs is the difference in elevation from discharge level to suction level (Eq. 8)

The difference in energy levels (differential head, H) between the suchv = Vd2-Vs2/2gc is the velocity tion (Hs) head and discharge (Hd) head component head of a pump can be expressed (Eq. 9) using Bernoulli’s equation and adding the friction loss. The friction loss is the loss in energy due to resistance to flow as it moves through a system. This resistance is due to viscous shear stresses within the liquid and turbulence at the walls of the pipe and fittings. It varies with liquid properties (viscosity, solids entrained), size Figure 5. A typical pump system The Pump Handbook Series

Velocity head is often ignored, since in most cases the velocity is low (in tanks maintained at constant level, it’s 0). However, if the discharge is coming out through a nozzle or other orifice, its velocity could be high and, therefore, significant if the other components are low. This should not be confused with entrance and exit head losses, which are accounted for in the Hf calculations. Velocity head (hv = V2/2gc) is a confusing concept to many—it’s the head (distance) required for a liquid to fall in order to reach a velocity, v. It costs energy to move (accelerate) from some low velocity to a higher velocity. All components of Eq. 6 must be consistent and usually in feet of liquid. Refer to Figure 5 for the following example. For the Pressure Component

It could have been:

Use either; just be consistent. For open tanks there is 0 psig, 14.7 psia (at sea level). For the Static Head Component h = hd-hs = 60-10 = 50’ Note that this is from liquid level to liquid level. This layout was for illustration; most industrial applications do not have the discharge pipe located beneath the liquid level because it would back siphon when the pump stops. If the pipe goes overhead and empties above the liquid line (or more often uses a vacuum breaker), use the centerline elevation of the discharge pipe at the exit for the hd. The pressure and static head components are constant with flow and plot as a straight line.

For the Velocity Head Components Assume the tanks are either being filled/emptied at the pumping rate or they are large enough that the downward and upward velocities are very low, or they are the same diameter and the velocity heads cancel. Therefore, the velocity head component is 0. The Friction Component Hf includes pipe losses and minor losses, as follows. Pipe Losses There are several ways to calculate these losses: 1. Published data such as the Crane Handbook, the Cameron Hydraulic Data Book or KSB’s Centrifugal Pump Design. 2. Hazen Williams formula: Hf = [3.022 (v)1.85 L]/[C1.85 x D1.165 ] (Eq. 10) Where: L = length of pipe in feet v = velocity ft/sec D = pipe diameter in feet C = constant for various pipe materials This is a simple way to go. The disadvantage is that it assumes constant NRe (Reynolds Number) in the turbulent range and a viscosity of 1.13 cs (31.5 SSU), which is water at 60°F. This can yield errors of 20% or more for the range of 32 to 212°F. 3. Using the Darcy formula: Hf = (Eq. 11) fLv2/2Dgc Where f = the friction factor based on Reynolds Number (NRe) NRe = D x v/vis (Eq. 12) Where: vis = viscosity in ft2/sec In laminar flow (less than Nre = 2000) f = 64/Nre (Eq. 13)

The Pump Handbook Series

In turbulent flow (NRe greater than 4000), f varies with Nre and the roughness of the pipe. This is expressed as “relative roughness” = e/D; where e is absolute roughness (in feet) found in tables and f is found in a Moody Friction Factor Chart using NRe and e/D. If the range of NRe is between 2000 and 4000 (the transition zone), use turbulent flow design. 4. Using computerized flow simulators. (You need to know limits and assumptions upon which the program is based.) Minor Losses These include obstructions in the line, changes in direction and changes in size (flow area). The two most common methods of calculating these losses are: 1. Equivalent lengths: uses tables to convert each valve or fitting into an “equivalent length,” which is then added to the pipe length in L, and Hf is calculated in Darcy (Eq. 11 above). 2. Loss coefficients: uses published resistance coefficients (K) in the formula: hf = Kv2/2gc. K values for many types of fittings are published. Loss coefficients are the way to determine entrance and exit losses and losses due to sudden reductions or enlargements. Control Valve Losses These losses are calculated using the Cv for the valve: Cv = gpm pressure drop (psi) The Hf increases as flow increases and is a parabolic curve. The Hf for the system is calculated using one of the above methods. Assume the suction pipe is 10” diameter, 20’ long, with two elbows and one gate valve. The discharge pipe is 8”, 1000’ long with eight elbows, one gate valve and one check valve. The calculations yields Hf values of: 1. Hf = 4’ @ 400 gpm

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2. Hf = 12’ @ 800 gpm 3. Hf = 27’ @ 1200 gpm 4. Hf = 37’ @ 1400 gpm With the system head calculated at several flows (the pressure and height components are not affected by flow), we can plot the results (Figure 6). The Total Differential Head at 1200 gpm is H = 11.5’ + 50’ + 27’= 88.5’. The pump’s head capacity curve is plotted on the same graph. Where they intersect is the operating point (“A”). Note the pressure and static head components (66.5’) are constant with flow, while the Hf component increases as flow increases. If a control valve in the discharge piping were used, the system head curve could be throttled back to a higher head and therefore lower flow (“B”).

What Can Go Wrong? A centrifugal pump will run where the system head curve tells it to run. The point where the centrifugal pump performance curve and system head curve intersect is the operating point. It’s important to understand where the operating point is at all times and to avoid the dangerous areas of pump operation. What are some of the dangerous areas to consider? Horsepower A pump must always have a driver with more hp than it consumes. With most centrifugal pumps, as the pump flow increases, the BHP increases to a maximum, and there must always be enough hp available from the driver to supply the needs of the pump. This means as TDH falls, the flow goes up and so does the BHP. Under-sizing overloads the driver. NPSH As the pump flow increases, the NPSHr increases. As flow increases, NPSHa decreases (hfs increases with flow). There must always be more NPSHa than NPSHr or the pump

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cavitates. Cavitation is where the pressure at a given point in the pump, prior to entering the impeller eye, falls below the vapor pressure of the liquid pumped, turning it to vapor. As the pressure increases through the Figure 6. System head curve impeller, the bubble of vapor collapses. This causes capacity reduction, noise, vibration, pitting of the impeller and, eventually, failure of a component. No Flow It is possible for the TDH to rise to a point where there is no flow. This is called shutoff and means no flow is going through the pump. At this point, the temperature will rise in the pump because the inefficiencies of the pump are causing heat to go into the pumpage. This raises the temperature and will eventually cause the pumpage to vaporize and the close clearances such as ring fits and mechanical seal faces will not receive lubrication. Eventually the seals fail or the pump seizes. High temperatures also cause the pump to distort and change clearances. Failure of some component will occur.

flow decreases, the efficiency falls and there is more BHP needed to drive the pump than is needed to move the liquid through the TDH. The excess BHP goes into the pump as heat. The flow is lower (less to carry away the heat) and therefore the temperature rise is greater as flow drops. It is possible to reach a point where the temperature rises to the point that it reduces the NPSHa below

Low Flow It is possible for the TDH to rise to a point of low flow where damage other than no flow occurs. Such circumstances include: • Low flow temperature rise—As the Figure 7. Low flow recirculation The Pump Handbook Series

the NPSHr and flashing occurs as above. This point is the “thermal minimum flow” and is a function of not only the pump, but also the system characteristics. • Low flow recirculation/separation damage—As flow decreases, the liquid begins to recirculate in both the impeller eye and impeller discharge (at the volute cutwater or diffuser). This recirculation will cause separation, eddy currents and turbulence that will cause the same noise, vibration and damage as cavitation (Figure 7). The minimum flow required to prevent low flow recirculation separation damage is called the “continuous minimum flow.” It is specified by the pump manufacturer based on the impeller eye geometry and Nss. • Radial or axial thrust damage—All pumps use bearings and balancing devices to prevent the rotating parts from contacting the stationary parts. These are pressure and flow related and they may limit the flow of the pump to a minimum or maximum. The pump manufacturer must make known the limits based on the pump design. ■

References

Nomenclature/Symbols:

1. Carver Pump Company. Performance Curves, Muscatine, IA 2. Lindeburg, PE, Michael. Mechanical Engineering Reference Manual, 8th Ed., Professional Publications, Belmont, CA (1990) 3. Karassik, Igor. Pump Handbook, McGraw-Hill, New York, (1976) 4. Westaway, CR. Cameron Hydraulic Data Book, 16th Ed., Ingersoll-Rand, Woodcliff Lake, NJ (1984)

H = Total differential head Hd = Total discharge head Hs = Total suction head S.G. = Specific Gravity Q = Flowrate NPSH = Net Positive Suction Head (r required, a available) BHP = Brake horsepower Ns = Specific speed Nss = Suction specific speed P = Pressure (1 suction, 2 discharge) hvp = Vapor pressure of liquid hs = Static suction head hd = Static discharge head h = (hd-hs) hfs = Friction losses, suction side hfd = Friction losses, discharge side Hf = (hfs + hfd) vd = Discharge velocity vs = Suction velocity v = Velocity (ft/sec) hv = Celocity head gc = Gravitational constant (32.2 ft/sec2) L = Pipe length in feet D = Pipe diameter in feet C = Hazen Williams constant for pipe materials NRe = Reynolds number f = Friction factor vis = Viscosity e = Absolute roughness K = Resistance coefficient for pipe fittings.

Doug Kriebel has presented pump seminars and workshops throughout his 30-year industrial career. He has extensive experience with pumping applications in a wide range of industries, including positions in the electric utility industry, as well as with a chemical equipment manufacturer and major pump OEM. Mr. Kriebel is currently president of a pump and equipment manufacturer’s representative/ distributor organization. He holds a B.S. degree in chemical engineering. He is an active member of AIChE and is a registered Professional Engineer in Pennsylvania.

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CENTRIFUGAL PUMPS HANDBOOK

Key Centrifugal Pump Parameters and How They Impact Your Applications – Part 2 Pumps and Systems: They Go Together By Doug Kriebel, PE, Kriebel Engineered Equipment

How to Examine a System Each system must be analyzed to determine what happens to the pump at not only design, but also every other possible point of operation—start-up, off peak, unusual modes, “best case”, “worse case.” The following examples illustrate how a system change can create either low flow, high flow or NPSH problems. Refer to Figure 1, which shows a pump system. We will first look at the suction side to determine the NPSHa. The suction side is 20’ of 5” pipe with two elbows and a gate valve. At 300 gpm the hfs is 1’. The pump selected is shown in Figure 2. Example 1 The tank is open to atmosphere, 0 psig, 34’ water = P1; pumping water at 70°F, hvp = .36 psi = .8’; from hs1 = 10’ then: The NPSHa = P1 + hs - hvp - hfs = 34 + 10 -.8 - 1 = 42.2’ Referring to the pump curve Figure 2, at 300 gpm, the NPSHr = 23’ so we have adequate NPSHa. Example 2 All conditions the same except the tank is drawn down to level hs2 = -10’

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Then the NPSHa = 34 - 10 -.8 1 = 22.2’ and the NPSHa is less than the NPSHr and the pump cavitates. Example 3 All conditions the same as Example 1 except the water is now 190°F; hvp = 9.34 psi = 9.34 x 2.31/.966 = 22.3’ Then NPSHa = 34 +10 - 22.3 - 1 = 22.7’ and the pump cavitates. Example 4 Same as Example 3 except from hs2 = -10’ Then NPSHa = 34 - 10 -22.3 - 1 = 0.7’ and the pump cavitates. Example 5 Conditions same as any above except the suction pipe is installed as 3” instead of 5”: hfs = 10’ (instead of 1’). All the above NPSHa would be reduced by 9’! The lesson to be learned on the suction side: Have as much static suction head as possible (hs), locate the pump as close as possible to the suction vessel (with a minimum suction side frictional pressure drop (hfs)) and be careful of liquids near their vapor pressure (hvp). The Pump Handbook Series

Let’s go further and use the Example 1 conditions and look at the discharge side: Assume 150’ of 4” pipe with six elbows, one gate valve and one check valve: hfd = 14’ @ 300 gpm. Example 6 Assume the levels in both suction tank and discharge tank remain constant with suction at hs1 then: H = (P1 - P2) + (hd - hs) + (V22 V21)/2gc + (hfd + hfs) = (50 psig - 0 psig) 2.31/1.0 + (45’ 10’) + 0 + (14’ + 1’) = 115.5’ + 35’ + 15’ = 165.5’ The pump selected is shown in Figure 3. The operating point is “A” What happens if the levels don’t remain constant and the suction tank pulls down from level hs1 = + 10’ to level hs2 = -10’ and the discharge tank fills up an equal 20’? The “h” component (hd - hs) would go up from 35’ to 75’ and the “P” + “h” components would rise from 150.5 to (115.5 + 75) 190.5’. This would cause the pump to “walk” back on the curve to some lower flow, point B. All this means is that the pump will pump a lower flow rate. Not necessarily a bad thing as long as it

meets the process needs and is above the minimum flow for the pump. If it does go below minimum flow, damage will occur. If the pressure (P2) is increased to 75 psig, the head goes up (75-0) 2.31/1.0 = 173 plus h = 35 ‘ to H = 208, which is above the shut off of the pump (205’). The pump would be running with no flow. This could be an upset condition that causes over-pressure to the discharge tank. It is also a situation where an inline filter is installed into an existing installation and the pressure drop is underestimated or the filter is not maintained (cleaned). This example cost an east coast utility company more than $45,000 in damage to a pump when maintenance installed a filter to protect some heat exchangers without taking into account the pressure drop when fouled!

Example 7 More often the pump is selected for a “worst case” condition, (sometimes called the “design condition”). This example uses the same system as above except use 75 psig for the design P2, assuming “worst case” situation (Figure 4, Point A). This accounts for an upset, over-pressure situation in the discharge tank or other circumstance. However, when the system actually operates at 50 psig, the pump runs out to point B. This could overload the motor, since more BHP is required. In addition, the increased

flow may cause the NPSHr to exceed the NPSHa (which decreases due to increased hfs). This may cause the pump to cavitate. Most systems that are designed for higher P2 than normal have a control valve that will increase the

Figure 1. Pump system for Examples 1-6

Figure 2. The selected pump, Example 1

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hfd and “walk” the pump back on the curve to the correct operating point (back to “A” via line C). A major chemical company using sealless pumps and no control valves (they were handling toxic material and did not want to risk leaks) designed a system for a specific H based on a P2 set by the process. After start up, which went without a hitch, the reactor pressure (P2) was decreased to improve yield. Without control valves, the pump ran out on the curve and destroyed itself by cavitation. It took $70,000 in repairs until someone ran the calculations to determine how the system change caused the pump failures! A similar situation occurs if excess safety factor is utilized (Figure 5). If the actual design point is calculated as 300 gpm at 165’ (point A) and the pump is selected for 345 gpm and 200’ head (point B) and the calculations are correct, the pump will run out to point C. This would pump more than required. So the pump would be throttled back to point D to pump the 300 gpm. This means 215-165 = 50’ f head is throttled out, which costs 5.2 BHP. Since you pay for BHP, this is wasted energy. In this case it would amount to $3100 per year at a rate of $.085/kW-hr for service 24 hrs x 360 days. Higher heads or flows would magnify the costs.

Figure 3. System head curve - Example 6

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There is nothing wrong with safety factors, and it is always better safe than sorry, but after installation, check the hydraulics and if gross over-sizing was the case, trimming the impeller will save considerable money in operations. Example 8 (variable speed operation) An economical alternative to throttling with a valve is to use variable speed to change the pump curve to meet variable demands of a system head curve. This is illustrated in Figure 6. The system head curve is plotted, a pump is picked to meet the worst case condition, and reduced speed is used to ride down the system curve to meet varying flows at reduced pressure. In some cases this can be very economical when compared to the wasted BHP of throttling control. If the normal point is the same as above (300 gpm at 165’ Point A on Figure 6) and with safety factors (345 gpm at 200’ Point B) the pump operates on the uppermost curve for 3500 rpm, where it intersects the system head curve at point C. By slowing down to 3200 rpm, it goes through point A without having to waste energy by throttling the head. Any point can be reached on the system head curve by reducing speed. This is a valuable method of operating when flows or heads have large changes in magnitude. The danger comes if you reduce the pump speed to a point where the system head curve causes the pump to operate at a point of too low flow or no flow (i.e., beneath the system head curve, such as 2700 rpm). It is possible to operate too The Pump Handbook Series

high flow where BHP or NPSH becomes a factor, or at a speed where other damage could occur. Examples of other speed related damages are if the pump reaches a critical speed, a speed where auxiliary equipment no longer functions (a shaft driven lube oil pump may not provide proper lubrication at too low a speed, or a motor fan motor may turn too slow to provide adequate cooling). Example 9 (parallel pumping applications) It’s not uncommon to use two or more pumps in parallel. The flows are additive at a constant head (Figure 7). (The flow, Q at head H for one pump at A is added to a second pump at the same head to get twice the flow, 2Q, at the intersection of the two pump curves, the system head curve, B. It’s important to draw the system head curve to make sure that if one pump operates alone it does not run out to point which will overload the motor or exceed the NPSH requirements. Point B is the operating point of two pumps in parallel and Point C is for one pump at runout (estimated here as 1.2 Q). The one pump runout must be checked to provide sufficient BHP, NPSHa, submergence against vortexing and mechanical design for the higher flow. This is more common in vertical cooling water service where several pumps are used in parallel and a control valve is not used on each pump. If one pump trips, the system allows the remaining pump(s) to run out. Another potential for parallel pump problems is where two dissimilar pumps (dissimilar curves) are used (Figure 8). The pump curve for pump A is added to pump B at constant head. The system head is drawn. Q1 is the flow with only pump A running, Q2 is the flow with only B running, Q3 is the flow with both A + B running. Note: This

is not Q1 + Q2. If the system is throttled back (by any means) the system head could cause the pump with the lower head to run below a minimum flow. Pump A is shut off at point C.

Figure 4. Example 7 - changes in head and throttling

Figure 5. System head curve: design/normal/runout/throttle

Example 10 (a pump for a given service is used for an alternate service) Figure 9 is the performance curve for a horizontal split case pump put in as a fire protection pump that is to run at 3000 gpm. The plant engineer lost his air compressor cooling water pump and decided to use the fire pump to supply cooling water for the compressor inner and after coolers. (If you need 30 gpm, a 3000 gpm pump will certainly do!) The pump had six failures of the radial bearings. The maintenance group replaced them, then redesigned the pump and added angular contact back-to-back bearings in place of the double row bearings. They then added oil lubrication to the existing grease bearings and were about to change to oil mist when someone finally asked why the fire pump was running when there was no fire! Point A is the design point, but point B is where it operated to supply the air compressor’s cooling water. This was too far back on the curve and caused high radial loads that were the cause of failure. There was no way any bearing or lubrication upgrade would enable this pump to operate at point B. In conclusion, the purpose of this article is to provide operations and maintenance people who do not normally do process calculations with the tools to analyze the pump systems they have to operate. This will enable them to operate the systems more economically and to determine the “real” root cause of failures. Failures are mechanical in nature, but the cause of the mechanical failure is a “system” incompatibility failure. In today’s business world, systems are often designed by process engineers who are not available when the system becomes operational and the system may not have been installed, operated or maintained the way it was intended to be. Or the correct method of operation is no longer the way it was originally designed. ■

References 1. Carver Pump Company. Performance Curves, Muscatine, IA 2. Lindeburg, PE, Michael. Mechanical Engineering Reference Manual, 8th Ed., Professional Publications, Belmont, CA (1990) 3. Karassik, Igor. Pump Handbook, McGraw-Hill, New York (1976) 4. Westaway, CR. Cameron Hydraulic Data Book, 16th Ed., Ingersoll-Rand, Woodcliff Lake, NJ (1984) Figure 6. Example 8 - variable speed operation The Pump Handbook Series

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Nomenclature/Symbols:

Figure 7. Example 9 - parallel pumps

Figure 8. Parallel pumps

H = Total differential head Hd = Total discharge head Hs = Total suction head S.G. = Specific Gravity Q = Flowrate NPSH = Net Positive Suction Head (r required, a available) BHP = Brake horsepower Ns = Specific speed Nss = Suction specific speed P = Pressure (1 suction, 2 discharge) hvp = Vapor pressure of liquid hs = Static suction head hd = Static discharge head h = (hd-hs) hfs = Friction losses, suction side hfd = Friction losses, discharge side Hf = (hfs + hfd) vd = Discharge velocity vs = Suction velocity v = Velocity (ft/sec) hv = Celocity head gc = Gravitational constant (32.2 ft/sec2) L = Pipe length in feet D = Pipe diameter in feet C = Hazen Williams constant for pipe materials NRe = Reynolds number f = Friction factor vis = Viscosity e = Absolute roughness K = Resistance coefficient for pipe fittings. Doug Kriebel has presented pump seminars and workshops throughout his 30-year industrial career. He has extensive experience with pumping applications in a wide range of industries, including positions in the electric utility industry, as well as with a chemical equipment manufacturer and major pump OEM. Mr. Kriebel is currently president of a pump and equipment manufacturers’ representative/distributor organization. He holds a B.S. degree in chemical engineering, is an active member of AIChE and is a registered Professional Engineer in Pennsylvania.

Figure 9. Example 10

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CENTRIFUGAL PUMPS HANDBOOK

Continuous Monitoring of Sealless Pumps It’s the next step to reliability. By Julien Le Bleu, Principal Mechanical Engineer, Lyondell Chemical, and James Lobach, Chief Developmental Engineer, Chempump Division of Crane Pumps

ll centrifugal sealless pumps, both canned motor and magnetic drive (Figures 1 and 2), should be monitored to determine mechanical condition. In sealless pumps, the pumped fluid is used as the cooling and lubricating medium for the pump bearings. If only intermittent monitoring is used, the chance of detecting pump damage caused by process changes is very small. Conventional vibration monitoring techniques used with sealed pumps have proven unreliable for detecting problems with sealless pumps. The effectiveness of conventional monitoring techniques is limited by the time interval between measurements, the relative isolation of the inner pump rotor from the outer measuring location and by the pumped fluid. Other factors such as fluid effects and process noises can make interpretation difficult. This article presents the synergistic combination of two relatively new methods of sealless pump monitoring. These methods (monitoring rotor position, vibration analysis) considerably enhance the range and magnitude of mechanical problems that can be identified on this type of pump.

A

Figure 1. Mag-drive pump

Rotor Position Rotor position is a measurement that has been requested by sealless pump users for a long time. As long as

Figure 2. Canned motor pump The Pump Handbook Series

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Figure 3. Pump test loop

Figure 4. Data acquisition block diagram

the bearings are not badly worn, serious damage caused by rotor-tostator contact cannot occur. Rotor position monitoring as a predictive tool is minimal when silicon carbide bearings are used. When bearings made of softer materials are used, such as carbon/graphite, the technique becomes predictive because it enables the user to track wear on the bearings and schedule maintenance before serious damage is done.

Vibration Analysis Overall High Frequency Tracking (OHFT) is a vibration technique used for detecting problems with rolling element bearings. The overall value is as an indicator of pump health or process problems. Experience has demonstrated that a narrow trace, especially at a low value, is desirable. This observation was validated during the course of testing.

Test Conditions The pump was installed in a test loop consisting of instrumentation, a supply tank and associated piping. A schematic of the test loop is shown in Figure 3. The pump was subjected to conditions that attempted to simulate what can be encountered during

Figure 5. Test pump curve

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Figure 6. Flow changes (20 gpm increments)

Figure 7. Large flow decrease The Pump Handbook Series

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plant operation. The pressure on the supply tank could be varied, giving testers the ability to induce or eliminate cavitation in the pump to measure its response with the sensors. Testing was done at a manufacturer’s facility. The pump was equipped with the manufacturer’s rotor position monitoring device, which monitors both axial and radial rotor positions. OHFT was measured using two accelerometers that were connected to a dual channel monitor. It was reasoned that measurements of rotor position, OHFT and power would provide sufficient information to determine the pump’s mechanical condition (Figures 4 and 5). We hoped this combination would also provide advance warning of process conditions that would adversely affect the pump’s health. Measurements were taken for the following pump operating conditions: • Pump capacity range from shutoff to 30% greater than BEP with data taken at 20 gpm intervals in the range • A sudden large increase in pump flow • Air leakage into the suction of pump (injected) • Dry pump operation (part of the “air leakage” test) • Best Efficiency Point (part of first item above, this was considered a base line) • Reduced NPSHA The following information was recorded for the operating conditions listed above: • Motor input power (watts) Labeled (D) (1 volt = 0 kW, 5 volts = 15.5 kW) • OHFT 0-16 gSE on the casing labeled (B) (1 volt = 0 gSE, 5 volts = 16 gSE) 0-5 gSE on the rear bearing housing labeled (C) (1 volt = 0 gSE, 5 volts = 5 gSE) • Rotor position (Axial and Radial) Axial labeled (E) (1 volt = .000” in, 5 volts = .100” in) Radial labeled (F) (1 volt = .000” in, 5 volts = .013” in)

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• Flow labeled (A) (1 volt = 0 gpm, 5 volts = 300 gpm) 3” Brooks MagFlowmeter • Suction pressure labeled (G) (1 volt = 0 psia, 5 volts = 30 psia) Absolute pressure transducer The data in parentheses are the plot scale factors for the data presented in the article. An increase in voltage for the rotor axial position data represents a movement of the rotor toward the suction flange of the pump. The underlined letters in parentheses represent the letters on the graphs for that data set.

Test Results Changes in Flow (Figure 6) The changes in data as a result of variation in flow over the range of 50 to 190 gpm are represented in Figure 6. Note that at capacities greater than the BEP, about 150 gpm, the rotor begins to move axially and OHFT begins to increase in value and width. Neither of these is desirable in sealless pumps. Large Flow Decrease (Figure 7) A large flow decrease from 200 gpm to shut off is represented in Figure 7. OHFT, indicated by (B) and (C) on the chart, is at a high value and a wide trace at 200 gpm, indicating cavitation and pump stress. The axial rotor position also has a wide trace during this part of the test, indicating the rotor was “hunting” to find its hydraulic balance. When the flow is decreased from 200 to 150 gpm, the rotor moves to a normal and more balanced axial location and OHFT decreases from 10 gSE to approximately 4 gSE. Large Flow Increase (Figure 8) Large flow increases are illustrated in Figure 8. At very low flows, the OHFT value is high and the rotor axial position is near center. At BEP, 150 gpm, OHFT is at a minimum level and rotor axial position is slightly toward the motor, relative to its position at BEP. When flow is increased to levels significantly higher than BEP, the onset of cavitation is indicated by an The Pump Handbook Series

increased width of signal in both the OHFT and axial rotor position data. Rotor axial position oscillates at the high flow level, and this “hunting” of the rotor is indicated by the wide data trace (E). The rapidly changing pressure balance on the impeller during cavitation causes this oscillation. The OHFT signal also becomes less stable and “nosier” when the pump is cavitating. Suction Side Restriction (Figure 9) The suction valve was closed at a constant flow to measure the response in terms of the measured parameters. The results are illustrated in Figure 9. The width of the trace representing axial rotor position and OHFT begin to increase. Both of these indicate instability. One is in the rotor position and the other is in the pumped fluid. This could represent a suction strainer plugging in the suction line if there were one or a valve that was not opened fully. It also can represent a fluid that has become too hot and is flashing in the suction of the pump. Failure of sealless pumps due to inadequate suction conditions or other cavitation-inducing operation can be minimized through the connection of the monitors to appropriate operator alarms. Air Injection and Dry Run (Figure 10) The air injection and resulting dry run test in Figure 10 is a graphic representation of the effects on axial rotor position when air leaks into the suction. OHFT immediately begins to decrease in amplitude. The radial position of the rotor changes because the loss of fluid in the radial bearings and around the rotor reduces the radial support stiffness. OHFT levels are very low because of the loss of coupling fluid and transmissibility between the pump casing and the rotor. The watt meter shows no load on the pump and the flow has fallen to zero. This test does show that a “dry run,” as with a tank pump out, does not result in an instantaneous

Figure 8. Large flow increase

Figure 9. Suction side restriction The Pump Handbook Series

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catastrophic failure. The effects of dry running and severe cavitation are cumulative in our experience. This is especially true of mag-drive pumps. Low Suction Pressure Induced Cavitation (Figure 11) While factory NPSH testing defines the onset of cavitation as a 3% loss of head, cavitation effects are sometimes seen well before a measurable head loss occurs. Continuous monitoring of OHFT and rotor axial position represents a practical method to measure the actual onset of cavitation through the direct measurement of pump response to hydraulic conditions. A test was conducted where the suction pressure was reduced with the pump capacity held at a constant 150 gpm to observe the effects on the monitored parameters. Figure 11 is the graphic representation of the test results.

Results Table 1 represents all of the data that was recorded during the testing of the pump for each measured parameter and test condition. The table can be used with whatever combination of sensors that exists in the user’s plant and will be helpful in the interpretation of indications. For example, if OHFT is added to a system that already has a power monitor, useful information can be obtained that should minimize maintenance costs and catastrophic failures.

•

Overall High Frequency Tracking (OHFT) During pump testing, several new items were noted regarding OHFT. • The mounting of the sensor was found not to be as critical as originally suspected. That is, if the sensor is mounted solidly to the pump casing, the orientation as to axial or radial did not significantly

Table 1. “Truth” table of results

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•

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•

•

change the magnitude of the overall readings. Lower baseline OHFT readings will result when the accelerometer is mounted on the upstream side of the cutwater. Higher baseline readings resulted when the accelerometer was mounted downstream of the cutwater. Presumably, the increased baseline noise was caused by possible turbulence or hydraulic noise associated with liquid passing by the cutwater. A rule of thumb when using OHFT is that less is better. The quieter the pump is, the better and more trouble-free it will operate. The exception to this rule is dry running. Wider traces of both OHFT and axial rotor position are indicative of operating conditions to avoid in pumps, especially sealless pumps. A time interval of one second or less should be used as a sample rate for

Figure 10. Air injection

Figure 11. Reduced suction pressure (at constant 150 gpm) The Pump Handbook Series

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capturing OHFT data. This will capture all of the fast-changing operating and mechanical data that can take place within the pump. • Mounting the sensor closer to the source of the stimuli is better, because the signal is stronger. Usually that means mounting on the pump casing. • Conditions that raised the OHFT level caused the rotor to move significantly in the axial direction.

Rotor Position Monitoring The rotor is monitored through a series of wound coils located outside the primary containment and protected from the process fluid by the stator liner. Electrical signals received from the coils are used to continuously monitor the actual running position of the rotor. The device detects any change in rotor position in the axial and radial directions simultaneously. By comparing the instrument’s output to the original factory test baseline of a new pump, the condition of the internal radial and axial bearings can be determined. After the initial calibration, radial bearing wear is determined by a change in output in the radial direction greater than the baseline data for new bearings. The amount of wear is proportional to the change in signal. It should be noted that normal operation of a sealless pump does not promote wear of the radial and axial bearing surfaces. The process upset conditions leading to lack of lubrication and rapid heat rise are the main causes leading to the wear of these surfaces. Continuous monitoring enables users to trend these damaging events to predict and improve the mean time between maintenance intervals.

Conclusions There are many benefits to monitoring sealless pumps continuously. Since the trend in process plants is to use distributed control systems, much of the plant equipment is being remotely

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operated. Presently, the board operator in a control room has little or no feedback on the operating condition of most of the pumps. These parameters are not instrumented with trip or alarm limits based on pump health. It is possible for the pump to be operating in an off-design condition or have mechanical damage and the operator not know it. The feedback presently comes in the form of failed equipment and expensive repairs. With continuous monitoring, it is possible to get immediate feedback on the condition of the pump and on the process conditions. Cavitation, dry running and extreme operating situations that will result in pump failure can be immediately detected. Armed with this information, the operator can make decisions to improve operating conditions that will prolong equipment life and maintain product quality. For new pump installations, use the latest technological advances being offered by sealless pump manufacturers. This includes rotor position monitoring on canned motor pumps and heat detection on the containment shell in the center of the magnet area of mag-drive pumps. It is best to utilize both radial and axial monitoring of rotor position when available. If you want to retrofit sealless pumps with OHFT monitoring systems, follow the guidelines below. • When metal-to-metal contact is detected, the pump should be stopped and scheduled for maintenance. • When OHFT is added to a pump, the condition of the pump should be known and a baseline representing that condition recorded. OHFT values should be fairly low, in the 10-40% of full-scale range. • A baseline set of readings should be taken with the pump operating at its normal pump curve capacity. This should be done even if the discharge valve has to be throttled to achieve this with the size of the impeller that is being used. • When an increase in OHFT is detected, the process should be The Pump Handbook Series

varied, if possible, to eliminate the “noisy” condition. This will help to determine if the increase is due to mechanical or process conditions. • If the OHFT levels are reduced to a relatively low value by adjusting process conditions to normal pump design values, the pump most likely does not have a mechanical problem but is probably not being operated on its curve. To prove this, allow the process to settle for a short while. Start the standby pump, if one exists, and look at its OHFT readings. If they are substantially lower than those of the recently running pump, leave the spare pump in service and put the other in standby mode. Schedule the recently stopped pump for maintenance. • If switching pumps cannot reduce the OHFT noise, it may be an indication that the pump is incorrectly sized for the process and will be a maintenance problem. The pump application should be investigated for proper sizing and adequate suction conditions. The greatest savings will come from detecting the conditions that will cause a problem in a pump early enough to eliminate them and thus prevent a failure. If early detection of off-process operation is not possible, the next best maintenance practice is to detect a problem at its inception and schedule the pump for maintenance at a point when the problem will result in minimal maintenance costs, business interruption and no leakage. ■

Nomenclature gSE - Dimensionless unit used in detecting problems with rolling element bearings.

Acknowledgments The authors would like to thank Crane, Chempump division, and Entek/IRD for the use of their equipment and facilities in conducting these tests. ARCO/Lyondell Chemicals should be thanked for their help and support in allowing this article to be written.

References 1. Shea, J. M. and Taylor J. K., 1990 “Using Spike Energy for Fault Analysis and Machine-Condition Monitoring”, IRD Mechanalysis, Inc. 2. Le Bleu, J., Jr., 1994 “Monitoring Sealless Pumps for Metal-to-Metal Contact of the Internal Rotor”, Proceedings of the 11th International Pump Users Symposium, Turbomachinery Laboratory, Texas A&M University, College Station, Texas. 3. Le Bleu, J., Jr. and Xu, Dr. M., 1995 “A New Approach for Monitoring Sealless Pumps”, Proceedings of the 19th Vibration Institute Meeting, Indianapolis, Indiana.

Julien Le Bleu is the principal engineer for rotating equipment for ARCO chemicals in Lake Charles, Louisiana. He has more than 25 years experience in the field of rotating equipment, and is responsible for all rotating equipment in the Lake Charles facility. He has authored several articles and has lectured at the Pump Symposium previously. Mr. Le Bleu is presently a member of the advisory board for the Texas A&M Pump Symposium. He received his bachelor’s degree from the University of Florida (1974). James Lobach is a Chief Developmental Engineer with the Chempump Division of Crane Pumps and Systems. He has had extensive experience in the design and application of high speed rotating machinery. For the past five years, he has been closely involved with canned

The Pump Handbook Series

motor pump design and innovations, including low specific speed pumping, pump hydraulics and performance, and monitoring equipment. He has provided field service engineering in the chemical and petrochemical industries for the past 15 years. Mr. Lobach received a B.S. degree in Mechanical Engineering from the University of Colorado (1969). He is a registered Professional Engineer in the state of Colorado. This article has been reproduced with permission of the Turbomachinery Laboratory from the Proceedings of the 15th annual Pump Users Symposium, Turbomachinery Laboratory, Texas A&M University, College Station, TX, pp. 125-132. Copyright 1998.

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Selecting Sealless Pumps and Circulation Systems for Difficult Pumping Applications Address concerns during the selection and purchasing process—not in the field. Ron Forsberg, Senior Project Engineer, Sundyne Corporation

any difficulties experienced by pump users could be eliminated through a good understanding of the application and better communications among the user, engineering contractor and pump manufacturer. Just as mechanical seals and their support systems represent the greatest challenge for the users and suppliers of sealed pumps, lubrication and cooling systems represent the greatest challenges for sealless pump users and suppliers. Taking time to learn basic principles associated with the selection and application of sealless pumps, though, could significantly reduce such challenges.

M

A sealless pump generally is used when there is a need to contain toxic, dangerous or valuable fluids, or where specific applications warrant their use. The application of the pump also may be dictated by environmental, safety, noise or space concerns.

The Fundamentals Where are Sealless Pumps Used? • Hazardous/lethal fluids • Heat transfer fluids • Expensive fluids • Fluids requiring high purity • Fluids with foul odors • Volatile fluids - solvents • Molten solids • Monomers & polymers • Replacement of expensive multiple mechanical seal systems

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Figure 1. Magnetic Drive Pump The Pump Handbook Series

How Do They Work? The rotating members are driven by a rotating magnetic field that is transmitted through a primary containment shell or liner. Two drive methods are used—Magnetic Drive Pumps (MDP) and Canned Motor Pumps (CMP). In

Figure 2. Canned Motor Pump (CMP)

both methods, a portion of the pumped precipitate out of the solution. fluid (pumpage) or an externally supplied • High melting point liquids & heat flush fluid is used to cool the drive transfer mechanism and lubricate the bearings. What are Canned Motor Pumps What are Magnetic Drive Pumps and When Do They Apply? and When Do They Apply? The Canned Motor Pump (CMP) has The Magnetic Drive Pump (MDP) a common shaft to link the pump and utilizes an outer ring of permanent motor in a single contained unit. The magnets or electromagnets to drive an pumped liquid or an external flush is internal rotating assembly consisting circulated through the motor but is of an impeller, shaft and inner drive isolated from the motor components member (torque ring or inner magnet by a corrosion-resistant containment ring) through a corrosion-resistant, shell. The canned motor pump is used non-magnetic containment shell. They when both primary and secondary are used when primary containment containment of the process fluid is a is a must and secondary control is must (Figures 2 & 3). desired. (Figure 1). • Lethal services • Mild acids • Extremely hazardous services • Solvents • High suction pressure services • Heat sensitive liquids • When advanced diagnostics are • Liquids containing moderate solids required • Fluids presenting difficult sealing • Where secondary containment vs. challenges: dissolved solids that secondary control is a must! The Pump Handbook Series

Two industry organizations, the Hydraulic Institute (HI), and the American Petroleum Institute (API), have recognized sealless centrifugal pumps as a unique pump type and have adopted a unique standard or specification for them. The HI standard is designated as: ANSI/HI 5.1-5.6 American National Standard for SEALLESS CENTRIFUGAL PUMPS. The API standard is designated as: API-685, SEALLESS CENTRIFUGAL PUMPS. Both HI and API have recognized the importance of the drive circulation systems, and have devoted a significant portion of their respective standards to describing the types of systems and the importance of each. Both have identified these circulation systems similar to their seal support and flush plans. However, neither specification goes into much detail about these plans. While they do

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Figure 3. Vertical In-Line Canned Motor Pump (CMP)

supply some graphic representations as shown on the accompanying figures, they do not provide a detailed description or direction. Therefore, caution must be taken when referring to these plans, as they do not use the same identification system, and not all plans are defined in each standard. Excerpts from these standards, used with permission of the appropriate organization, are shown in Figures 4 & 5 and API Figures D-1 & D-2.

Why Do We Have to Be So Careful? For most fluids, the care required is fairly obvious. If the fluid is going to be circulated through the drive section and bearings, it must be clean enough not to damage the components. If the fluid is too hot to provide cooling of the drive section, other cooling means

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must be provided. If the fluid is corrosive, then the correct materials of construction must be provided. Let’s consider this fundamental principle. In order for a pump to operate properly, the fluid must remain in the liquid state. For years pump manufacturers and users have experienced the negative effects of operating pumps in the unstable situation where the process fluid changes phase and vaporizes. This vaporization occurs when the combination of temperature rise of the fluid coupled with insufficient pressure rise or pressure drop moves the fluid to the wrong side of the vapor pressure curve. In sealed pumps, this vaporization can result in several undesirable conclusions, including, pump cavitation due to insufficient net positive suction head available (NPSHA) and The Pump Handbook Series

dry running liquid seals resulting in seal failure. For sealless pumps, the problem is even more serious. The vaporization of the fluid can result in significant drive system failure with resultant parts damage and potential fluid release to the environment. In the conventional sealed pump, after suction conditions are satisfied, the primary heat concern is hydraulic inefficiency. That portion of the input power that is not converted to the head and flow of the product is inefficiency. That inefficiency becomes heat. For sealless pumps, in addition to the heat from hydraulic inefficiency, there are several other heat sources, including rotor/stator inefficiency, liner losses, and fluid circulation parasitic losses. The parasitic losses include the auxiliary impeller, viscous drag and re-circulation losses. In order

EXCERPTS FROM HI 5.1 – 5.6 2000 (Used with permission of the Hydraulic Institute)

5.3.3.3 Circulation piping plan selection Selection of an appropriate circulation piping plan depends upon knowledge of liquid properties such as: cleanliness, volatility, specific heat, viscosity, specific gravity, toxicity, melting point, temperature, corrosive-ness and any tendency to form solids.The following should also be considered are: flow rates, NPSHA, frequency of starts, cooling or heating availability and potential loss of suction liquid. Typical circulation plans are shown in Figure 5.9 and are grouped by application type considering pumped liquid cleanliness, volatility and temperature. The manufacturer may also offer additional plans specific to design and application requirements. A detailed analysis should be conducted for each application.

5.3.3.3.4 Volatile liquids Circulation to suction vessel or pressurized circulation may be used to avoid thermal effect of drive heating on pump NPSH requirements. Consideration of vapor pressure increase with temperature and of specific heat of liquid shall be required. Use of a separate low-volatility drive external flush liquid is also possible. Liquid cleanliness and temperature shall also be considered.

5.3.3.3.5 Liquids that solidify Jacketed pumps may be required for high melting point liquids and easily polymerized or crystallized liquids. External flush liquids may also be used. Figure 5.9 does not include plans for jacketed pumps.

5.3.3.3.1 Clean liquids Clean liquids are those with no solid particles, non-volatile, moderate temperature, sufficient NPSH and a moderate degree of hazard. This description fits the majority of sealless pump applications and can be handled by various circulation piping plans.

5.3.3.3.6 High viscosity Viscosities that would cause objectionable drag losses in the drive section or inadequate bearing cooling flow (generally above 200 mPa-S (200 centipoise)) may be handled with external flush liquid. Start-up and operating viscosity should be considered.

5.3.3.3.2 Dirty liquids Dirty liquids include solid particles. Centrifugal separation, mechanical filtration, or a separate clean external flush liquid may be used. Also, volatility and temperature shall be considered.

5.3.3.3.3 High temperature The temperature of motor windings or magnetic drive components can be controlled by using a variety of circulation piping plans. Volatility and cleanliness should be considered in selecting a plan

5.3.3.3.7 External flush External flush should be used where there is potential loss of suction, zero flow, or entrained vapor. Separate compatible external flush liquid supply with appropriate cooling may be used to provide lubrication and cooling of the drive section. Precautions still apply for normal centrifugal pump operation.

Figure 4. Excerpts from HI 5.1-5.6 2000 (Used with permission of the Hydraulic Institute)

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EXCERPTS FROM API 685 U.2 CIRCULATION PLAN SELECTION AND APPLICATION (See also API Figures D-1 & D-2) It is recognized that the product lubricated bearing design and application considerations are essentially the same for canned motor pumps and magnetic drive pumps. Selection of an appropriate circulation plan depends upon knowledge of fluid properties such as cleanliness, volatility, specific heat, toxicity, melting point, and tendency to form solids or polymerize. Also to be considered are intended operation, flow rates, NPSH, frequency of starts, and cooling or heating availability. Factors internal to the unit design such as pressures, temperatures, flows and heat transfer characteristics within the drive section as well as hydraulic performance of the pump end must be understood in order to properly select circulation plans and assess application questions. Possible advantages and limitations of available plans must also be understood. The circulation plans shown in Figures D-1 and D-2 coupled with detailed knowledge of individual unit design allow for the handling of most applications. Comments on individual considerations are as follows: U.2.1 Clean, non-volatile, moderate temperature fluid with sufficient NPSH. This description fits the majority of sealless pump applications and can be handled by variations of circulation plans shown. U.2.2 High Temperature: Temperature of motor windings or magnetic drive components can be controlled by a variety of circulation plans shown in the grouping for high temperature. U.2.3 Volatile Fluids/limited NPSH Available: Reverse circulation and pressurized circulation plans may be used to avoid the thermal effect of drive heating on pump NPSH requirements. Consideration of vapor pressure increase with temperature and of specific heat of fluid is required. Use of a separate low volatility drive buffer fluid is also possible. U.2.4 Venting and Cool Down: When pumping cold fluids which are volatile at atmospheric temperature use of a separate vent line back to the supply vessel is necessary to cool the pump and piping to near pumping temperature prior to start-up. U.2.5 Fluids Containing Abrasive Particles may cause objectionable wear. Centrifugal separation, mechanical filtration, or separate, clean buffer fluid may be used to remove particles from the circulation fluid. U.2.6

Jacketed designs may be required for high melting

point fluids and easily polymerizing or crystallizing fluids. Buffer fluids may also be used. U.2.7 High Viscosity: Viscosities that would cause objectionable drag losses in the drive section or inadequate bearing lubrication (generally above 100 CP) may be handled with an external source of circulation fluid. (CPS = CS X SG, SSU = 4.64 x CS). Start-up as well as operating viscosity must be considered.

CIRCULATION SYSTEM APPLICATION: CLEAN PUMPAGE API 685 1-S/HI 101: An internal system where a portion of the process fluid is diverted from a high pressure region of the pump, circulated through the drive section, and returned to a lower pressure region of the pump, usually the suction. This system is typically the most reliable as it requires no maintenance or external systems. The application engineer and user must evaluate the effects of process fluid temperature rise through the drive section throughout the complete operating range of the pump. For volatile fluids, with little NPSH margin, the temperature rise combined with a return to suction pressure could result in flashing of the fluid. This Vapor Pressure Margin Analysis is supplied as a standard by some manufactures and will be a requirement of the API standard. API 685 1-SD: An internal system similar to 1-S above, but including an auxiliary impeller and returning the circulated fluid to discharge. Again a Vapor Pressure Margin Analysis must be performed. This system is susceptible to fluid flashing if the pump is operated at the end of curve. API 685 11-S/HI 111 & 112: Circulation from pump discharge through the drive section back to suction. This system is also susceptible to problems at end of curve operation. A drop in differential pressure would reduce flow and cooling through the drive section. API 685 13-S, 13-SE/HI 113: A reverse system where a portion of the process fluid is diverted from a region of high pressure within the pump cavity, circulated through the drive section, exits the pump through a restriction orifice and is returned to either the pump suction or the suction vessel. This system is useful when the temperature rise through the drive section prevents the return of the fluid to the pump suction. The fluid can be returned to the suction line at a sufficient distance to allow flashed vapors to re-condense, or returned to the vapor area of the suction vessel. The restriction orifice must be sized to maintain the fluid as a liquid throughout the drive chamber.

Figure 5. Excerpts from API 685

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API 685 21-S/HI 121: An external system where a portion of the process fluid back pressure is diverted from the pump discharge, through a heat exchanger, and circulates it through the drive section, to pump suction. These systems are typically used when the process temperature is only moderately high, but too hot to provide adequate cooling for the drive section or for volatile fluids where temperature rise through the drive section could flash the fluid. The affects of cooling the process fluid must be carefully considered. With many fluids, cooling can result in a significant increase in viscosity or the precipitation of solids. As with the API 685 1-S/HI 111 & 112 plans above, these systems are also susceptible to problems at end of curve operation. API 685 23-S/HI 123: This system also uses an external heat exchanger. However it re-circulates a single mass of fluid using an auxiliary impeller.Its advantages are that it restricts the amount of fluid transferred between the pump and drive section, and it only has to remove slightly more than just drive section heat. This can greatly reduce the cooling water requirements. It helps prevent particulate from being circulated through the drive section, and it can be used on very hot applications. Again, the affects of cooling on the process fluid viscosity must be carefully considered. In addition, this system will typically require a venting sequence during startup.

DIRTY FLUIDS

is not recommended. In addition to its use with dirty fluids, this system is often used where there is a potential of operating the pump dry, such as, in tank unloading and transfer operations. API 685 41-S. This system is similar to the 31-S with the addition of a heat exchanger. The application is for mildly hot dirty fluids. The same limitations apply with the added consideration to the affects of cooling on the viscosity and precipitation potential. API 685 53-S/HI 153. This system is essentially the combination of the API 685 23-S/HI 123 and the API 685 32S/HI 132. This system uses an internal auxiliary impeller to circulate the cooling fluid through an external heat exchanger, with the addition of an external pressurized buffer fluid. Some systems use a reservoir tank in addition to the heat exchanger or use cooling coils within the reservoir. This system is particularly effective where it is desirable to reduce the amount of process fluid dilution by the buffer. Some systems use close clearance throttle bushings or internal mechanical seals to restrict the amount of buffer transfer between the drive section and the pump. API 685 54-S. This partial flush system is similar to the API 685 32-S/HI 132 with the exception that most of the pressurized buffer fluid is returned to the buffer source. Close clearance throttle bushings or internal mechanical seals are used to restrict the amount of buffer transfer between the drive section and the pump.

The objectives of these plans is to provide cooling and lubrication to the drive section, yet prevent the introduction of contaminates that can damage bearings, penetrate containment liners or block flow passages. API 685 31-S & 31-SE/HI 131. A system where a portion of the process fluid is diverted from the pump discharge through and external centrifugal separator or filtration system, circulated through the drive section, to the pump suction. As with the API 685 1-S/HI 111 & 112 plans above, these systems are also susceptible to problems at end of curve operation. Contamination and blockage of flow with these systems is always a consideration. API 685 32-S/HI 132. Similar to API or HI Plan 32 this full flush system uses an external supply of clean, product compatible, fluid to cool the drive section and lubricate the bearings. The customer must be able to accept this continued dilution of the process fluid by the buffer fluid. Particular care must be exercised in applying multiple pumps to a single buffer supply system. Balancing of the required buffer flow to each pump can be difficult and this type of system

See Appendix D and Appendix U in API Standard 685 Sealless Centrifugal Pumps for Petroleum, Heavy Duty Chemical, and Gas Industry Services, First Edition, October 2000, reprinted courtesy of the American Petroleum Institute.

To order API publications call (800) 854-7179, fax to (303) 397-2740, or order on-line at www.global.ihs.com.

Figure 5. continued

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Internal recirculation through drive chamber back to suction

Internal recirculation from high pressure through the rotor chamber to an auxilliary impeller and back to pump discharge

Dead-ended rotor chamberwith no circulation of flushed fluid (used for temperature thinning fluids)

Recirculation from pump discharge through an orifice to the rotor chamber and back to suction

Recirculation from pump discharge to the rotor chamber, through an orifice and back to pump suction

Reverse circulation from pump discharge through the rotor chamber to an external suction vessel

Recirculation from pump discharge through an orifice and cooler to the rotor chamber and back to pump suction

Recirculation from rotor chamber through a cooler back to rotor chamber using an auxilliary pumping device

API Figure D-1. Circulating Pump Arrangements—Clean Pumpage See Appendix D and Appendix U in API Standard 685 Sealless Centrifugal Pumps for Petroleum, Heavy Duty Chemical, and Gas Industry Services, First Edition, October 2000, reprinted courtesy of the American Petroleum Institute. To order API publications call (800) 854-7179, fax to (303) 397-2740, or order on-line at www.global.ihs.com.)

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Recirculation of cleaned fluid from a separator through the rotor chamber

Circulation from pump discharge through Purchaser’s filter system to rotor chamber

Injection to rotor chamber from external source

Recirculation from pump discharge of cleradned fluid from a separator through a cooler to the rotor chamber

Circulation of an externally pressurized fluid through a cooler and the rotor chamber using an auxilliary impeller

Circulation of external source fluid through the rotor chamber, and back to the fluid source

API Figure D-2. Circulating Fluid Arrangements—Dirty or Special Pumpage See Appendix D and Appendix U in API Standard 685 Sealless Centrifugal Pumps for Petroleum, Heavy Duty Chemical, and Gas Industry Services, First Edition, October 2000, reprinted courtesy of the American Petroleum Institute. To order API publications call (800) 854-7179, fax to (303) 397-2740, or order on-line at www.global.ihs.com.) The Pump Handbook Series

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Figure 6. Vapor Pressure Margin Curve Model-HP-C3 Splitter Feed-Item C Figure 6. Vapor Pressure Margin Curve Model-HP-C3 Splitter Feed-Item C

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Figure 7. Vapor Pressure Margin Curve Model HP-C3 Splitter Feed-Item D Figure 7. Vapor Pressure Margin Curve Model HP-C3 Splitter Feed-Item D The Pump Handbook Series

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Figure 8. Vapor Pressure Margin Curve Model HT-C3 Splitter Feed-Item D2 Figure 8. Vapor Pressure Margin Curve Model HT-C3 Splitter Feed-Item D2

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to calculate these losses and evaluate the condition of the fluid throughout the circulation path, manufacturers have developed proprietary equations that require specific knowledge of the fluid.

How Do You Select a Sealless Pump? In order to properly size the sealless pump, select the proper circulation system and evaluate the vapor pressure margin, the pump manufacturer must have the specific fluid properties data. This not only includes the standard head, flow, specific gravity and NPSHA data that customers normally provide, but also includes suction pressure, suction temperature, specific heat, viscosity and a vapor pressure curve. The vapor pressure curve is required to evaluate what happens to the fluid as the temperature increases. A single point is not enough. Sundyne Corporation has developed a sealless pump sizing program which not only selects the best pump based on head, flow and efficiency, but uses those proprietary equations to calculate the fluid temperature rise and pressure change throughout the circulation path. Figures 6, 7 and 8 are computer-generated Vapor Pressure Margin curves of vertical in-line canned motor pumps. They show the temperature rise, the actual pressure and the vapor pressure at the elevated temperature at selected locations within the pump circulation flow path. Figure 6 shows an API Plan 1-S circulation system with adequate vapor pressure margin. Figure 7 also shows an API Plan 1-S system, but the temperature rise and the corresponding vapor pressure exceeds the fluid pressure with the motor. This results in the fluid flashing to a vapor and eventual pump failure. Figure 8 shows the same pump but with an API Plan 53-S cooled circulation system with make-up buffer flow. This again provides adequate vapor pressure margin.

Customer/Vendor Communications The selection and application of the sealless pump, with its circulation system requires an understanding by the user and the manufacturer of the pump, the process and the fluid. Moreover, it requires complete exchange of information, particularly with respect to fluid properties and the effect of the pump, drive section and circulation system on the fluid. With inadequate or erroneous fluid information, the wrong circulation system might be applied, which could result not only in performance problems, but in equipment damage and potential release of the fluid to the environment. It should be the goal of the user and the manufacturer to address concerns during the selection and purchasing stage—not in the field after the system has failed. ■

Ron Forsberg is a Senior Project Engineer with Sundyne Corporation in Arvada, CO. Since joining the company in 1973, he has held various positions within its quality assurance, maintenance, product engineering and product development departments. He has a BS degree in Marine Engineering from the U.S. Merchant Marine Academy at Kings Point, NY. E-mail him at: [email protected] . (Editor’s Note: This article was adapted from the Official Proceedings of the PumpUsers Expo 2000 Technical Conference.)

References: 1. ANSI/HI 5.1-5.6 American National Standard for SEALLESS CENTRIFUGAL PUMPS Hydraulic Institute 9 Sylvan Way Parsippany, NJ 07054-3802 2. API-685 SEALLESS CENTRIFUGAL PUMPS American Petroleum Institute 1220 L Street, Northwest Washington, DC 20005 (See Appendix D and Appendix U in API Standard 685 Sealless Centrifugal Pumps for Petroleum, Heavy Duty Chemical, and Gas Industry Services, First Edition, October 2000, reprinted courtesy of the American Petroleum Institute. To order API publications call (800) 854-7179, fax to (303) 397-2740, or order on-line at www.global.ihs.com.)

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Installation: The Foundation of Equipment Reliability

The road to trouble-free service begins with these steps.

Joseph F. Dolniak, Reliability Engineer, Eli Lilly and Company

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Initial installation is a key factor in the lifetime cost of almost any equipment. When properly installed, operated and maintained, many years of trouble-free service can be expected. However, when rotating equipment is not installed correctly, years of maintenance and operational problems will impede its performance. While the following steps are designed for the proper installation of ANSI pumps, they also can be adapted for use with API and larger equipment.

1. Pre Alignment Check A "rough" alignment of the equipment should be done as soon as the equipment assembly is received. This rough alignment should be done to final tolerances. Straightedge alignment is not acceptable. Doubledial (indicator reverse) or laser alignment should be used. The reason to do the rough alignment to final tolerances is to ensure that the equipment assembly can be aligned to specification. If the equipment assembly cannot be aligned now, it indicates a problem with one of the components. If caught in time, there still should be time to correct the problem before the system is installed. This will save time and money. If a good alignment is achieved, the pump and motor can be removed from the base to ease the installation and help prevent damage to critical components.

2. Site Location Often, the site location of equipment is pre-determined because of physical constraints, such as being in a building, a column location, etc. When there is a choice, choose a place that will be accessible and

aesthetically pleasant. You will hear these two terms over and over again because they are so important to achieving reliability. When a piece of equipment is easy to reach and maintain, mechanics will be able to perform better work on it and operators will be able to activate and control it more efficiently. This is not saying that a mechanic intentionally will do poor work if the equipment is hard to access, or that an operator intentionally will mis-operate the equipment. Obstacles to accessing equipment do hinder their performance. Also, if equipment is kept clean, painted and well arranged, it will be treated with more care than equipment that is dirty, rusty or poorly maintained. There are also dollar savings when maintenance is required. The time and safety factors involved in maintaining hard-to-reach equipment are greater than on equipment that is accessible. The accessibility and aesthetics of equipment are important. If these two factors can be considered during the design stage of a plant, the plant will likely be a smoother-running facility with higher profits than one where these factors have been neglected.

3. Pump-Pad Location The pump-pad location is often less constrained than the site location. Existing equipment, kettles, piping, and such, will determine where the pumps must be located, but there is usually room to finetune the exact location of the pump pad. Items to remember when choosing the exact location will be allowing room (for walkways, existing piping, new piping), operator and mechanic accessibility and aesthetic considerations. Items that can

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usually be altered are the direction the pumps face, their closeness to walls, and the height and depth of the pump pads. If the pump height or location is altered, ensure that the operating parameters of the pump can still be met (i.e., NPSHR, straight pipe run to the suction, etc.). If installation standards or guidelines already exist, consult the standards committee before any guidelines are altered.

4. Pad Dimensions The pump pad has two specific purposes. It will be the support for the pumps to operate on in a safe manner, and it's mass will transmit vibrations away from the equipment to allow the equipment to operate "vibration free." Pump-pad dimensions must be determined from the size of the pump base. For ANSI pumps, consult the current ANSI B73.1M-1991, which is currently being revised (revisions being made do not alter the pump base dimensions). The author prefers to add about 6 inches to the sides and front of the pump base to derive the pad dimensions. An exception to this can be made when reusable base forms, adaptable to many size bases, are available. If one pump pad comes out to 58 inches, and another pump pad comes out to 60 inches, the 60-inch pump-pad would be used for both, to minimize the number of pump-pad dimensions. The number of pumps per pad is limited to two. If there are two pumps per pad, there will be 10 inches between the pump bases. The maximum pump pad height will always be 18 inches, unless operational parameters preclude. The distance between the pump bases on the pad and pad height can vary, somewhat. Once determined, though, these dimensions should be used throughout the plant for uniformity. It is important that the pump pad have sufficient mass to absorb and transmit the vibrations of the pump and motor. Accepted figures for pad mass are as follows: • For centrifugal types of pumps, the mass of the pad should be at least three times the weight of the pump, motor, base and liquid. • For reciprocating machines the pad should have a mass of five times the

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weight of the pump, motor, base and material, and permits for digging. It usually takes less than a one day to dig the average liquid.

ANSI pump-pad hole. Having standards for the pump bases also allows easier pipe prefabrication with 7. Install Pump-Pad Forms fewer mistakes. Your pump site now should have a trued hole to be filled, but many steps still 5. Pump Orientation need to be accomplished before the In choosing the orientation, note items cement or epoxy can be poured. The first such as aisles, piping flow, available room, step is to install the pump-pad forms. valve location, etc. Because some pump These can be reusable forms or one-time pads are almost square, it can be easy to use forms. It is important that the forms have the pump pad facing the wrong way. are true, solid, well braced and liquidIf this mistake happens, there is a chance tight. If an epoxy material is being used, for pad failure because the pump base liquid-tight is especially important, as hold-down bolts may be too close to the epoxies will seep through small holes ends of the pad. much more readily than cements will. When installing the forms, ensure that 6. Dig the Hole they are level and square. They must be Once the orientation of the pump pad securely fastened to the floor or ground, is determined, the next step for proper as the weight of the cement or epoxy installation is to dig the hole for the pump could easily shift the forms. If the forms pad. The depth is determined by three shift during the pour, there is a good chance that the job will need to be termifactors: nated and started over. This would be a • Is the pump inside or outside? big waste of material and man-hours, and • Is there a seasonal climate change? • What is the condition of the soil at the could further delay a large project. It pays to take the time and energy to do the job pump site? correctly the first time, rather than to have Inside pump pads can normally be 18- to redo it. For epoxy pours, the insides of the inches deep. Outside pump-pad sites that forms need to be coated so that they do have winter freezes are normally 36not stick to the epoxy. If the forms are inches deep, or go below the frost line. In not coated, they cannot be removed from areas where there is no winter freeze, 18 inches is adequate, but keep in mind the the epoxy without destroying them. Wax rule-of-thumb for pump-pad mass. works well for this. It is easier to coat the Another factor that must be considered forms with the wax before they are assemis the soil condition. Loose soils may need bled and installed. more of a base than more packed soils do. Items that will be needed for this step As a general rule, if the pump-pad height are transportation (for the forms and is maintained at a maximum of 18 inches, associated hardware) to the job location, a inside pads are 18-inches high and 18- Hilti gun, drill, anchor bolts (or other inches deep, while outside pads are 18- means of securing the forms to the floor or inches high and 36-inches deep. ground), bracing material, a square, a level, You now have the location of the pump necessary wrenches, bolts and nuts to pad, the dimensions of the pump pad, the assemble the forms, and a means to seal orientation of the pumps, and you know the form to the floor or ground etc. Some the climate in which the pumps will procedures may require drilling or other operate. Use that data to mark the location, types of permits. and dig the correctly sized hole to the correct depth. You will usually need a 8. Rebar & Pump Hold-Down Hardware cement saw, jackhammer, air supply, Now that the forms are squared, level wheelbarrow, shovels, a small front-end and secured to the floor or ground, it is loader, a disposal method for the excavated time to install the rebar and pump holdThe Pump Handbook Series

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down bolts. First, install the rebar. For an average-size ANSI pump pad for two pumps, one that measures about 60"L X 60"W X 36"D, there are usually three rings of rebar tied to eight posts of rebar. A 60"L X 60"W X 54"D pad would have four rings connected to eight posts. The rebar should be several inches in from the sides, top, and bottom of the pad, and equally spaced from top to bottom. A number four (#4) rebar is usually sufficient. Once the rebar is installed, install the proper hold-down Photo 1. Forms, rebar and hold downs–ready to poor bolts. An acceptable hold-down bolt for most ANSI pumps would consist the hold down bolt cannot properly of a 5/8-inch j-bolt mounted in a 1-1/2 inch stretch, the bolt could snap, and/or the pipe. The pipe should be at least six inches pump base may not have the proper long. The j-bolt should extend past the top torque to hold it properly (Photo 1). of the pipe by a minimum of the pump base height plus one and one half inch. The j- 9. Pour Preparation bolt should extend past the bottom of the The pour preparation is no more than pipe an inch or two before the "j". The assembling all of the needed material bottom of the pipe needs to be sealed by neatly at the pour site. If pouring concrete, welding on a washer that the j-bolt will pass this may only mean providing access for through. Then the j-bolt is welded to the the cement truck and supplying a hoe, a washer. Stainless steel all-thread with a trowel and a bucket of water. If the pour washer, double nutted to the bottom of it, site is inaccessible for a cement truck, is also a good combination. The hold-down include a wheelbarrow and a ramp (and bolts must be mounted true. When 2 boards to run the wheelbarrow on, if the pumps are on a pad, ensure that both sets ground is soft). More precautions will be needed if of hold-down bolts are true to themselves, and to each other. You can do this by having pouring epoxy, because of the need to pour a single bracket that secures all of the hold- the entire base in a timely manner, and down bolts. This also ensures that because of the floor area consumed by prefabbed piping will fit as designed. staging the material. Using the previously Usually wood 2 X 4s, mounted across the mentioned example of a pump pad with top of the forms, are sufficient for secur- the dimensions of 60"L X 60"W X 36"D, ing the hold-down bolts to the form on the epoxy and aggregate material alone single-pump pours. Reusable forms will would consume an area of about 70 square feet. This would not include the area for have a more elaborate system. After the j-bolts are installed, proceed mortar mixer or dumpsters. Before the to fill the space between the hold-down pour can begin, the aggregate will need to bolt and the pipe that surrounds it, so that be acclimated (for up to three days if the the grout does not fill the cavity. Expand- material is chilled) to 70-80°F. This is so ing insulation foam works quite well for the epoxy reaction can occur as designed. this. When this cavity is filled, the threads If outside in a cold climate, a structure with above the pipe on the hold-down bolts a heat source will be needed around the need to be isolated from the grout so that pour area. Staging for an epoxy pour the bolt can stretch during the final torque- should include a mortar mixer, the epoxy, ing. Tape, "rubber" tubing or grease will aggregate, a dumpster for waste disposal, suffice. This is important. If the threads of a power supply for the tools (air tools are 338

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recommended), a shovel and trowel and a covering for the floor for spillage (cardboard). A bucket of water is useful for keeping the tools clean during the pour. There needs be a way to clean the mixer (a small hydroblaster does a fine job), as some of the epoxy might begin to set before the pour is complete. A means of transporting the mortar mixer to a cleaning area also will be needed if the mixer cannot be cleaned at the job site. Remember, for both cement and epoxy pours, that the freelength areas of the threads on the hold-down bolts need to be coated so that the cement or epoxy does not stick to them. For epoxy pours, assure that the forms are coated before proceeding.

10. Pouring the Pad When pouring a cement pad, have the driver of the cement truck slowly fill the hole while vibrating the cement. If the job site is not accessible for a cement truck, the cement will need to be either pumped to the site, or loaded into a wheelbarrow and transferred to the site. In either instance, trowel-finish the top of the pad when the cement is ready for finishing. For an epoxy pour, there is more work involved, but the finished product is more durable, and does a better job of transmitting the vibrations away from the equipment. First, dedicate at least four people to the job. These four must know the routine of pouring epoxy. The first person will be mixing the epoxy resin parts A and B as detailed in the directions. He/she will also ensure that all of the part-A and part-B cans have been intermixed to eliminate problems of disposal. When mixed together, the epoxy is nonhazardous. If left separate, the material may need special disposal precautions. The second person will be operating the mortar mixer, a job that involves adding the mixed resin into an empty mortar mixer, and then adding the aggregate. The aggregate should consist of pure silica, some with the texture of sand and some with the

texture of pea gravel mixed at a specific promptly. ratio. The ratio can vary depending on the weather, temperature etc. Pure silica is used 11. Post-Pour Clean-Up in lieu of sand and pea gravel because of its The post-pour clean-up will consists of superior heat-sink capabilities. This also two parts--cleaning the equipment and adds to the overall strength of the pour. cleaning the job site. When the pour is Once the silica aggregates are added to the completed, the tools must be cleaned mixed resin, they are mixed until a uniform before the cement dries, or before the consistency is achieved. The mixing must epoxy starts to cure. A small hydroblaster be done slowly to ensure that no air is works quite well. entrained in the batch, as air is detrimenRemember that some epoxy resins are tal to the overall strength of the pour. A considered hazardous material before parts spiral-blade mortar mixer works best for A and B are mixed together. Once mixed, this. Follow the specific directions listed on the material is not hazardous. Some resins the containers. The mixed epoxy is then are also water-soluble. Knowing these two poured into the hole, and the process is factors greatly simplifies the clean-up repeated until the pour is complete. procedures. The resin buckets simply can The third person needs to keep the area be crushed and recycled, or thrown away. clean. This may sound menial but, consid- The mortar mixer and tools can be hydrobering the above pour will have about 150 lasted, without the worry of capturing the bags and 55 buckets, the area could epoxy contents for disposal. When all of become too cluttered to work in a short the tools are cleaned, return them to their time. proper location so they are ready for the The fourth person keeps the products next pour. staged where they are being used, and The other part of the post-pour cleanrelieves the others as needed. The needed up will be the removal of the pump-pad material should be conveniently staged so forms, and cleaning up the job site after the that the pour can be made in a quick and forms are removed. This step can usually efficient manner. An average-size pad of begin as soon as the next day with cement, 60W X 60L X 36H can be poured in less but one may prefer to wait an additional than 4 hours when properly administered. day. The form removal and clean-up can Skreeting is all the finishing needed on an epoxy pour, with minor trowelling around the hold-down bolts. Finishing needs to be completed before the epoxy cures. Some may consider pouring an entire pad out of epoxy excessive, but there are many benefits. Initial costs can be far outweighed by the superior vibration dampening aspects of epoxy, and by it's chemical resistance. This may be more true on large high dollar equipment, but it also applies to less costly, but critical, equipment. Another major benefit is the time Photo 2. Complete, accessible and asthetic saved on installing equipment. If you only have a one-week shutdown to definitely begin the next day with epoxy. get equipment installed, the cement will Removing the forms from a cement pad will not be adequately cured to apply the forces not impose any difficulties, as the forms will needed to attach the equipment. The not stick to the cement. Start by removepoxy will be cured in 24 hours, and allow ing the hardware from the forms. Then, pry the equipment installation to begin or use a dead-blow hammer to break the The Pump Handbook Series

forms free from the cement pad. When the epoxy-pad forms are properly prepared (coated), an average-size dead-blow hammer should release the forms from the epoxy pad, once all of the hardware is removed. If, however, the forms are not properly prepared, it will be quite difficult to remove the forms from the epoxy pad, and the forms will likely need extensive repairs. The epoxy adheres to the forms. If the forms are reusable, they should be repaired as needed and stored in their proper location. Then sweep up the area and leave it in a clean condition.

12. Set/Grout the Pump Bases This step begins with staging the needed tools and material, and preparing the pump bases. If cement is used, this step should take place no sooner than 7 days after the pour, but preferably longer, to allow sufficient cure time for the cement. If epoxy was used, this can be started the next day. The first step is to ensure that the pump-base bottom is properly prepared, so a good bond occurs between the grout and base. For metal bases, grit blast to remove all rust, paint, and other foreign material. This should be done within 24 hours of the grouting procedure so that the bases do not have time to re-rust. Non-metallic bases can be lightly sanded to roughen up the bottom to ensure a good bond. For metal bases, fill all of the needed holes on the pump base so that the grout does not fill them. Using the properly sized bolts in the holes, and isolating the treads on the bottom of the base from the grout, will allow the bolts to be removed, and will leave a cavity for the original bolts to fill. Do not fill the vent holes. Now, the pump bases can be installed over the hold-down bolts. Using shims, leveling screws (again with threads protected), or wedges, level the bases. Eliminating any sharp corners on the wedges or shims will help eliminate stress risers. Check level in two directions, perpendicular to each other. Do not use nuts on the bottom side of the base on the hold-down bolts for leveling, as this will not allow the bases to be properly 339

PS0601PG22

torqued down, and it will negate many of the benefits of proper equipment installation. When the bases are leveled, it is a good idea to bolt them down from the top, tightening the nuts about a 1/4 turn after hand tight. This helps keep the base secure in case it is bumped. Then recheck the level. If it is not level, re-level and proceed. Seal off the ends and sides of the pump base. Duct tape works well for sealing the bases before grouting. With the bases sealed, you should now be ready to pour the grout. Cement types of grout are not recommended. Epoxy grouts are superior in bonding strength and chemical resistance, and they shrink less. Tools that will be needed are a hoe, a wheel barrow, a bucket of water, a large funnel, a five-gallon and a one-gallon pail, floor covering, a trowel, epoxy resin parts A and B and sand-sized silica. Mix parts A and B of the resin per instructions, and pour the mixed resin into the wheelbarrow. Pour the proper amount of silica in the wheelbarrow and continue stirring it with the mixed resin until a uniform consistency is reached. Depending on accessibility of the pump bases, it is usually easier to fill the five-gallon bucket using the one-gallon bucket. Then pour out of the five-gallon bucket into the funnel. An orange road cone is an ideal funnel, as the opening is sufficient and the height gives about a two-foot head-pressure to assist the grout flow. This system will completely fill a one-half inch gap, 15" wide by 52" long and more, using only the center grout hole. Pour the grout material and allow it to fill the base cavity by pouring it in through the grout hole(s). Vent holes are needed to allow any trapped air to escape. If your base has bracing underneath, vent holes should be placed on both sides of the bracing to allow for a full bond. If any grout spills on the top of the pump base, it should be cleaned up immediately after the grout pour, before it cures, as it will be much more difficult to clean off the pump base after it cures. Allow the grout to full-cure before proceeding to remove any forms. After full-cure, remove the sealing material (duct tape) and the wedges or shims, if possible. Use silicone caulking to fill the shim holes. Then, tighten the base down to the recommended torque. Clean up any mess before proceeding to the 340

next step.

13. Installing the Pump Once the grout has cured and the pump base is clean and ready to receive the pump and motor, the bolts that were used to seal the needed bolt holes can be removed. Mount the pump as recommended by the manufacturers' installation directions. The pump should be in operation-ready condition when it is installed on the base.

14. Installing the Motor Install the motor per the manufacturer's specifications. After the motor is wired, do the rotation check. Be sure that the pump and motor are uncoupled for the rotation check! Pumps, seals or mag-drive bearings can be ruined if operated dry or in reverse. If hour meters or power monitors are to be installed on the system, install them at this time. Some equipment installed to the above procedures has run so quietly that operators have been unaware that the pumps were operating. Indicator lights could be added to the local start/stop switch to avoid this problem, and would be installed at this time. After the rotation check, de-energize the motor and lock out the breaker. Then, reinstall the coupling, but not the coupling guard.

15. Pre-Alignment

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engineering practices for proper piping. Pipe up to the pump. Install the piping to the pump with no pipe strain. Have sufficient piping support, separate from the pump pad. Allow a straight, unobstructed flow of at least 10 pipe diameters into the suction of the pump. The suction line needs to be at least as large as the suction of the pump, and going one size larger is recommended. Use an eccentric reducer. Do not construct the suction piping so that air pockets can form (i.e., eccentric reducers should have flat side up). Use strainers where needed. Have discharge and suction gauges installed on the pump side of the isolation valves. Use pump connectors if your standards allow. Sharp bends in the piping for buffer pots for double/tandem mechanical seals should not be allowed. Following proper piping procedures may take more time and dollars up front, but the lifetime cost of the equipment will be considerably less, and the reliability higher.

17. Final Alignment After all piping is complete, do the final alignment. If proper procedures were followed throughout this process, there is a good chance that this will only be an alignment check. Once again, straightedge alignment is not acceptable. File the alignment records in the equipment file, along with the pump repair manuals, installation manuals, original pump curves and copies of the original purchase order. This information could be valuable down the road if changes are made to the system. Verify that the necessary lubricants, guards, plugs and other components needed for operations are in place. After verifying all this, baseline vibration data can be recorded. If all components pass, you can proceed to the next step.

For this step, perform a "rough" alignment, aligning the pump-motor assembly to final tolerances. Again, use double-dial, or laser-alignment methods. An alignment is performed at this time in case the pump has shifted. If the pump did shift, the odds are that the motor would become boltbound during the final alignment, after all of the piping was installed. It is much easier to correct the problem before the piping is attached to the pump. Altering the pump location after the piping is attached would 18. Release to Production increase the amount of pipe strain on the The pump should now be ready for pump casing, especially if there are no operation and for release to production. Be pump connectors in the system. present for the initial run. There are two reasons for this. First, if there is a problem, 16. Piping Up the Pumps this is the most likely time that it will occur. Now that the pump and motor are Observing any abnormality first-hand is installed on a solid base, the plumbing can much more informative than hearing about be piped up to the pump. Follow good it second-hand. The second reason is to The Pump Handbook Series

obtain the initial operation-condition baseline vibration readings. If these installation procedures have been followed, the low levels of the initial readings may surprise you. Do not forget to get the proper repair items stocked into stores. Also be sure that equipment files AND maintenance get the needed copies of the maintenance and repair manuals. If the equipment was installed correctly, though, it should be some time before they are needed.

Conclusion The old saying about “an ounce of prevention…” is appropriate to pump installation. Taking the time and making the effort to do things right at the beginning can avoid “pounds” of operational and maintenance headaches down the road (Photo 2). ■ Joseph F. Dolniak has been a Reliability Engineer at Eli Lilly and Company's Tippecanoe Manufacturing facility for the past 2 years. He previously was employed as a Maintenance Engineer at a specialty chemical manufacturer for 10 years, and as a Reliability Engineer at a coal-fired Electric Utility for 7 years. Mr. Dolniak is a current member of the ANSI B73.1 committee. He is active in the International Pump User Symposium, having lectured, presented a short course, and assisted with Discussion Groups. He also has presented at PumpUsers Expo, and has had numerous articles published on the topic of equipment reliability. E-mail him at [email protected] . (Editor’s Note: This article was adapted from the Official Proceedings of PumpUsers Expo 2000 in Louisville, KY.)

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POSITIVE DISPLACEMENT PUMPS HANDBOOK Table of Contents NPSH Required for Reciprocating PD Pump .............................................................1 Centrifugal and Positive Displacement Pumps in the Operating System.................3 Positive Displacement Two Screw Pumps ..................................................................5 Expansion Joints and Air Chambers ..........................................................................9 Selecting a Progressing Cavity Pump ........................................................................11 A Review of Positive Displacement Pumps..............................................................15 Handling Abrasives and Corrosives with Positive Displacement Pumps ...............19 Fluid Metering System Options.................................................................................22 Pulsation and Surge Control .....................................................................................27 Chemical Additive Pumps for Paper Mills ...............................................................31 Air Operated Diaphragm Pumps - Nineties Style ....................................................35 Positive Displacement Pump Vibration....................................................................39 Specifying Air-Operated, Double-Diaphragm Pumps ..............................................42 Pipeline Screw Pump Efficiency ...............................................................................43 In the Pipeline: PD Screw Pump Valving.................................................................48 Valve Dynamics Affect Recip Pump Reliability – Part 1.........................................52 Valve Dynamics Affect Recip Pump Reliability – Part 2.........................................55 Reducing Pulsations in Metering Pumps .................................................................60 Maintaining and Operating Positive Displacement Rotary Gear Pumps ...............64 Peristaltic Pumps Offer Custom Fluid Solutions.....................................................67 Selection Guide: Rotary Gear Pumps .......................................................................71 Gear Drive Options for Rotary Pumps .....................................................................75 Rotary Pump Startups...............................................................................................77 Rotary Pump Troubleshooting..................................................................................82 Selection Guide to Metering Pumps .........................................................................86 Reciprocating Metering Pumps in Leak-Free Design ...............................................90 Chemical Injection: Simplex or Complex? ...............................................................96 Built-in Relief Valves: The Case For and Against...................................................100 Applying the NPSHR Standard to Progressing Cavity Pumps.............................103 The Canned Rotary Pump - Circa 1997 ...............................................................108

POSITIVE DISPLACEMENT PUMPS HANDBOOK Table of Contents Metering System Design Requirements....................................................................111 Sealing Technology for VOC Control........................................................................114 A Users Guide to Rotary Pumps .......................................................................................118 Rotary Pump Inlet Pressure Requirements .............................................................124 Rotary Pump Overhaul Guide...................................................................................127 A Guide to High Pressure Reciprocating Pumps ....................................................................129 Sealing Hazardous Fluids with Dry Seal Technology ................................................133 Well Pump Applications for Mine Dewatering...............................................137 Sealing Positive Displacement Pumps: What Are Your Choices?...................141 Flow Control With Metering Pumps .....................................................................................................146 Want to Go From 0 to 100 in 3 Seconds?.......................................................150 Maximize the Performance of Progressive Cavity Pumps ............................153 Minimizing Pressure & Flow Pulsations from Piston/Diaphram Metering Pumps ...........................................................................................................161 Installation, Start-up and Operation of a Reciprocating Pump.....................165 Extending the Life of Positive Displacement Pumps – Part 1: Gear Pumps ..........172 Extending the Life of Positive Displacement Pumps – Part 2: Multiple Screw Pumps.............................................................................................176 Extending the Life of Positive Displacement Pumps Part 3: Progressive Cavity Pumps............................................................................................179 Peristaltic Pump Technology for Industrial Applications.............................182 Pulsation Control for Reciprocating Pumps ............................................187 Precision Solutions for Low-Flow Handling ....................................................................194 Metering Pump System Design for Dependable Performance ....................199 Twin-Screw Pumps vs. Centrifugal and Reciprocating Pumps ...................205 Using PC Pumps and Other Rotary PD Pumps for Metering ..........................213 Multphase – The Final Pumping Frontier ......................................................................................219 All materials © 2002 Pumps & Systems, LLC. No part of this publication may be reproduced without the written consent of the publisher. The publisher does not warrant, either expressly or by implication, the factual accuracy of the articles or descriptions herein, nor does the publisher warrant the accuracy of any views or opinions offered by the authors of said articles or descriptions.

POSITIVE DISPLACEMENT PUMPS HANDBOOK

NPSH Required for Reciprocating PD Pump BY JIM MILLER Many times I am asked to predict the flow of a positive displacement pump operating under conditions with insufficient inlet pressure to ensure complete filling. In these circumstances the mixture may be considered two-phase, with either fluid vapor or some other entrained gas (air) present. Is there any literature dealing with the theory necessary to evaluate the filling capability? A: Adequate suction pressure for a reciprocating positive displacement pump is an often misunderstood concept. Pump manufacturers are asked the minimum suction pressure that can be used for both continuous and intermittent service for a particular pump. The net positive suction head (NPSH) required may or may not be provided. NPSH required takes into account the fluid properties of water (usually), pump dynamics, and pump design. When analyzing a pump’s operation at very low suction pressure, several pump parameters must be studied.

Q:

NPSH REQUIRED The subject of NPSH for such pumps is also frequently misunderstood. The definition in the Hydraulic Institute Standards of minimum NPSH required is established for a specific pump, pump speed, and discharge pressure by reducing the suction pressure until one of the following conditions occur: 1.

a 3% drop in capacity

2.

clearly audible cavitation noise

Several things are wrong with operating a pump at the NPSH required suction pressure. By definition, the pump is cavitating when condition 2 occurs. Loss of capacity according to condition 1 means that the liquid chamber is only partially filling as the result of cavitation.

= v1 -

p1 __ v1 p2

b) gas saturation pressure

= v1 -

p1v1 _______ p1 + ∆p

c) fluid compressibility

∆liquid volume = βυ0∆p

Mechanical design

∆volumeplunger =

NPSH required accounts for 1) Fluid properties a) vapor pressures

2)

a) chamber volume b) valve design

3)

= v1 -

p1v1 _______ + βυ0∆p p1 + ∆p

c) piping system

Solve for ∆p, the pressure

Operating conditions a) pump speed

(βυ0)∆p2 + (βυ0p1 + v1 ∆volumeplunger) ∆p + (-∆volumeplungerp1) = 0

b) discharge pressure c) fluid temperature The meaningfulness of the NPSH required will become questionable if any of these parameters change significantly for a given application.

a = βυ0 b = βυ0p1 + v1 - ∆volumeplunger c = -∆volumeplungerp1 ________ -b ± √(b2 - 4ac) ∆p = ________________ 2a

PUMP LIQUID CHAMBER It is difficult to predict whether a pump will operate with some degree of vapor or gas breakout. Given a known volume and pressure of the gas or vapor escape, the solution is relatively easy to solve. The difficult issue is the amount of gas initially formed in the pump chamber in the first place. For the pump to continue to operate, the cylinder pressure must exceed the discharge manifold pressure at the end of the plunger’s discharge stroke. The following approach has been used to solve for pressure at the end of the discharge stroke: Given: ∆volumeplunger = plunger displacement ∆volumeplunger = ∆gas volume + ∆liquid volume ∆gas volume = v1 - v2

The Pump Handbook Series

For a typical 2-in. plunger by 6-in. stroke pump the following values have been calculated where: ∆volumeplunger = 18.85 in.3 β = 0.000003/psi for water liquid chamber volume = 100 in.3 υ0 = 100 (1 - gas fraction), (in.3) v1 = 100 x gas fraction (in.3) p1 = 30 psi As can be seen in Figure 1, this pump operating at 1,000 psi discharge pressure will become vapor locked when the gas volume is equal to 19% of the chamber volume. The graph suggests that a reciprocating pump will operate at very low volumetric efficiencies, but in fact it will stop pumping when it falls to the 75-85% volumetric efficiency range. This is

1

CONCLUSIONS Minimum NPSH required suction pressure should not be used for a continuous-duty pump application. As general practice, suction pressure should be 10 psi above the minimum NPSH required for the pump. This additional pressure will significantly reduce the potential for pump cavitation problems such as damaged plungers, valves, and pump liquid ends. The minimum NPSH required could be used as a conservative suction pressure to assure that the pump will continue to operate in an intermittent service. Jim Miller is president of White Rock Engineering in Dallas, TX. He has degrees in chemical engineering and business administration from the University of Texas at Austin. WRE offers reciprocating pump hydraulic analysis consulting and maintenancefree pulsation products.

2

FIGURE 1 100000

Maximum Cylinder Pressure - psig

probably because when the gas or vapor phase develops, it very quickly exceeds the 19% volume of the liquid chamber.

10000

1000

100

10 0

0.05

0.1

0.15 0.2 Liquid Chamber Gas Fraction

0.25

Maximum cylinder pressure versus liquid chamber gas fraction.

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0.3

POSITIVE DISPLACEMENT PUMPS HANDBOOK

Centrifugal and Positive Displacement Pumps in the Operating System BY ROSS C. MACKAY

I

n the many differences that exist between centrifugal and positive displacement pumps, one which has caused some confusion is the manner in which they each operate within the system. Positive displacement pumps have a series of working cycles, each of which encloses a certain volume of fluid and moves it mechanically through the pump into the system, regardless of the back pressure on the pump. While the maximum pressure developed is limited only by the mechanical strength of the pump and system and by the driving power available, the effect of that pressure can be controlled by a pressure relief or safety valve. A major advantage of the positive displacement pump is that it can deliver consistent capacities because the output is solely dependent on the basic design of the pump and the speed of its driving mechanism. This means that, if the required flow rate is not moving through the system, the situation can always be corrected by changing one or both of these factors. This is not the case with the centrifugal pump, which can only react to the pressure demand of the system. If the back pressure on a centrifugal pump changes, so will its capacity. This can be disruptive for any process dependent on a specific flow rate, and it can diminish the operational stability, hydraulic efficiency and mechanical reliability of the pump.

CENTRIFUGAL PUMP PERFORMANCE CURVE The interdependency of the system and the centrifugal pump can be easily explained with the use of the pump performance curve and the system curve. A centrifugal pump performance curve is a well known shape which shows that the pressure the pump

can develop is reduced as the capacity increases. Conversely, as the capacity drops, the pressure it can achieve is gradually increased until it reaches a maximum where no liquid can pass through the pump. Since this is usually a relatively low pressure, it is rarely necessary to install a pressure relief or safety valve. When discussing the pressures developed by a centrifugal pump, we use the equivalent linear measurement referred to as “head,” which allows the pump curve to apply equally to liquids of different densities. [Head (in feet)=Pressure (in p.s.i.) x 2.31+ Specific Gravity of the liquid]

When the pump curve is superimposed on the system curve, the point of intersection represents the conditions (H,Q) at which the pump will operate. Pump Curve

System Curve

Pumping conditions change ONLY through an alteration in either the pump curve or the sysSYSTEM CURVE tem curve. When considering possible The system curve represents the movements in these curves, it pressures needed at different flow rates should be noted that there are only to move the product through the sysa few conditions which will cause tem. To simplify a comparison with the pump curve to change its posithe centrifugal pump curve, we again tion and shape: use the ‘head’ measurement. The sys• wear of the impeller tem head consists of three factors: • change in rotational speed • static head, or the vertical eleva• change in impeller diameter tion through which the liquid • change in liquid viscosity must be lifted Since these conditions don’t nor• friction head, or the head required mally develop quickly, any sudden to overcome the friction losses in change in pumping conditions is the pipe, the valves and all the fitlikely to be a result of a movement tings and equipment in the system curve, which means • velocity head, which is the head something in the system has required to accelerate the flow of changed. liquid through the pump (Velocity Since there are only three ingrehead is generally quite small and dients in a system curve, one of often ignored.) which is minimal, it follows that As the static head does not vary either the static head or the friction simply because of a change in flow head must have changed for any rate, the graph would show a straight movement to take place in the sysline. However, both the friction tem curve. head and the A change in the static velocity head head is normally a result of will always a change in tank level. If vary direct- Head System the pump is emptying a Friction & ly with the Curve Velocity Head tank and discharging at a capacity. The fixed elevation, the static combination Static Head head against which the of all three Capacity pump must operate will be creates the gradually increasing as the system curve. The Pump Handbook Series

3

suction tank empties. This will cause the system curve to move upwards as shown.

An increase in friction head can be caused by a wide variety of conditions such as the change in a valve setting or build-up of solids in a strainer. This will give the system curve a new slope.

When the operating conditions of a system fitted with a centrifugal pump change, it is helpful to consider these curves, focus on how the system is controlling the operation of the pump, and then control the system in the appropriate way. ■ Ross C. Mackay is an independent consultant located in Tottenham, Ontario, Canada. He is the author of several papers on the practical aspects of pump maintenance, and a specialist in helping companies reduce their pump maintenance costs.

Both sets of events produce the same result: a reduction of flow through the system. If the flow is redirected to a different location (such as in a tank farm), it means that the pump is now operating on an entirely new system which will have a completely different curve.

Thus, it is clear that regardless of the rated capacity of the centrifugal pump, it will only provide what the system requires. It is important to understand the conditions under which system changes occur, the acceptability of the new operating point on the pump curve, and the manner in which it can be moved.

4

The Pump Handbook Series

POSITIVE DISPLACEMENT PUMPS HANDBOOK

Positive Displacement Two Screw Pumps BY ROBERT A. PLATT he two screw pump is a positive displacement, rotary screw type distinguished by its high flow rate, suction lift and differential pressure capabilities (Figure 1). Based on a simple design principle, these pumps operate without internal contact between the pumping elements. This enables them to run dry and pump virtually any fluid, regardless of its viscosity or lubricity, and also gives them excellent flexibility and longevity. These rugged and versatile pumps are especially well-suited for difficult fluid handling situations and applications above 125 centistokes, where their efficiency and cost advantages are most apparent.

T

FIGURE 1

HOW IT WORKS As a double-ended pump, fluid enters the suction port and divides into left and right flow streams. The fluid flows outward toward the ends of the pump, joining the rotating screws. As the screws rotate, they form chambers without contacting each other. These chambers capture the incoming fluid and push it axially through, simultaneously pumping from left-to-center and right-to-center. The flow streams develop pressure as they approach each other. They merge at the center and exit the discharge port at discharge pressure, like two opposing half-flow pumps on a common shaft.

INTERNAL DESIGN The two screw pump consists of two shaft and screw assemblies. The screws are integral with the shafts or separate and attached during pump assembly, depending on the manufacturer. The bearings are oil or grease lubricated and separate from the fluid being pumped. Some manufacturers also offer an internal bearing (i.e., bearings exposed to the fluid). However, this design requires clean fluids with some lubricity and has limited flow and differential pressure capability. Beyond a certain shaft diameter, soft packing cannot be used due to

Cutaway view of a two screw pump the required large sealing surface area; a mechanical seal must be used instead. Since these pumps are designed for single-, double-, tandem-, or cartridge-type seals, this is rarely a problem. The seals are on the suction side of the pump and, with some designs, directly in the fluid flow stream. With this arrangement, fluid flushes and cools the seals as it flows around them, extending their service life in the process. The symmetrical two screw design results in screws balanced in the axial direction. Some manufacturers also balance the shafts and screws radially, so that only the hydraulic forces (during high-pressure operation) are significant. As a result, the mean time between failure of the bearings and seals often exceeds two years. By operating at standard driver speeds, the space, cost and maintenance of speed reducers are also avoided. With symmetrical screws and casings, bi-directional pumping is possible. The suction and discharge lines are reversed by shifting the driThe Pump Handbook Series

ver rotation. No other valve or piping changes are necessary. As a precaution, make sure the mechanical seal and inlet flange ratings are sufficient for the expected discharge pressure. Another distinguishing feature of the two screw pump is its external timing gears. These transfer torque from one shaft to the other and time the pump screws. Timing the screws means adjusting the timing gear backlash to position one shaft/screw set relative to the other. When properly timed, the screws interlock with uniform clearances between them and with no physical contact. Timing gears are at the inboard or outboard end of the pump, depending on the manufacturer. They have their own lube oil reservoir and rarely require external lubrication or cooling.

CASING MATERIALS AND OPTIONS Casings are available in cast or nodular iron, bronze, cast or welded steel and various grades of stainless steel. For dirty fluid applications some manufacturers offer hard-face coated casing bores and chrome- or stellite-coated screw outside diame-

5

ters (ODs). Some also have a replaceable casing liner to offset wear. Screws and shafts are typically made of cast or nodular iron, bronze and various stainless steels, but they are available in other materials. When selecting screw and shaft materials, make sure they are compatible with both the application and the pump casing or liner material. Two screw pumps are also available with cartridge-type inserts, safety relief valves and integral heating jackets. A cartridge insert makes it possible to remove the shafts, screws, liner, gears, bearings and seals as one unit. The entire pump can then be inspected quickly without disturbing the piping system or driver. Relief valves protect the pump from over-pressurization by diverting discharge flow to the suction side of the pump. These valves are not for continual flow or pressure regulation, and are usually for one-direction flow only, although bi-directional valves are available. For high temperature fluids, heating jackets are used with steel and stainless steel pumps. The jackets are either a partial or full type, and steam or hot oil is usually the heating medium. Full-heating jackets encompass the entire pump casing, as well as the relief valve passages (when provided) and mechanical seals. Some manufacturers offer partial-heating jackets and electric immersion-type heating coils for applications not requiring full-pump casing heating. These alternatives save the expense of a full-heating jacket while providing sufficient heat for a trouble-free start.

MOUNTING OPTIONS Two screw pumps can be mounted in several ways. The most common is horizontally with side inlet and side or top discharge connections. The casing is usually symmetrical about its vertical centerline with the inlet flange centerline above the screw set centerline. This keeps the screws wetted after a shutdown so the pump can reprime quicker on subsequent restarts. It also protects the seals from dry running during priming, tank stripping or suction line evacuations. Other designs place the inlet centerline below the vertical

6

A multiphase two screw pump centerline to reduce sufficiently the NPSH Required. A second mounting option is vertical with in-line suction and discharge connections. The casing is symmetrical about its vertical centerline, permitting suction or discharge connection on either side. The connections are also above the screw set horizontal centerline to prevent air entrapment and vapor locking of the upper screw set.

SELECTING A PUMP Two screw pumps are sized for an application based on their casing, shaft, and screw diameter sizes. Casing size is based on the screw and shaft diameters. Screw diameters vary from 50 to 400 mm. After selecting a screw diameter, choose the screw pitch and speed to fine tune the flow, pressure, and viscosity requirements. The screw pitch is the axial distance a screw travels as it rotates through one full revolution, ranging from 10 to 300 mm. As the screw pitch increases, so does the size of the pumping chamber. This increases the flow per revolution but also means fewer chambers fit in the pump. Since differential pressure capability is related to the number of chambers, The Pump Handbook Series

fewer chambers mean less differential pressure capability. Most two screw pumps can develop 150 psi per chamber.

PERFORMANCE CHARACTERISTICS For any positive displacement pump, output flow is only marginally affected by discharge pressure. The output flow of the two screw pump is a function of the pump geometry (screw diameter, screw pitch and internal clearances), rotational speed, viscosity and the differential pressure across the pump. Clearances in the pump exist between the screw OD and casing ID, between the screw OD and adjoining shaft root diameter OD, and between adjoining screw flanks. These clearances are areas of potential leakage, or slip, paths for the higher discharge pressure to cause some fluid to flow backwards toward the lower pressure suction end. Flow equations vary among manufacturers, but all follow the basic form: Qactual = Qtheoretical – Qslip Qtheoretical = AN [D]2 Qslip = B [∆P]1/2 [1/ν]1/3 Qactual = AN [D]2 – B [∆P]1/2 [1/ν]1/3

FIGURE 2 1.00

0.95

0.90

Efficiency (%)

VOLUMETRIC EFFICIENCY MECHANICAL EFFICIENCY

0.85

OVERALL EFFICIENCY

0.80

0.75

0.70

2000

1900

1800

1700

1600

1500

1400

1300

1200

1100

1000

900

800

700

600

500

400

300

200

100

0.65

Vicosity (Centistoke)

Typical two screw pump efficiency curves where Q is the flow rate A is the screw pitch N is the pump speed D is the screw OD B is an empirical flow slip factor ∆P is the differential pressure ν is the fluid viscosity As flow slip increases, the total output flow decreases. The volumetric efficiency, a key performance benchmark, decreases as well. Volumetric efficiency is defined as: ηvol = [Qtheoretical – Qslip] Qtheoretical = Qactual Qtheoretical Two screw pump volumetric efficiencies range from 70 - 97% and are typically from 85 - 95%. Theoretical power is a function of the flow rate and differential pressure across the pump. The actual power required is the theoretical power plus the power to overcome the mechani-

cal losses. Mechanical losses in a two screw pump result from the viscous drag of the rotating screws and the friction of the seals, bearings and timing gears. Equations for determining power requirements follow the general form: Wactual = Wtheoretical + Wlosses Wtheoretical = C [Qtheoretical] ∆P Wlosses = F [N]4/3 ln[n] Wactual = ∆P[Qtheoretical][C] +F [N]4/3 ln[n] where W is the power required C is a power conversion factor Q is the theoretical flow rate ∆P is the differential pressure F is a friction loss factor N is the pump speed ln is the natural logarithm ν is the fluid viscosity

The Pump Handbook Series

A second benchmark of two screw pump performance is mechanical efficiency, comparing the theoretical to actual power requirements. Mechanical efficiency is defined as: ηmechanical = Wtheoretical [Wtheoretical + Wlosses] = Wtheoretical Wactual Mechanical efficiencies range from 60 to 90% and are typically from 70 to 80%. A third measure of pump performance is overall efficiency, measuring overall performance and defined as: ηoverall = [Qactual Wtheoretical] [Qtheoretical Wactual] = [ηvol] [ηmechanical]

7

Overall efficiencies range between 55 to 75%. Typically they are from 65 to 70%. Figure 2 shows the typical volumetric, mechanical and overall efficiency curves for a two screw pump. Another advantage of two screw pumps is their high suction lift capability. The Net Positive Suction Head Required (NPSHR) is a function of the viscosity and axial velocity through the pump. Equations for calculating two screw pump NPSHR follow the general form: NPSHR = 5.0 + K1[Va]2 + K2 [Va]3/2 [ν]1/2 where K1 and K2 are empirical adjusting factors Va is the fluid axial velocity ν is the fluid viscosity Like the flow and power equations, these equations originate from classic fluid dynamic theory. Manufacturers then modify them to coincide with their experiences and test stand measurements.

APPLICATIONS One result of the pump’s low fluid velocities is laminar flow (i.e., NRe < 2000) when handling viscous fluids. This makes it a low fluidshearing pump and ideal for polymers, emulsions and other shear-sensitive fluids.

8

Two screw pumps can handle some abrasives, but are not specifically designed for dirty liquid services. If sand or solids are present, their size, hardness, and size distribution must be known. If the solids are soft, or consistently small enough to pass through the screw/casing clearances (typically 1 to 2 mm), their contribution to pump wear will be small. In all cases, a 1/8” to 1/4” perforation suction strainer should be used. One emerging application for two screw pumps is multi-phase flow pumping (Photo 1). This is a benefit to oil and gas producers operating in remote areas, where a water/oil/gas mixture can be moved by a single pump to a downstream separation and processing center, rather than doing it on site. Some pumps can even handle gas volume fractions up to 100%, making them a compressor as much as a pump. Because of their simple design, few moving parts and lack of any internal contact of wearing parts, two screw pumps have many applications: • Gathering and boosting on offshore platforms and crude oil pipelines. •

•

Off-loading viscous products from barges and ocean-going vessels. Transfer and blending in refineries, tank farms, and petrochemical plants.

The Pump Handbook Series

•

Low shear transferring in chemical plants.

•

Feeding polymers and resins to extruders, mixing tanks and pelletizers.

•

Pumping multi-phase fluids with gas volume fractions as high as 100%.

•

Stripping operations in storage tanks and tank cleaning operations.

•

Liquefied coal/water slurry transport.

•

Any large volume or difficult transfer application.

SUMMARY Over the last 60 years the two screw pump has been a heavy-duty, dependable fluid transfer pump. Its gentle, low shear pumping and compatibility with a wide range of fluids distinguish it from other pump types. Few pumps can match its design features, materials and mounting arrangements. Often considered one of the pump world’s best-kept secrets, the external bearing two screw pump is a high-quality solution to a variety of problems. ■ Robert A. Platt is President of Bornemann Pumps Inc., Monroe, NC.

POSITIVE DISPLACEMENT PUMPS HANDBOOK

Expansion Joints and Air Chambers BY BOB STOVER What determines the need for expansion joints and/or air chambers at the pump suction and discharge? I’ve seen applications, large and small, that don’t have these components, and I’ve seen some that do. All seem to be nearly the same in reference to the intent of the system, leading me to suspect that the need for these devices is in the eye of the beholder; some people believe they are necessary, while others do not. Because of the expense of this equipment, some

Q:

definitive information on the topic could save as much as several hundred dollars per pump. Expansion joints and accumulators correct very common but specific problems. Systems containing any of the flaws discussed below – and most systems do – would benefit from their use. In systems without these flaws, such accessories would be harmless but superfluous. Typical applications and general guidelines for these products include:

A:

FIGURE 1 Pressure Gauge 0 - 100 PSI Snubber Diaphragm Seal Ball Valve

Compound Gauge 0 - 30" Hg; 0 - 30 PSI

ACCUMULATOR TANKS Distinct pulsations in flow are common with positive displacement pumps. With some pumps, these pulsations are more pronounced. AOD pumps, for example, have clear, strong pulsations, while gear pump pulsations overlap and are therefore less distinct. Nonetheless, some degree of pulsation is almost always present. Centrifugal pumps, on the other hand, can pulsate during the onset of recirculation cavitation, or given certain clearance ratios between the splitter and the impeller periphery. But these situations are rarely troublesome or even noticeable. Water hammer can be another cause of strong pulses or shocks. It can do great damage due to the high pressures developed, estimated by: 62.4av ∆p = ———— 144g

Snubber Diaphragm Seal

where

Ball Valve

p = hammer pressure (above existing line pressure) in psi

Discharge Air chamber

a = speed of a pressure wave in the customer’s liquid in ft/sec (note: water = approximately 4,000 ft/sec) v = velocity before incident causing hammer in ft/sec

51.25" approx.

Suction Air Chamber 8" DIA. x 36" Long 1800 Cubic iinches TYP.

DISCHARGE 4" STD. 125# Flange Drilling 15 5/16"

By-pass Pipe SUCTION 4" Std. 125 # Flange - Drilling 2" Gate Valve for Draining

19.5"

1.5" 1/16" CL BASE

18.5"

21" REF.

A plunger-type pump with both suction and discharge air chambers.

The Pump Handbook Series

PHOTO COURTESY CARTER PUMP, INC.

Shaft Pump Base

g = acceleration of gravity (32.2 ft/sec2) Water hammer is usually caused by a valve closing too quickly for the system. A rule of thumb for the minimum closing speed of a valve is: 2L S = —— a where S = minimum closing speed in seconds L = length of pipe in feet

9

An accumulator tank (also called an air expansion chamber or pulsation dampener) can smooth out the pulsations or damp a certain amount of the water hammer, prolonging pump life and helping equipment such as flow meters to read more accurately (Figure 1).

ACCUMULATOR APPLICATION GUIDELINES A typical accumulator mounts in the line (ideally as close to the pump discharge as possible) on a tee, although in-line designs exist. Accumulators come in several general configurations, but the common distinction is whether a bladder (diaphragm) is necessary. A bladder will allow the accumulator to be charged with a gas to assist in the damping, without exposure to the danger of the gas escaping into the line should system pressures drop below the gas charging pressure. A bladder also allows certain accumulators to function as storage tanks. They are applied where a pump is sized for average demands but pump output can occasionally fall short. In these systems the accumulator can be sized in flow and pressure so that at times of peak demand, the pump output will be assisted by the liquid stored in the accumulator. This feature can also be of value when the accumulator is mounted on the suction side of a pump (usually a reciprocating positive displacement pump), where the reversing action of the piston causes sudden pressure drops and the possibility of cavitation. In this instance the accumulator provides fluid under pressure to help fill the pump cavity. Accumulators without bladders are less expensive and are suited for systems without wide swings in pressure or flow. These designs typically call for a gas (often air) to be injected into the vessel as often as a sensor (usually mechanical) indicates. Acoustical filter systems are available for applications with sophisticated or complicated patterns of pulses due to multiple pumps or sources of pressure. They consist of two or more vessels with a choke line connecting them that is sized so that the relative volumes will provide a damping

10

effect when the system experiences extreme pressures. Accumulators are available from many manufacturers, all of whom can help with a selection once given flows and pressure extremes.

FIGURE 2

EXPANSION JOINTS Expansion joints (also called flexible connectors) (Figure 2) An elastomeric spool-type expansion joint. are typically used to: pump designs allow rigid mounting— • Avoid transmitting stresses some close-coupled sanitary pumps, caused by pipe misalignment, or for example, must be mounted in a the stresses encountered when fashion that allows cleaning below temperature variations cause their mounting pads—expansion expansion or contraction of lines. joints are one of the few effective If allowed to distort a pump casmeans of insuring minimum pipe ing, these strains can cause prestrain if the pump or lines are jostled. mature wear of seal faces, early To avoid causing any further bearing failure, and, perhaps, stress to build up in the lines, expanwear in close tolerance parts, sion joints should be mounted with such as those found in many posan anchor placed between the joint itive displacement pumps. and the pump port(s). If connected • Damp a certain amount of vibradirectly to the pump, a misaligned tion, often inevitable even in cenexpansion joint could help develop trifugal pump installations. forces to a magnitude of: • Isolate and therefore prevent F = AP noise from being transmitted or magnified through solid pipe or where other parts of a plant. F = force at the pump port in lbs

EXPANSION JOINT APPLICATION GUIDELINES

A = expansion joint cross section in inches2

Expansion joints are usually of a slip-joint or corrugated diaphragm design, although flexible lines are sometimes used to similar effect when looped at pump ports to allow for line expansion. In systems where lines have been installed correctly, with proper supports, expansion joints are often a subjective matter, determined largely by a customer’s tolerance of noise levels or his willingness to subject associated equipment to vibration. However, anticipated movement of as little as 0.01–0.02 in. caused by heat expansion usually necessitates expansion joints. And because not all

P = line pressure in lbs/in.2

The Pump Handbook Series

This is often sufficient to fracture the casing. Both accumulators and expansion joints are accessories that add to the cost of a system. But because they address specific problems encountered in many operations, they protect a larger equipment investment and pay for themselves many times over through reliable operation. ■ Bob Stover is western regional manager for Tri-Clover, a manufacturer of pumps, valves, and fittings.

PHOTO COURTESY RM–HOLZ INC.

a = speed of pressure wave, as above

POSITIVE DISPLACEMENT PUMPS HANDBOOK

Selecting a Progressing Cavity Pump BY ALAN G. WILD rogressing cavity (PC) pumps are unique among positive displacement pumps for their ability to pump a wide range of fluids. Clean, water-like liquids; delicate products such as maraschino cherries and viscous, solids-laden fluids are all appropriate applications for PC pumps. While PC pumps perform well in applications where other pumps won’t, PC pumps should not be considered a pump of last resort. They can be valid choices even for less-demanding applications such as baby food, salad dressing or lubricating oils. By understanding the mechanical design of PC pumps, you will be able to match a pump to your pumpage.

PHOTO COURTESY OF ROBBINS & MYERS, INC.

P

DESIGN BASICS PC pumps are not new. Rene Moineau invented the progressing cavity pump in 1929. In 1934, Robbins & Myers licensed Moineau’s technology and introduced PC pumps to the United States. At that time, four companies licensed the technology; now more than a dozen companies manufacture PC pumps. Progressing cavity pumps are positive displacement pumps. But they are a unique type of postive displacement pump. In all positive displacement pumps displacement is proportional to the number of strokes, cycles or revolutions. The similarity of PC pumps to other positive displacement pumps ends there, however.

FIGURE 1 Dr

STATOR Ps Dr+4E

4E

ROTOR

Pr

Dr E

A ROTOR & STATOR

E

E

A

A-A

The positive displacement Moyno® 2000 progressing cavity pump is a model 21115-SSQ-AAA pumping whole grapes at a winery in California. The simplest PC pumps have a single threaded screw, called the rotor, turning inside a double threaded nut called the stator (Figure 1). The stator’s pitch length is twice that of the rotor, so this configuration is called a 2:1 profile (two leads on the stator, one on the rotor). Other profiles are used (e.g., 10:9) but the 2:1 is most common. Only the 2:1 profiles are discussed here. As the rotor rotates inside the stator, two cavities form at the suction end of the stator, with one cavity closing as the other opens. The cavities progress from one end of the stator to the other. The result is a flow with relatively little The Pump Handbook Series

pulsation. Since the longest path through the elements is a spiral path not far from a straight line, shear rates are low compared to those in other types of pumps. In most PC pumps, the stator is made with an elastomeric material that interference-fits on the rotor. This compressive fit between the rotor and stator creates seal lines where the two components contact. The seal lines keep the cavities separated as the cavities progress through the pump with each rotation of the rotor. Elastomeric materials make the pump suitable for abrasives and for large solid particles in suspension. The materials also enable the pump to be self-priming to 28 ft.

11

Q = 4ErDrPs

where Er is the eccentricity of the rotor, Dr is the minor diameter of the rotor and Ps is the pitch length of the stator. The displacement can be altered by changing any of these variables,but certain E r :D r and Dr:Ps ratios provide optimum flow. Most manufacturers offer rotor and stator designs optimized according to these ratios (and further constrained by manufacturing considerations and other requirements such as minimal torque and resistance to abrasion). Rotor and stator geometry also affect flow rate and differential pressure, two fundamental considerations

PHOTO COURTESY OF ROBBINS & MYERS, INC.

The manner in which the rotor rotates within the stator complicates the mechanical design of PC pumps (Figure 2). As the rotor turns in the stator, the centerline of the rotor orbits about the centerline of the stator. This eccentric motion means the pump must be fitted with universal joints to transmit power from the concentric rotation of the drive shaft to the eccentrically rotating rotor. These joints must transmit torsional and thrust loads. This novel requirement is responsible for today’s wide assortment of PC pumps. Designs range from simple ball-and-pin mechanisms to heavy-duty sealed gear couplings. The design determines whether the pump will be suitable for a given application. Although the mechanical design of PC pumps appears complex, the basic calculations of displacement, internal velocity and shear rates are simple. For example, displacement is defined by:

Viscous industrial hand cleaners containing heavily abrasive perlite are easily handled with this progressing cavity pump.

their pumps with elastomeric stators to 90 psi per stage, discounting this value for certain properties (e.g., abrasiveness) of the fluid to be pumped. Slippage is also proportional to the number of stages. For a given pressure, a three-stage pump will slip less than a single-stage pump (Figure 3). In fact, a three-stage pump will be at three times the differential pressure of the onestage before it slips as much as a one-stage pump.

for selecting any pump. Flow rate is determined by displacement per revolution multiplied by the number of revolutions. Although fluid properties do not directly affect pump displacement, they do affect the maximum speed at which the pump should be operated. Therefore they affect the maximum capacity of a specific rotor/ stator geometry. Pressure capability is proportional to the number of stages (within the mechanical limitations of the drive components). Each set of seal lines formed between the rotor and stator comprises a stage. A three-stage pump can develop three times the pressure of a one-stage pump. Manufacturers typically limit-

SHEAR-SENSITIVE FLUIDS Slippage is an important consideration for shear-sensitive fluids. PC pumps have long been favored for

FIGURE 3

FIGURE 2

3 Stage

Capacity - GPM

1 Stage

Theoretical Displacement 50 X GP

M 3 Stage

1 Stage 0

12

The Pump Handbook Series

X

100 200 300 Differential Press - PSI

GPM

VISCOUS FLUIDS PC pumps also are a good choice for pumping viscous fluids. In some fluids (i.e., non-Newtonian fluids), viscosity depends on agitation and shearing (Figure 4). Highly viscous fluids resist flow to such an extent that the liquid will not completely fill the cavities in a PC pump (especially with increased speed of the pump). Theoretically, the pump is cavitating

TABLE 1 Displacement gpm/100 rpm

Shear rate sec-1/100 rpm

0.26 0.86 2.02 5.20 12.00 22.00 36.00 65.00 115.00 175.00

93.9 93.1 93.3 78.3 64.2 64.2 57.1 57.8 64.6 63.8

Typical shear rates for progressing cavity pumps.

whenever its cavities are only partialTABLE 2 ly filled. Cavitation is usually to be avoided in most pumps, but PC Specifications: pumps can operate at 50 percent volCapacity ……………… 10 gpm umetric suction (occasionally even Pressure ……………… 80 psi lower) with no detrimental effects on Viscosity ………………100,000 cP the pump or process. Abrasion ……………… none Still, the maximum speed of the pump depends on viscosity, so users Performance: should know the viscosity of their Speed…………………… 10 rpm pumpage to select the proper pump. Slip ……………………… 0 gpm Data are available to predict volumetShear rate……………… 6.5 sec-1 ric efficiency according to fluid Power req’d ………… 4.9 hp viscosity and pump speed. To determine fluid viscosity, Moyno ® Pump Model we recommend the Brookfield 2K115G1-CDQ Viscometer. Having used that instrument to evaluate the rheologiculty is determining the abrasiveness cal properties of fluids since 1957, of a fluid. The abrasiveness of dry we have developed thousands of materials can be measured, but the plots of viscosity vs. shear rates. methods for doing so do not translate Most viscometers read at low to liquids and slurries. Users instead shear rates. Data from the viscosity must choose units based on the meter must be extrapolated to higher mechanics of wear in PC pumps. shear rates to determine the viscosity Several wear phenomena at various points in the system accubetween the rotor and stator account rately. We have been successful in for abrasive wear in PC pumps. The sizing pumps based on such extrapotwo principal effects are velocity of lated data. But specifying pumps rubbing between the rotor and stator according to “apparent” viscosity can and erosive wear associated with slip. lead to an oversized pump. Wear associated with centrifugal and For example, suppose viscosity compressive forces between stator data were obtained for shear rates and rotor is also involved, but the ranging from 0.1 to 1 inverse seceffect is small compared to the first onds. The viscosity for this fluid two and can be neglected. would be reported as 100,000 cP and Rubbing velocity is a function of the users would likely specify a large rotor and stator geometry: pump, operated slowly (e.g., Table 2). Such a pump would cost about (6.2834 (N/60) ((Dr/2) ± 2Er)) $27,000 (Table 2). The shear rate in this pump is where N is the rotational speed, Dr is 6.5 sec-1 and the corresponding visthe minor diameter of the rotor and cosity is 25,000 cP. An iterative Er is the eccentricity of the rotor. selection process would show that a Some manufacturers attempt to minimuch smaller pump (Table 3) is ademize rubbing velocity by altering the quate. This pump would operate at rotor and stator geometry. A far more 77 rpm with a shear rate of 49 sec–1. The corresponding visFIGURE 4 cosity would be 8,000 cP. The bottom line is that this pump would cost $7,000 instead of Dilatant or the $27,000 pump selected on Rheopectic the basis of the 10,000 cP Newtonian apparent viscosity.

ABRASIVE FLUIDS While users can select PC pumps for shear-sensitive and viscous fluids objectively, abrasive fluids call for a good deal of subjective judgment. One diffiThe Pump Handbook Series

Viscosity (cp)

shear-sensitive applications because of their low shear rate, especially at low speeds (Table 1). But a large pump run slowly can’t provide the pressure needed for some applications. That’s when the relationship between slip and shear becomes important. Minimal slip (i.e., more stages) minimizes shear rate. Slip rates of 10 to 15 percent of throughput yield excellent results for shearsensitive pumpage. Shearing can also be decreased by increasing the compression between the rotor and stator (which decreases slippage) or by using a smaller pump (which increases volumetric efficiency). Minimal slippage is also important in metering applications. With increased slip, flow becomes disproportionate to pump speed (Figure 3) and metering cannot be repeatable without a simple relationship between speed and flow. Variable speed drivers are usually recommended for metering applications. Thus, high volumetric efficiency and variable speeds provide optimum metering.

.01

Pseudoplastic or Thixotropic .1

1 10 Shear Rate - SEC.

100

13

significant reduction in wear can be achieved by reducing pump speed. If the viscous fluid described above were also abrasive, we would select a four-stage rather than the twostage pump (Table 4). The additional stages would reduce slip, and the reduced slip would allow reduced operating speed. The lower speed will result in reduced shear rate as well as increased fluid viscosity. Although proper selection of progressing cavity pumps can entail a lengthy process, following a few general guidelines can avoid improper equipment selection. 1. In shear-sensitive applications, minimize slip and maximize volumetric efficiency. 2. For viscous fluids obtain as much information as possible about the rheological characteristics of the fluid and select the pump and operating speed based on actual viscosity of the fluid as it passes through the pump.

14

TABLE 3

TABLE 4

Specifications: Capacity ……………… 10 gpm Pressure ……………… 80 psi Viscosity ……………… 8,000 cP Abrasion ……………… none

Specifications: Capacity ……………… 10 gpm Pressure ……………… 100 psi Viscosity ……………… 9,500 cP Abrasion ……………… heavy

Performance: Speed……………………77 rpm Slip ………………………5 gpm Shear rate……………… 49 sec-1 Power req’d ………… 3 hp

Performance: Speed…………………… 56 rpm Slip ……………………… 1.4 gpm Shear rate……………… 35 sec-1 Power req’d ………… 4 hp

Moyno ® Pump Model 2E022G1-CDQ

Moyno ® Pump Model 4F022G1-CDQ

3. For abrasive fluids, select the proper materials. Lower speed and slip for higher abrasive fluids.

Alan Wild is Manager, Fluid Systems for Robbins & Myers. He has been with the Fluids Handling Group of Robbins & Myers for 24 years and has been directly involved with the design, development, manufacture and application of Moyno Progressing Cavity Pumps and Downhole Motors.

More users are now selecting PC pumps for applications previously handled by other pumps.When selected and applied properly, PC pumps are versatile and reliable. ■

The Pump Handbook Series

POSITIVE DISPLACEMENT PUMPS HANDBOOK

A Review of Positive Displacement Pumps BY BRENT ROLAND AND GREGORY ZIMMERMAN entrifugal pumps are velocity machines. The impeller imparts kinetic energy to the pumpage. Containment in the pump casing converts the kinetic energy to pressure. If everything goes well, the result is a continuous flow of liquid from inlet to outlet. But sometimes everything doesn’t go well. Imagine trying to impart all that kinetic energy to an extremely viscous fluid. Or consider the effects that kinetic energy would have on a shear-sensitive fluid. Or imagine how much energy would have to be added to achieve the high pressures required in some applications. And sometimes continuous flow isn’t desired. In these and other cases, the alternative to the ubiquitous centrifugal pump is a positive displacement pump. Rather than jamming energy into the pumpage, these pumps move fluid directly by means of gears, lobes, vanes, flexible impellers, diaphragms, screws, rollers, pistons, or other mechanisms. Reflecting the diversity of mechanisms used, the Hydraulic Institute Standards for Centrifugal, Rotary and Reciprocating Pumps, 14th Edition lists literally scores of designs for positive displacement pumps. To simplify the array of choices, positive displacement pumps can be placed in two broad categories: rotary and reciprocating.

C

directly contacts the second gear. Thus the first gear can drive the second gear directly. To allow larger tolerances, other designs use timing gears to drive the gears. The timing gears (also called pilot gears) keep the two gears in mesh without requiring direct contact. In any case, the operating principle is the same: the meshing creates a vacuum at the inlet of the pump and atmospheric pressure pushes the liquid into areas between the gear teeth. The rotation of the gears traps the pumpage in the cavity between the gear teeth and moves the pumpage from the suction to the discharge side of the pump. When the liquid passes the junction of the casing and the discharge port, it is released from the cavity. Tight clearances between the gears and the casing prevent recirculation of the pumpage back to the suction side, so the fluid follows the path of least resistance, i.e., out of the dis-

charge port. The result: a continuous, repeatable flow — even when the discharge line is blocked. Without the possibility of recirculation, a block in the discharge line can cause pressure to rise until the pump or some other system component fails. The moral of the story: gear pumps require a pressure relief system. The relief valve may be integral to the pump or external. The basic components of a typical gear pump are: • the casing — similar to the casing of a centrifugal pump, but for external gear pumps it is typically oval instead of round, to accommodate the two round gears • the gears—which may have straight (i.e., spur), curved (i.e., helical), or herringbone teeth • the shaft — which rotates the gear(s) and to which the gears are attached by a key or other device

Gear pumps represent many of the principles and system requirements of rotary pumps. In an external gear pump, two gears, both independently supported by shafts, mesh with each other (Figure 1). The term “external” refers to the fact that the teeth are cut on the outside — they project out from the gear. Internal gear pumps use one external gear rotating within a gear with teeth projecting inward (Figure 2). In internal gear pumps and in one design of external gear pumps, one gear is driven by the motor and

PHOTO COURTESY OF MGI PUMPS INC.

A BASIC ROTARY PUMP

Rotary lobe pump in operation at a major paper maker in the midwest, feeding abrasive clay coating to a short dwell-on machine coater 24 hours a day, 7 days a week.

The Pump Handbook Series

15

• the bearings — which support and position the shafts, and can be internal (i.e., lubricated by the pumpage) or external, and of roller design or fluid film design (e.g., sleeve or journal) • wear plates — lubricating devices that fit between the gear ends and the housing and through which the shaft passes; designed to absorb any thrust movement of the gears and lubricate the region of contact between the gears and the casing • end covers — which close off each end of the case and support the bearings, shaft ends, wear plates (if applicable) and relief valve (if applicable) • stuffing box — part of one end cover through which the motor shaft enters the pump; also holds the seals or packing • seal — which prevents pumpage from leaking out of the pump where the motor shaft enters the pump through an end cover; as with centrifugal pumps, the seal may be a mechanical seal, packing and a gland follower, or the sealless design (which uses magnetic drive)

DO’S AND DON’TS FOR GEAR PUMPS The design of gear pumps makes them good candidates for pumping highly viscous fluids; but users must be sure to match the speed of the pump to the viscosity of the fluid being pumped. More viscous pumpage requires a slower pump speed. The reason is that NPSHR increases with viscosity of the

FIGURE 1

Borneman Series MPC multi-phase pump operating in the field for an international oil company. pumpage because viscous fluids experience more pressure drop getting into the pump than thin fluids do. To keep NPSHA greater than NPSHR (and thus prevent cavitation), the rate at which fluid enters the pump must be reduced. And because a gear pump displaces a set amount of fluid per revolution, the only way to reduce the flow rate is to reduce the pump speed. Users thus will typically have to specify a larger, slower pump for pumping viscous materials compared to pumping water. The picture gets complicated with fluids for which viscosity changes with shear rate. Thus a major “do” for gear pumps (and all pumps) is to check viscosity vs. speed tables before selecting the size of the pump.

TWO OTHER CRUCIAL DO’S: External gear pump

16

• filter out solids unless the pump is specifically designed to handle them The Pump Handbook Series

• run the pump only at a speed compatible with long pump life

SOME IMPORTANT DON’TS: • don’t throttle a gear pump • don’t starve the suction • don’t operate a gear pump for a long time with the internal relief valve in relief mode • don’t forget to compensate for all the characteristics of the fluid and all operational criteria when specifying and using a gear pump These guidelines apply to other types of rotary pumps as well.

OTHER ROTARY PUMPS Similar to a gear pump, a lobe pump (Figure 3) traps fluid between interlocking protrusions from the rotors, but a lobe pump uses only a few (one to five) large, smooth, rounded fingers (lobes) instead of the sharp teeth of a gear. Lobe pumps

FIGURE 2

Internal gear pump use external gears (i.e., no process contact) to drive the lobes. This design keeps the lobes close together but not contacting. Lobe pumps are well suited for highly shear-sensitive fluids. The smooth surfaces of the lobes also make lobe pumps a good choice for sanitary applications because the surfaces are easy to clean. Most sanitary lobe pumps are specially designed to facilitate disassembly and cleaning. Other sanitary pumps can be cleaned or steamed in place to meet increasingly stringent standards up to and including straight pipe cleanliness. For sanitary service, surfaces are finely polished to remove all possible cracks or crevices where pumpage could reside. Noncorrosive materials are also important for sanitary pumps. The 3A Food and Dairy Standards suggest design and define specifications for cleanability. A vane pump (Figure 4) also looks something like a gear pump, except there is only one rotor and it has long vanes instead of teeth. The

FIGURE 3

Lobe pump

vanes (usually five or nine) fit into slots cut lengthwise into the shaft. When the shaft is not turning, the vanes rest in the slots. The ends of the vanes may not touch the inside of the casing; but when the shaft turns at sufficient speeds (i.e., about 700 rpm), centrifugal force slides the vanes outward along the slots until the tips of the vanes touch the casing. With the bottom of the vane held by the slot and the top passing by the pump casing, a positive displacement effect is created with little or no slip of the liquid past the vanes. The vanes provide positive contact against the casing until they wear out. Most vane pumps have replaceable vanes and casing liners to let

FIGURE 4

Vane pump users restore “like-new” performance to their pumps. Some vane pumps use springs between the shaft and bottom of the vane to force the vane out from the shaft even at slow speeds. Gear, lobe, and vane pumps can be suitable for pumping slurries, especially when medium to high discharge pressures are required. Of course, internal clearances must be sufficient for passing the particulates in the slurry; but because larger internal clearances reduce the efficiency of the pump, users must carefully specify the clearances to balance the size and efficiency requirements. Another problem is abrasion due to hard particulates. If sufficiently hard materials can be used in the pump, the pump can be designed to crush the particles. Otherwise, materials may be chosen such that the particles imbed in the rotating elements. In either case, the pump should not be run so fast as to aggravate abrasive wear by the particles; but slurries also require a minimum speed to prevent settling of the particles. Users The Pump Handbook Series

must therefore know the characteristics of the slurry thoroughly to specify the correct speed of the pump.

UNIQUE ROTARY PUMPS A rotary pump that departs significantly from the gear-like design is the progressing cavity (PC) pump (Figure 5). PC pumps feature an axial flow driven by an externally threaded rotor turning within an internally threaded stator. As the rotor turns, cavities form between the rotor and stator. The cavities progress from the suction to the discharge side of the pump, moving the material along the length of the pump. Tight clearances between the rotor and stator form seal lines and prevent fluid from moving back toward the suction side. Rotors are typically metallic and stators are made of elastomers to prevent wear. Also, stators are generally replaceable. (For more information about PC pumps, see “Progressive Cavity Pumps,” Pumps and Systems, May 1993). Progressing cavity pumps offer good suction lift; smooth, pulseless flow; low speed operation; good slurry handling; the ability to handle high viscosities; low shear rates; and high efficiency. But the choice of construction materials is somewhat limited, and costs of repair in some cases is greater than other pump types. Another atypical rotary pump is the peristaltic or flexible tube pump (Figure 6). This design consists of an arc of flexible tubing set in a circular pump housing. Rollers, driven by an electric or an air-operated motor, compress the tubing very tightly against the casing, pushing the pumpage through the pump. The rollers flatten the tubing down to practically zero cross-sectional area, preventing recirculation. When the

FIGURE 5

Progressing cavity pump

17

FIGURE 6

commonly used in pharmaceutical and analytic chemistry “dosing” applications; but peristaltic pumps with flow rates greater than 100 gpm are also available.

RECIPROCATING PUMPS

Peristaltic pump tubing recovers from compression (i.e., regains its former shape), the resulting vacuum draws the fluid into the suction side of the pump. Usually, the same piece of tubing acts as suction, pump, and discharge tubing. Pumps designed for large flows sometimes use a rigid tube for suction and discharge and a largediameter flexible tube inside the pump. In either design, manufacturers use highly flexible tubing in the pump to achieve a long service life; but eventually the tubing loses its flexibility and must be replaced. Peristaltic pumps offer the advantage of sealless design and very precise flows of small quantities (e.g., fractions of a mL) of fluid. They are

FIGURE 7

Diaphragm pump 1. diaphragm, flat; 3. suction valve assembly; 4. discharge valve assembly; 5. liquid end head; 6. plunger; 7. hydraulic oil.

18

Reciprocating pumps are the other general classification of positive displacement pumps. In a piston pump, the motion of the piston away from the head creates a vacuum in the chamber, and the vacuum draws in pumpage. Motion of the piston toward the head compresses the pumpage and forces it out of the chamber. Check valves in the inlet and outlet prevent the fluid from being drawn from the discharge or pushed back into the inlet. Piston rings minimize passage of fluid past the piston. Diaphragm pumps (Figure 7) are a sealless design in which the plunger (piston) moves a diaphragm that in turn increases and decreases the volume of the pumping chamber and thus draws in and forces out the pumpage. As in piston pumps, check valves ensure one-way flow. As an alternative to direct mechanical drive, the plunger can also be driven by magnetic force. In magnetic designs, the plunger rod is surrounded by a coil. When the coil receives an external electrical impulse, it magnetically moves the plunger assembly a specific distance, which in turn moves the diaphragm. Typically, the electrical signal energizing the coil terminates just before the plunger is fully extended. The plunger comes to a stop and the pumpage is at its maximum compression. The plunger is spring-loaded and thus returns to its original position. The electronic actuation makes for easy control of pumping rate. Other drive mechanisms for diaphragm pumps are hydraulic fluid and compressed air. Reciprocating pumps create pulses in the flow of the pumpage: the fluid accelerates during the compression phase of the pump stroke and decelerates during the suction phase. Calculations done for specification of a pump must include these factors. The resulting pressure pulses cause inefficiencies and can lead to serious vibrations that can damage parts of the discharge system, so reciprocating pumps often require pulsation dampeners. These devices consist of a chamber into which the fluid flows during the pump’s compression stroke. A flexible diaphragm separates The Pump Handbook Series

the fluid chamber from an air chamber. The air chamber is pressurized with air or nitrogen to about 50% of the system’s discharge pressure. As the pump compresses the fluid into the discharge piping, some fluid enters the pulsation dampener and compresses the air in the dampener. When the pump goes into the suction phase of its stroke, pressure in the discharge piping falls below the pressure in the air chamber of the dampener and thus the compressed air forces fluid out of the dampener. This cycle allows flow to continue during the suction phase of the pump cycle and provides a smoother flow.

METERING APPLICATIONS The repeatable and controllable flow of positive displacement pumps makes them good choices for metering applications. The most common metering pumps are piston or diaphragm pumps with flow rates adjusted by mechanically or hydraulically altering stroke length and/or mechanically or electronically changing the stroking rate (for more information about metering applications, see “Fluid Metering System Options” and “Metering Pump Designs,” Pumps and Systems, April 1993). With the wide range of designs available for positive displacement pumps, users are likely to find a pump that can meet practically any requirement. High viscosity, shearsensitivity, high solids content, precision metering, low flows, high pressures, and many more applications are well suited for positive displacement pumps; but as with any pump, users must provide the pump vendor with adequate information about the pumping application in order for the vendor to suggest the right pump and the necessary system components. ■ Brent Roland is Director of Marketing for Gelber Industries, Elk Grove Village, IL. Gelber Industries provides fluid handling solutions using many designs of centrifugal and positive displacement pumps manufactured by a wide range of companies. Greg Zimmerman is associate editor for Pumps and Systems.

POSITIVE DISPLACEMENT PUMPS HANDBOOK

Handling Abrasives and Corrosives with Positive Displacement Pumps BY RODGER JACOBY AND JOHN PETERSON

A

number of pumping principles apply to pumping abrasives and corrosives with both centrifugal and positive displacement pumps. Each pump type has its own set of problems when handling these liquids, but this discussion will be limited to general areas concerning positive displacement (PD) pumps. An abrasive liquid is one that has particles in it. Some, like inks, have very fine particles, while others, like some paints, contain much larger particles. Handling abrasive liquids is a difficult application for any pump, because they promote pump wear. Likewise, corrosive liquids, by nature, attack the materials the pump is constructed of. The strength of a corrosive liquid depends on its concentration and temperature. The effects of moving both corrosives and abrasives are similar— pumps wear more quickly. Both corrosion and abrasion remove some of the material the pump is constructed of. Evidence of corrosion is different from indicators of abrasion. Corroded parts show even wear and possibly some pitting (Photo 1). Abrasion, however, causes uneven wear that follows the mechanics of the pump. On the outside diameter of a gear, for example, wear causes a scoring along the path of rotation (Photo 2). Typical corrosive applications can be found in almost every industry, but they are particularly common in the chemical and paper industries. Typical abrasive applications are found in the paint and coatings industry, the printing industry, magnetic oxide tape coating, and a variety of other processes. The first consideration when applying a PD pump to any application is to try and determine how abrasive and corrosive the product to be pumped is. For abrasives, finding the type of material, the size of the parti-

Another major consideration in cles, and how concentrated they are is successfully applying a pump to cora good starting point. Some printing rosive, abrasive liquids is to keep the inks have minimally abrasive characdifferential pressure as low as possiteristics, with pigments that tend to be ble. While this is primarily a system softer and finer, while some paints consideration, it will go a long way have harder, coarser pigments and toward extending pump life. Many extremely abrasive properties. manufacturers limit the differential A quick way to get an idea of the pressure for abrasive liquid pumps to abrasive nature of a product is to put about 60% of the pressure allowed a small amount of the liquid between for their standard pumps. For corrotwo glass slides and rub them togethsive liquids the lower differential er. Highly abrasive properties result pressure will reduce the amount of in a grinding, scratchy sound. slippage in the pump, and conseAdmittedly, this is a very subjective quently reduce the related liquid test, but with a little experience it can velocities that tend to increase the be related to the potential for pump aggressiveness of many corrosives. wear. A test for corrosives is someAs mentioned above, careful what more straightforward. Wafers material selection plays an important of materials under consideration for role in moving difficult liquids. pump construction can be immersed Materials come in various hardnesses in a sample of the liquid to be and have different levels of corrosion pumped, and weight loss recorded resistance. Each pump component over time. should be matched to the nature of As mentioned previously, abrathe liquid being pumped. sive liquids wear pumps by nature. The materials of construction of This wear can be slowed dramatically bushings exposed to the pumpage is by slowing the pump down. It is not one area of concern. One common unusual for pump manufacturers to option is carbon graphite, which is recommend speeds from one third to modestly priced and has excellent one half of rated speed to retard wear. This depends on how abrasive the product is and the economics of using a larger pump and slowing it down, but it sometimes costs less to use a larger, slower pump that lasts longer, rather than replacing a smaller, faster pump (Figure 1). When pumping corrosives, operating speed is less important than the selection of the right materials. When considering materials, pay particular attention to the temperature of the liquid. Most corrosive materials become more aggressive at higher temperatures, so a lower temperature will help Photo 1. An idler with corrosive wear. extend the life of the pump. The Pump Handbook Series

19

corrosion resistant properties, but its softness does not work well with abrasives. Bronze is harder and less expensive, but it has limited corrosion resistance and needs a lubricating liquid to prevent wear, again a drawback for use with abrasives. Cast iron generally has a modest cost and can be easily handled for field replacement. For mild abrasives, some users have found that a cast iron bushing works well; but with corrosives it has very limited value. All of these bushing materials are common to many manufacturers and, at best, offer only a minimal resistance to abrasives; and, other than carbon graphite, they are not normally considered for use with corrosives. There are alternative hard bushings available that can serve to improve the life of a pump in abrasive service. The least costly option is to use hardened cast iron. This material offers a significant improvement in life span over normal cast iron in abrasive service; however, it has no advantage in corrosive applications. One note of caution: make sure that all cast iron bushings have an initial start-up lubricant in the pump. These bushings are subject to rapid initial wear without lubrication, and some manufacturers pre-lube these bushings to assist in start-up. Be sure to check your pump’s start-up requirements. After the initial run-in,

cast iron bushings give a very long and reliable life. The next option in cost and life expectancy are coated bushings. There are many of these available, one example being Colomony-coated bushings. Colomony is a hard coating that resists abrasive wear and has excellent corrosion resistance, provided that the material the coating is applied to can resist the chemical attack. Unlike the materials mentioned above, a Colomony-coated bushing cannot be used with a non-hardened shaft. The bushing would rapidly wear into a standard unhardened shaft, causing an immediate reduction in service life of the pump. The coated bushing can be run with a coated shaft, giving an excellent life span and a high degree of abrasion resistance. Again, the coating material is only as good as the base material when it comes to resisting corrosion, so coated parts are not commonly used in corrosive applications, except where a harder part is needed for wear resistance. Colomony coatings also need surface lubrication supplied by the liquid being pumped, so they are seldom used for low-viscosity applications. Another option is the ceramic bushing. This material, while able to resist abrasion and corrosion, is a good heat sink, and care must be taken to ensure that the bushing is properly cooled. As with coated bushings, a hard shaft must be used to prevent premature shaft wear. Ceramic bushings also tend to be a poor choice for thin liquids, which limits their use with corrosives. Another consideration with this type of bushing is that its coefficient of thermal expansion is quite low, so it often requires a heatshrink fit when temperatures are elevated. One of the most superior abrasion-resistant materials is tungsten carbide, but its properties come at a higher cost. Again, as with coatPhoto 2. An idler with wear from a large ed bushings, tungsten abrasive particulate. carbide bushings must be

20

The Pump Handbook Series

used with a hard shaft material. A tungsten carbide-coated shaft is one option. Tungsten carbide running against tungsten carbide has been used successfully with thin liquids, and would certainly be a prime choice for thin, abrasive liquid applications, especially if more than modest differential pressure was required. Tungsten carbide does well in liquids with a pH higher than 4, while liquids below that attack the binders and cause the tungsten to “come apart.” A major wearing point in any pump is the area of shaft sealing. This is even more critical in abrasive or corrosive liquid applications. Standard shaft packing needs a liquid film to lubricate the shaft, and if the pumpage contains abrasives this film will cause additional wear, not lessen it. Corrosive materials, while able to provide some lubrication for the packing, are not the sort of material that should be allowed to exit the packing gland while providing that lubrication. As a result, packed pumps are usually shunned for corrosive applications. For packed pumps in abrasive service, special packing and hardened shafts are often used. The other option, that of a mechanical seal, is frequently used for abrasives and almost always for corrosives. The material of the face seal is an important consideration, and selection follows the guidelines for other pump materials discussed earlier. The harder the face, the more it will resist damage by abrasives. An area of the seal that deserves special attention related to corrosive applications is the type of elastomer used. Viton is good for many applications and is relatively inexpensive. Teflon is better in many cases and somewhat more costly, while Kalrez provides excellent corrosion resistance at a high price. In addition to the bushings, shaft material, and seals, consideration must also be given to materials used in the rotating elements and housing. Cast iron is the least expensive option. As mentioned, it has a degree of abrasion resistance, but little corrosion resistance. It has the added benefit of being a low-cost replacement part if the application is temporary. The next step up in abrasion resis-

overlook the composite materials, which generally cost less than even 316SS. Most of these materials have corrosion resistance that is best restricted to particular types of pumpage. Talking to pump manufacturers and experts on the material being pumped is a good way to find a material that will stand up to a specific service. ■

FIGURE 1

John E. Petersen is the Vice President of Engineering for Viking Pump, Inc. He received his P.E. in 1975 and has worked for Viking for 24 years. Mr. Petersen started as a Project Engineer and was promoted to Chief Design Engineer before receiving his present assignment.

A pump fitted for abrasive liquids. tance and cost is hardened cast iron or steel. While they have little value for use with corrosives, they can better resist abrasives and can be selectively applied to various pump components. As these harder components are incorporated into the pump, the cost goes up dramatically. The harder the parts become, the more resistant they are to abrasives; but at the same time, they also become more difficult for manufacturers to machine. This trade-off has given rise to coating materials for pump gears and casings. Work has been done on adding an even layer of tungsten carbide to the surfaces of these components, with good success. There are also other coatings that have proven their worth in resisting abrasion, and new coatings are being developed and tested. Corrosives, on the other hand, call for different materials, and hardness is not normally a factor. 316 stainless steel is the most universally selected material for use with corrosives. It has a wide resistance to corrosion, and because many manufacturers have made this a standard material, it is available at a reasonable cost (Figure 2). Beyond stainless steel for corrosion resistance, things get a little more difficult. The most serious con-

cern is that cost goes up dramatically. PD pumps constructed of Alloy 20, titanium, and Hastelloy are available. Hastelloy is the most universally corrosion resistant material, but the cost is considerable. Also, do not

Rodger A. Jacoby is the Marketing Manager for Viking Pump, Inc. and has worked there for 20 years. He started in inside sales in at the main plant, then was promoted to District Manager before receiving his present assignment.

FIGURE 2

A corrosion-resistant 316 SS pump.

The Pump Handbook Series

21

POSITIVE DISPLACEMENT PUMPS HANDBOOK

Fluid Metering System Options BY FRIEDRICH MÜLLER AND WILLIAM E. NEIS etering systems can improve quality and decrease costs of chemical process by precisely controlling the rate at which feed stock enters the process. But all parts of the system must be selected properly or quality and cost of the product will suffer as will the reliability (e.g., safety and environmental performance) of the process. Proper selection involves considering the characteristics of the fluid, the type of processing required, the skill of the operating personnel, and safety, environmental and quality assurance requirements. Designers and users of fluid metering systems must understand these relationships. Admittedly, all the required data are often not available so assump-

FLUID PROPERTIES Many metering system design features are dictated by the nature of the pumpage. One important characteristic is corrosiveness of the fluid. Obviously, the wetted parts of the system must be able to stand up to corrosion in the pumpage. Corrosivity inturn depends on the operating temperature and other conditions. A compound that’s considered only slightly cor-

rosive at “standard” conditions can be very corrosive when hot. Or an elastomer rated as compatible with a process fluid can become brittle or swell when placed under mechanical load. Unfortunately, compatibilities shown on materials selection charts assume “standard” conditions. The charts also assume that the description of the fluid is accurate. Too often, what a user thinks is a pure liquid isn’t pure at all but contains impurities that change the chemical aggressiveness of the fluid. Vapor pressure of the liquid is another characteristic that must be considered. Vapor pressure determines how much pressure must be available at the suction side of the

Photo courtesy of America Lewa, Inc.

M

tions must be made; sometimes the data needs to be padded and safety factors must be inserted into designs. If the process is adequately understood, though, padding and safety margins can be kept to a minimum and correctly sized equipment can still be selected.

Fully automated continuous system blending up to 16 components in various recipes to produce liquid detergent products at 20 tons/hour. Utilizes a 17 head metering pump driven by one motor.

22

The Pump Handbook Series

find it difficult to get reliable data extreme measures to overcome about high viscosity and high prespressure loss, you should first sure conditions. Performance data determine that all fittings, gauges for non-newtonian fluids and and other equipment causing presthixotropic liquids are even harder sure losses on the suction line are to come by. Thixotropic liquids actually necessary. You may find, decrease in viscosity with rising for example, that simply moving a shear rate because the internal flow meter from the suction side to structure of the fluid is partially the discharge side will reduce presdestroyed. At rest, the structure sure loss sufficiently. slowly rebuilds and viscosity Viscosity of the pumped fluid is increases. Viscosity might become another basic parameter for selecting too high if the system is shut down components in a metering system. for an extended period. In that Flow of low viscosity fluids (e.g., case, an automated control may be water) is generally turbulent and necessary to start the pump after a thus viscosity contributes little to given time and keep the fluid in pipeline pressure loss. Without vismotion, thus keeping its viscosity cosity-related losses, metering syslow. tems can be selected from virtually Another important fluid characthe entire range of designs. teristic is elasticity, the reduction Systems for high viscosity fluids in volume due to increased must be designed more carefully. Laminar flow is the rule in high viscosity liquids and pressure loss increases linearly with viscosity. With the possibility of substantial pressure loss, particular kinds of equipment may not be usable. For example, you may not be able to find a suitable place to install a flow meter: it may cause too much pressure loss if installed upstream and it may not be able to withstand the pressures of the discharge flow. Like other fluid characteristics, viscosity depends on temperature so be sure to design for the temperature at which viscosity is maximum. Viscosity also increases with pressure, which makes highpressure metering diffiSemi-automatic batch system to produce car cult to achieve. Because perfor- care products using a mix of metering pumps mance data are gener- and flowmeters. Batch makedown is to the ally based on water top tank. When complete, batch is delivered pumping, you may to the bottom tank to feed the packaging line. The Pump Handbook Series

23

Photo courtesy of American Lewa, Inc.

system. If pressure at the suction side falls below the vapor pressure of the liquid, cavitation can occur and accurate, reliable metering will be impossible. Thus, static head must be sufficient to overcome pressure loss on the suction side of the pump due to friction and inertia. Pressure losses in the suction pipeline also must be covered. When considering the vapor pressure, don’t forget effects of temperature on vapor pressure. The system must be designed to handle any temperature extremes the process will be subjected to. This point is important to remember for outdoor applications. To maintain adequate pressure on the suction side, the suction vessel should have sufficient geodetic height. If the pump cannot be lowered or the vessel raised, the pressure losses themselves must be reduced. For example, friction can be reduced by lowering the velocity of the liquid, by increasing the nominal width of the pipeline and fittings or, for reciprocating pumps, by installing a pulsation damper. If none of these alternatives is possible, another approach might be to pad the pressure in the suction vessel with an inert gas. But be careful. This approach will not work with liquids that absorb substantial quantities of the inert pad. Subsequent pressure losses in the suction pipeline will cause the absorbed gas to reappear in the liquid and lead to cavitation-like problems. Liquified gasses are prone to such problems. It may be difficult to predict the amount of gas absorbed by the liquid because the equilibrium solution depends on the pressure level in the suction vessel. Another alternative is to take advantage of the relationship between vapor pressure and temperature — that is, heat the suction vessel or cool the suction line at the pump inlet. A 10° to 20° F gradient is usually sufficient. If these procedures don’t work, the only way to increase suction side pressure is with a booster pump functioning on the NPSHA. Of course, before taking any

pressure. Every liquid is at least somewhat elastic. Therefore, in reciprocating pumps a portion of the stroke does nothing but raise the pressure in the pump head from suction pressure to discharge pressure. The liquid has not moved, but the internal pump components have already accelerated to a particular velocity. When the discharge valve opens, the liquid is abruptly coupled to the movement of the pump. Theoretically, the liquid must accelerate instantaneously from standstill to the speed of the pump parts. The acceleration

FIGURE 1

Continuous blending with individual metering systems results in a pressure shock that is transmitted to the piping. If the internal diameter of the pipe is smaller than the diameter of the plunger, this problem is accentuated.

SPECIAL CASE FLUIDS Other fluid characteristics that can cause trouble in metering systems include slurries, foams, supercritical fluids and melts. An important consideration for slurries is maintenance of fluid velocity. If the fluid slows down, sedimentation can occur and accurate metering will be difficult. Experience with the slurry will be your best guide to decide what flow velocity is necessary to keep particles in suspension. Fluid velocity can be maintained with proper diameter of piping and the pump. If you’re planning to use a reciprocating pump, be sure to account for the pulsation velocity. Slurries require careful design of piping and other components. Abrupt changes in direction and in

24

Liquids close to their melting points require special design criteria. Some liquids are sensitive to overheating, which can occur with strip heaters wrapped around the pipes. Such techniques might be suitable for a lab test or pilot project, but are unsuitable for a production system. (This underscores the fact that plant engineers and Continuous blending with multihead chemists must understand pump the problems involved in scaling up a process from the cross sectional area of the pipe the lab bench to a plant!) at joints must be avoided because CONTINUOUS VS. BATCH PRODUCTION turbulence may cause sedimentation or clogging by forming bridges. The design of a metering system The suction pipe must be mounted also depends on whether producso that the slurry is drawn from the tion will be continuous or batch. agitated area of the suction vessel; With continuous processing, comvalves must be installed so that ponents are metered simultaneoussedimentation does not occur in ly into the process stream where front of the valve seat; pulsation they are mixed and treated continudampeners must be designed to ously. prevent settling of solids in the Continuous processing is advandampener. tageous for large production quantiFoams (i.e., fluids with 50 percent ties and for processes in which the or more air) are highly compressible recipe and other production and therefore difficult to meter voluprocesses don’t often change. metrically. Metering pumps can be Automated systems can assure conmore accurate than a flow-meter-andstant product quality in continuous control-valve system but the pressure processing by keeping the process level must be raised on both sides of parameters and characteristics the pump to keep pressure within narrow limits. fluctuations low. (Relative Continuous processing can pressure fluctuations cause for solids, too, as Start-up and work variations in mass flow.) long as the solid can be disSupercritical liquids shut-down solved or converted into a (i.e., liquids above critical slurry before metering. temperatures and presOne method of continuare sures — for example, CO2 ous processing is to use difespecially ferent metering systems for at temperatures above 87o F and pressures greater component and then challenging each than 1015 psig) can also combine components with a be highly compressible. control system (Figure 1). Or in They cannot be treated a multihead pump can be like regular fluids. continuous used, with one pump head Pressure and temperature component and each processing. per conditions must be kept pump drive coupled by a constant because fluctuamechanical connection tions in pressure and tem(Figure 2). Multihead pumps perature affect pump efficiency offer the advantage of a steady ratio and accuracy. For example, comof recipe components regardless of pression elevates fluid temperature. the speed of the pump. Thus, the pump head must be cooled Whether components arrive in to a constant, lower temperature to the process stream from multiple keep the system trouble-free. metering systems or a multihead

FIGURE 2

The Pump Handbook Series

viscosity η(mPas)

pump, the components may is impossible for simultaneous FIGURE 3 need to be mixed. For metering. The optimal method for example, continuous prosome processes is a combination cessing may not offer suffiof simultaneous and sequential cient reaction time for some metering. chemical processes. Inline The flexibility of batch processmixers induce high shear ing is also illustrated by doublerates which render quick vessel processing. This method material exchanges and uses a tank with upper and lower thus cut reaction times. compartments. A recipe can be Multiple mixers can be placed in the upper compartment placed in the system, for of the tank. When the recipe is example to premix initial complete, a valve is opened and components of the recipe the fluid runs into the lower combefore secondary ingredi- Typical use of mixers in a continuous partment. When the valve is ents are added and mixed process closed, the upper compartment is by another mixer (Figure 3). then available for use for another Several types of mixers are batch or even another recipe. The available. As with all system comprocessing is also not advantageous previous process may still be runponents, the best choice depends for processing that requires long reacning in the lower compartment. on the characteristics of the fluid tion times. This multiple batch idea can be and the process. Also, the choice of Batch processing — the placement taken further by parallel producmixer affects selection of other of metered chemicals into a reaction tion in which the components are components. (See “A Closer Look at container — is the preferred alternametered into a funnel instead of Mixers,” page 25.) tive for processes that require long directly into a vessel. The funnel Start-up and shut-down are espereaction times. It also offers high can be alternatively positioned over cially challenging in continuous accuracy. Individual metering sysany of a number of vessels and processing. On start-up, several tems can be laid out optimally, even thus the metering system can supminutes may be required before if some components require a wide the process reaches a steady state range of adjustment. FIGURE 4 (e.g., the recipe may need to be In addition, production may be adjusted or the reaction may 10000 more flexible with batch than with require some amount of time to continuous processing. New 5°C begin). During this start-up time, a recipes can be run without requirt=2 1000 considerable amount of out-of-specifiing changes to the system. Batch 0°C t=4 cation product may be produced. processing may also simplify the °C t=60 Shut-down and cleaning procedures sequencing of components into a 100 °C t=80 produce similar out-of-spec product. reaction. And lab tests can be run Because of cost of materials or to test the quality of the final prod10 environmental and safety regulauct while the product is held. If tions, this product cannot simply the quality is unacceptable, correcbe thrown away. Instead, it is coltions can be made to the batch. 1 1 1000 2000 lected in a separate vessel and later Batch processing is often the prepressure p[bar] reinjected a little bit at a time into ferred method for pharmaceuticals the process stream. This “rework” and other products that require A closed parallel production system method necessitates an additional documented lab verifications. pump. It also requires that the Components can be introprocess run long enough so that all duced into a batch process simultaply several downstream production of the out-of-spec product can be neously or sequentially. lines running different recipes. A reinjected at its low input rate. Simultaneous metering requires parallel production system can be When this method cannot be used, each component to have its own closed by connecting the vessels the collected waste must be stored metering system, but minimizes the with solenoid valves operated by a for later rework or batch treattime required to get all the compocontrol system (Figure 4). ment. nents into the reaction. Sequential Metering systems are complex Start-up and shut-down difficulties metering requires more time, but and their design must incorporate can make continuous processing dissaves the expense of separate many factors. To ensure that all advantageous for plants that change metering systems for each compofactors are considered adequately, recipes too often for new conditions nent (assuming the components are all disciplines in the plant should to be accommodated by the range of compatible). Sequential metering participate in the design. Project adjustment of the pumps. Continuous can also be based on weight, which and process engineering, operations The Pump Handbook Series

25

and maintenance, quality, finance, management, even suppliers, vendors and outside specialists should be involved. Other requirements for successful design include: • clear data •

a mutually accepted scope of supply

•

realistic design goals based on sound theory of operation

•

well defined responsibilities

•

open and ongoing communications.

Communications are especially important given the inevitability of changes in the process and requirements. These five steps can help ensure a smooth flow of design: 1. Prepare a clear design goal and scope of study. Include data on any current system and any

required improvements, such as level of automation. 2. Define accurately all fluids to be handled. 3. Select the metering concepts to be used and outline the desired theory of operation. 4. Understand the chemical reaction and mixing requirements. 5. Optimize for the most costeffective approach. These steps provide a solid foundation for developing a metering system that will meet your needs. The clear specifications that come out of such a process will also foster the creation of a partnership with your systems supplier. There are no shortcuts to this process. It will require an investment of time and effort. But the result will be an accurate metering system — one that provides consistent and reproducible quality at an optimal price. ■

Friedrich Müller heads the Applications Techniques Department at LEWA Herbert Ott GmbH, Leonberg, Germany. He holds a Dipl. Ing. degree in chemical engineering from Technische Univeritat Stuttgart in Germany. In his current position, he oversees the supply of metering systems and packages with various levels of automation for diverse industrial process applications. He was formerly head of LEWA’s Application Lab, where he developed expertise in liquid metering under a wide range of fluid characteristics and operating conditions. William E. Neis, P.E., is a technical marketing specialist for American LEWA, Holliston, Massachusetts. He holds a BSChE from Villanova University, has conducted graduate work in mechanical engineering at Pennsylvania State University and has fifteen years experience with several metering pump manufacturers.

A Closer Look at Mixers Several types of inline mixers are available for use in continuous processing. Motor driven dynamic mixers build up high shear rates either by pumping the fluid through annular apertures at high pressure or by high speed rotor/stator systems. Some motor driven mixers can increase the pressure in the system by an effect similar to a centrifugal pump. Static mixers have no moving parts. They are simple devices, inexpensive and easy to install. Some are just a piece of pipe. Static mixers are effective for high viscosity fluids, but when designing your system, remember that the energy for static mixers comes from the pressure generated by the metering system. The pressure requirements may be substantial, especially for high viscosity fluids. A mixing valve is a normal pressure release valve used to mix highly miscible fluids. Inline mixers hold only a small volume of liquid. Thus, little longitudinal mixing (mixing in the direction of flow) occurs. The mixing is across the direction of flow. Cross-mixing will be adequate if recipe components are metered into the mixer at a constant rate (all components are present in a cross-section of pipe). A multihead metering pump phased for simultaneous discharge will also bring all components into a cross-section of pipe. But in this case, a short static mixer may not provide adequate mixing since the mixing action depends on velocity and velocity fluctuates during the stroke cycle. A long static mixer may be sufficient. If the individual heads of a multihead metering pump do not discharge simultaneously or if they operate at different speeds, static mixers may not be adequate. Mixing problems may also arise with solenoid pumps used for proportional metering. If pulse rate cannot be directly controlled by varying stroke length, the controller varies the stroke rate and the changing interval between strokes interferes with mixing. Mixers cannot simply be added to a metering system without regard to their effects on the total system. Mixers cause backpressure. If the metering pump causes pulsating flow conditions,the backpressure can become severe and dampers may be necessary to take up that pressure.

26

The Pump Handbook Series

POSITIVE DISPLACEMENT PUMPS HANDBOOK

Pulsation and Surge Control BY GARY CORNELL ulsation and surge create vibration—potentially the greatest hazard to a plant’s liquid transfer system. What are pulsation and surge? Why do they occur? And what can be done to control their damaging effects? These are important questions for anyone using or involved with pumps. Pulsation is the acceleration and deceleration of uncontrolled energy units. For this discussion, these units of energy are transmitted through liquid flowing in a system of pipes. Pulsation in these systems is most often created by the start/stop or reversing action of a reciprocating type pump. Observed as vibration, pulsation can be measured by using a simple pressure gauge or using a transducer and oscilloscope for greater accuracy. Pulsation is generally expressed as a sinusoidal curve or wave of peaks and valleys of pressure spikes. A related phenomenon, surge (water hammer) is created by rapid changes in a liquid’s velocity. Velocity changes can be caused by a rapid or abrupt change in direction, valves or ells in pipe systems, a change in pipe diameter, or differences in elevation. Start-up and shutdown of the primary pump itself creates an instantaneous change in liquid velocity. Doubtless, the single biggest contributor to surging is the quick-closing valve. When a valve closes rapidly, fluid flow comes to an abrupt stop and reverses direction. The pressure spike created can be as great as five times the working pressure in the system, traveling at up to 4,000 feet per second. This rapid change in velocity creates a compression wave that travels the length of the pipe run and continues to oscillate until friction dissipates—or worse—a system component fails.

P

WHY DOES THIS OCCUR? For all practical purposes, liquids are not compressible. Liquid neither

Pulsation and surge become efficiency’s enemies when they cause system failures that mean downtime and loss of production. Pulsation and surge are a major cause of these failures in liquid transfer processes. Pipes, joints, and other connections can leak and fail from the vibrations caused by pulsation and surge. In many situations where economics or the specifics of a given application would indicate that plastic materials are appropriate, pulsation in the system dictates use of a stronger and more costly material. When liquid is metered, pulsation will cause pressure fluctuations well beyond the acceptable range of 10 to 15 percent of mean operating pressure. Excessive pulsation prevents filters, which require a steady, non-pulsing flow, from operating efficiently. It causes excessive wear on pump components such as seals, valves, fasteners and pistons, leading to premature failure. Many applications such as filling, coating and spraying simply cannot tolerate a changing flow. Manufacturers have always been concerned about protecting the environment by minimizing spills and product leakage. However, new EPA guidelines and reporting requirements include a long list of hazardous and toxic materials, making it more vital than ever to ensure that piping systems and components do not fail.

were thin, a centrifugal type pump would do the trick. Even if the liquid were thick but free of entrapped solids, a rotating positive displacement pump, such as a gear or progressive cavity pump, would do the job. The fact is, however, that a great majority of liquids don’t match these ideals. Liquids can be thick, abrasive, and shear sensitive. They may have high solids content. You may need to measure them. Pumping liquids like this requires a positive displacement pump. Unfortunately, these pumps inherently have reciprocating components that generate a pulsating, nonsteady state flow, and we are back to where we started from. Virtually every commonly used pump does create pulsation. However, centrifugal, gear and similar designs produce a very high frequency (rapid) but low amplitude (narrow pressure band) pulse. The flow of fluid is a near steady state with minimal peaks and valleys of pressure spikes. These types of pumps, especially centrifugals, generally only create surge (water hammer) when starting up or shutting down. Reciprocating positive displacement pumps move liquid by trapping and then expelling discrete units, or slugs, of liquid. As we’ve seen, these slugs of uncontrolled energy generate the pulsation. In reciprocating pumps, the pumped fluid actually comes to a complete stop every time the pump shifts, which can occur up to several hundred times per minute. This start/stop action creates harmful pressure spikes in the form of high amplitude, low frequency pulses. The worst offenders are: • Piston pumps—generally used for high-pressure, clean but higher viscosity liquids. A single pump has one or more pistons.

PULSATION STARTS WITH THE PUMP

•

absorbs nor cushions the pulse or surge but rather transmits the energy. According to Newton, energy is neither created nor destroyed, but only transferred from one form to another. Thus, we must realize that pulsation and/or surge will occur when liquid is pulsated, stopped, or turned.

CONSEQUENCES OF UNCONTROLLED PULSATION

Why not use a non pulse-generating pump to avoid these problems? Unfortunately, solutions are not always that simple. If every liquid The Pump Handbook Series

Air operated diaphragm pumps— very versatile, used for thin, thick, clear or solid-laden liquid. These usually have two diaphragm pumping chambers.

27

•

•

Peristaltic pumps—larger models are used for thick and abrasive liquids while smaller models are used for low viscosity, clear liquids. Basically, these contain a hose within a chamber, and two or three lobes progressively squeeze and release as they rotate.

large extent, surge does not depend on pump type. Surge is most damaging at higher flow rates with large pipe diameters or long pipe runs. The larger the volume of liquid involved, the greater the force exerted on the system. Uncontrolled energy mass is the culprit here.

Metering pumps—used to transfer many types of liquid in a precisely measured flow. Normally, these employ piston or mechanically driven diaphragm designs.

SURGE DEVELOPS IN THE SYSTEM Unfortunately, eliminating pumpinduced pulsation does not necessarily remove the potential for piping system damage, for a change in velocity can also create destructive pressure spikes. Quick-closing valves are certainly a major culprit, but less obvious situations such as direction changes in the system, the ells involved, changes in pipe diameter and lengthy vertical rises can also contribute to surge. To a

•

Surge tanks—must be oversized to be effective. Eventually the air cushion will be absorbed into the liquid, and the tank will become water- logged.

•

Alternate pump types—can be a solution, but there is not always an alternative for certain types of fluids.

Several methods are available to control pulsation and surge. However, most of them simply address the effect, not the cause: • Rigid brackets—can be used to brace against vibrations. However, they are expensive, restrict pipe line location and eventually wear out and/or break.

•

Timer controlled valves—will eliminate surge but cannot always be substituted for quickclosing valves.

•

Flexible hose—may work. However, the hose can rupture, and varying pressures cannot be controlled.

•

PULSATION DAMPENERS

METHODS OF CONTROL

•

Flexible couplings—generally address mechanically-induced vibration and misalignment rather than fluid-induced vibration. Regulating valves and restricting devices—can be effective but generally increase power requirements by increasing flow resistance.

Pulsation dampeners use some form of potential energy to absorb pressure spikes. They act directly to equalize or balance the pulses generated by the pump, and in general provide the most economical, efficient and safe method for the control of pulsation and surge.

FIGURE 1

A

Air Supply

B

Air Supply

GAS

GAS

From Pump

To Discharge

Compressed gas is used to charge the top of the dampener at a predetermined pressure (usually 80% to 90% of liquid pressure) as pumped fluid flows, it takes the line of least resistance and enters the dampener.

28

C

From Pump

Air Supply

GAS

To Discharge

As fluid fills the dampener, the bladder compresses the gas to equal maximum fluid pressure. At the point of pump shift, fluid pressure drops below compressed gas pressure.

The Pump Handbook Series

From Pump

To Discharge

With fluid pressure now less than gas pressure the bladder is forced down discharging accumulated fluid back into the pipeline filling the void created during pump shift. The result is a continuous, vibration free flow of fluid.

Industry has had pulsation dampeners in some form practically since there have been pumps. At first they were merely gravity types, where a piston was placed inside a tube attached to the liquid line and weight was added to the top of the piston to minimize pulsing. Over the years designs have changed and improved. Adjustable springs and air-tight pressure vessels shrunk dampeners and made them less cumbersome. Around the turn of this century, the first hydro-pneumatic-type dampener using a rubber separating membrane was patented. Since gas is compressible and liquid is not, Boyles law of P1V1 = P2V2 is the principle involved. That is, volume is inversely proportional to pressure. Air or gas under pressure is introduced into the top section of a sealed pressure vessel. Liquid is allowed to enter into the lower half, but some form of separation membrane (usually a bladder, bellows or diaphragm) separates the liquid from the gas. The pressure of the gas is set at a predetermined level based upon calculated or observed fluid pressure in the pipeline. When pulsation creates pressure spikes, the trapped gas in the dampener absorbs the spikes as an automobile’s shock absorber handle potholes in a road (Figure 1). To achieve desired results, you must properly size the dampener. Pulsation dampeners are available in capacities from less than a cubic inch to well over 100 gallons and pressure more than 30,000 psi. Often you can reduce pulsation up to 98 percent, depending upon the characteristics of your particular application. The following information is generally needed to size a dampener properly: •

volume per stroke of pump (usually in cubic inches)

•

type of pump (i.e. number of pumping chambers, and whether it is single-or double-acting)

•

minimum and maximum operating pressure

•

type of fluid being pumped, including viscosity, specific gravity, corrosiveness, and temperature.

While sizing should be left to the experts, Figure 2 shows a basic sizing formula so you can estimate.

FIGURE 2

Information Necessary 1. Volume per stroke of pump (in cubic inches) One gallon equals 231 cubic inches Volume is determined by the following: .7854 x (bore diameter)2x (stroke length) 2. Type of pump, i.e., simplex (single piston), duplex, triplex 3. Maximum operating pressure 4. Minimum operating pressure 5. Average operating pressure (minimum and maximum divided by 2) Capacity (c) equals: (V) x (K) x (Pav/Pmin) Y 1-(Pav/Pmax) Y Where: V = volume per stroke in cubic Pmax = maximum discharge pressure Pmin = minimum discharge pressure (usually 50% of the maximum pressure in a double diaphragm pump) Pav = maximum pressure added to minimum pressure divided by 2 K = number from pump type table Y = 1 if charge in suppressor is compressed air = .714 if gas charge is nitrogen

{

K Factor Table: Pump Type K Factor Simplex single acting .60 Simplex double acting .25 Duplex single acting* .25 Duplex double acting .15 Triplex single acting .13 Triplex double acting .06 * Air operated double diaphragm pumps are this type. Example: V K Pmax Pmin Pav Y C

= = = = = = =

.33 gallons or .33 x 231 = 76.23 cubic inches .25 from K Factor Table (typical Air diaphragm pump) 60 psi 30 psi 60 + 30 divided by 2 = 45 1 for compressed air charge 76.23 x.25 x (45/30)1 1 - (45/60)1 = 19.06 x (45/30)1 1 - (15/60)1 = 19.06 x 1.5 1 - .75 = 28.59 .25 114.36 cubic inches suppressor capacity

Basic dampener sizing formula for pulsation control The Pump Handbook Series

29

FIGURE 3

Cushions the shock of high inlet pressure (over 10 PSI)

Reduces vibration in long pipe runs Absorbs water hammer of high vertical discharge head

Maintains accuracy of inline meters provides cushion for shock of closing valves

Works as an accumulator for suction lifts

Typical dampener installation for various types of applications

DAMPENER DESIGNS AND APPLICATIONS

are used in different areas to accomplish different results (Figure 3):

•

•

Pulsation dampeners—used to dampen energy pulses produced by reciprocating pumps.

•

Surge suppressors—used to absorb pressure surges caused by quick-closing valves, long vertical runs and pump start-up/shutdown.

•

Suction stabilizer—used to control pressure fluctuations at the pump inlet. These minimize cavitation, increase efficiency, and reduce pump component wear.

•

Accumulator—used as a source of energy/liquid storage for fluid make up as well as momentary power source.

•

Appendage type—usually installed in a tee as close to the source as possible for maximum effectiveness. Appendage types are generally more compact and less complex. Flow-through type—installed directly in the pipeline as close to the pump as possible. These can be slightly more effective than the appendage design, depending on the application, but they are also more cumbersome and complicated, require more space and must generally be braced due to their offset installation.

Although the basic design is similar, different types of dampeners

30

The dampener supplier needs to understand how you intend to use

The Pump Handbook Series

the dampener because sizing formulas are different and more complex when used for surge or suction control. Pulsation dampeners are available in precharged designs, where the manufacturer either recommends or initially charges the unit to a predetermined pressure. Charging is usually done with nitrogen and for pressures over 150 psi. Adjustable designs, in which a regulator is supplied with the dampener, are also available. These allow the user to make pressure adjustments depending on fluid pressure. A third type, the so-called automatic design, allows the pressure to change automatically with the changes in the liquid pressure of the system, as in a filter press application. Adjustable and automatic models are generally used under 150 psi and with compressed air instead of nitrogen. They require an air line attached to the dampener and a continuous source of compressed air.

A FINAL WORD Reciprocating pumps can effectively transfer the most difficult of solids-laden and viscous fluids. Unfortunately, they expel liquid in slugs rather than in an uninterrupted laminar flow. With pulsation dampeners, however, these pumps can provide the steady, non-pulsing flow of a centrifugal pump. I encourage you to discuss your specific needs with quality manufacturers of pulsation dampeners. Many types and designs are available to meet specific application requirements and help you eliminate pulsation’s potentially damaging effects. ■ Gary Cornell is president and founder of Blacoh Fluid Control, Inc. He has worked in the reciprocating pump industry for 19 years and holds a bachelor’s degree from California Polytechnic University.

POSITIVE DISPLACEMENT PUMPS HANDBOOK

Chemical Additive Pumps for Paper Mills BY DAVID J. DAMATO AND ROBERT SCIASCIA

M

ost paper mills have a chemical additive system as a part of their process. When planning installation of such a system, or modifications in an existing system, you need to consider the type of additive used, its initial form, the application amount, the addition point and how to control the addition rate. Not all additive systems will employ the same type of pump to transport the chemical. Some additives use several different pump types, depending on their specific application. Even for the same chemical, the unloading, makedown, and feed pumps may use different pump technologies. Consequently, you should note the effectiveness of the particular pumping application and design and the economics of each delivery system. Piping systems will also vary with the specific application and function of the additive. Piping systems require special fittings, cleanouts, and elbow radii.

consider the following steps. Taking these design and selection considerations in this order, you should obtain the best results. 1. Determine the required flow rate. 2.

Determine the liquid viscosity at the operating temperature.

3.

Determine the suction piping size, which may be set by the storage/supply tank or net positive suction head (NPSH) requirements.

4.

Determine the discharge piping size by noting the flow rate and viscosity of the process fluid.

5.

Select the type and class of pump based on the application and process.

•

Select the pump size based on flow, pressure, NPSH and viscosity from the selected manufacturer’s data.

fluid properties (e.g., vapor pressure and pumping material compressibility)

•

mechanical design of the chemical additive pump (e.g., the volume of the pumping chamber)

•

the suction piping design

•

operating conditions (e.g., pump speed, fluid temperature and flow rate)

6.

7.

DESIGNING A CHEMICAL ADDITIVE PUMP SYSTEM There are several key process criteria that you should address when designing a chemical additive pump system. These considerations are equal in importance and should not be overlooked when selecting the pump for a chemical additive system. •

fluid to be pumped

•

flow requirements

•

pressure

Determine the horsepower requirements by noting the flow rate, total differential head, efficiency of the pump, and specific gravity and viscosity of the process fluid.

8.

Select the materials of construction for the pump by noting the fluid characteristics of the application, (i.e., fluid acidity or alkalinity, pH, abrasives, etc.).

9.

Consider the operating temperature of the process fluid because this may affect the materials of construction and the seal design.

PUMP SELECTION CRITERIA

10. Select the pump mounting and drive arrangement. Consider the physical arrangement and geometry of the suction piping as well as the discharge configuration, pump/motor orientation, filters, cleanouts and sample valve connections to ensure adequate accessibility for maintenance.

After looking at the process criteria, examine your pumping requirements. When selecting the proper pump for your specific application,

The NPSH for the pump application must be sufficient to support the process requirements. The NPSH required takes into account:

•

chemical unit weight

•

specific gravity

•

viscosity

•

temperature

•

pH

The Pump Handbook Series

Photo 1: Rotary lobe pump

Many chemical additive pump applications fail because of improper determination of the process variables and the design of the supporting equipment, and consequently the ability of the pump to supply the chemical to the pump suction adequately.

PIPING SYSTEM DESIGN The additive system must also be supported by well-designed piping. There are several methods of piping system design for any given application. Carefully study the details of the specific application and design the best piping system to support that application. To select a piping system: 1. Choose the type of dye system addition: neat (concentrated) or dilute. Neat dye system flow rates are typically in the milliliter per hour range, while dilute dye system flow rates are usually in

31

2.

3.

the gallons per minute range. The piping requirements for these two systems are totally different. Line sizes, fitting sizes, and materials of construction may vary depending on the dilution or concentration of the chemicals. In-line devices, such as flow meters and control valves, must be placed in the optimum locations for proper operation and maintenance accessibility. For polymer addition systems, carefully analyze the piping system prior to design. The viscosity of polymer changes greatly as the concentration is reduced. Also review the process piping variables in detail. For acid/alkaline addition systems, such as those used for pH control, note that the concentration of the chemical additive affects the corrosion properties. Obviously, the pH will dictate the piping material specification.

SPREADSHEET OF PROCESS VARIABLES Develop a spreadsheet to contain all the known process variables and supporting tank data. The spreadsheet will provide a concise list of the applications for each chemical additive in your paper mill application. Design the spreadsheet to calculate the tank volume, pump out rate required to support the chemical loading on the paper machine, tank retention time and chemical delivery volume and required ordering schedule. The spreadsheet can also serve as a reference for specifying other process equipment, such as tanks and agitators. Further customize the spreadsheet to track actual chemical usage in order to optimize delivery schedules. Table 1 shows a sample chemical additive spreadsheet.

ADDITIVE PUMP SERVICES & THEIR APPLICATIONS The following is a brief listing and description of the various pump service applications used in paper mill chemical additive systems.

UNLOADING PUMPS As their name implies, the main function of unloading pumps is to unload a chemical additive from a rail car, tank truck or drum. In many cases the pumps employed for rail

32

car and tank truck unloading must handle several different chemical additives. Economic consciousness prohibits the use of a dedicated unloading pump for each additive. Each unloading system must be equipped with a means to flush the unloading loop. Two common methods used today are: 1. a quick-disconnect type fitting method, which uses hoses to convey the chemical from the delivery vehicle to the storage vessel

Photo 2: Progressing cavity pumps in chemical additive applications

2. a hard-piped system that has the flush medium, usually water, hard piped into the unloading loop Isolation valves provide unloading or cleaning of the loop. Either method requires a flush capability that provides adequate cleaning of the unloading pump, line, and strainers. Centrifugal pumps, with stainless-steel wetted internals and mechanical seals, have been used with excellent results. Some chemical additives are shear sensitive and must use a two-vane centrifugal pump or a diaphragm pump. A conventionally packed pump may work; however, the introduction of seal water into a leaking packing gland can dilute the chemical additive to a point of poor performance, or even worse, dangerous concentrations of acids or caustic cleaning solutions can destroy the pump. Today the mechanical seal packages available perform well if they are properly installed and maintained. Some chemical additive unloading systems do not use pumps in their loops. Air padding systems (which use clean, dry air to pressurize the rail car or tank truck) transfer the chemical out of the shipping vessel and into the chemical storage tank. These systems should still have the same type of straining elements, to filter large contaminants, and also a type of cleansing flush loop.

MAKEDOWN PUMPS The main function of a makedown pump is to transfer a chemical additive from the storage tank to an intermediate process step for concentration reduction or ”makedown“ purposes. More precise control of the additive flow rate is required. The accuracy of the delivery system often The Pump Handbook Series

requires a positive displacement pump. When employing a positive displacement pump, design the process piping system carefully. Include several levels of pressure relief. Use lobe pumps, with stainless-steel lobes/rotors and mechanical seals, for systems where you need a relatively high flow rate (above 25 gpm), such as services with mildly abrasive pumpage. High-abrasive process fluids can cause excessive wear to the pump rotor and wetted parts, which can then reduce the pumping efficiency and throughput of the pump. For abrasive fluid applications, you may want to consider using specialized 32 RMS internal finish; however, this finish may be costly. To prolong the life of the rated pumping capacity, you can also use hardened rotors on abrasive applications. If you have moderate-viscosity process fluids, with little or no abrasive properties, use a synthetic rubber-covered or other nonmetalcovered rotor. Where flow rates operate below 5 gpm, use chemical metering pumps. These pumps are very accurate and repeatable. Limit the use of large flow rate chemical metering pumps because they can be costly.

MIXING PUMPS At times you may need to use a chemical additive pump to provide some mixing capabilities. This is a very difficult task for a ”non-vane“ pump. A centrifugal pump offers good mixing characteristics, however. Mixing pumps are not often used in paper processes because chemical additive applications usually ”inject“ the chemical into a tank or process pipe line. Mixing is accomplished by an in-line mixer, sized to fit the flow

Photo 3: Gear pump rate and the viscosity and specific gravity of the fluids to be mixed.

CHEMICAL FEED PUMPS The chemical feed pump is the ”fine control“ pump of the chemical additive application pumps. This type of pump application administers the chemical additive to the main process. The chemical feed pump can include a progressive cavity pump, rotary lobe pump, gear pump, diaphragm pump or chemical metering pump, depending on the process fluid flow rate requirements and fluid properties. These pumps are usually a positive displacement type pump. The hydraulic fluid relief method must be included in the design of the pump system. Several pump types have internal relief built-in; other pump types must have external relief methods included in the piping design. Precise, repeatable control must be easily achieved through the use of conventional means, including the piping system itself, fittings and valves, in-line flow (mass flow) meters, devices and controls. The process requirements will normally dictate the accuracy requirements.

ADDITIVE PUMP TYPES & THEIR APPLICATIONS PROGRESSIVE CAVITY PUMPS Progressive cavity pumps have been used in chemical additive systems for years with great success. Like all pump types, proper application, design, and operation techniques must be properly orchestrated to ensure a well-tuned pump system. Progressive cavity pumps are used for thick, viscous and abrasive chemicals and for various applications and flow rates. The single largest cause of progressive cavity pump failure is operating the pump dry, or without the process fluid flooded in the pump suc-

tion. Extreme friction, due to an interference fit (by design), generates an excessive amount of heat, causing the pump-casing liner to fail. Certain chemicals will attack the stator and cause a general breakdown in the pump internals. Carefully design controls for this application. For example, incorporate built-in time delays between a pump valve opening (to allow the process fluid to flood the pump suction) and starting of the pump motor. Other design considerations include the use of variable-speed motors to allow a ”soft“ start or slow initial rotation speed. Variable-speed motors applied to these pumps provide a wide range of service to any application. The rotation of the pump rotor conveys the process fluid at the desired point. The pump will ”see“ a process hydraulic back pressure during normal operation. If this back pressure is excessive, internal damage to the pump can occur. One common mistake is to install a pipe elbow at the discharge of the pump. During pumping, a viscous fluid is being positively displaced and the system pressure must eventually be equalized. Pipe elbows too close to the discharge restrict flow and reduce the life of the pump stator material. ”No-flow“ or low-flow rate monitors can be installed in the piping system to provide additional safety and prevent pump damage. Progressive cavity pumps are manufactured in a wide range of flows with a variety of stator liner materials available with both chemical and abrasion resistance properties. This pump type is a good application for the hard to handle viscous, abrasive or high solids process fluids.

pump through the suction port. The liquid fills the space between the teeth of the rotor and the idler. The unique shape of the pump head splits the flow of liquid as it is moved toward the discharge port. The rotor, supported and driven by the pump shaft, carries the liquid between its teeth, the casing and the outside surface of the pump head. The idler, supported on the idler pin projecting from the pump head, carries liquid between its teeth and the inside surface of the head. Gear pumps have excellent features and application for low abrasive levels of chemical additives for the pulp and paper industry. They are relatively low in cost and maintenance.

GEAR PUMPS

i = number of pump cylinders for multiple cylinder applications; this value can be fixed or variable, depending on the specific design Most chemical metering pumps use either the diaphragm or piston (or plunger) type of chemical propulsion. The diaphragm is flexed to pump the chemical. The piston (or plunger) type is stroked to pump the chemical. Chemical metering pumps can be configured several ways to increase their variability and resolution of control. For a variable stroke length, adjust the plunger stroke length to vary the quantity of fluid that will be

Gear pumps are part of the positive displacement, chemical feed pump category. The gear pump is a simple but accurate means of pumping precise amounts of a chemical additive into a process stream. The operating principle is unique in that there are usually less than four moving parts within this pump. The meshing of a rotor and idler gear teeth (running under close operating tolerances) causes the positive displacement of the additive chemical. With each shaft revolution a precise amount of liquid is drawn into the The Pump Handbook Series

METERING PUMPS Chemical metering pumps are usually regarded as the most precise and repeatable of the pump categories discussed here. The mathematical formula for calculating a metering pump size is expressed as: V= A * h * n * i where V = theoretical volumetric flow from the metering pump A = plunger area, which is usually a constant value, depending on the specific pump design h = pump stroke length, which varies from zero to a maximum value, depending on the design; the pump stroke length is equal to the radius of the plunger shaft to the center of the rotating mechanism of the specific chemical metering pump n = the speed or stroke frequency of the metering pump

33

displaced (pumped) with each stroke. This amount can range from zero to the maximum stroke length of the plunger. For variable speed, the variable-speed motor drives the stroke of the plunger, which then regulates the quantity of fluid being pumped per unit time. You can also significantly increase the range of the metering pump by including both a variable stroke length and variable speed. This feature can be useful for both batch and continuous chemical feed systems. In a chemical metering pump system several pump heads can be mounted on the same motor drive. This feature, which provides an even greater flexibility in chemical delivery flow rates, is especially valuable for dye addition or makedown, and polymer addition and makedown systems. Subprocesses can be serviced by the same multihead pump configuration. Multiple chemicals can be pumped by the same motor that drives a multiheaded pump. During the design phase of a chemical metering pump system, you will need to address several concerns. The pump’s internal orifices can be rather small, depending on the pump capacity or design. This makes the pump potentially vulnerable to contamination, so strainers or filters must be installed on the pump suction. Use careful design practices to prevent affecting the NPSH of the pump system. Filters with a very fine (200 mesh or greater) filter media may limit the NPSH available to the pump. This stresses the requirement of filtering the chemical additive stor-

age tank during both the unloading and delivery of the chemical to the process. There are two concerns with positive displacement pumps of this type: pulsation and overpressure. Pulsation can introduce unwanted hydraulic noise into the process. Overpressure can cause damage to process equipment or Photo 4: Pulp and paper chemical metering system personnel. Pulsation dampers resource in the application of pumps are available to smooth out inherent and support systems. pulsations from the pump discharge. There are many useful pump The dampers act as a hydraulic shock types for the addition of chemical absorber. Pressure sustaining valves additives in the paper industry, but to are available to maintain a relatively experience the best fit for your constant pressure or a chemical feed process application, choose the right stream to a main process. Overprespump for your job. sure of a process system can easily Support systems and equipment occur when a positive displacement play an equally important role in pump supplies a fluid to a closedchemical additive pump applications. process loop. The system must be proTake care to ensure that you apply tected against overpressurization with these items well. Plan ahead to maxia ”closed loop“ relief valve, which mize space, capacity, flexibility and returns excessively pressurized system expandability for your chemical fluid to the desired point, such as a additive pump and support system. storage tank. ■

CONCLUSION Chemical additive systems come in a variety of types and application requirements in the paper industry. Study each application and use good, sound engineering practices in the design. Equipment suppliers are quite valuable in supplying the appropriate equipment and application data for your specific use, and they are a key

David J. Damato is a Consulting Engineer in the OPTEC Division of RUST Engineering Company in Birmingham, AL. Robert Sciascia, a Staff Engineer for RUST, is also a member of the Pumps and Systems Editorial Advisory Board.

TABLE 1. CHEMICAL ADDITIVES Stra’t Wet End Chemicals Wall Chest Name Height Data Ft. Retention Aid Mix 10 Retention Aid Supp. 10 Dye Storage Tank 10 Dye Head Tank 4 Defoamer Storage 12 Defoamer Head Tank 6 Wet Strength Agent 18 Mach. Caustic Tank 6 Alkaline Size Mix Tank 10 Felt Wash-Batch 10 Wet End Cat. Starch 12 Starch Slurry Tank 9 Size Press Paste Stg. 12

34

Dia. Ft. 6 8 12 4 12 4 12 6 10 6 12 6 8

Volume Gal. 2,115 3,760 8,460 376 10,152 564 15,227 1,269 5,875 2,115 10,152 1,903 4,512

Chem Weight Lbs./Gal. 9.2 9.2 9.6 8.8 10.0 10.0 11.0 12.7 12.7 12.5 12.5 11.1 11.1

Spec. Grav. 1.10 1.10 1.15 1.05 1.20 1.20 1.32 1.52 1.52 1.50 1.50 1.33 1.33

pH 8.0 8.0 8.0 7.5 6.0 6.0 4.5 13.0 9.5 12.0 8.0 9.0 9.0

The Pump Handbook Series

Visc. cp. 75.0 75.0 75.0 65.0 15.0 15.0 50.0 40.0 40.0 25.0 25.0 25.0 25.0

Max Usage Ret’n. Content lb/g per min Hours Wgt. (Lbs) 0.50 70 19,405 10.00 6 34,497 0.01 14,099 81,147 13.00 0 3,293 0.25 677 101,610 1.00 9 5,645 7.50 34 167,656 60.00 0 16,088 5.00 20 74,482 25.00 1 26,461 48.00 4 127,012 27.00 1 21,116 100.00 1 50,052

POSITIVE DISPLACEMENT PUMPS HANDBOOK

Air Operated Diaphragm Pumps - Nineties Style BY ALAN D. TUCK, JR. ir operated diaphragm (AOD) pumps are a popular choice in a wide range of applications for several reasons, including low initial cost, simple installation, exceptional tolerance to variation in operating conditions/procedures, and the ability to handle problem media. AODs will pump virtually anything that flows, from water to viscous, corrosive, and abrasive sludges, and even some dry powders. Sealless, self priming, and inherently variable speed, they offer the plant engineer a reliable and easy-to-use alternative for difficult pumping applications.

A

HISTORICAL PERSPECTIVE Since the introduction of compressible fluid operated diaphragm pumps in the 1940s, the concept has experienced continuing adaptation and Photo 1: Aro Fluid Products’ 1/4” diaphragm pump in transfer service innovation. Early pumps were fabricated steel with fabric- reinexpelled. The air control mechanism Perhaps the largest single application forced rubber diaphragms, and were functions alternately to pressurize for AOD pumps is in the fast food marketed in the mining and construcand exhaust the air side of the industry where small pumps, runtion industries. During the 1960s diaphragm. ning on bottled gas, dispense soft industrial users became aware of the This principle may be incorporatdrink syrups and bulk wines. unique operating characteristics of ed in both simplex and duplex pump PRINCIPLE OF OPERATION the AOD pump and pressured manudesigns. Simplex designs, which facturers to produce cast iron and require a mechanical device to reposiThe basic principle of AOD stainless steel pumps with chemically tion the diaphragm after the dispump operation is displacement of resistant diaphragms. The 1970s saw charge stroke, have more complex the fluid being pumped by comthe development of the first noncontrol systems. The more common pressed air. Because the volume per metallic pumps and Teflon diaduplex design incorporates two stroke may vary under differing load phragms, which further expanded the diaphragm chambers working in tanconditions, the AOD is more cormarket. In the 1980s a number of dem with a common shaft. As one rectly classified as a quasinew manufacturers entered the marchamber discharges, the shaft reposipositive displacement pump rather ket, and more reliable air valve systions the diaphragm in the second than a positive displacement pump. tems that operated without lubchamber creating a suction condition, Its structure consists of a pump housrication were developed. Today, air and it also sends a signal to the air ing, diaphragms, fluid valves, and an operated diaphragm pumps are being control mechanism to reverse the air control mechanism. The housing, manufactured in every major indusmotion. Duplex designs are generally plus the diaphragm and fluid valves, trial market in the world. They are more versatile and operate at higher form a containment vessel. As the available in countless materials, sizes, efficiencies. diaphragm reciprocates, the volume and design configurations, and serve of the vessel alternately expands and ADVANTAGES OF AOD PUMPS a remarkable number of industries contracts, filling with fluid and then from computer chip manufacturing The structure of the pump housexpelling the fluid. The valves control to waste treatment plants, from pharing has a number of operating advanthe flow to ensure that the cavity fills maceutical production to mining. tages. All sealing surfaces are static from the pump inlet and the fluid is The Pump Handbook Series

35

with no relative motion between components in contact with other elements, and high displacement ratios are possible without precision fits or close tolerances between components. With no mechanical seals or close-fitting components requiring lubrication, the AOD pump can run dry indefinitely without damage—an important advantage where operating conditions or operator error may result in interruptions in the supply of fluid to the pump inlet. The absence of closefitting parts and low internal fluid velocities make the AOD pump an excellent choice for abrasive fluids. High displacement ratios generate the vacuum required to handle high-viscosity materials and to ensure troublefree priming. The AOD pump’s diaphragm is a separation medium isolating the compressed air that is doing the work from the fluid being pumped. The differential pressure across the diaphragm and corresponding internal Photo 2: Versa-Matic Elima-Matic stresses are low in comparison with similar stresses in mechanically driven certain materials. Ball valves have diaphragm pumps that use the better vacuum capability and are diaphragm to generate pressure. available in all materials. The choice Discharge pressure capability of valve design will depend on the depends on the compressed air presmaterials pumped and application sure available, up to 300 psi in some requirements. models. The advantages are longer The air control mechanism is the diaphragm life and higher discharge functional heart of the AOD pump. It pressures. must direct incoming compressed air Fluid valves in AOD pumps are to the appropriate chamber of the available in two primary types: ball duplex pump while exhausting air valves and flap valves. Each style has from the other chamber. Without advantages. Flap valve pumps can centering or failing to shift, the air pass larger solids, but they may not valve system must operate from full be as reliable and are not available in throttle to a stalled condition when

36

The Pump Handbook Series

the pump is dead-headed. It should function correctly using either dry laboratory compressed air or the wet dirty air found in many plants. Most duplex pumps rely on a mechanical signal generated by the pump shaft, or attachments to the shaft, to shift either a pilot valve or the primary air valve. A limited number of models are available with externally controlled solenoid valves. In response to environmental concerns, several manufacturers offer valve systems that do not require external lubrication. Other designs rely on filters to remove the oil from the exhaust, or simply direct the exhaust to a safe area. Some valve designs are made with precision elements and are reliable for clean dry air, but do not accommodate the scale and detritus found in many plant air systems. Others feature loose fits and large clearances, but may not be as reliable in every circumstance. The reliability and durability of valve systems should be a primary consideration in selecting an AOD pump. A plant engineer should also consider air quality and maintenance experience in the facility when selecting an AOD pump. AOD pumps are available in a wide range of corrosion resistant metals and plastics, including polished stainless steel models that have received USDA approval for use in meat and poultry processing. Diaphragms are available in a variety of chemically resistant polymers including Teflon. For pump applications demanding high purity or

involving hazardous materials, sealless AOD pumps offer many options.

USE OF COMPRESSED AIR

prevents stalling even under low air inlet pressures. The ARO design incorporates an unusual octagonal shell, giving the pumps a distinctive new look. The pump is available in a variety of materials, including a patented conductive Acetal for use with flamma-

The use of compressed air as a power source for a pump has both positive and negative effects. Compressed air is easily controlled by a variety of simple, inexpensive devices permitting variable output volume and pressure without costly motor controls. It is easy to install and completely portable, and eliminates the requirement for motor couplings and precision alignment. Compressed air is inherently explosion proof and allows AOD pumps to be dead-headed without damage. These factors give AOD pumps the advantages of low initial cost, simplicity of installation and operation, and higher system reliability by protecting against operator error or other anomalies. However, compressed air is a relatively costly utility in many plants, requiring compressors, Photo 3: Wilden Pump M20 in sludge service air distribution systems, and higher energy ble solvents and other volatile costs. The plant engineer must determaterials. mine if the many advantages of the Versa-Matic Pump Company’s AOD pump offset total costs. new Elima-Matic Air Valve (Photo 2) DESIGN INNOVATIONS was recognized at the 1993 National Plant Engineering and Maintenance The pace of innovation continues Show as the “Best New Product in to accelerate as AOD manufacturers the Fluid Handling Category.” This respond to “market requirements.” valve has features designed to miniTwo new AOD designs received mize air valve “icing” and to improve industry recognition in 1993. ARO reliability. “Icing” is a problem for Fluid Products’ new 1/4” diaphragm AOD pumps in plants with poor qualpump (Photo 1) was selected as the ity lubricants. The Elima-Matic solution “Best New Product of ‘93” in the fluid uses a series of progressively larger power category by Design News magexhaust passages that control expanazine. The ARO pump delivers 5 galsion of the compressed air in the lons per minute and has +20 feet valve and minimize cooling in this suction lift. The unit features univerarea. The pump center section, sal mounting, and can be plumbed in which contains these passages, is a four distinct fluid inlet/outlet configufinned aluminum die casting with rations. The pump features ARO’s improved heat transfer properties unbalanced air valve design, which The Pump Handbook Series

that further delays the onset of “icing” conditions. The Elima-Matic Air Valve system incorporates two air valves, a main air valve and a pilot valve, activated by the final 1/2 inch of travel of the shaft connecting the diaphragms. The main air valve has small magnets mounted in each end of the spool and in the cylinder end caps. These magnets hold the spool, which is shifted by the pilot valve, in position until there is sufficient air pressure to shift completely, minimizing stalling or centering.

WASTE TREATMENT PUMPS Wilden Pump and Engineering Company and Warren Rupp, Inc. have introduced new pump models designed to meet waste treatment industry requirements. Wilden’s M20 (Photo 3) and Rupp’s wastewater pumps (Photo 4) have externally accessible fluid valves that can be cleaned without disconnecting the pump from the system plumbing or disassembling the pump. The Rupp wastewater pump incorporates a swing valve that will pass 3-inch spherical solids. Wilden’s M20 design uses ball valves with clearance to pass 1 3/8-inch solids. These new models also have applications in other industries, pumping slurries with even larger solids or fibrous materials that tend to clog conventional AOD pumps.

NEW AOD PUMP ACCESSORIES Most manufacturers now offer accessories to facilitate or improve the performance of AOD pumps in new applications. Wilden and Rupp have developed state-of-the-art, microprocessor-controlled computer flow control systems for metering and batching applications. These make it possible for cost-effective

37

AOD pumps to be used in applications that previously required more expensive metering pumps. Wilden’s controller can be utilized with a standard AOD pump by attaching a pulse sensor to the air discharge, or with Wilden’s solenoid controlled pumps. If the system needs to ensure a constant differential between pump inlet pressure and discharge pressure, the AOD will deliver a very consistent flow rate. Pulsation dampeners, now offered by virtually every manufacturer, are increasingly being installed for several AOD pump applications. Vibrations and surges produced by the reciprocating character of the AOD pump can take a heavy toll on piping systems. Correctly installed dampeners provide a smoother continuous flow that is critical for proper performance of batching controllers or flow meters. It also protects piping and valves and increases the life of pump components. Research by Blacoh Fluid Controls, a manufacturer of dampeners, suggests that installing a pulsation dampener on the inlet side of AOD pumps used in high flow rate

Photo 4: Warren Rupp AOD pump with speed control system

TABLE 1. AIR OPERATED DIAPHRAGM PUMP ADVANTAGES Dry, self-priming suction up to 20 feet Variable speed Variable pressure (up to 300 psi – with HP model) No mechanical seals Explosion resistant Submersible Completely portable No rotating parts (or ball bearings) Can run dry without damage No couplings required Dead head capability (without bypass valves) No efficiency loss due to wear No efficiency loss due to changing pressure or flow Complex controls not required Does not generate heat Handles abrasive fluids Handles abrasive, high viscosity fluids Handles non-abrasive, high viscosity fluids Handles abrasive highly corrosive fluids Handles non-abrasive, highly corrosive fluids One unit handles many types of materials Operating temperature from -30°F to over 300°F Ideal for shear sensitive products

38

applications may dramatically improve diaphragm life. In these situations, the fluid velocity in the inlet stream can produce a differential load (e.g., water hammer) on the diaphragm during the start-and-stop action of the pump, leading to premature diaphragm failure.

THE BOTTOM LINE Air operated diaphragm pumps don’t cost a lot. They are easy to use, difficult to abuse, and they get the job done. The continuing response to market requirements and innovation by AOD pump manufacturers offers the pump user an increasing range of cost-effective solutions to difficult pumping problems. ■

Alan D. Tuck, Jr. is President of Fluid Systems Engineering, an independent consulting firm specializing in pump design and troubleshooting in Bellingham, WA. His career includes 12 years as a marketing manager and engineering specialist with Wilden Pump and Engineering Company and Discflo Corporation. He received a bachelor’s degree from the US Air Force Academy.

The Pump Handbook Series

POSITIVE DISPLACEMENT PUMPS HANDBOOK

Positive Displacement Pump Vibration BY: KEN ATKINS

PULSATION High vibration levels resulting from inadequate pulsation control usually correlate with high maintenance and poor machinery reliability. In addition, pressure pulsation acting on unbalanced areas such as elbows and closed valves will generate dynamic shaking forces that can cause high vibration levels in spite of adequate mechanical supports. A less obvious effect of high pulsation is the potential for cavitation even when the static pressure has a sufficient margin above the vapor pressure. Pulsation may cause the pressure to drop instantaneously to the vapor pressure, (Figure 1) resulting in severe pressure spikes from bubble collapse as illustrated by Figure 2, which represents actual field data. These pressure spikes can damage pump internals, such as valve plates, plungers and working barrels. With severe pulsation, this phenomenon has been documented even in the discharge of pumping systems. A feature article in the June 1994 issue of Pumps and Sytems magazine (Ref. 1) discussed gas-charged pulsation dampeners as a means of pulsation control, and a comprehensive discussion of how these devices work can be found in reference 2. Since the gas charge offers an acoustical compliance characteristic that yields an effective volume many times larger than the same volume of liquid, these devices can be very compact. The effective volume can be computed as follows:

KliquidVgas = (ρc2)liquidVgas Kgas (ρc2)gas

•

bladder stiffness

from the gas-charged device. The gascharged device acts simply as a large compliance (volume). The all-liquid device is usually designed as a lowpass filter to attenuate pulsation levels at frequencies above a specified cutoff frequency. Acoustic elements such as volumes, choke tubes, and orifice plates are used to design the filter characteristics to accomodate particular applications. The filter itself actually creates an acoustic resonance (Helmholtz frequency). But, the device works to attenuate pulsation levels at frequencies well above this resonance. Consider a triplex pump operating at 250 rpm. The plunger frequency can be calculated as follows:

•

absorption of gas in non-bladder devices

Plunger Frequency =

•

sensitivity to charge pressure

•

bladder fatigue failures

•

permeability of bladder materials to certain liquids.

V'=

where: K = bulk modulus (psi) c = speed of sound (ft/sec) ρ= density (lb/ft3) V = volume (ft3) V' = equivalent liquid volume (ft3) Unfortunately, the performance of gas-charged devices can be degraded due to several factors, including: • neck restriction (which reduces the compliance effects)

3 x 250rpm = 12.5Hz 60rpm/Hz

An all-liquid filter can be designed with a Helmholtz frequency of 6 Hz. This device would create an acoustical resonance at 6 Hz; but since the lowest excitation frequency is 12.5 Hz, pulsation levels at the plunger frequency and its harmonics would be effectively attenuated. Generally, the

Another common method of pulsation control is the "all-liquid filter." This method differs conceptually

FIGURE 1. CAVITATION DUE TO PULSATION PRESSURE Instantaneous Pressure (PSI)

I

nadequate pulsation control and poor valve performance, along with faulty mechanical piping design, are the primary causes of excessive vibration in reciprocating positive diplacement (PD) pump systems. Therefore, a working knowledge of how pulsation and valve dynamics influence PD pump vibration is essential for the design and operation of safe and reliable systems.

Pd Ps Pvp P=O Time If Pd > Ps - Pvp : then cavitation will occur. Ps = Static Pressure Pd = Dynamic Pulsations, O - p Pvp = Vapor Pressure

The Pump Handbook Series

39

sharply as the volume-choke-volume filter does. However, it is effective in pump systems since, unlike compressor systems, higher pressure drops may usually be tolerated. As a result, this volume-choke configuration often provides a good compromise between initial cost and pressure drop concerns. All-liquid filters are relatively maintenance free when compared to gas-charged devices. They may also provide more effective pulsation control and are easily fabricated from typical piping components. The main disadvantage of all-liquid filters, relative to gas-charged devices, is their larger size.

Pressure (PSI)

FIGURE 2. FIELD DATA SHOWING CAVITATION SPIKES

Cavitation Spikes

Vapor Pressure

0.00

0.01

0.02 0.03 Time (Seconds)

lower the Helmholtz frequency, the better the pulsation filter. All-liquid filters can be configured to achieve various characteristics. A symmetrical volume-chokevolume arrangement is shown in Figure 3. The following equation gives the relationship between filter frequency and the dimensions of the filter arrangement. c f=

2πL

(

d D

)

where: f = frequency (Hz) c = speed of sound (ft/sec) d = choke diameter (ft) D = diameter of each bottle (ft) L = acoustic length of bottles and choke (ft) As indicated by the equation, lower filter frequencies require either larger bottle chambers or smaller diameter choke tubes. Larger volumes cost more to build initially, but smaller choke tubes result in higher pressure drops and therefore higher operating costs. However, this equation does not take into account the effect of attached piping networks on the filter performance. Because this effect does not generally follow a simple mathematical relationship, critical systems should be simulated using digital or

40

0.04

0.05

VALVE PERFORMANCE Valve dynamic effects can also cause large deviations in the pressures acting on a pump's working parts. The typical valve consists of a spring and a valve plug or plate that seals against a seat. An overpressure may occur once per cycle, producing dynamic stresses on pump components. Since the pressure required to open the valve is controlled by the spring forces and the differential pressure across it, the valve design can influence the magnitude of the overpressure. The forces acting on a closed valve that can result in overpressure (illustrated by Figure 5) are described by the following equations:

analog techniques to ensure adequate pulsation control. This modeling also allows the trade-offs between installation costs (volume bottle size), pressure drop, and pulsation attenuation to be optimized, and it should be done regardless of whether gascharged devices or all-liquid filters will be used. A commonly employed all-liquid filter configuration for PD pump systems is illustrated by Figure 4. This configuration is referred to as a volume-choke all-liquid filter and utilizes a single volume (bottle) near the pump flange and a choke tube that may be internal or external to the bottle. The choke tube connects directly to the larger suction or discharge piping through a reducer. This configuration does not cut off higher frequency pulsation components as

1. Spring force Fs = Fpreload 2. Pressure forces F1 = PcAc (in cylinder)

FIGURE 3. SYMMETRICAL VOLUME-CHOKE-VOLUME ALL-LIQUID FILTER

L

L d D

D

L

The Pump Handbook Series

FIGURE 4. VOLUME-CHOKE ALL-LIQUID FILTER Suction or discharge flange

CL Pump

F2 = PmAm + Fpreload (in manifold) The valve will open when: F1 > F2 or: PcAc > PmAm + Fpreload Pc > Pm(Am/Ac) + Fpreload/Ac where: Ac = the area of the valve in contact with the cylinder pressure Am = the area of the valve in contact with the manifold pressure The differential area, required at the valve seat to seal the liquid, may cause an overpressure. For the valve to open, the cylinder pressure must be greater than the manifold pressure by the ratio of the areas Am/Ac plus an additional factor to overcome the spring force. The area ratio Am/Ac is typically 1.1-1.5 to yield a seating area sufficient to control valve impact stresses. In addition to the spring, mass and differential pressure effects, there is a "sticktion'' effect that opposes the separation of two lubricated flat surfaces (Ref. 3). This sticktion force is influenced by the initial fluid film thickness, the viscosity of the fluid and the geometry of the surfaces. The sticktion results in ”overpressure“ spikes on the discharge valves and ”underpressure“ spikes on the suction valves. Large overpressure spikes can cause various problems such as:

1.

working barrel failure

REFERENCES

2.

crosshead guide and case failure

3.

bearing damage

4.

crankshaft and connecting rod failures

5.

reduced valve life.

1. Cornell, Gary, "Pulsation and Surge Control," Pumps and Systems Magazine, June 1994, pp. 32-36. 2. Wachel, J.C. and Price, S.M., "Understanding How Pulsation Accumulators Work," Pipeline Engineering Symposium-1988, PDVol. 14, Book No. 100256. 3. Bauer, Friedrick, "The Influence of Liquids on Compressor Valves," 1990 International Compressor Engineering Conference, Purdue University, West Layfayette, Indiana, July 1990. 4. Tison, J.D., et. al., Vibrations in Reciprocating Machinery and Piping Systems, EDI Seminar Manual, Chapter 2, May 1992.

Another phenomenon related to valve performance is valve lag. The time required for a valve (suction or discharge) to return from a fully open position to its seat is dependent on the valve mass and return spring properties. Therefore, as the speed of a pump increases, the fixed finite time required for valve closure to occur results in a greater relative valve lag in relation to crank rotation. The lag of the discharge valve closing can actually cause backflow through the discharge valve. Likewise, a lag in the suction valve can allow backflow through the valve. This problem not only reduces pump capacity, but also changes the flow excitation characteristics of the pump. Since the pulsation levels are directly proportional to the flow modulation amplitudes, valve dynamic effects can increase the pulsation levels in a pump-piping system. Valve dynamics effects can be computed using a time-stepped integration method (Ref. 4). This computation involves calculation of spring and mass properties, sticktion effects, and valve lag as well as calculations of pressure versus time to predict overpressure and underpressure spikes and valve lift versus time to define valve lag. Valve parameters (e.g., spring stiffness, mass, surface geometry, lift) can be optimized using this technique to minimize pressure spikes and valve lag.

CONCLUSION Pulsation and valve dynamic effects should be of primary consideration when designing and troubleshooting PD pump systems. Careful attention to these factors, as well as to the mechanical support of the attached piping, will ensure a more reliable system. ■

Ken Atkins is Senior Project Engineer for Engineering Dynamics Incorporated in San Antonio, TX.

FIGURE 5. FORCES ACTING ON A CLOSED VALVE Pm Am

x

Ac Pc cylinder F2

F1

The Pump Handbook Series

41

POSITIVE DISPLACEMENT PUMPS HANDBOOK

Specifying Air-Operated, Double-Diaphragm Pumps BY: GARY J. BOWAN ir-operated, double-diaphragm pumps (AODDP) are employed for a host of difficult pumping applications ranging from municipal sump to high purity operations. These easy to maintain pumps offer the ability to vary speed and pressure, self-prime, run dry and dead-head with the absence of leak prone dynamic seals. Proper pump specification will result in high return on investment. Many AODDP manufacturers provide performance curves, exploded views, parts listings, engineering and maintenance manuals and videos to assist end users in the specification process. In addition, the following key points can serve as a guide to establish specification criteria for your AODDP application.

A

MATERIALS OF CONSTRUCTION When specifying materials of construction for AODDPs, consider the following factors. • Chemical compatibility Wetted materials and elastomers, including liquid inlet and discharge manifolds, liquid chambers, outer diaphragm piston plates, diaphragms, valve balls, valve seats, and O-rings, must be chemically compatible with the process fluid as well as solutions used to clean the system. Consult a chemical resistance chart available from the AODDP manufacturer to assist in your selection efforts. If the process fluid is not listed in the chart, contact the chemical manufacturer for assistance. • Abrasion Because the AODDP doesn’t exhibit any metal to metal contact and uses two pumping chambers with low internal velocities, it is able to handle viscous, abrasive slurries by design. Abrasion, however, can occur in these pumps, particularly in the discharge ball cage area. To prevent abrasion in applications involving abrasive slurries, specify mate- rials exhibiting good abrasion resistance such

42

as cast iron for pump housing and polyurethane for elastomer components. • Temperature limitations If operated outside stated temperature limitations, ”cold flow,” a change in molecular structure, can occur in plastic pumps. Under this condition the pump may become brittle or soft, creating a hazardous condition. Metallic pumps are generally limited in temperature range by the pump elastomer specified. Elastomers may increase or decrease in size when operated outside their temperature range, creating unacceptable working tolerances. Consult the AODDP manufacturers for specific material temperature limitations. • Initial investment Pump prices can vary by as much as 1000% depending on the choice of materials. When selecting materials, consider all aspects of the application and compare the predicted return on investment to the initial cost.

AIR DISTRIBUTION SYSTEM The air distribution system, the heart of the pump, includes the air valve and related components. This system reciprocates the pump by directing the air supply to one air chamber and redirecting it to the other upon completion of the stroke. There are many different types of air distribution systems on the market. Some air valves are externally serviceable, employing only one moving part which operates strictly on pressure differential. Others require bearings, springs, magnets and mechanical assistance to operate. Consider the ease of cleaning, repair time and component replacement cost incurred after diaphragm failure.

AIR CONSUMPTION Calculating the Total Dynamic Head (TDH) for the specific application and plotting this figure versus the desired flow rate (gpm) on the pump performance curve supplied by the AODDP manufacturer will allow you to determine the amount of air required. This calculation should account for pressure losses attributed The Pump Handbook Series

to viscosity, specific gravity and pipe friction. Once these points are plotted on the curve, the air supply pressure (psi) and volume (scfm) can be derived. The following simple calculations provide a basic guideline to estimate horsepower based on the volume of air required to operate the pump. For pump supply pressure under 40 psi: .125 (scfm) = Horsepower For pump supply pressures between 41-125 psi: .25 (scfm) = Horsepower

CHECK VALVES Most AODDPs use four check valves. The ball/seat and flap valve type configurations are the most common. Of these, the ball/seat configuration performs better, providing greater vacuum capabilities and longer component life. The ball/seat design wears evenly because the valve ball spins when released and assumes a different position on the valve seat during each stroke. The flap valve, hinged in a fixed position, flexes continuously at the same point, which leads to fatigue and failure. Particles in the process fluid can easily be trapped between the flap valve and seat, resulting in damage to the flap and leading to pump performance concerns. The flap check valve should be used only if the application requires the passage of large solids that cannot be passed with a ball/seat configuration.

COST OF MAINTENANCE Contrary to popular belief, AODDPs do require maintenance to optimize performance. Replacement parts, maintenance costs, and down time should be taken into consideration when specifying an AODDP to assure long-term satisfaction. ■ Gary J. Bowan is market support coordinator for Wilden Pump & Engineering Company in Grand Terrace, CA.

POSITIVE DISPLACEMENT PUMPS HANDBOOK

Pipeline Screw Pump Efficiency Screw pumps can offer efficiencies of better than 80% in dealing with medium and high viscosity fluids. By James R. Brennan fficiency is just as important in screw pumps as it is for other kinds of pumps. In fact, screw pumps, like other rotary positive displacement pumps (gear, vane, lobe or other designs), are often selected because they offer excellent efficiencies at medium and high viscosities. Boosting and pipelining heavy crude oil is a classic example in which screw pumps offer cost benefits over conventional centrifugal pumps. With today’s oil prices, production and transportation operating costs are more important than ever, and they are likely to remain that way in the future. Screw pumps are uniquely suited to this service because they offer pump efficiencies in the 80% range and require little, if any, additional crude oil heating or dilution. They are available in large flow rates as well as high pressure designs, providing the power range usually needed for pipeline services. Multiple screw pumps have been used in heavy oil pipeline services since the 1950s. Their capabilities and capacity range have improved steadily, and such pumps are now able to provide reasonable life in this demanding service. They also offer excellent operating efficiencies. Multiple screw pumps for crude oil pipeline service are produced in two basic configurations. One of these, the twin-screw double suction design pump, is illustrated in Figure 1. Each shaft is independently supported with bearings external to the pumped liquid. The mesh of the screw set is synchronized through the outboard timing gears;

E

the screws do not touch each other. Figure 2 shows the construction of a three-screw pipeline pump of the single suction design. It incorporates a replaceable liner and single,

external support bearing. The smaller outside screws, called idler rotors, are driven by and are in rolling contact with the center screw, called a power rotor. Because there is metal-

Lube Oil Reservoir

Inlet

Outlet

Timing Gears FIGURE 1. Double suction twin-screw pump THE PUMP HANDBOOK SERIES

43

Inlet

Outlet

to-metal contact, however, these designs cannot be used with high gas content or for 100% water. Water-oil emulsions, on the other hand, do not present a problem to this type pump. Each design has its merits depending on pressure, flow and crude oil conditions. Screw pumps are positive displacement machines; every pump shaft revolution causes a specific volume of space to be opened to the system inlet pressure environment and then closed off from the inlet. The volume moves in an axial direction and is expelled from the pump by the next succeeding volume. Flow is very smooth and almost completely free of any measurable pressure or flow pulsation. At constant speed the theoretical displacement of screw pumps depends upon the size and geometry of the screws and the screw pitch or lead. The pumps, obviously, have internal running clearances and will not be able to deliver 100% of their theoretical flow when pumping against a differ-

44

Pressure Rise

FIGURE 2. Single suction three-screw pipeline pump

Wraps FIGURE 3. Pressure distribution over screw stages

ential pressure.

SLIP FLOW Slip flow occurs through the running clearances; it is a function of differential pressure and fluid viscosity. Increasing differential pressure and decreasing viscosity will increase slip. Slip flow is the volumetric inefficiency. A pump with a theoretical flow of 432 gpm operatTHE PUMP HANDBOOK SERIES

ing at 1000 psid and 20 centistokes (100 SSU) might have a slip flow of 59 gpm. Thus, the pump would deliver 432–59 or 373 gpm. The volumetric efficiency of the pump would be 373/432 or 86%. At higher viscosities, common on crude oils with an API gravity less than 20° (specific gravity greater than 0.934), the volumetric efficiency of a multiple screw pump can be 90 to 95%. To limit the slip flow characteristics of multiple screw pumps, higher pressure pumps of both two- and three-screw design use more “wraps” of screw thread than lower pressure designs. Each wrap acts as a barrier to slip flow, effectively causing the pump pressure rise to occur in stages (Figure 3). The staging effect lowers the loading on rotating pump components and provides greater resistance to slip flow which improves volumetric efficiency. At low viscosities, slip flow is the major contributor to multiple screw pump inefficiency. At increasing viscosity, the slip flow is reduced, sometimes to a negligible level. However, as viscosity increases, more power is required to rotate the pumping screws within their close clearance stationary boundaries. Viscosity is defined as a liquid’s resistance to shear. The pumping screws shear the liquid that is within the running clearances, and this is the major contributor to inefficiency when operating at high viscosity. Proper selection of pump size and speed can keep these viscous shear losses within reason and usually does not require any pump speed reduction devices except under the most severe conditions. The low shear characteristic of screw pumps is particularly desirable when pumping shear sensitive crude oil-water emulsions. Figure 4 is the performance curve of a typical crude oil pipeline screw pump at constant speed. The effects of viscosity and differential pressure are clearly evident. Note that, like all machines, the pump efficiency is zero at zero differential pressure. This is the point where the machine would pump its theoretical flow, but since there is no pressure rise, there is no power output, yet it requires about 15 hp minimum to keep the pump rotating. Also note that there is no “best efficiency” point as with a centrifugal pump.

MECHANICAL LOSSES The last category of losses within a multiple screw pump is mechanical losses due to bearing friction, timing gear inefficiency and mechanical seal drag. Most modern pipeline screw pumps use antifriction, externally lubricated bearings; thus friction losses are very low. Precision timing gears can operate in the 98% efficiency range. The mechanical seal drag has a component of loss for the body rotating within the fluid as well as a component due to the shearing of the liquid film between the rotating face and stationary seat. All these losses are normally very small and might contribute only a few percentage points of inefficiency. This is true unless pumps are operated at very low hydraulic power levels, where these fixed losses can become a significant portion of the total power requirement. As with most rotating equipment, larger machines are more efficient. In the case of screw pumps, the reason is that the theoretical flow rate is a function of the cube of the screw size while slip flow, everything else being constant, is a function of the square of the screw size. Figure 5 illustrates this effect for an 8 wrap, 1000 psid, 1200 rpm screw pump handling 100 centistoke (500 SSU) crude oil. To a lesser degree, smaller size pumps have disproportionately larger running clearances due to manufacturing accuracy limitations. They therefore exhibit lower volumetric efficiencies than larger machines.

PUMP EFFICIENCY CALCULATIONS Because pumps are frequently sized to operate over a range of pressure and viscosity, cost calculations should use the power required at normal operating pressure and viscosity (see Table 1). Screw pumps require their maximum input power at maximum viscosity. The minimum flow will be delivered at the minimum viscosity. Do not use the minimum delivered flow and the maximum required power to calculate overall pump efficiency. This method understates the efficiency as simultaneous operation at these conditions is not possible.

GPM 450 425 400

300 cst

100 100 cst

90

20 cst

80

375

70

350

60 400

50

300 cst

350

40

100 cst

300

30

20 cst

250

20 200 10

300 cst (1400 SSU)

150

0

100 cst (500 SSU)

100

EFF. %

20 cst (100 SSU)

50 BHP 150

300

450

600

750

900

1050 1200 1350 1500

Differential Pressure PSI FIGURE 4. Pipeline screw pump performance curve

100

90

80 Overall Pump Efficiency %

Multiple screw pumps have a rapidly rising efficiency curve which then holds fairly high throughout its design pressure range.

70

60

50

1

2

3

25

50

75

4

5

100 125 Pump Screw Size

6

7 Inches

150

175 MM

FIGURE 5. Efficiency improvement with pump size The Pump Handbook Series

45

E0 = Power out/Power in = (QD x PD x 100 /k)/W E0 = EV x EM x 100 EV = QD /QT x 100 where: E0 EV EM QD PD k W QT

= = = = = = = =

overall pump efficiency pump volumetric efficiency pump mechanical efficiency pump delivered flow rate pump differential pressure conversion constant pump input power pump theoretical flow rate

units: % % % gpm psi 1714 hp gpm

TABLE 1. Pump efficiency calculations

Photo 1 shows a three-screw crude oil emulsion shipping pump on a California offshore platform. There are three pumps on each of two platforms. Each pump delivers 800 gpm (27,500 B/D) at design discharge pressures to 1190 psig. The 800 hp, 1150 rpm electric motor drivers were sized to handle a maximum pumping viscosity of 350 centistokes (1610 SSU). The overall pump operating efficiency at this point is 82%. Photo 2 is a twin-screw pump for pipeline service in Venezuela. It handles 2552 gpm (87,500 B/D) of Orimulsion, a shear-sensitive emulsion of 30% water and 70% bitumen plus surfactants that is exported as a power plant fuel. Design differential pressure is 531 psid and the pumping viscosity range is 215 to 970 centistokes (1000 to 4500 SSU). The pumps are driven by 1250 hp, 1150 rpm electric motors and operate in the 78 to 82% efficiency range.

ENERGY COSTS Energy cost evaluations can be extensive or simple depending on a company’s accounting practices. On a simplified basis, let us assume a comparison between a screw pump and a centrifugal pump on typical heavier crude oil pipeline service:

PHOTO 1: Three-screw crude oil emulsion pump on California offshore platform

PHOTO 2: Twin-screw pump for pipeline service in Venezuela

46

The Pump Handbook Series

Requirements: Flow: 583 gpm (20,000 B/D) Pressure: 1000 psi (70 BAR) Crude Viscosity:1000 SSU (200 cst) Screw pump Efficiency: 82% Power required:415 hp (309 kW) Centrifugal pump Efficiency: 45% Power required:756 hp (563 kW) The power difference is 563–309 or 254 kW. With 8760 hours in a year, the annual energy difference is 254 × 8760 or 2,225,040 kW-Hr. At an energy cost of $0.07 / kW-Hr, the annual direct energy cost difference is $155,000. Add the cost of carrying money, currently about 10%, and the difference is more than $171,000 per year for one pump. Most pumping stations include two 50% capacity running pumps plus a third in standby mode. Driver costs also can be substantially different, as are the costs of ancillary items such as motor starters and cabling. The differential cost of each motor for the preceding example is approximately $33,500 each, or

BAR 140

100

PSI 2000

1500

Three Screw 55

400 Two Screw

20

300

0 US gpm

0 0

500

1000

B/DX1000

0

20

40

M3/H

0

2000 60

250

3000 80

100

500

4000 120

140

750

5000 160 1000

6000

180

200 1250

7000 15000 220

515 1500

3400

FIGURE 6. Capacity and pressure range for crude oil screw pumps

$100,500 premium in initial cost for a three pump station using centrifugal pumps with 800 hp motors vs. screw pumps using 450 hp motors. Maintenance expense differences are difficult to judge because each operator has different practices and expectations. Often, the operator does not have enough of each type of pump operating on the same crude oil at the same differential pressures to make valid comparisons. On heavy crude oils, substantial wear can physically take place before the increase in slip flow suggests that an overhaul be conducted. The annual differential energy cost savings can

be typically several times the replacement cost of a screw pump, so total cost of operation—energy and maintenance—is lower while capital investment is usually equal or lower using screw pumps. Figure 6 shows the range of multiple screw pump sizes available for transport of crude oil. To maintain high efficiencies over longer periods when pumping crudes with sand, carbonates and sediment, screw pump manufacturers use a number of techniques to enhance the life and prolong the running clearances within these pumps. Hardened or hard coated screws, hard chrome plated

The Pump Handbook Series

liner bores, hard/soft and hard/hard combinations of running surfaces and erosion resistant inlays and overlays all can be used to contribute to longer useful pumping life between overhauls. ■ James R. Brennan is Manager, Crude Oil Pumps, for three operating units of Imo Industries Inc., Monroe, NC. His responsibilities include worldwide marketing and technical support for crude oil pumping applications. He is a 1973 graduate of Drexel University and a member of SPE.

47

POSITIVE DISPLACEMENT PUMPS HANDBOOK

In the Pipeline: PD Screw Pump Valving Proper screw pump valving and instrumentation boost pipeline yields. By James R. Brennan ultiple screw pumps are often used in pipelines to transport heavy, viscous petroleum liquids such as crude oil, residual fuels, bitumens and their emulsions. These pumps have a record of high efficiency and reliable operation. Valving and instrumentation used with them can range from the very simple to the complex, including computer controlled SCADA (Supervisory Control and Data Acquisition). The latter is commonly used in remote control of large, long pipeline operations, but it is beyond the scope of this article. Figure 1 shows the typical piping components at a pump station using rotary, positive displacement pumps. Gate valves A and H are used principally to isolate the pump for maintenance purposes. They are kept either fully open (normal pump operation) or fully closed (pump not running). Unlike the valves in a centrifugal pump system, these valves must never be used to throttle or restrict flow to or from a positive dis-

M

placement pump. Manually operated gate valves can partially close due to flow turbulence. A positive lock down method such as a chain is recommended for use in preventing accidental closures.

ISOLATING PUMPS FOR MAINTENANCE Like most everything, there is a right and wrong way to isolate pumps for maintenance. After shutting down the pump, the discharge isolation valve is closed first; then the inlet isolation valve is closed and the bleed or vent valve opened (know where any drained liquid will be going). The reason for this sequence is to prevent any discharge check valve leakage from pressurizing the inlet side of the system. If the inlet isolation valve is closed first, a very small amount of leakage past the check valve will bring the pump inlet, inlet piping and inlet instrumentation to system discharge pressure. Most pipeline screw pumps are designed to resist discharge pressure only on their discharge casing end, not their inlet side. Inlet strainers and gauges are typically not rated anywhere near system discharge pressure.

F •

J

E

K PUMP

A

B

G C

D

H

C

FIGURE 1. Valves installed in a rotary, positive displacement pumping station 48

The Pump Handbook Series

Strainer B can be a simple cone strainer, a simplex strainer or a duplex strainer with valving. An inlet strainer is a decidedly mixed blessing unless its pressure drop is monitored. The purpose of the strainer is to protect the pump from such debris as weld bead, weld rod, pipe scale and pig bristles, which can do serious damage to rotating machinery. However, if the strainer accumulates significant content, the pressure drop across it can reach a level that will cause it to collapse, enter the pump and probably cause catastrophic damage. Always instrument and alarm the strainer. It is difficult to make specific particle size removal recommendations. A strainer fine enough to keep out all solids would not be affordable, either to buy or to keep clean. On the other hand, having no strainer leaves equipment so vulnerable that inevitably something will come down the inlet piping and cause extensive damage. Photo 1 shows inlet strainers installed at the Jose’ marine export terminal in Venezuela. The fifteen 800 hp, parallel operating screw pumps each deliver 2900 gpm of Orimulsion, a boiler fuel composed of an emulsion of 70% bitumen and 30% water. The strainers provide pump protection from some 330 kilometers of upstream pipeline and storage facilities. They also assure the buyer that the product is reasonably clean. The strainer design should make periodic cleaning relatively easy to accomplish. Valves C are bleed valves used to depressurize and drain the system between the isolation valves in order to conduct maintenance. These

PHOTO COURTESY OF WARREN PUMPS

Photo 1. Inlet strainers (foreground) on 15 marine terminal screw pumps in Venezuela

would typically be one-inch ball or plug valves. Valve K, similar to C, is an air bleed to be located at a high point in the pump discharge system upstream of the check valve. The air bleed valve is used to allow air to be purged on initial pump startup after service. Rotary pumps are not necessarily good air compressors, and the bleed valve provides a low pressure vent for the air. It is closed after slowing essentially all liquid flow. Valve G is a check valve (nonreturn or one way valve). It prevents back flow of liquid through a nonrunning pump and allows multiple pumps to operate in parallel—the most common pipeline pump arrangement. Valve E is a pressure relief valve. It is normally set to open at a pressure just slightly above the maximum expected pump discharge pressure. It protects the pipeline, pump, pump driver and discharge system instrumentation from damage due to over pressure should the discharge isolation valve accidentally be left closed, or some other downstream blockage or valve closure occur. A relief valve is an important protection component. Positive displacement pumps should never be operated without one. The exhaust or low pressure side of the valve should not have any shutoff valves between it and the tanks or, if connected as shown by alternate path J, between it and the pump. A blockage in this line defeats the pur-

pose of the relief valve.

BYPASS FLOW CONTROL Valve F is a bypass flow control valve, frequently motorized. Its function is to allow pump flow to be bypassed either back to the tanks or back to the pump inlet piping. On pump startup, this valve is normally open. It is gradually closed, allowing controlled buildup of pump discharge pressure. For example, if one pump is operating at design discharge flow and a second similar parallel pump is started, the system discharge flow will double. To do so, the flow velocity must accelerate to a speed twice as high as it was before the second pump was started. The bypass avoids shock and allows this acceleration to take place in a gradual, controlled manner. Any flow bypassed either back to tankage or to the inlet side of the pump represents wasted pumping energy. The system should be designed so that bypass rates in excess of about 50% of pump rated capacity are avoided. Otherwise, the wasted energy recirculated will be converted to temperature rise in the pumped liquid. At high recirculation rates and high power levels, the temperature rise can be excessive and put pumps at risk of damage. If tankage is reasonably close, the preferred routing for relief and bypass valving is to the tank(s), which provide a greater heat sink. Pumping stations usually have The Pump Handbook Series

three 50% capacity pumps, two for operation and one used as an installed standby. Some stations have four 1/3 capacity pumps. This arrangement, together with the bypass control valves, provides reasonably good flow flexibility. Photo 2 shows a crude oil pumping station in Indonesia with 16 rotary screw pumps operating in parallel. Pumps were added over the years as oil production increased. All the pumps and motors are the same, as they are at nearby stations, providing full pump as well as repair part interchangeability. Where intermediate pumping stations are installed without buffer storage tanks (Figure 2), stations 2 and X, a bypass control valve is installed around the station. It senses station inlet pressure and bypasses whatever flow is necessary to maintain station inlet pressure at its set point, usually around 50 to 75 psig. This is essentially a station trim control and does not normally bypass large percentages of station flow. On decaying station inlet pressure, the valve opens, reducing the station throughput. When station inlet pressure rises, bypass flow is reduced. This system is normally tied to starting or stopping one or more of the station’s pumps if the valve flow range is not capable of maintaining station inlet set pressure. For clarity, the individual valves shown for each pump in Figure 1 were omitted from Figure 2.

PRESSURE GAUGES Regardless of how sophisticated the instrumentation and control system are, all pumps should have inlet and discharge pressure gauges for routine observation. Any inlet strainers should have differential pressure gauges indicating the condition of the strainers. Although this gauging provides information to on-site personnel, it does little to control operations or protect system equipment. It also requires the periodic presence of station attendants to read the gauges and take appropriate actions. Twenty-four hour per day coverage is thus advisable. The next level of control is that of pressure and temperature switches, with one or more set points at which electrical contacts open or close and initiate either alarms or equipment shutdown. These should

49

include a low pressure switch immediately before the pump inlet, a high temperature switch at the same location and possibly a high pressure switch in the pump inlet line if the upstream source of flow is not storage tanks but upstream pumps. Switch set points should be established not only to protect equipment, but also to prevent inappropriate switch actuation. Transient conditions at the moment of pump startup normally require that some switch actuation be ignored for several seconds after startup. For example, pump inlet pressure will momentarily drop to nearly a full vacuum while inlet fluid accelerates from zero velocity to full flow velocity. This is especially true with an electric motor drive and a typically low inertia screw pump. The low inlet pressure alarm and/or shutdown should be disabled briefly until full flow is established. Pump shaft seal leak detection switches (float, flow and pressure switches) should be considered, especially if pump operation is unattended. A shaft seal failure can spill a lot of liquid over several hours.

PHOTO COURTESY OF WARREN PUMPS

SWITCHING INSTRUMENTATION

Photo 2. An Indonesian crude oil pumping station using 16 rotary screw pumps operating in parallel

Switching instrumentation must be carefully thought-out and made compatible with all station operating modes. An operator’s ultimate solution to repetitive false alarms and shutdowns is bypassing or disabling the instruments. However, this leads to increased risk of personal injury, equipment damage and pipeline outage. Many users apply vibration and/

STATION INLET PRESSURE CONTROL VALVES

INITIATING TANK FARM

PIPELINE STATION NO. 1

PIPELINE STATION NO. 2

FIGURE 2. Pump station inlet pressure trimming with bypass control 50

The Pump Handbook Series

PIPELINE STATION NO. X

TERMINAL TANK FARM

or temperature detectors to pump and motor bearing housings as applicable. If the driver is not an electric motor (diesel, for example), additional driver-related parameters —such as cooling water temperature, lubrication pressure/flow/temperature and perhaps others—may need to be monitored. Some high power screw pumps may have a force cooled lubrication system that should be instrumented to shut down the pipeline pump should lubrication pressure, flow or temperature fall outside the range of normal operation. Consult equipment suppliers for recommendations that can enhance the operation or safety of their products. The highest level of instrumentation will incorporate transducers that provide an electrical signal that is proportional to the measured parameter. Pressure, flow rate, temperature, valve position and speed transducers are all readily available and can be linked to recorders for logging purposes, or tied into local or remote manual

or automatic control systems. They provide local and/or remote locations with continuous detailed data regarding pump and station operation. Integrated into an appropriate computer control system, these transducers can substantially automate and centrally control pipeline operation, providing reduced operating costs to pipeline owners and operators. The recommendations in this article are a general overview to help the system designer produce a good working installation. The presentation is not intended to cover every aspect of component selection, sizing, rating, noise suppression, shock pressure, temperature variations, corrosion allowances or other aspects of a fully and professionally engineered pumping station. Provisions must be made for compliance with national and local codes as well as owner/operator standards and specifications. ■

NC. His responsibilities include worldwide marketing and technical support for crude oil pumping applications. Brennan is a 1973 graduate of Drexel University in Philadelphia and a member of the Society of Petroleum Engineers. He has 26 years of service with Imo Industries.

James R. Brennan is Manager, Crude Oil Pumps, for three operating units of Imo Industries Inc., Monroe,

The Pump Handbook Series

51

POSITIVE DISPLACEMENT PUMPS HANDBOOK

Valve Dynamics Affect Recip Pump Reliability - Part 1 Valve design modifications help improve reciprocating pump reliability. By Stephen Price, Donald Smith, Jim Tison ompared to advanced technologies such as high-speed turbomachinery, reciprocating pumps may appear to be simple. The technology uses a reciprocating plunger to draw liquid from a low-pressure manifold through suction check valves and expell it into a high-pressure manifold through discharge check valves. However, interaction of flow, valve dynamics, and pump/piping system acoustics can generate high amplitude pressure pulsation. This causes severe vibration and reliability problems in some systems. We will examine the effects of improper valve operation and how such problems can be solved by modifying the valve design. Since the pump valves are operated by fluid action, valve components (including the springs, valve body, valve disc shape and sealing surfaces) must be selected carefully to ensure that the valves open and close at the proper times. Problems generally associated with valve dynamics include over-pressure spikes at the opening of discharge valves, under-pressure spikes at the opening of the suction valves, high noise levels and excessive valve component wear. Severe over-pressure spikes can cause fatigue failures of pump working barrels, connecting rods, bearings, and even drive-train components. Under-pressure spikes can cause cavitation, which can lead to plunger and valve failures due to pitting damage. High level valve noise usually indicates severe impacts associated with valve open-

C

52

ing and closing. Impacts at valve closing, known as valve hammer, can damage the sealing surfaces. Impacts at the valve opening, often identified by damage to the backside of the valve disc, are due to over-pressure and under-pressure spikes that can cause fatigue failures of the components. Excessive valve wear also can be experienced when the valve disc material is not suitable for the pumping conditions. We can analyze valve dynamics using field data and a computer model. These tools enable us to observe the effects of various valve modifications such as changes in sealing surfaces, valve lift, spring preload and spring stiffness. We will examine the instrumentation and data analysis techniques to evaluate these problems. In addition, we will look at the effects of valve modifications (including changes in lift, spring preload, spring construction and stiffness and valve disc geometries) using data from actual systems.

PUMP VALVE PROBLEMS Many problems with reciprocating pumps are caused by improper valve operation. In an optimum system, the suction and discharge valves open and close at the precise instant to facilitate the suction or discharge stroke. However, the plunger can change velocity faster than the valves can respond, resulting in “valve lag.” If a discharge valve lags on opening, this creates higher than normal pressures in the pump cylinder (working barrel) since the plunger continues to compress the liquid in the cylinder until the valve opens. When the valve finally opens, The Pump Handbook Series

the pressure in the cylinder is higher than the discharge line pressure. The rapid build-up and release of this pressure at the beginning of the discharge stroke is called an over-pressure spike. Similarly, if the suction valve lags on opening, the pressure in the cylinder is rapidly reduced until the suction valve opens. This rapid reduction and return of pressure at the beginning of the suction stroke is called an under-pressure spike. Cavitation can result if the pressure falls below the vapor pressure. As the suction pressure returns to normal, vapor bubbles formed during cavitation collapse, creating high amplitude “cavitation spikes.” Even with properly operating valves, lag of a few degrees is inevitable. However, spring preload, stiffness, valve disc mass, valve lift, and pump speed can significantly affect valve lag. Pressure-time data taken in the working barrel from a triplex pump that had an over-pressure spike problem are shown in Figure 1. The over-pressure spikes act on the plungers, and the resulting forces are transmitted to the rods, crankshaft, bearings and frame. The added dynamic pressure therefore increases the load-induced stresses on the power end and liquid end components, and it can contribute to fatigue failure of these parts. The over-pressure spike results from several combined effects including: • Viscous adhesion (“sticktion”) of the valves to the seat, the intensity of which depends upon sealing area, surface finish, disc flexibility and fluid properties. • Plunger-side/line-side dynamic pressures

•

Differential area (unbalanced valve area) • Acceleration of valve disc (due to changes in running speed) • Spring preload and stiffness • Valve mass Data obtained from numerous field measurements show that when significant over-pressure spikes occur, they are most often due primarily to the sticktion effect. In addition to causing potentially damaging forces and cavitation, this phenomenon creates a reduced pressure area near the center of the sealing surface. Here, cavitation may result on both the suction and discharge valves. Resulting cavitation pits on the discs and seats are usually concentrated near the center of the sealing surface. These pits are often mistaken for foreign object damage.

INSTRUMENTATION Several types of instruments are needed to quantify pump behavior. A typical installation on a vertical triplex pump is illustrated in Figure 2. Pressure Transducers (PT). Pressure-time data should be obtained in one or more pump cylinders (working barrels), in the suction and discharge manifolds, and at locations in the piping where high vibration levels occur. These data are usually obtained with dynamic pressure transducers. Sensitivities of 5-10 mV/psi are usually the best choices for measuring pulsation in the piping and manifolds. Since the change in cylinder pressure from suction to discharge pressure fluctuates over a wide range (typically 3000 psi), low sensitivity (1 mV/psi) transducers should be used for measurements in the cylinders. Static pressure measurements are sometimes needed to evaluate system cavitation, or to quantify certain process conditions that are important to pump operation. Straingage type transducers are used most often for this purpose. These require more elaborate signal conditioning and calibration than piezo-electric transducers. The static transducers also can be used for dynamic measurements if the conditioning amplifiers have sufficient frequency response. Accelerometers (XL). With

RECIPROCATING PUMP FIELD TESTS Field data are used as the primary diagnostic tools to analyze pump valve behavior. The data ultimately can be used as the basis for performing analytical calculations. However, the data must be of high quality to be useful. The usual technique for field data acquisition combines instrumentation, acquisition software and data analysis to deliver the desired characteristics. Each of these aspects is discussed in the following sections.

ENGINEERING DYNAMICS INCORPORATED

SAN ANTONIO TEXAS Pressure-Time History Original Valve Design Pump Pressures: Discharge = 2400 psig Suction = 400 psig

1

Cylinder Pulsation 1) Plunger #2 2) Plunger #3 2000 psi / Div

Over-pressure Spike (apx 1100 psi)

2 3)Microphone

Acceleration 20g / Div 4) Suction Manifold 5) Discharge Manifold

3 Noise ‘bursts” at Discharge Valve Opening

6) Timming Mark (reference Plunger #1)

4 Impacts at Discharge Valve Openings

5

0-0.5 sec (6000 pls / sec) 250 RPM 6000 pls/sec

6 0

0.05

0.1

0.15

0.2

0.25

0.3

0.38

0.4

0.45

0.5

FIGURE 1: Pressure time histories – triplex pump with over-pressure spike The Pump Handbook Series

appropriate signal conditioning, accelerometers can be used to measure low-frequency piping vibration. However, high-frequency energy should be measured, too. Cavitation, over-pressure spikes and other impact energies often result in highfrequency vibration in the pump. Accelerometers mounted on the pump manifolds provide acceleration data that can be used to identify problems with specific cylinders or valves. To obtain the required highfrequency response, the accelerometers should be mounted rigidly (i.e., not attached with magnets). Strain Gages. Where failures have occurred, it is sometimes useful to install strain gages to measure strain levels at the location of the failures. Although these instruments require a good deal of effort to install, the resulting data are usually worth it. Acquiring strain data on reciprocating or rotating components may require that the data be telemetered (although directly wiring gages to reciprocating parts has been somewhat successful, too). Strain gages attached to the plunger also can be used to infer cylinder pressures when it is not possible to install cylinder pressure transducers. Note, however, that the data can be distorted due to excitation of the plunger’s lateral natural frequencies, friction and inertial effects. Microphones. Some problems are discovered because workers hear strange noises emanating from the pump. Microphones installed in the near-field can be used to correlate perceived noise with other phenomena that might be occurring. Thermocouples/RTD’s. For problems involving cavitation, it is often important to measure the fluid temperature at the pump. Thermocouples and RTDs (resistance thermal devices) can be easily installed to provide this information. Timing Mark. Much of the acquired data is analyzed in the time domain. Thus, a timing mark is required to help identify specific portions of the pumping cycle and locate improperly operating valves. Magnetic pick-ups, optical pick-ups or proximity probes can be used to provide a once-per-revolution pulse. The pick-ups can be installed to sense a keyway or reflective tape on an exposed shaft. An optical sensor often is used to sense passage of a

53

reflective tape on the side rods. Tape Recorder. A multi-channel FM or digital recorder should be used to record data for later playback. This approach is invaluable for capturing data from a start-up, a shut-down and “trip” situations. The recorder should be adjusted to provide at least a 5 kHz bandwidth.

Location of Working Barrel Pressure Transducer

The data acquisition system should be able to collect data in both the time and frequency domains. Time domain data are used to evaluate pump operation (i.e., valve lag and pressure build-up). Frequency domain data are usually used to evaluate system resonances. Given the ability of portable computers to manipulate and store information, time domain data are acquired most efficiently using analog-to-digital (A/D) conversion hardware coupled with a computer. Most reciprocating pumps operate at speeds below 300 rpm, which would seem to indicate that only low-frequency data is required. However, the rapid pressure build-up and reactant flow bursts from each cylinder can generate energy to frequencies of 3000 Hz or more. Therefore, sampling rates as high as 6000 Hz per channel may be required. A typical test will usually involve 10–15 channels of data, which should all be acquired simultaneously for data comparison. Since most A/D boards multiplex their input, the sampling rate requirements for 15 channels of data could be as high as 90,000 Hz. Typically, data will be acquired for periods of 1 sec up to 1 minute. The acquisition system should be capable of manipulating as much as 10 Mbytes of data (15 channels at 6000 Hz sampling rate per channel, 1 minute of acquisition) per captured event.

DATA ANALYSIS Analyzing acquired data can be a time-consuming task. Acquisition software can ease the burden of manipulating large amounts of data, however. A system developed at Engineering Dynamics Incorporated has proven effective for acquiring the necessary data. The system capabilities include: Channel Management. A facility is provided to manage and display transducer sensitivities, descriptive text and channel numbers. Acquisition Control. This

54

Timing Mark

Strain Gage

DATA ACQUISITION

Optical Pickup XL

XL

XL

PT

XL RTV or Thermocouple

PT

PT

PT

Microphone Shaft

FIGURE 2: Typical reciprocating pump instrumentation

facilitates control of parameters such as sampling rates, styles and acquisition time. Acquisition Triggers. Acquisition of data is sometimes desired only when specific events occur. Typical events include a pulse from the timing mark, a peak-peak dynamic signal level or a certain static pressure. A pre-trigger (beginning acquisition before the trigger occurs) is provided to add flexibility to the triggering. Filtering. When the signals include unwanted “noise,” digital filter algorithms are used to remove it. Low-pass, high-pass, band-pass and band-stop filtering may all be required. Data Display. Some or all traces may be presented simultaneously on the “page.” The orientation and placement of each trace is easily adjustable. Capabilities for detailed documentation and annotation of data (to note specific test conditions, time and date) are also provided. Hardcopy can be obtained using various devices including both aster and vector devices. Data Storage/Retrieval. Acquired data may be stored for later recall. Facilities are available to import/export data from/to other software packages. This capability allows further manipulation of data (e.g., to compute a single channel of principal strain data from a rosette of three strain gages). The Pump Handbook Series

Next month we will show how the field data acquisition techniques described here have been used to address different types of pump and system problems, and we will present case histories of actual pump installations. ■ Stephen M. Price is a senior project engineer for Engineering Dynamics Incorporated, San Antonio, TX. He specializes in solving equipment failures traceable to dynamic phenomena. Donald R. Smith is a senior project leader at the same company. For the past 25 years he has been active in field engineering services, specializing in vibration and pulsation analysis and noise problems with rotating and reciprocating equipment. James D. Tison is a senior staff engineer at Engineering Dynamics and has been actively involved in field measurements and computer modeling of rotating and reciprocating equipment for more than 17 years. Editor’s Note: Reproduced with the permission of the Turbomachinery Laboratory from Proceedings of the 12th International Pump Users Symposium, Turbomachinery Laboratory, Texas A&M University, College Station, TX, pp 221–230, Copyright 1995.

POSITIVE DISPLACEMENT PUMPS HANDBOOK

Valve Dynamics Affect Recip Pump Reliability - Part 2 Using field data acquisition techniques, the authors show how modifying valve design improves pump reliability. By Stephen Price, Donald Smith and Jim Tison sing the field data acquisition techniques described in Part I of this article many types of pump and system problems can be addressed. In some cases it is necessary to utilize analytical capabilities to understand a problem fully. Case histories of actual pump installation are described in the following sections.

U

PUMP COMPONENT FAILURES DUE TO OVERPRESSURE SPIKES

manifolds, and with accelerometers mounted on the manifold. A nearfield microphone was also installed. Initial data showed that high amplitude (40-50%) over-pressure spikes were occurring at the beginning of the discharge stroke. Cavitation due to an under-pressure spike was also observed at the instant the suction valve opened. Sound measurements showed noise bursts at the instant of discharge valve opening. Accelerometers also showed impact energy that correlated with the opening of the

Several triplex pumps experienced drivetrain and A fluid-end component failures 16 Equally that, when the pumps were spaced grooves inspected, appeared to be the result of excessive loads. When the pumps were operA ating, high amplitude impact noises could be heard, as if internal components were knocking. Ground-borne vibration could be felt by people standing near the pump. Typically, valve life was short. After only a few hours of operation, the valve discs and/or seats became pitted as if foreign object damage had Seating Area occurred. Several times, the Seating Area valve disc or seat developed a series of pits at the center of the sealing surface, and these joined to form a ring that Hole appeared to have been machined into the surface. One of the triplex pumps was instrumented as depicted in Figure 2 of Part I (May 1995 Pumps and Systems) A-A with pressure transducers in the pump cylinders and FIGURE 1. Groove pattern to reduce sticktion. The Pump Handbook Series

discharge valves and sometimes with the opening of a suction valve. The over-pressure spikes generated impact forces. Knocking noises occurred as the impact force traveled through the power end. These forces were apparently responsible for the damage that had been experienced. Cavitation in the suction side of the pump resulted when the under-pressure spike caused the suction pressure to fall below the vapor pressure. It was thought that the delay in opening of the suction and discharge valves could be due to sticktion. As discussed earlier, sticktion is a phenomenon that produces a force that holds the valve disc onto the seat, retarding the opening of the valves. The magnitude of the sticktion force is a function of the width of the sealing surface of the valve disc. A wide seal will produce a much higher force than a narrow seal. Therefore, to reduce the sticktion force, the valve sealing surface area must be reduced. (Considering only the potential for sticktion, the ideal seal surface would be a knife edge.) However, if the valve seating surface area is too small, impact stresses in the valve disc and seat will be too high, resulting in valve damage. We determined that a groove pattern cut into the valve disc or seat can be used to reduce the sealing area, which can sometimes reduce the sticktion force without significantly affecting the

55

discharge into the working barrel at the beginning of the suc6.125″ 0.4375″ tion stroke. Similarly, Seating Area “A” suction valve lag causes backflow from the cylinder back into the suction as the discharge stroke begins. The suction and discharge valve lag reduce volumetric Seating efficiency–i.e., they Area “B” produce a decrease in flow. The pumps were operated for approximately one year with the valve seat modification. No additional 0.4375″ failures were report2.75″ ed. However, when 3.625″ the valves were inspected, the valve discs appeared to be Spring “forging” themselves Disc into the valve seat. Although the valves Seat were operating satisfactorily, it was felt that the effective groove pattern was reduced. Therefore, the discs were FIGURE 2. Typical pump valve disc and seat. replaced with new seating area. A typical groove pattern ones. Even though the one-year for a disc is shown in Figure 1. valve life was much improved over A set of valve seats was that in the original design, alternamachined with the groove pattern tive valve disc materials are currentshown in Figure 1 and installed in a ly being investigated to extend their pump. It was decided to machine the life. grooves into the valve seat rather USE OF VALVE DYNAMICS ANALYSIS than the disc because it was felt that such a modification would be more It is possible to reduce the permanent. When the pump was over/under-pressure spikes by cutstarted, knocking noises were no ting grooves in the valve. It is also longer present, and those standing possible to evaluate other valve near the pump indicated that the design parameters (e.g., fluid viscosiground-borne vibration could no ty, valve disc mass and spring rates) longer be felt. Data acquired in the to affect over/under-pressure spikes. pump cylinder showed that the overWhile the specific modifications pressure spike had been nearly elimmaybe arrived at by trial and error, it inated. Cavitation spikes in the can be more effective to use an anasuction side were reduced. Impact lytical tool (Valve Dynamics Analyenergy (acceleration) measured at sis) to evaluate the effects of the pump manifolds was also signifidifferent designs before implementcantly reduced. ing them. As will be demonstrated, Flow rates with the modified field data combined with analytical valve seats were 6.5% higher than analysis can be used to achieve betwith the original valve seats. This ter results in actual operation. initially surprising result was found A computer based dynamic simto be due to a decrease in valve lag ulation technique developed by from 16.4 to 11.7 degrees. Discharge Engineering Dynamics, Inc., has valve lag causes backflow from the been used to determine the sensitivi7.00″

56

The Pump Handbook Series

ty of the valve displacement-time history and the cylinder pressuretime history to the various pump and valve parameters. The valve motion and cylinder pressure are determined by numerical integration of the governing differential equation of motion. Valve dynamics simulation techniques have been presented previously in numerous sources. However, most of these models do not include the viscous adhesive force (sticktion) acting on valve due to velocity of the separation of the disc from the seat. Sticktion effects can be modeled by considering the case of two parallel surfaces immersed in a liquid. The viscous adhesive force is [1] µb3 de FS=–   L (y+eo)3 dt where: FS =sticktion force acting on seat µ = liquid absolute viscosity b = width of seat e = film thickness eo = initial film thickness y = displacement of valve disc from seat, and L = circumferential length of seat This effective sticktion force is due to the pressure profile created over the width of the seat as the fluid fills the void created by the separation velocity. This effect is also referred to as the Bernoulli effect. From this equation it can be seen that the valve seat width b has a strong influence on the sticktion force. Therefore, this dimension is an extremely important design parameter. The sticktion force is also proportional to the fluid viscosity, and the force is strongly dependent on the initial effective film thickness eo, which in turn is influenced by the surface finish of the disc and seat and the degree to which the surfaces are in intimate contact. The eo, parameter can change as the valve wears and with pressure differential. Different valve disc materials may conform more or less to the seat, changing the effective eo even though the design may be dimensionally identical.

Therefore, it is extremely difficult in practice to evaluate eo, accurately, making model normalization using measured data a pre-requisite to obtaining valid simulation results. Note also that since the separation velocity appears in the equation, it would be expected that higher pump speeds would result in higher sticktion-induced pressure over-spikes. Field data have indeed shown this to be the case.

COUPLING ANALYSIS WITH FIELD DATA A crude-oil pump system in operation for a short period of time experienced fatigue failures of small bore piping (vents and drains) attached to the main pipe. Due to the critical service of the pump, a field study was done to determine the cause(s) of the piping failures. Instrumentation was installed similar to that shown in Figure 2 (Part I, May Pumps and Systems). After the piping was repaired, strain gages were attached at the locations of the failures, and the pump was operated. The data indicated that the failures were the result of excessive vibration of the piping “stubs” at their structural natural frequencies (which were in excess of 200 Hz). This seemed unusual since under normal conditions little energy is generated by the pump at these frequencies. Further data acquisition and analysis showed that the energy exciting the resonant vibration was not from pulsation, but was an impact energy being mechanically transmitted from the pumps through the piping and support structures. The source of the high-frequency energy was determined to be impact forces generated by high amplitude over-pressure spikes in the cylinders. One obvious solution to eliminate the piping failures was to add braces to reduce the differential vibration between the vent/drains and the main piping; however, there were many similar fittings in the piping system, and these made it impractical to brace all of the fittings. Therefore, it was decided to reduce the over-pressure spikes by modifying the valves, a measure which, it was hoped, would also reduce the high-frequency vibration of the stub piping. Initially, the groove pattern shown in Figure 1 was cut into the

valve discs. When the pump was operated, reductions in over/underpressure spike amplitudes were not significant. Rather than proceed on a trial and error basis, it was decided to utilize valve dynamic analysis to determine a groove pattern that would be effective. Numerous valve designs were analyzed. A sketch of the valve disc is shown in Figure 2. The pump characteristics are outlined in Table 1. The results of the computer analysis were presented in time domain graphs of the plunger pressure, suction valve displacement and velocity, and discharge valve displacement and velocity. The simulation results at 300 RPM predict an overpressure spike of approximately 285 psi (1000 psi spike minus 715 psi static discharge pressure). Under-pressure spikes of 175 psi caused the pressure in the cylinder to fall to the vapor pressure, initiating cavitation spikes. The valve displacement graphs indicate that the valve discs would impact the valve stops during opening. This agreed with the damage that had been experienced on the back side of the valve discs and the broken springs that occurred after a short operating time. For reduced speed operation of 150 RPM, the peak pressure was reduced to approximately 825 psia, which is equivalent to an over-pressure of 175 psi. The under-pressure spike was computed to be 70 psi above vapor pressure. Field data obtained for these two cases correlated well with the simulation results. To reduce the over-pressure spike, it was proposed that the discharge valve discs be modified as shown in Figure 3. To reduce the under-pressure spike, the suction valves were similarly modified. The concept of the modification was to reduce the width b of the sealing surface. This parameter directly influences the sticktion force by its value cubed. Both the over-pressure and under-pressure spikes were significantly reduced at 300 RPM. In addition, the valve vertical displacement (lift) was no longer hard against the stop. This would reduce the impact force against the valvestop and reduce the damage to the springs. Similar improvements were obtained at 150 RPM. The Pump Handbook Series

Field data acquired after the valves were installed showed that in the pressure-time histories, the amplitudes of the over-pressure and under-pressure spikes were significantly reduced. Strain gage data also showed a significant decrease in strain amplitudes at the stub piping. A summary of the measured and computed over/under-pressure spike amplitudes at 300 RPM is shown in Table 2. Other designs analyzed showed a further reduction in over-pressure spike amplitude. Although the computer analyses indicated that additional improvement could be obtained with the other groove patterns, these designs were not tested because there was concern that the remaining seat area was insufficient to provide adequate load bearing for the pressures experienced. Additionally, analytical and field data showed that at some point further reductions in seat area did not cause further reductions in over-pressure spike amplitudes, indicating that other effects were predominating the over-pressure spike generation. Therefore, another type of valve would probably need to be considered to provide further improvements.

OPERATIONAL AND DESIGN PARAMETERS AFFECTING STICKTION Other pumps were tested that had similar over/under-pressure spike problems due to sticktion forces. As has been shown, cutting grooves into the valve disc or seat will not always eliminate the spikes. Several tests showed that pump operational parameters and valve design parameters other than seat width could also affect the sticktion forces.

VALVE DISC MATERIAL A pump with flat, dual-ported steel valve discs was experiencing over-pressure spikes of approximately 50% of discharge pressure. For reasons unrelated to the over-pressure spike, the user decided to change valve disc material to PEEK plastic. No changes were made in the valve disc or seat designs. With the PEEK valve discs, over-pressure spikes were reduced to 16% of developed pressure. Cutting grooves into the valve disc further reduced overpressure spike amplitudes to 11%,

57

which was not as dramatic a reduction as had been previously experienced at some other installations. Computer analysis showed that the reduction in mass of the PEEK valve disc could cause a reduction in over-pressure spikes, but by an amount smaller than what was actually experienced. One possibility for the discrepancy is that the valve disc was probably not in complete contact with the seat due to the increased flexibility of the PEEK material (Figure 4). The imperfect contact would have effectively reduced the sealing area, which in turn reduced the sticktion effect. This phenomenon would also account for the grooves being less effective.

Parameter Operating Speed Liquid Density Liquid Viscosity Stroke Bore Suction Pressure Discharge Pressure Suction and Discharge Valve Disc Weights Suction and Discharge Spring Constant Suction and Discharge Valve Lift Valve Spring Preload Valve Disc Material

30 µreyns 7.0 in 4.75 in 230 PSIA 715 PSIA 1.7 lbs 112 lb/in 0.46 in 54 lbs Delrin

TABLE 1 Original Design Data

PUMP SPEED Theory predicts that the valve disc velocity also affects the magnitude of the sticktion force. Since the velocity of the disc is directly related to pump speed, the sticktion forces and resulting over/underpressure spikes should be higher. When data were acquired on variable speed pumps, this was indeed what was found. For example, the over-pressure spike amplitudes were increased from 20% to 45% of discharge pressure when the pump speed was increased from 150 to 300 RPM.

Value 150 & 300 RPM 61 lb/ft3

Over-pressure

Modified Design

Under-pressure

Over-pressure

Under-pressure

Field

295 psi

190 psi

120 psi

100 psi

Analytical

285 psi

175 psi

96 psi

88 psi

TABLE 2

A 16 Equally Spaced Grooves

A

FLUID DIFFERENCES It was found that virtually identical pump valve designs operating with similar pressures and speeds but different fluids could have dramatically different sticktion characteristics. Field data have shown that sticktion forces are greater for fluids having higher viscosity and molecular cohesion. For instance, a pump that operated with water experienced much less over-pressure than a similar design operating with Amine (a substance somewhat like automotive anti-freeze).

Seating Area Hole

A-A

CONCLUSIONS

FIGURE 3. Double-row groove pattern.

Extensive field testing of reciprocating pumps that have experienced failures in the working barrels, valve and piping have shown that the valve behavior strongly influences the vibration and resulting failures. Over-pressure spikes have previously been attributed to area ratio, valve disc mass

and preload. Perhaps a more important factor is sticktion (the Bernoulli effect), which is primarily related to seat area. Fluid effects such as viscosity and molecular cohesion, as well as pump speed, valve disc and seat surface finish and valve disc stiffness also affect sticktion, but to a lesser degree.

58

The Pump Handbook Series

Sticktion delays the valve opening, which results in over-pressure spikes, valve disc impacts at valve opening and localized damage (cavitation pits) to valve seats. High amplitude impact noises are often an identifying characteristic of overpressure spike problems. These over-pressure spikes can also cause

enth Edition, Mc-Graw Hill Book Company, New York, 1979. 2. Bauer, Friedrich, The Influence of Liquids on Compressor Valves, 1990 International Compressor Engineering Conference at Purdue University, West Lafayette, Indiana, July, 1990.

Disc Valve Seat

Flat Seating Surface A. Idealized flat seating surfaces are more sensitive to sticktion.

Disc Valve Seat

“Knife Edge” Seat B. Large deflection with thin disc reduces sticktion.

lems. This information can be used alone or combined with valve dynamical analyses to yield solutions to problems. The groove pattern on the valve discs and seats has been shown to be effective in reducing the sticktion effects and associated problems. ■

FIGURE 4. Imperfect valve seating due to disc flexibility.

excessive loads to be transmitted to pump components, and these can result in drivetrain component failures and working barrel failures. In some cases the high frequency mechanical impact energy was shown to be structurally transmitted throughout the pump and piping system, exciting structural resonances that ultimately caused piping failures. In the design phase it is useful to perform detailed valve dynamic analyses to assist with valve geometry selection. When problems occur, use of instrumentation and data analysis hardware/software will lead to an understanding of pump prob-

ACKNOWLEDGMENTS The authors wish to thank the staff of Engineering Dynamics, Inc., and especially Mark Broom for his assistance in producing many of the figures presented in this article.

Stephen M. Price is a senior project engineer for Engineering Dynamics, Inc., San Antonio, TX. He specializes in solving equipment failures traceable to dynamic phenomena. Donald R. Smith is a senior project leader at the same company. For the past 25 years he has been active in field engineering services, specializing in vibration and pulsation analysis and noise problems with rotating and reciprocating equipment. James D. Tison is a senior staff engineer at Engineering Dynamics and has been actively involved in field measurements and computer modeling of rotating and reciprocating equipment for more than 17 years. Editor’s Note: Reproduced with the permission of the Turbomachinery Laboratory from Proceedings of the 12th International Pump Users Symposium, Turbomachinery Laboratory, Texas A&M University, College Station, TX, pp 221-230, Copyright 1995.

REFERENCES 1. Streeter, Victor L., and Wylie, E. Benjamin, Fluid Mechanics, Sev-

The Pump Handbook Series

59

POSITIVE DISPLACEMENT PUMPS HANDBOOK

Reducing Pulsations in Metering Pumps Zero leakage is one advantage of using diaphragm style metering pumps for hazardous fluids, but they produce pulsating flow and thus pulsating pressure. By Ed Warwick etering pumps are frequently used to move hazardous liquids. Accordingly, the system designer and user should be aware of some important characteristics of metering pumps that can help ensure successful operation. Metering pumps and accessory equipment should be selected to avoid effects of slug feeding, inaccurate flow meter indications and “water hammer” pressure pulses that can cause pipe stress, joint leakage, significant unexpected overpressures and overfeeding and underfeeding of chemical. Various types of pumps are used to meter liquids, including vane pumps, progressing cavity pumps, roller/tube pumps and reciprocating plunger pumps. This article focuses on the reciprocating plunger type, which includes plunger actuated diaphragm designs. These are used to pump virtually any liquid. Some of the more common are: • Hydrofluorosilisic acid • Hydrochloric acid • Sulfuric acid • Sodium hydroxide (caustic) • Sodium hypochlorite • Anhydrous ammonia • Pesticides • Corrosion inhibitors.

M

principle of operation, consider Figure 1. The drive end usually consists of an electric motor, gears and a mechanism attached to the plunger to make it move back and forth (reciprocate) with a stroke length ranging from a fraction of an inch (a few millimeters) to more than 4″ (100 mm). Typical stroking speeds range from 10–200 strokes per minute. The liquid end of the metering pump directly contacts the liquid being pumped. Its valving is almost always ball style check valves. The check valve consists of a ball and a ball seat as shown in Figure 3. The ball rises up off the seat to allow liquid flow in one direction, and then sits on the seat to block (check) flow back through it. As the drive end draws the plunger back, liquid cannot flow into the liquid end from the discharge line because the discharge check valve seals against its seat, preventing flow from that direction. The suction ball can rise, and liquid from the chemical supply tank flows through the suction line, through the

suction check valve and into the liquid end. After reaching its backdead-center position, the piston begins a forward, or discharge, stroke. It pushes liquid in front of it, and now the suction ball returns and seals against its seat, preventing flow from going back into the suction line. The flow pushes past the discharge ball, forcing it off its seat, and continues into the discharge pipe. Note that during the suction stroke there is no flow in the discharge line, and that during the discharge stroke there is no flow in the suction line. The sequence repeats during every stroke of the pump, resulting in start-stop, start-stop flow alternating between the suction pipe and discharge pipe. The liquid end shown in Figure 1 to illustrate this principle of pumping is of the packed plunger variety. It uses packing material similar to valve stem packing to seal the chemical along the moving plunger. By its nature, the packing will leak chemical. So this type of liquid end is seldom used to pump dangerous

Chemical Supply Tank

Drive End

METERING PUMP CONFIGURATION Reciprocating plunger pumps are used in process chemical industries, water treatment, wastewater treatment, and petroleum recovery and refinement. Flow rates range from less than 1⁄10 of a gph (a couple of cc’s per minute) to more than 3000 gph (about 10M3 per hour), at pressures from near atmospheric to 10,000 psi (700 kg/cm2). To understand a metering pump’s

60

Liquid End

FIGURE 1. Metering pump drive end and liquid end The Pump Handbook Series

Ball Chemical Supply Tank

Flexible Diaphragm

Seat FIGURE 3. Check valve ball and seat arrangement

FIGURE 2. Flexible diaphragm style liquid end

chemicals. Instead, a diaphragm style liquid end is used. It has the distinct advantage of having zero chemical leakage to the “outside world,” and is the preferred style to use when pumping dangerous chemicals.

DIAPHRAGM METERING PUMP Like the packed plunger design, it uses suction and discharge check valves, but it places a flexible disc (usually of Teflon or Teflon-faced rubber) between the chemical and the plunger. Figure 2 shows one version of the diaphragm style liquid end that has the diaphragm physically attached to the reciprocating plunger. The front side of the diaphragm contacts the pumped chemical; the plunger side is vented to the atmosphere. As the plunger moves back and forth, it pushes and pulls the diaphragm, creating pumping action just like that of the packed plunger. This design is usually used at discharge pressure under 150 psi (10 kg/cm2). The forces created by chemical pressure on one side of the diaphragm, and the plunger on the other side, create stress on the diaphragm at the point of attachment, thereby increasing the risk of a diaphragm tear with higher pressure. A design was developed to allow the diaphragm to operate with pres-

sure on both sides in excess of 3000 psi (200 kg/cm2). In this design, shown in Figure 4, the piston is not attached to the diaphragm, and the space between the diaphragm and plunger is sealed and filled with hydraulic oil. As the plunger moves, it displaces oil, which moves the diaphragm, which displaces an equal

amount of chemical. Pressure is distributed evenly across both sides of the diaphragm, providing long life at both low and high pressure. This is accomplished with no moving seals in contact with the chemical and therefore no chemical leakage to the environment. Alternating pulses of flow in the

Chemical Supply Tank

Oil Oil

Oil Oil

FIGURE 4. Diaphragm filled and sealed with hydraulic oil The Pump Handbook Series

61

suction and discharge pipes still occur, just as in the packed plunger liquid end example. Figure 5 is a graph of the flow at any given instant in the pipeline. Note that the peak flow in the pipe for every stroke is greater than the average flow. For example, if the average flow in the pipe (the flow delivered to the process) is 100 gph, then the peak flow that the pipeline must carry each stroke of the pump is π times the average flow, or 314 gph! The pressure loss (frictional pressure drop) in the pipeline, therefore, should be calculated with a flow of π × average flow, and usually (depending on which pressure drop equation is used) results in a frictional pressure drop through the pipeline valves, fittings and straight pipe of more than 8 times the drop from equal but steady, nonpulsating flow from a centrifugal pump. (Note: for metering pumps, peak flow in the pipeline is more than 3 times the average flow. Peak pressure drop is more than 16 times that of a centrifugal pump for the same flow rate!) Remember that this pressure drop, or pulse pressure, adds to the average pressure in the discharge line and subtracts from the average pressure in the suction line. Higher frictional pressure than expected in the discharge line can cause: • Pump motor overload • Pressure gauge damage • Pipe joint leakage • Pipe failure.

PULSE PREDICTIONS Higher frictional pressure drop than expected in the suction line causes NPSH (starved pump) problems. The pulse pressure caused by pulse flow friction can be predicted either by substituting the value of π × flow rate for the flow rate used in the steady flow pressure drop equations, or by using p=14.3 x 10–6QµLp  d p4 where: p = Friction pressure drop, psi Q = Average pump flow rate in pipeline, gallons/hour µ = Chemical viscosity, centipoise

62

Peak Instantaneous Flow Rate

Instantaneous Flow Rate Average Flow

Discharge

Suction

Discharge

FIGURE 5. Graph of pipeline flow

Lp =Pipeline total equivalent length, feet dp =Pipeline inside diameter, inch Because of the pulsating nature of flow in the pipelines, there is a source of pressure pulses, in addition to the frictional pressure, that is usually much greater than the friction. It has various names: “acceleration” pressure, “inertia” pressure or sometimes just “pulse” pressure. Regardless of what it is called, it is caused by the chemical in the pipe starting and stopping (pulsating flow) with every stroke of the pump. Recall that when the pump is on a suction stroke, liquid in the discharge line is at rest— not flowing. When the pump now begins a discharge stroke, it must overcome the at-rest inertia of the entire mass of liquid in the discharge line and get it moving. It takes extra pressure at the pump to do this. For example, a discharge line (or a suction line) that is 1⁄2″ diameter and 300’ long, contains about 50 lbs if it is holding sulfuric acid. A pump that operates at 120 strokes per minute might well be selected for this service. This is one complete stroke every 1⁄2 second, or only 1⁄4 second to perform a discharge stroke. Imagine the force the pump must exert to accelerate that 50 lb mass of liquid from rest to a velocity—in less than 1⁄4 second—and continue doing this at a rate of 120 times The Pump Handbook Series

per minute. In fact, the pump must generate more than 900 psi at every stroke just to pump sulfuric acid at 5 gpm through a 1⁄2″ by 300’ long pipe. If the end of this 1⁄2″ pipe discharges into a 50 psi reactor or water line, the metering pump, gauges, piping and pressure relief valves (most likely set at 75–100 psi) will be subject to 900 + 50 psi. If this is a 1⁄2″ x 300 suction line, then the pump will have to “pull” on the liquid in the pipe with a “vacuum” of 900 psi— which is impossible. Thus, the liquid will cavitate, the pump will appear to have a “starved” suction, and only a fraction of the liquid will be pumped because of inadequate NPSH. This acceleration, or inertia pressure, can be predicted with: p=NQ(SG)Lp  27,700 dp2 where: p = Acceleration pulse pressure, psi N = Stroke speed of pump, strokes per minute Q = Average flow rate, gallons/hour SG= Specific gravity of chemical - no units Lp = Physical (not the equivalent) length of pipe, in feet dp = Inside diameter of pipe, inches

Remember, this equation applies to both the suction line and the discharge line. This acceleration pressure is usually responsible for what is called “water hammer.” Depending on the degree of piping support and restraint, an acceleration pressure of 5 psi can cause the pipe to move at every stroke of the pump.

MINIMIZING PULSATIONS Fortunately, good options are available for reducing the effects of pulsating flow. The two most obvious ones can be identified from the equations. Shorten the pipe: make “Lp” smaller. Locate the pump closer to the tank. Increase the pipe diameter. This is very effective. Because of the Dp4 effect, doubling the pipe ID will reduce the frictional pressure drop by a factor of 16. Acceleration pressure is reduced by a factor of 4 when the pipe ID (dp4) is doubled. Other approaches include using a pulsation dampener of the type that separates the chemical from the gas pad with a bladder or diaphragm. Figure 6 shows a pump system outfitted with full pulse dampening.The pulse dampener is most effective in the discharge line; it eliminates pulsations and their effects. It should have a volume (size) greater than the volume of 15 pump strokes. Locate it within a few feet of the pump. Pressurize it with an inert gas (usually nitrogen) to 75–85% of the final expected operating pressure. A simple standpipe in the suction line will often solve the NPSH problems. It should have a diameter twice that of the pump plunger or more. It won’t work with suction lift. Locate it within a few feet of the pump. The top must be a few feet higher than the highest level of chemical in the tank. Also, it should be vented back to the air space in the storage tank to protect against spillage and allow it to breathe in operation. Another strategy to reduce or eliminate pulsations is to use a pump that drives two or three plungers from the same motor. If the pistons are properly synchronized out of phase, Figure 7 indicates that pulsation effects are reduced by at least 1⁄2 if a duplex (two head) pump is used, and are reduced to a negligible

Pulsation Dampener Vent Line

Pump

Chemical Tank

Standpipe FIGURE 6. Pump system with full pulse dampening

INSTANTANEOUS FLOWRATE Pumps with equal total flowrate Average Flowrate

Simplex

Duplex

Triplex

FIGURE 7. Pump pulsations reduced by duplex and triplex heads

amount if a triplex (three head) pump is used.

OVERVIEW AND STRATEGIES Diaphragm style metering pumps can be very effective in pumping dangerous liquids with no environmental leakage. They do produce pulsating flow, which in turn creates pulsating pressure. Pulsating flow generates 8 × the frictional pressure drop of steady flow, and it causes an acceleration pressure not found in steady flow from centrifugal pumps. These effects can cause unexpected overpressure in discharge lines and underpressure (NPSH) problems in suction lines. Equations are available to predict the pulse pressure of both types—frictional and acceleration. Strategies to reduce or eliminate effects of pulsating flow include: The Pump Handbook Series

• Shorten the pipe length • Increase the pipe diameter • Use a pulsation dampener in the discharge line • Use a standpipe in the suction line • Use a multiplex pump configuration. Metering pump manufacturers recognize that there are unique considerations in applying their products, so they encourage the designer or user to discuss specific applications. ■ Ed Warwick is a graduate mechanical engineer from the University of Florida. He has worked in various design and applications engineering positions for Milton Roy Company since 1965.

63

POSITIVE DISPLACEMENT PUMPS HANDBOOK

Maintaining and Operating Positive Displacement Rotary Gear Pumps he gear pump is one of the oldest pumps in existance—with drawings of them dating back to the 18th century still in existance. It is also one of the most common pumps, with tens of thousands in use today. A gear pump traps fluid between the gear teeth, pushes it around the casing, and discharges it against the system back pressure. It delivers a smooth, pulse-free flow directly proportional to speed and only marginally affected by discharge head. It is very compact, with the ability to efficiently handle a wide viscosity range and develop high suction lifts.

T

BASIC GEAR PUMP TYPES Gear pumps have been used for many years in metering, unloading, transfer and blending services. A better understanding of them would be beneficial to specifiers and users alike. There are two types of gear pumps; internal (gear within a gear), or external (two gears, OD’s meshing). Internal gear pumps employ an overhung shaft design with the driven larger gear (rotor) contacting and driving the smaller inner (idler) gear. A pinion and bushing attached to the front cover position the idler gear. External gear pumps employ two equal-sized gears meshing about their outside diameters. Bearings support the shaft ends with the gears located between them. Pumpage circulates through the bearings by venting the outside end of each to the suction side of the pump. The packing (or mechanical) seal vents in a similar way.

INSTALLING AND OPERATING A GEAR PUMP The precautions and practices that apply to gear pumps are the same as those found with other pump types. If the pump is not installed immediately, store it in a warm, dry place. If not done by the manufacturer, spray all internal parts with a light oil to inhibit rust and provide lubrication during initial start-up.

64

Rotate by hand all pumps in storage to keep the oil coating evenly spread among the gear teeth and seal faces. Suitable foundations are mandatory for all pump baseplates, and gear pumps are no exception. They should be rigid enough to support the pump, motor and baseplate without vibrating. To mount the pump correctly, all mounting surfaces must be flat and parallel. Correct coupling alignment is also essential. A flexible coupling will not compensate for a misaligned pump and motor. Routinely check alignment for compliance with manufacturer’s specifications. New installations must be checked because assemblies aligned at the factory will misalign during the journey from their facility to yours. When piping to the pump remember that the pump is a pump— not a pipe anchor. Never use the pump to support the weight of the suction or discharge piping. Doing so can distort the casing, altering the internal clearances and causing accelerated wear of the internal components. Suction piping should be at least one diameter from the pump suction port. It should be as short and straight as possible, to decrease suction piping losses. When in doubt, use a suction strainer to protect against foreign matter. Despite claims to the contrary, there is no such thing as a perfectly clean system. Check all pumps for correct direction of rotation before starting them up. Unlike a centrifugal pump, which will only work in one direction, the gear pump is bidirectional. If rotated in the reverse direction, it will pump that way regardless of the consequences. Because it is a positive displacement pump, the gear pump will pump against any back pressure imposed upon it. Operating against a closed discharge valve, it will continue building pressure until it damages itself or overloads the motor. A pressure relief The Pump Handbook Series

valve must be in the system. Most gear pumps are available with integral relief valves. While these safety devices are a wise investment, they are not a substitute for external system relief valves. Like the parking brake on an automobile, they are a secondary safety device of last resort and not intended for continuous duty. While this may all seem painfully obvious, it would be impossible to list all the “pump” problems encountered that were really due to something else. Improper foundations, shoddy baseplate work, flawed pump mountings, and misaligned couplings are common

Figure 1. Basic internal gear pump

Figure 2. Basic external gear pump

Internal Wear and Flow Output Classic theory tells us that flow between two plates follows the general equation: Qs = K (∆P) (w) (d)3 / (υ x l) where Qs = slip through the clearance K = a constant ∆P= differential pressure w = width of clearance d = clearance υ = absolute viscosity l = length of clearance Assume we have a pump with gear tooth length (l) and width (w). As long as the viscosity (υ) and differential pressure (∆P) are constant, flow slip is only affected by the cube of the clearances (d3). A typical plot of output flow as a function of wear is shown below. Up to a certain critical point wear doesn’t affect flow that much. Beyond this point, which varies from case to case, performance deteriorates rapidly and a complete overhaul may be necessary to bring the pump back to original specifications.

25

15 10 Critical Point

0.012

0.010

0.008

0.006

0.004

0

0.002

5 0

Outputflow (GPM)

20

Figure 3. Gear pump with integral relief valve

examples of external problems typically blamed on the pump.

SPECIAL GEAR PUMP CONSIDERATIONS With either internal or external design, each shaft revolution displaces a finite amount of fluid. In actual operation, clearances must exist between the pumping elements to lubricate the rotating parts. A small percentage of fluid, known as “slip,” will leak back through the clearances from the higher discharge pressure to lower suction pressure. The slip paths are the space between the gear teeth and casing wall, the sides of the gears and pump casing, and the meshing gear teeth themselves. The gear pump relies on close running clearances to seal between the suction and discharge pressures. This requires a trade-off to handle a wide viscosity range. The clearances must be tight enough to yield an acceptable volumetric efficiency with thin fluids, yet large enough to facilitate internal lubrication with higher viscosity fluids. A drop off in output flow is the first sign of internal wear in a gear pump. With a centrifugal pump the impeller wear will affect flow and discharge head together, and can shift the point where the pump performance curve and system curve interact. With the gear pump, flow is directly proportional to speed— and only marginally affected by discharge head. So, unless one of these changes significantly, a decrease in flow is a reliable sign of internal wear. To verify this, compare current to original gear and casing dimensions and replace any worn The Pump Handbook Series

Figure 4. Internal slip paths in a gear pump are between: A. The sides of the gears and the pump casing B. The ends of the gear teeth and pump casing C. The meshing gear teeth themselves

components. Because the gear pump provides a smooth, pulse-free flow and operates at modest speeds, noise and vibration are rarely a problem. If a pump vibrates and the problem is not cavitation, it may be a bad bearing or an out-of-balance gear or shaft due to a large solid passing through the pump. In either case, inspect the pump and replace faulty parts as necessary.

SUMMARY The rotary gear pump is a versatile solution to a wide range of pumping applications. Understanding its characteristics will help you get the most out of the pump. Use a suction strainer and only operate the pump at the selected speed. Never throttle the pump suction or discharge side, and only operate it in with an external relief valve in place. Users following these simple guidelines report the gear pump to be one of the simplest, most versatile and dependable pumps in operation today. For a pump that’s been around as long as this one, that’s saying something! ■

ABOUT THE AUTHOR: Robert A. Platt is Sales Manager, Rotary Pumps for Ingersoll-Dresser, Chesapeake, VA. He is a registered professional engineer with 16 years of experience in positive displacement pumps.

65

Speed Considerations Speed is important when sizing a gear pump. Since flow is directly proportional to speed, it can be tempting to increase the pump speed to obtain more flow. Do not do this without consulting with the manufacturer. Besides overloading the driver and increasing the suction piping losses, consider one other item when pumping viscous fluids. As the gears rotate, the gear teeth present a void for the incoming fluid to fill. This void is available for a fixed amount of time, and the fluid in the suction chamber must fill this void in that available time.

In the case of viscous fluids in pumps running at high speeds, if the fluid cannot fill the void in the time available, a partial vaporization will occur. The vapor will carry through the casing and condense back to liquid once it sees discharge pressure. The result is cavitation, though not in the usual way, along with a loss in flow and increase in noise and vibration. To safeguard against this, manufacturers limit each pump’s maximum speed as a function of viscosity—all else being equal. Following these guidelines will extend the pump’s life and reduce the likelihood of operating problems.

Drive Gear Space to be filled with liquid

Inlet

Outlet

Figure 1. Liquid must completely fill the space between the gear teeth to avoid cavitation.

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The Pump Handbook Series

POSITIVE DISPLACEMENT PUMPS HANDBOOK

Peristaltic Pumps Offer Custom Fluid Solutions Peristaltic pump selection tips help solve fluid transfer problems. By Larry Van Bogaert he most important advantage of peristaltic pumps is their use of tubing as the pump chamber. The fluid is inside the tubing and does not contaminate the pump, nor does the tubing contaminate the fluid. Cleanup requires only a change of tubing. This saves costly breakdown and maintenance time. Peristaltic pumps transfer fluids successfully in industries such as food processing, pharmaceutical manufacturing and chemical processing, as well as in laboratory research, agriculture and water treatment. Specific applications include dispensing, sampling, metering, filtering, fermentation (Photo 1) and general fluid transfer. In general, if the fluid flows, a peristaltic pump can pump it. In many situations, peristaltic pumps are ideal for pumping abrasive as well as aggressive fluids. The key to a successful peristaltic pump application is to optimize the selection of pump head, tubing and drive. Application requirements thus dictate the features of pump products being selected for purchase. Many different types and shapes of pump heads, tubing materials and tubing sizes are available to meet specialized needs. Each different combination provides a special fluid transfer capability.

T

As a roller passes over the tubing, it is first occluded (squeezed), then released. The progression of this squeezed area forces fluid to move in front of the roller. The tubing behind the rollers recovers its shape, creates a vacuum, drawing fluid in behind it. As the roller moves faster, vacuum pockets are created more quickly and the fluid

moving through the system picks up speed. The rollers act as check valves to prevent siphoning or loss of prime. The distance between the rollers creates a “pillow” of fluid. This volume is specific to the ID of the tubing and the geometry of the rotor. Flow rate is determined by multiplying pump head speed by

PERISTALTIC PRINCIPLE Peristaltic pumps operate on a simple principle. The alternating pattern of squeezing and releasing the tubing moves the fluid through the pump.

PHOTO 1. Peristaltic pumps are popular in fermentation systems. The Pump Handbook Series

67

the size of the pillow by the number of pillows per revolution. The pillow volume stays very constant except with highly viscous fluids. Among pumps with the same diameter of rotor, pumps with large pillows will deliver higher volumes of fluid per revolution. A greater degree of pulsation exists with these higher flow rates, not unlike the pumping profile of a diaphragm pump. Pumps with small pillows deliver small volumes of fluid per revolution. With many of these small pillows developing in rapid succession, the motion of the fluid appears smoother, similar to that seen in gear pumps.

ADVANTAGES AND LIMITATIONS Isolation of the fluid within the tubing, flexibility in quick tubing changes, reversibility of flow and ability to run dry are the primary competitive features of the peristaltic pump. See Table 1 for a complete listing of feature benefits. Some potential limitations are service life, chemical compatibility and pressure. Consider these limits when evaluating tubing options.

MAJOR MARKETS AND APPLICATIONS Laboratory Research and Development. Tubing pumps offer excellent repeatability in low volume dispensing and metering applications. The valveless design eliminates clogging and siphoning of fluid under most conditions. The modular design of the system allows one pump drive to be used for many 1. 2. 3. 4. 5. 6. 7. 8. 9. 10. 11. 12. 13. 14. 15. 16.

applications. Common laboratory research and development applications include cell tissue transfer, destaining, perfusion, liquid chromatography and transfer of acid and base solutions. Pharmaceutical Manufacturing. The non-contaminating and non-invasive design of tubing pumps makes them very popular. Sterilizing the USP class VI tubing is easy. A wide variety of drive options fit the pump into many different applications. Pumping nutrients or pH adjusters for fermentation, filtration of media and dispensing of cosmetics are just a few of the many applications for this market. General Industry. Peristaltic pumps offer predictable service in continuous-duty applications while solving many tough pumping problems. These pumps will handle waste water, suspended solids, harsh chemicals and other challenging fluids up to 45 L/min. (12 gpm). Many drive designs handle rugged plant environments. The self-priming, run dry capabilities of the pump prevent catastrophic failures in many process systems. A consistent service schedule prevents pump downtime. Some common applications include the pumping of dyes, etching chemicals, printing inks, laundry chemicals, lapping fluids and lubricating oils. Commercial Food Processing. A wide range of tubing materials is available to meet USP, FDA, NSF and 3A requirements. Tubing

Fluid does not contact any part of the pump except the tubing. No seals to leak. No valves to clog or wear. Self priming (up to 30' or 8.8m in some models). Pumps liquids, gases, solids or mixed phases. Can use one continuous piece of tubing from inlet to outlet. Some tubing materials can be easily sterilized. Easily cleaned at the end of the day, saves time, no corners or fluid holes to collect material or bacteria. Easy and fast product change--simply change tubing--no cross-contamination. Operates in any position (orientation insensitive). Wide range of flow rates. Many types of tubing are available. Wide selection of drives/motors. Easily repaired. Fewer parts to inventory. Multiple channel applications.

TABLE 1. Feature benefits of peristaltic pumps

68

The Pump Handbook Series

pumps easily handle viscous fluids and small or soft particulates without clogging, and fast pump cleanups make them a particularly attractive labor saving tool. Popular applications include dispensing juices, yogurts, condiments, molasses and other food products on production lines or into individual servings.

CUSTOMIZING FOR A SPECIFIC APPLICATION Peristaltic pumps can be adapted to specific applications and marketplaces by considering the primary features and benefits required by the user. Tubing pumps consist of three major components (pump head, tubing, drive), all of which are interchangeable in the most popular models. To make the best choices, evaluate each component individually. When evaluating pump components, it is important to know the critical limits of the fluid transfer application. Understand the flow rate, type of fluid, service life requirements, pressure, viscosity of the fluid, size of any particulates present, shear rate of fluid and degree of pulsation acceptable.

PUMP HEAD SELECTION The selection of a pump head for any application is a critical step. The use of the proper pump head eliminates unnecessary problems that can develop during setup and operation. The following are features of the pump head to consider: Flow Range. Flow range requirements dictate the size of tubing and ultimately the type of pump head for a specific application. Materials of Construction. Chemical resistant materials of construction and shielded bearings help pumps withstand exposure to aggressive fluids or rugged environments. Pump heads with high performance plastic bodies provide a lightweight, chemically resistant product at an attractive price. Number of Rollers. Fewer rollers on a given size rotor allow quicker fluid transfer, but at greater pulsation. More rollers reduce pulsation and improve dispensing accuracy but decrease flow rate and tubing life. Vacuum and pressure performance of a pump improve as the number of rollers occluding the tub-

TUBING SELECTION

PHOTO 2. Easy loading pump heads can translate into labor savings.

ing at one time increases. Ease of Loading. The method of tube loading greatly affects user satisfaction with the pump. Pharmaceutical, food service and printing applications frequently require several tubing changes over the course of a shift or day. In view of time saved during cleanup or changeovers, easy-loading pumps translate into hundreds of dollars in labor savings (Photo 2). Number of Tubing Sizes Accepted. A pump head design that accepts only one tubing size can maximize the performance (e.g., pressure, vacuum and flex life) for that one tubing size. A pump head designed to handle a range of tubing sizes has some averaged design features. This gives the user greater flexibility in the range of flow rates with one pump head. Fixed or Variable Occlusion. How much control of occlusion is enough to adapt the pump? Fixed occlusion pumps optimize the performance for repeated uses and reduce the chance of operator error. When operated with precision extruded tubing, they deliver excellent repeatable service. Adjustable occlusion pumps allow the tubing to be over occluded to facilitate priming or reduced to improve tubing service life in low pressure applications. Minor adjustments in flow rates are possible. This is helpful when synchronizing

flow rates in multichannel pump heads. Stackability. Design a multichannel pump with one drive system. The number of flow channels possible depends upon the additional torque required for each tube channel up to the limits of the drive. Select from a stack of individual pump heads or a more compact cartridge pump head. Individual pump heads offer a wider range of performance options. Cartridge pump heads have become increasingly popular for their small overall package size. Several individual cartridges mount on one pump body. They are ideal for applications requiring very low flow rates and/or synchronous flow. Specialty Pump Heads. Unique pump head designs meet specific market needs. Pump heads are now available for smooth fluid transfer, fast dispensing volumes, long pump operation and extraordinary chemical resistance. These pump heads have special features that enable them to maximize a particular benefit. Smoother fluid transfer is accomplished by combining two flow channels that have offset pulsation. Extraordinary chemical resistance is possible by using PTFE tubing. A special tube set is matched with a special pump head to provide chemical resistant or high purity fluid transfer.

The Pump Handbook Series

Proper tubing selection is as important as selecting the optimal head. General purpose pump heads accept a wide range of nominal tubing sizes. Specialty pumps usually require special tubing profiles, tubing sets with collars or special fittings. Many types of flexible tubing materials are available on the market at a wide range of prices (Photo 3). Only a few of these materials are suitable for pump tubing. Similar looking materials can deliver vastly different pump performance characteristics. A good pump tubing possess good tensile and compression capabilities. Most pump manufacturers offer a range of prequalified tubing for use in their pumps. These formulations offer consistent flex life and flow rate. Select tubing materials based on the requirements of the application and the preferences of the operator. Pay special attention to the following criteria: Chemical Compatibility. When considering tubing pumps for pumping aggressive fluids, it is critical to select the correct tubing material. The wrong tubing can lead to a hazardous situation with potential to damage equipment and harm people. Consult the pump and tubing manufacturer’s chemical compatibility charts for every new application. With new or unrated chemicals, test the tubing in the fluid before testing in the pump. Immerse a short section of tubing in the fluid. Check for changes in tubing size, color, weight and strength. If possible, test the tubing in the pump before extended use. The flexing of the tubing during the occlusion process works the chemical into the tubing wall and accelerates any decomposition that may take place due to weakness in the tubing material. Non-contamination. Isolation of the fluid is essential for many applications in the laboratory, pharmaceutical and biotechnology marketplaces. Countless studies are available on the inertness of silicone tubing and basic silica material with various biological materials. Several other tubing materials are available as well that meet USP Class VI criteria. Some new pumps use special

69

PHOTO 3. Pump tubing selections

tube sets constructed from inert PTFE tubing. NOTE: Small particles of pump tubing materials will break off into the fluid after extended pump operation. Industry experts refer to this as “spallation.” In some applications these particles are visible to the naked eye. To minimize such degradation, one should change the tubing more frequently. Flex Life in the Pump Head. Different tubing materials have differing abilities to withstand the repeated squeezing action of the rollers. In general, each tube size, tube material, pump head style and operating speed in combination has its own life characteristics. Service life, or flex life in the pump, is the primary concern in a new application. Today greater than 5000 hours of operation is commonplace in systems employing thermoplastic elastomer (TPE) materials. Pump performance is very con-

70

sistent in a specific application. Maximize the life of a pump system by selecting a tubing material that offers long flex life, using thicker wall tubing, and/or by operating a larger pump at slower speed. Clarity. Visibility of fluid movement in the tubing is valuable for many applications to confirm pump priming or a run dry condition. Opaque tubing materials can, on the other hand, limit the exposure of light-sensitive fluids. Durometer. The stiffness of the tubing wall is important in determining the pumpability of a specific tubing. If the tubing is too soft, it collapses. If it is too hard, the pump cannot operate. Durometer (e.g., Shore A, etc.) is the measurement of this physical characteristic for a particular tubing material. Pressure. Pressure performance is usually a limitation of the pressure handling capabilities of the tubing. Most peristaltic pumps use unsupported tubing, which has pressure limits of 2–3 bar (30–45 psi). Some larger pumps on the market use a supported (braid reinforced) tubing immersed in a lubricating fluid to generate pressures up to 15–20 bar (300 psi).

ing the tubing size. Variable Speed Drives. When flow rate flexibility is crucial, consider a variable speed drive. Pumping the right amount of fluid in the right amount of time can be tricky. Variable speed drives usually provide infinitely variable flow from minimum to maximum speed. Motor speed, turn-down ratios and percentage speed regulation are key factors to consider in selecting a drive. Reversibility facilitates purging of the suction lines before changing tubing. Digital Display. Microprocessor-based control improves the level of pump operation. Tight control of line and load regulation provides stable fluid transfer. Simple calibration of your pump system allows digital display of flow rate and other important setup guidelines. Remote Control. Integration of pumps into automated systems for research and process control has become increasingly important. Controllable functions include start or stop, reverse, prime, dispense or copy, and speed. Select from analog (e.g., dc voltage, current or contact closure) or digital (e.g., RS-232-C) interfaces.

DRIVE SELECTION

SUMMARY

Choose a drive motor based on the necessary type of control. Most pump manufacturers offer some degree of interchange between pump heads and drives. This can be very important in the initial stages of pump testing and various types of research. Fixed Speed Drives. These represent the simplest method of fluid transfer. The pump system operates at a single constant speed for the entire time of operation. Flow rates can sometimes be changed by chang-

Peristaltic pumps offer a high degree of customization to meet the specific requirements of many applications. New developments are constantly expanding the capabilities of this versatile pump design. ■

The Pump Handbook Series

Larry Van Bogaert is the marketing manager for the Masterflex product line at Cole-Parmer. He has more than 12 years experience with peristaltic pumps.

POSITIVE DISPLACEMENT PUMPS HANDBOOK

Selection Guide: Rotary Gear Pumps Timely advice for readers considering the purchase of these versatile pumps. By C. M. Kent Whitmire

PRINCIPLE OF OPERATION Gear pumps are rotary, positive displacement pumps in which the main pumping action is caused by relative motion between the pump's rotating elements (gears) and stationary elements (the case and endplates). In simpler terms, the meshing of two or more gears provides the pumping action. It is also characteristic that one of the gears be capable of driving the others. Their positive displacement nature distinguishes them from general centrifugal pumps, in which liquid displacement and pumping action is dependent, in large part, on the velocity imparted to the fluid. The ability of a gear pump to hold a constant flow, not depending on the system pressure curve, differentiates it from a centrifugal pump, which varies the flow range as a function of the system curve change. Being positive displacement, a gear pump's flow

is directly proportional to speed and is marginally affected by differential pressure. Originally designed for clean, lubricating fluids, gear pumps have found success in a wide variety of applications. Standard industrial pumps are available with flow rates to 2,000 gpm and differential pressure to 1,500 psi. Improved bearing materials and rotor dynamics have widened their application range all the way from high viscosities (250,000 SSU and greater) to low viscosities (<30 SSU). Interestingly, Hydraulic Institute data, reflecting mainly centrifugal pump experience, does not cover the range above 10,000 SSU.(Ref. 1)

GEAR PUMP APPLICATIONS Gear pumps are well-suited for both metering and general fluid transfer applications. Metering

pumps tend toward the smaller sizes and are built to closer clearances than transfer pumps. Although most applications are at fixed speed, variable speed drives are becoming more popular, and this brings to gear pump applications the added advantage of selective flow variability. The uses for gear pumps are many. Specific industries and applications include the following: •Chemical and petrochemical processing – for mixing, blending, transferring and metering fluids in process applications. •Lube oil – for circulating lubricating oil in compressors or small and large gasoline or diesel engines. •Food and beverage – for transferring chocolate from processing vats, for circulating hot oil in commercial deep fryers, and for removing heated corn

PHOTO COURTESY ROPER PUMP COMPANY

Rotary gear pumps, both the external and internal variety, belong to the group of rotary pumps that also includes vane pumps, screw pumps, progressing cavity pumps and lobe pumps. Often overlooked in commercial areas and industrial plants, gear pumps play important roles in moving many of today's more difficult to handle fluids. Recently, gear pumps have found new applications in chemical plants, utilities and other areas due to their reliability records and their maintainability and simplicity. Much of this new acceptance can be attributed to better seal and bearing life.

Photo 1. Heavy-duty stainless steel industrial gear pump The Pump Handbook Series

71

syrup. •Fuel delivery – for moving and metering of fuels to single or multiple burner systems, and transferring high pressure fuel to gas turbines in power generation facilities. •Power transmission – for generating hydraulic power to operate off-road machinery, cargo elevators and power tools. •Construction – for handling hot asphalt in road paving and for sealant transfer and metering in general. •Pipeline operations – for crude oil transfer and lube and seal oil supply systems. •Bulk transfer – for loading and offloading fluid cargo from ships, rail cars and trucks. •OEM's – for fuel, lube and seal oil supply, hydraulic power systems, metering systems, and proprietary fluid handling equipment. •General industry – for handling machine coolant or cutting compounds on metal working machinery, for lacquer metering and transfer of beverage can coatings and liquid cattle feed, and for metering and moving printer's ink in printing equipment.

GEAR PUMP DESIGNS All gear pumps have a set of gears, casing, endplates, bearings and a stuffing box or seal chamber. In an external gear pump, the center of rotation of each gear is usually external to the major diameter of an adjoining gear, and all gears are of the external tooth type. In an internal gear pump, the center of rotation of at least one gear is inside the major diameter of an adjoining gear, and at least one of the gears is an internal tooth type. Figure 1 shows a typical external gear pump design. The gears can be of the spur, helical or herringbone design. Used when differential pressures are high, spur gears minimize the loading on the endplates from any thrust reaction. Helical gears are used to reduce noise levels and reduce hydraulic trapping between the meshing of the gear teeth. Her-

72

pump with a specially designed seal chamber for enhanced sealing abilities.

GEAR PUMP SELECTION

Figure 1. General design of an external gear pump

ringbone gears provide the noise reduction of helical gears and the low end thrust of spur gears into one gear; however, their cost is prohibitive for use in most applications, and fluid trapping may be a problem with them. Manufacturers normally select the design of the gears based on years of pumping experience and the internal hydraulic forces generated by the pumping action. Gears are available in many materials. Iron, carbon steel, stainless steel and engineered plastics are the most common. Aside from considering the fluid's effects on the gears, the materials selected must be suitable to prevent galling between the gears. Cases and endplates are available in the same basic materials as gears, and the same considerations must be applied in their selection. Most gear pumps have one or more sleeve type bearings operating in and lubricated by the application fluid, although some are available with all of the bearings located external to the fluid. Bronze, carbon, iron and composite materials are the most frequently used bearing materials. The stuffing box or seal chamber, located in one of the endplates, can be fitted with conventional packing or mechanical seals. Most types of mechanical seals can be fitted, and, as with other pump types, common considerations for seal selection apply. Photo 2 shows an industrial gear The Pump Handbook Series

Flow rate is a function of the gear geometry, fluid viscosity, differential pressure and rotating speed. Comparing different gear pumps directly is difficult because there are no industry standards defining performance, connections, mountings or speeds, as there are for end suction centrifugal pumps (ANSI B73.1). But Hydraulic Institute standards(Ref. 2, 3) provide definitions and requirements for types, testing, general design and other parameters. The spaces between the gear teeth and closely controlled clearances between the gears and the casing and endplates form the pumping chambers. These clearances are controlled by the manufacturer to minimize slip. Slip is the portion of fluid that "slips" back from the high pressure side of the pump to the low pressure side. This effect is similar to leakage flow through the wear clearances in centrifugal pumps. Manufacturers' technical data includes information to predict slip, and therefore delivered flow, for various combinations of speed and viscosity. The flow rate of a centrifugal pump is usually controlled by a throttling valve at the discharge of the pump or, in some cases, through bypass recirculation. In contrast, a gear pump's flow rate is controlled by varying the speed of the pump. With modern driver controls, gear pumps offer opportunities to automate and optimize fluid delivery for enhanced process and batch control. The elimination of a throttling valve or recirculating bypass also reduces the amount of energy transferred to the fluid, so it improves efficiencies. Gear pumps can operate over a wide range of application temperatures and viscosities. Application temperatures up to 450ºF (230ºC) are common, and special designs or modifications are available for temperatures ranging to

PHOTO COURTESY ROPER PUMP COMPANY

Photo 2. Sectional view of an industrial gear pump-note enlarged seal chamber

Thus, their relief valves must be able to pass the rated flow without raising the pressure above a set design point. The bypass flow from the relief valve should be piped back to the suction source, rather than the pump suction. Bypassing the fluid back to the pump suction creates a short loop for the recirculating fluid, allowing temperatures to rise too quickly.(Ref. 7) Many gear pumps fitted with an integral relief valve are available. Photo 3 shows a pump with an integral, spring actuated, poppet type relief valve. Integral relief valves should not be depended upon to protect the system, however. Some standards such as API Standard 676 prohibit their use.

The quantity and quality of application information is the most significant factor in determining pump installation success or failure. Most pumping problems, with both centrifugal and gear pumps, can be traced to conditions at the pump inlet. A minimum amount of absolute pressure, Net Positive Inlet Pressure Available (NPIPA), is required to supply fluid to the pumping chambers. Gear pumps generally require less absolute pressure than do centrifugal pumps. NPIPA is calculated in the same manner as NPSHA, but is normally expressed in pressure units: psi, kPa, bar, etc. The Net Positive Inlet Pressure Required (NPIPR), usually at the inlet connection of the pump,

PHOTO COURTESY ROPER PUMP COMPANY

600ºF (315ºC) and higher. The temperature of the application determines the fluid's viscosity. For general transfer applications, the lowest temperature to be encountered should be used in calculations to determine the pump's highest power requirements. In extreme cases, such as outdoor transfer of corn syrups or heavy fuel oils, a means of heating the fluid will have to be furnished.(Ref. 4) Many gear pumps can be obtained with integral jacketing to help in maintaining a fluid's temperature for pumpability. Generally, centrifugal pumps are not used on fluids with viscosities much above 4000 SSU because they exhibit significant reductions in capacity and head and lose efficiency as viscosity increases. In contrast, gear pumps can maintain high efficiencies throughout the viscosity range. Centrifugal pumps and gear pumps do not normally operate at the same speeds, although there are exceptions. In most industrial applications, gear pumps operate at speeds of 900rpm or less. Gear pumps that are large or handle more viscous fluids normally operate at lower speed.(Ref. 5) Their generally reduced speed and lower shaft deflections in the sealing area offer improved reliability and long term cost benefits for users. Centrifugal pumps can operate briefly without damage against a closed discharge and only generate a pressure equal to its shutoff head. Gear pumps can not operate against a closed discharge. Gear pumps create flow; if that flow is blocked, pressure in the pump's downstream system builds rapidly. If there is no relief valve, the peak pressures generated will depend on the power available to drive the pump. Pressures several times the designed operating limit can be reached in seconds. Relief valves must be sized to bypass the full flow of a gear pump. This is different from a centrifugal pump, in which relief valves need to be sized only for partial flow; as pressures increase, flow decreases so at higher pressure, there is less flow to relieve. Gear pumps exhibit constant flow.

Photo 3. Sectional view of an industrial gear pump showing integral relief valve The Pump Handbook Series

73

is the rating of the total inlet losses occurring within the pump at rated conditions. These consist of loses in pressure due to friction along the inlet passage to the pumping chambers, the loss in pressure caused by friction as the pumped fluid enters the pumping chambers, plus any reduction in pressure resulting from a change in elevation from the inlet connection to the pumping chambers and acceleration of the fluid to the velocity of the pumping chambers. The NPIPA must always be more than the NPIPR for a successful pump installation.

ADDITIONAL BENEFITS Aside from their application benefits, gear pumps offer significant opportunities to reduce operational costs. The low number of parts in gear pumps combined with their relatively low operating speeds add up to a reliable pumping solution that is simple and economical to maintain. Slow speeds translate into lower pressurevelocity (PV) values for mechanical seal faces, resulting in excellent seal longevity. The design of most gear pumps allows service, if required, to be performed without disconnecting any piping to the

74

pump. The high mechanical efficiencies of gear pumps offer energy savings as well, and these can be quite substantial as fluid viscosities increase.

SUMMARY With pump users refocusing on reliability issues, extended life and improved plant economics, gear pumps are beginning to draw renewed interest. At the same time, new technological developments from gear pump manufacturers have narrowed the traditional distinction between centrifugal and gear pump applications. In fact, new opportunities exist in the application of all rotary pumps, including gear, progressing cavity and others. ■

REFERENCES 1. Engineering Data Book, Hydraulic Institute, 2nd edition, 1990. 2. American National Standard for Rotary Pumps, for Nomenclature, Definitions, Application and Operation, Hydraulic Institute, ANSI/HI 3.1-3.5-1994 3. American National Standard for Rotary Pump Tests, Hydraulic Institute, ANSI/HI 3.61994.

The Pump Handbook Series

4. How to Solve Pumping Problems, Roper Pump Company, 1965. 5. Purcell, John, Gear Drives for Rotary Pumps, Roper Pump Company, 1995. 6. Nelik, Dr. Lev, Bearing Life Extension and Reliability Features of Modern ANSI Pumps, 2nd International Conference on Improving Reliability in Petroleum Refineries, Chemical and Natural Gas Plants, Houston, TX 1993. 7. Parker, David, Positive Displacement Pumps - Performance and Application, Warren Pumps, 1994. Kent Whitmire is a Project Engineer with Roper Pump Company (Commerce, GA), where he has worked for the past 18 years. His responsibilities include engineering activities for gear, progressing cavity and diaphragm pumps. Kent also has several years of experience with large centrifugal and axial flow pumps, and he is currently a member of the 3-A Sanitary Standards Committees.

POSITIVE DISPLACEMENT PUMPS HANDBOOK

Gear Drive Options For Rotary Pumps By John E. Purcell otary positive displacement pumps use a wide variety of drivers including electric motors, steam turbines, internal combustion engines, air motors and hydraulic motors. The ideal operating speed of the driver, however, does not always match the ideal operating speed of the pump, or the motor may have a horizontal drive shaft while the pump shaft is vertical. Also it may be advantageous to operate several pumps from the same drive. Gear drives can solve all of these problems. They can change the speed or the direction of the driver’s output, and they can split this output among several pumps. Gear drives are commonly used to change the speed of the driver, usually to a lower speed. This is because most rotary pumps operate best at speeds below the most efficient speeds for electric motors, internal combustion engines or turbines. Large pumps or those that operate on more viscous liquids normally operate even slower. Typical operating speeds for rotary positive displacement pumps range from 3600 to less than 300 rpm. Gear drives are also used to increase drive speed, but this happens mostly on high speed, high head centrifugal pumps.

R

For speed reduction, the parallel shaft gear reducer is the most common. It contains either a helical or spur gear and pinion with the input and output shafts parallel but offset. It changes the output speed to a fixed ratio of the input speed, which can be as much as 1:8 for a single set of gears. Helical gears are quieter and have higher power capacities than similar spur gears. Adding sets of gears, with the reduction in stages, can produce ratios in the range of 1:300 in three stages. Two- or threestage reducers are sometimes designed as concentric shaft, instead of parallel shaft, reducers. This means that the input and output shafts are in a direct line with each other, not merely parallel. Singlestage reducers are most frequently used because they can reduce the speed of a four-pole electric motor to the lower end of the speed range for rotary pumps. These reducers are the simplest and least expensive gear drives. A significant feature of gear drives is that they allow the use of standard motor speeds, and by selecting the proper reduction ratio, one can tailor the pump’s speed to the application. Some pumps feature a single reduction gear drive built in as an integral part of the pump. This

Types of Gear Drives Type of Gear Drive

Single Reduction Speed Ratios Available (Output:Input)

Changes Direction of Drive

Multiple Output Shafts Available

Features

spur gear

1:8

no

yes

least expensive

helical gear

1:8

no

yes

quieter & higher capacity than spur gears, most common

epicyclic

1:10

no

no

high power ratings, small size

bevel gear

1:8

yes

no

right angle drives for vertical pumps most common use

worm gear

1:100

yes

no

high speed reduction ratios

The Pump Handbook Series

Electric motor driving a gear pump through an integral helical gear speed reducer mounted on the pump drive shaft

allows the unit to be simply connected to a motor. Free-standing gear drives offer a wider variety of available ratios and power capabilities, but they require alignment and coupling with the shafts of both the motor and the pump. In cases where the gear drive must transmit large amounts of power and the space available for the drive is small, epicyclic, or planetary, reducers are the best choice. These are more expensive than standard helical gear reducers, but they are very compact for the power rating. They are composed of a sun gear in the center, several planet gears around the sun gear, and a ring gear around the planet gears. The planet gears are held in position relative to each other by a planet carrier, and the ring gear is an internal gear, which means that it has teeth on the inside instead of the outside. Epicyclic gears can be used in several ways, but the most common arrangement is for the high-speed input shaft to be connected to the sun gear, and for the planet carrier to be connected to the low-speed output shaft. In this setup the ring gear is fixed and does not rotate. This type of gear drive can transmit large amounts of power because of the multiple planet gears. In an ordinary helical gear reducer, there is only one pair of teeth in contact at a time. This means that in the epicyclic gear reducer the

75

power is split into several paths instead of being concentrated on one pair of gear teeth at a time. Epicyclic reducers naturally have a concentric shaft arrangement that lends itself to mounting directly on either the motor or the pump. Another type of gear drive commonly used to achieve large speed reductions is the worm gear reducer. This uses a screw shaped gear, called a worm, as the high-speed input gear, and a special type of gear, the worm gear, as the low-speed output gear. Output to input speed ratios from 1:10 to more than 1:100 are readily available in single-stage reduction units. Worm gear reducers have a slightly lower efficiency than helical gear reducers, so they produce more heat. Many have built-in fans to help dissipate this heat. The input and output shafts of a worm gear reducer are at right angles to each other. Consequently, this type of drive changes the direction as well as the speed of the driver. The center lines of the two shafts do not intersect; they are offset from each other. Another characteristic of worm gear reducers is that on most units the output shaft will not turn the input shaft. This means that when the pump is not operating, pressure on one of the pump’s ports can not cause the pump to rotate and drive the motor back through the gear drive. Sometimes brakes or check valves have to be added to the system to prevent this. These may not be necessary when using a worm gear drive, but check with the drive’s manufacturer because under certain conditions this may not hold true. The other type of gear drive frequently used to change direction

Pump driven through a concentric shaft gear reducer

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Gas turbine fuel pump composed of two gear pumps mounted on a helical gear reducer.

from the driver to the pump is a bevel gear drive. Bevel gears have teeth cut into the surface of a cone instead of a cylinder as in standard spur or helical gears. Straight and spiral bevel gears are the counterparts of spur and helical gears respectively. The input and output shafts of this type of drive are at right angles to each other, and the centerlines of the two shafts intersect. Bevel gear drives can be used to change direction only, or to change both direction and speed. Ratios comparable to a helical gear reducer are possible; 1:8 is the largest reduction that is practical in a single stage. These drives are most frequently used as drives for vertical pumps. The drives have built-in thrust bearings to support the weight of the shaft going to the pump. They allow an ordinary horizontal motor or engine to be connected to a vertical pump mounted in a sump, a tank or even an oil well. Gear drives are also used to power several pumps from a single source. These types of drives are usually composed of helical gears and consist of a single pinion driving several gears. The output speeds can be lower, higher or the same as the input speed. Different output shafts can even have different reduction ratios, allowing pumps to be driven at different speeds from a single motor or engine. This type of gear drive is frequently used with hydraulic pumps on equipment that has several independent hydraulic circuits that have different pressure and flow requirements and different loading cycles. Gear drive manufacturers use a system of service factors to help match the capacity of the drive to the The Pump Handbook Series

application. The service factor is a number multiplied by the power transmitted by the drive to find the power rating that is necessary for the application. The service factor depends on how smoothly the driver and the driven equipment operate. The presence of uneven or shock loading raises the service factor. Rotary pumps have a constant torque requirement with constant differential pressure. This smoothrunning characteristic means that they rate a service factor of 1.0. Electric motors and steam turbines also have a 1.0 service factor. Piston engines are not so smooth, and most manufacturers of gear drives recommend a service factor of 1.25 for multi-cylinder engines and 1.5 for single cylinder engines. Gear drives are useful components of pumping systems because they offer flexibility in connecting a pump to its power source. They can change the speed and direction of the drive, and they can split the power input among several pumps. They transmit a lot of power for their size. And they operate reliably with very little maintenance for long periods of time. ■ John E. Purcell is a project engineer at Roper Pump Company. He has been with Roper for 10 years and is a Registered Professional Engineer in the state of Georgia. He holds a Bachelor of Mechanical Engineering from Georgia Tech and a Master of Business Administration from Brenau University in Gainesville, GA.

POSITIVE DISPLACEMENT PUMPS HANDBOOK

Rotary Pump Startups A pre-startup guide that will help keep gremlins away from your system. By James R. Brennan out forcing piping into position. There should be a flange-to-pump gap not exceeding the greater of twice the flange gasket thickness or 1/16" (3 mm). If the gap is greater than either of these values, rework the piping until the gap is this width or less. Positive displacement pumps will normally have a system pressure relief valve installed from the discharge piping to either the source of the pumped liquid, such as a supply tank, or to the pump inlet piping (a less desirable point due to the potential for temperature buildup during relief valve operation). This valve will usually be set slightly higher than the maximum anticipated normal system operating pressure. If possible, verify that it has been properly set. If this cannot be verified, consider adjusting the relief valve to a very low pressure and changing it upward after pump startup. Consult relief valve vendor's technical data to be sure valve adjustment is done in the correct (to lower pressure) direction.

Piping and valving installation (Figure 1) should probably be considered first. Be sure all required valves have been installed. Verify that none is installed backwards. An absent or reverse mounted check valve, foot valve or relief valve can cause very serious damage. Piping should have been inspected during fabrication to insure that weld bead, weld rod and scale have been completely removed. Such hard particles can cause catastrophic pump failure should they lodge in the wrong pump clearance. Temporary, if not permanent, pump inlet strainers should be considered if not present. They should start in a clean condition so that accumulation of dirt can be monitored. The piping system should be pressure tested. Avoid imposing on

A Caterpillar diesel drives a three screw pump via a reduction gear. The pump operates at 1,200 rpm, 200 gpm, 14,000 psig and 10,200 SSU. It was recently placed in service on a new crude oil pipeline at Xan, Guatemala.

M

PHOTO COURTESY OF WINN CZERNY, IMO INDUSTRIES, INC.

PIPE AND VALVES

any system component pressures in excess of its design limits. Many pumps can withstand discharge pressure only on their discharge side. Inlet piping systems are frequently suitable only for low pressure. The pressure test medium should be compatible with the components/ system to be tested. Don't use water if the system is not a water system. A low pressure (15 psig, 1 Bar g) compressed air test may be adequate to find missing flange gaskets or other obvious leak sources. Check and tighten all flange bolts to specified torque. Pump inlet and discharge piping should have been made up from the pump for a distance of perhaps 20 ft (6 meters) to minimize pipe strain on the pump. Piping should be independently supported. Close internal clearance positive displacement rotary pumps do not make very good pipe anchors. When pipe flanges are unbolted from the pump, flange bolts should be able to be installed/removed with-

any pump startups are the culmination of months if not years of work designing the process, machine or system, specifying components, instrumentation, protective devices, and reviewing and qualifying suppliers. It is also the most vulnerable time for any pump. This article describes cautions, reviews and inspections that should be conducted before startup to help insure that all those many gremlins of pumping systems are found out and eliminated in time. Thoroughly read the technical manuals and instructions from the pump, driver and all auxiliary equipment suppliers to learn of requirements that may be specific to their equipment design. While this is the easiest method to protect the system, it is overlooked more often than not.

The Pump Handbook Series

77

Ideally, the entire piping and valve system will be thoroughly flushed to remove all dirt and fabrication debris. This is customarily done using a flush pump – not the normal system pump. Strainers and or filters are installed at appropriate locations, and their dirt accumulation is monitored until they show no accumulation for a period of 24 hours. Flushing usually uses light, fairly hot (150ºF, 65ºC) oil delivered at flow rates higher than system design. The higher flow rates cause higher liquid velocities within the piping system and are more likely to dislodge debris. Some systems will use vibration equipment to impose mechanical "shaking" on the piping, again to maximize the dislodging of dirt. Very extensive piping systems have been known to show debris accumulation even after 30 days of flushing. Because of their long distances and relatively huge holding volumes, pipeline systems will frequently use "pigs," bullet shaped devices, sometimes equipped with wire bristles, which are propelled ahead of a flush or initial product batch of liquid to scrub debris and dirt from the inside of the pipe. Before final startup, be sure valves are open or closed as required. Pump inlet and discharge valves are normally left fully open. Manual pump bypass valves are also normally left open on startup. An air bleed valve in the discharge piping at a high point near the pump will significantly improve the pump's ability to self prime. The valve is left open during startup until liquid flows. It is then shut. Be sure to know where this flow will be directed to avoid inadvertent discharge to atmosphere or spillage. Steam turbine steam valving is very important. Turbine startup procedures should be thoroughly reviewed as there are personal injury issues associated with this equipment if it is started or operated improperly.

FOUNDATION, ALIGNMENT AND ROTATION If horizontal pumps are used, be sure the foundation is level, that hold-down bolts are tight and that grouting, if used, has completely filled the baseplate (no hollows or voids) and has cured. If the pump will be handling liquid above about

78

D

E F

A

B

A. Discharge Block Valve B. Check (one way) Valve C. Drain Valves D. Pressure Relief Valve E. Bypass Valve F. Air Vent Valve G. Pump H. Inlet Strainer J. Inlet Block Valve K. Foot (one way) Valve L. Atmospheric Vent

C

G

C

H

J L

K

Figure 1. Typical piping and valve installation Material

inches/inch/F°*

mm/mm/°C*

Cast Iron Ductile Iron Cast Steel 316 Stainless Steel

6.0 6.6 6.5 9.4

11.0 12.0 12.1 17.0

Table 1. Thermal growth coefficients (*x10-6)

150ºF (65ºC) or a steam turbine is used as the driver, an estimate of the centerline growth in height of the hot machine must be made. Shaft to shaft alignment (cold) should incorporate a deliberate, compensating offset, so that alignment is more nearly correct when equipment is up to operating temperature. Coefficients of thermal expansion for common pump case materials are provided in Table 1. The coefficient is applied to the centerline height of the shafts and the difference in temperature between that at which the unit was aligned and temperature of expected operation. The cold machine should be shimmed high by this calculated amount. The purpose of any shaft aligning procedure is to align the centers of the machine shafts with each other, NOT to align the flexible coupling hubs. At temperature, alignment should be within 0.003 inches (0.076 mm) Total Indicator Reading (TIR), both angular and parallel. Consult a good aligning procedure to achieve or verify this degree of precision. The fact that the coupling may be rated to a much greater misalignment capability has nothing to do with the shaft-to-shaft alignment of the equipment. Survival and longevity of the machinery, NOT the couThe Pump Handbook Series

pling, are the objectives. If hot pumps and/or drivers are used, after they are at nominal operating temperature long enough for thermal growth to have stabilized (probably one hour or more), shut down the equipment and verify that alignment is within prescribed limits. Never rely on the alignment that was produced where the pump and drive train were assembled. Transportation, lifting and handling as well as foundation irregularities will impact alignment, always in an undesirable direction. Final alignment should be achieved as nearly the last step before actual starting of the pump. If equipment is to be dowelled in place, do so to the pump ONLY after several hours, if not days, of good operation and hot alignment checks. The use of resilient mounts is sometimes desirable to reduce vibration being transmitted into the underlying foundation. If used, such mounts must not be deployed beneath the pump or driver but between the pump/driver baseplate or bracket and the foundation. The pump and driver must be rigidly aligned, not resiliently aligned, since the resilient mounts will not maintain adequate alignment under torsional reactions from the transmitted torque.

Coupling

Direction of rotation is critical for most equipment. It is usually indicated by arrow nameplates. Remember that some gearing will reverse rotation from input shaft to output shaft. Most engines and turbines must be purchased for a specific direction of rotation. This is also true of most pumps. Standard AC electric motors are frequently bidirectional; their direction of rotation will depend upon how the power cables are connected. It is normally not possible to predict their direction of rotation before- hand. It is recommended that the flexible coupling at the motor shaft be disconnected and the motor momentarily energized (jogged on, then immediately off) to see if its rotation is correct for the rest of the driven equipment. If not, two of the electric power cables will need to have their connections reversed. Verify correct rotation after reversing, if necessary, before reengaging the flexible coupling.

Pump

Driver Gear Bearings Figure 2. Typical lubrication points

with cooler, filter and instrumentation. Be sure to verify that any lubrication required has been addressed. Equipment that has been in storage may require draining and addition of fresh lubricant or even flushing out. Any gearing present (pump timing gears or reduction drive gears, for example) should be reviewed for the presence of the correct type and quantity of lubricant. Constant level oilers should be filled to their mark with clean, fresh lubricant of the correct type. Some flexible couplings are grease lubricated and should also be checked. Most electric motors will have grease lubricated antifriction bearings that should be checked as well. A person should be able to turn

over almost all rotary pumps by hand. Pumps should generally turn over smoothly, with no catches or uneven rubbing. Very large pumps may need a helper bar but should not be at all difficult to turn. If one is, consult the pump vendor. Partial disassembly may be advisable to determine the cause of difficulty encountered (foreign material, pipe strain, rust) before starting.

STARTUP SPARES With care and planning, startups will generally go smoothly, without significant problems. However, it is prudent to have key spare parts on hand in the event they are needed quickly for correction after

PHOTO COURTESY OF IMO INDUSTRIES, INC.

Ideally, the entire piping and valve system will be thoroughly flushed to remove all dirt and fabrication debris. This is customarily done using a flush pump – not the normal system pump.

LUBRICATION Most rotating machinery has some form of lubrication for its bearing systems (Figure 2). It may be as simple as a permanently grease packed, sealed ball bearing or as complicated as a separate lubricating oil pump system complete

Twelve stage pump to DC motor alignment check using dial indicator The Pump Handbook Series

79

some unanticipated problem, minor damage, or need to disassemble a piece of equipment for inspection. For rotary pumps this would normally be a set of shaft seals, gaskets, o-rings and bearings, frequently available as a minor repair kit. For other rotating equipment, spare bearings, grease and oil seals and gaskets should be on hand so as to

If the pump will be handling liquid above about 150ºF (65ºC) or a steam turbine is used as the driver, an estimate of the centerline growth in height of the hot machine must be made. avoid delay in the startup. More extensive spares will depend on availability from the vendor, criticality of pump operation, plant practice and, perhaps, other issues specific to the installation. If the startup goes well and the spares are not consumed, it is appropriate to keep them on hand for future routine inspections and service.

RESOURCES Be sure electric power, steam, cooling water, hot oil, instrumentation power or air or any other auxiliary resources are available and ready before start. Be sure adequate pressure and temperature gages are in place so observations can be made during startup. Without them, you are working blind. Speed indication (tachometers) may also be needed if the drive is not a fixed speed one such as an AC electric

80

ROTARY PUMP STARTUP CHECK LIST Project:__________Location:__________Unit No.:__________Tag No.:_________ 1. PIPING ❏ Clean ❏ Bolts Tight ❏ Strain Removed ❏ Gaskets in Place ❏ Pressure Tested ❏ Flushed ❏ Other__________ 2. VALVES ❏ Not Backwards ❏ Clean ❏ Bolts Tight ❏ Gaskets in Place ❏ Correct Position (Open/Close) ❏ Relief Valve Set Pressure ❏ Other__________ 3. FOUNDATION ❏ Level ❏ Solid (No Voids) ❏ Bolts Tight ❏ Other__________ 4. ALIGNMENT ❏ Angular (Cold & Hot)__________ ❏ Parallel (Cold & Hot)__________ ❏ Other__________ 5. ROTATION ❏ Verified (CW or CCW) ❏ Other__________

7. SPARES AVAILABLE ❏ Pump ❏ Driver ❏ Gear ❏ Other__________ 8. RESOURCES AVAILABLE ❏ Electric ❏ Steam ❏ Cooling Water ❏ Hot Oil ❏ Auxiliaries ❏ Gages in Place ❏ Other__________ 9. LAST MINUTE ❏ Pump Filled with Liquid ❏ Shaft Seals Wetted ❏ Key Contact Phone Nos. ❏ Inlet Liquid Supply ❏ Discharge System Ready ❏ Air Bleed Valve Open ❏ Pump Hand Rotated ❏ Other__________ 10. COMPANY/INDUSTRY SPECIFIC ❏ _______________ ❏ _______________ ❏ _______________ ❏ _______________ ❏ _______________ ❏ _______________ ❏ _______________ ❏ _______________ ❏ _______________

6. LUBRICATION ❏ Pump ❏ Driver ❏ Gear ❏ Other__________ Figure 3. Rotary pump startup checklist

motor. If the pump will be handling hot liquid, preheat the pump as necessary so it is not exposed to thermal shock when otherwise hot liquid reaches an ambient temperature pump. Rotary pumps may be somewhat more sensitive to thermal shock due to their close internal running clearances.

LAST MINUTE It is good practice to fill the pump and as much of the inlet pipThe Pump Handbook Series

ing system as possible with the liquid to be pumped. This will assist in priming and reduce the risk of pump damage during an otherwise dry start. A rotary pump will prime more quickly if internal pumping elements are at least wetted. Priming is nothing more than pumping air from the inlet system to the discharge system. The ability of a rotary positive displacement pump to act as an air compressor is very much related to having some liquid

present internally. Pump shaft seals, especially mechanical seals, should never be operated dry. Immediate, or at best premature, seal failure is the inevitable result. Again, filling the pump with the liquid to be pumped and hand rotating the pump a few times helps insure that liquid is present at the shaft sealing mechanism to carry away frictional heat during startup. If the particular pump has a seal chamber access plug, remove the plug, fill the chamber with liquid and reinstall the plug. Have phone numbers on hand of key vendor service departments, fire brigade and medical emergency services in the event they are needed. When handling petroleum and other flammable liquids, there are both pollution and fire hazards present. Insure that there is an adequate supply of liquid in the pump inlet system (no half empty supply oil tanks or the like). It is also prudent to confirm where pump discharge flow

will be going to be sure the discharge system is ready. Loud or erratic noise at startup is an indication of cavitation (inadequate pump inlet pressure) or air being drawn into the pump inlet system. It is frequently accompanied by increases in or excessive vibration. If mild, troubleshoot the cause. If severe, shut down the pump and find the source of the problem. Use the Rotary Pump Startup Check List accompanying this article or a similar control to help insure that all contingencies have been addressed.

CONCLUSION Our discussion cannot be considered all-inclusive since each pumping system has unique features and requirements, some of which may interact with each other or with other aspects of the overall plant operation. In addition, no allowance has been made here for regulatory requirements, special-

The Pump Handbook Series

ized industry or company guidelines and the like. Where values are recommended, they are intended for use in the absence of vendor or specifically engineered information. Always use the more stringent of either the recommendations herein or the vendor or engineering guidelines. ■ James R. Brennan, currently group manager for three pump divisions of Imo Industries specializing in crude oil transport pumps, is a 1973 MIE graduate of Drexel University, Philadelphia, PA, USA. He has more than 25 years of experience with screw pumps at Imo Industries, is a member of the Society of Petroleum Engineers, and was engineering manager of a pump division for five years. He has authored many papers and articles as well as spoken at a number of industry conferences. He is also a frequent contributor to Pumps and Systems.

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POSITIVE DISPLACEMENT PUMPS HANDBOOK

Rotary Pump Troubleshooting Here is a framework for reducing the time it takes to develop troubleshooting skills. By James R. Brennan rouble shooting rotary pumps in systems is a skill that takes some time to develop. This article will attempt to reduce that learning time while providing a framework for and guidance in determining the cause of a problem. In almost any pumping system the pump is the most vulnerable component. Regardless of what may be wrong in the system, the symptoms frequently indicate a "pump" problem, another problem results in damage to the pump, which is then viewed as the culprit. Reputable pump manufacturers produce very few defective pumps. The problem is usually caused by a system component malfunction, inadequate control of the liquid or a change in operating requirements, burdening the system or pump with conditions in which it cannot perform.

T

scope of this material. Some rotary pump systems are life supporting or operate on critical or sanitary services in which malfunctions can have dire consequences. Be sure to understand the repercussions of trouble shooting activities on the system as well as personnel involved. Before beginning the trouble shooting process, be sure pressure gages are available for a pump's inlet and discharge. A temperature gage at the pump inlet can also provide valuable information, depending on the application or problem. There should be some method of verifying pump speed through either a handheld tachometer or strobe tachometer if the system is not already so equipped. This review will seek to explain problems that may come up in existing systems that have been operating satisfactorily. The approach is some-

what different from a first time startup of a new system where foul-ups like reverse installed valves, missing components, wiring errors, extensive fabrication debris and pipe strain are the norm. Be sure that the drivers are absolutely locked out before removing guards or conducting any system inspections. If the system includes accumulators, pulsation dampeners or other spring or compressed gas energy storage devices, be certain they are fully discharged of liquid before doing any system work. Observe all company and industry safety regulations and guidelines to minimize any inherent hazards to personnel or the system.

INFORMATION GATHERING Make note of anything that has changed since operation was last satisfactory, regardless of how unrelat-

Identifying which of these factors is the cause of a problem is not always easy, but an open mind certainly helps. As the following topics are covered, there will be a mechanical aspect and a hydraulic one. For example, loss of flow from the pump could be caused by a broken driver shaft (mechanical) or by cavitation caused by excessive inlet pressure losses due to cold, viscous liquid (hydraulic). This article makes no provisions for special industry or regulatory practices. Some pumping systems handle toxic, corrosive, flammable or other dangerous fluids and need special precautions that go beyond the

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PHOTO COURTESY OF IMO INDUSTRIES, INC.

INTRODUCTION- CAUTIONS AND RECOMMENDATIONS

Timing gear inspection being performed on a 5,000 gpm twin screw cargo pump. The Pump Handbook Series

ed to the problem it seems. Was the system undergoing routine maintenance? Were any new or repaired components changed out? When was the pump last serviced? Of what did that service consist? What was the appearance and condition of the pump's internal parts? From where were replacement parts obtained? For recirculating systems such as for lubrication, was new or additional liquid added? For one-time-through systems such as fuel oil burners, has the supplier of the fuel, fuel grade or fuel temperature changed? How long did the pump operate before the problem presented itself? Describe the problem in as simple and straightforward a manner as you can. Write down the normal expected pump operating conditions as follows: Inlet pressure Outlet pressure Flow rate Pump speed Liquid Minimum and maximum liquid temperature Minimum and maximum liquid viscosity Continuous or intermittent duty cycle Note the presence or absence of a change in pump noise or vibration. A change in either or both of these characteristics points to a number of very specific things to check as possible causes.

FLOW LOSS OR LOW FLOW Be sure pump direction of rotation is correct, an obvious but sometimes overlooked problem. (See the "Loss of Suction" section of this article, next, for a discussion of the impact of this problem.) Be sure pump operating speed is correct. This is especially true for drivers that can operate at more than one point, such as direct current (DC) motors, steam turbines, air motors and engines. Flow loss is normally accompanied by a reduction in system pressure. Either the pump is delivering less flow, or the system is bypassing it – such as through a defective or worn relief valve or pressure control valve. The pump could be worn and internally bypassing (slipping) flow so that less flow reaches the system. In that event, pump repair will be necessary. A partial inspection of the pump internals will usually provide

a good indication of wear condition. If the operating viscosity of the liquid has been reduced (new liquid or higher operating temperature), a rotary pump's rated flow will be slightly reduced, more so for higher pressure operation.

LOSS OF SUCTION Loss of suction can be minor, causing little short term damage or major enough to cause catastrophic damage. Loss of suction means that liquid flow is not reaching the pumping elements or not reaching it at a sufficiently high pressure to keep the pumpage in a liquid state as the pumping elements capture a volume of liquid. Loss of suction can be interpreted as a pump's inability to prime, as a sign of cavitation or an indication of a gas content problem. Most rotary, positive displacement pumps are self priming. This means that, within bounds, they can evacuate (pump) a modest amount of air from the inlet (suction) system into the discharge (outlet) system. Rotary pumps are frequently not very good air compressors, and the pump discharge should be temporarily vented to allow the inlet side air to escape from the discharge side of the pump AT LOW PRESSURE. If possible, filling the inlet system with liquid – or at least filling the pump (wetted pump elements) – will make a major improvement in the pump's priming capability. Cavitation is insufficient system inlet pressure to the pump to prevent the liquid from partially changing to a gas. This can be caused by an inlet system restriction, excessive liquid viscosity or excessive pump speed. Inlet restrictions can include dirty or clogged inlet strainers, debris floating in the liquid supply that covers the inlet piping intake, rags or port blanking flanges that have gotten into the system or not been removed, especially after maintenance. If the pumped liquid is cooler than design temperature, its viscosity (thickness) may be too high, causing excessive friction (pressure loss) in the inlet piping system. In the latter case it may be necessary to increase the pumping temperature of the liquid. If the liquid being pumped has changed, a change in its viscositytemperature characteristics or its vapor pressure-temperature characteristics should be examined. CavitaThe Pump Handbook Series

tion is caused by operating the system inlet pressure-temperature in a combination such that the liquid's vapor pressure (a pressure at which the liquid converts to a gas at pumping temperature) is reached. The liquid begins to turn partially into gas, and the pump is unable to handle this compressible gas-liquid mixture. Cavitation is frequently accompanied by noise, vibration and a significant increase in discharge pressure pulsation. Modest cavitation will cause pitting to appear on pumping elements not unlike that found near the root of marine propeller blades. Gas in the inlet flow has the same impact on pump operation and the same symptoms as cavitation. It can be caused by vortexing (whirlpooling) of the liquid in its supply source, an action that drags air into the liquid. Many wellhead flows of crude oil are already a mixture of gas and liquid. The natural gas normally must be separated before the oil is pumped unless multiphase flow pumps are used. If a pump operates at an inlet pressure below local atmospheric, it is quite possible that air is being drawn into the inlet piping through a loose piping or pump casing joint, leaky suction valve stem, or a defective, cut, folded or otherwise damaged inlet system joint gasket. In recirculating systems, such as a lubrication system where the liquid pumped is continuously returned to a supply source or tank, if the tank and return lines are not adequately designed, located and sized, air is easily entrained in the oil and immediately picked up by the pump inlet system. Be sure liquid level at its source is at or above minimum operating levels. Lines returning flow to a supply tank should terminate below minimum liquid level. Internal tank baffles are usually necessary to provide full tank volume retention time so any entrained air can more readily dissipate.

LOW DISCHARGE PRESSURE Low discharge pressure can be caused only by loss of flow. Pump discharge pressure is caused ONLY by the system's resistance to the flow provided by the pump. Either the pump is not providing the flow expected, or the system is not offering the expected resistance to that flow. It is possible that flow into the pump is being restricted (cavitation

83

vibration mode caused by operating pressure, flow rate and the valve design. Resetting or a change in an internal valve component is usually sufficient to solve the problem. Consult the valve supplier. If the drive system includes gearing, belts or chain drives, sheave and sprocket alignments become very important and should be checked.

PHOTO COURTESY OF IMO INDUSTRIES, INC.

EXCESSIVE POWER USAGE

The complex hydraulic pump/motor/valve package of a rotary pump system depends on many interactive components.

or suction starvation). This phenomenon is usually accompanied by noise and vibration. Or it could be that the pump is not producing its rated flow (pump worn or damaged), or that the pump flow is bypassing rather than being delivered into the system as intended (open, improperly set, damaged or worn discharge system valve). If the pump is relatively new and not being used in abrasive service, it is most probable that discharge flow is bypassing. The most likely paths for such unwanted bypass are the system pressure relief valve (sometimes built into the pump), a bypass pressure regulator leaking (typical of a fuel oil burner system), an inadvertently open bypass valve, or any of these valves having worn valve seats, incompletely closed stems, incorrect signal control or broken springs. Many pumps can receive a quick, if incomplete, inspection in place without disturbing piping or pump alignment. If the pump does not turn over by hand or with a little leverage assistance and in a smooth manner, the pump itself may be the problem. If one or more of the pumping elements can be visually inspected without major tear down or pump removal, do so. Enough wear to cause a pressure reduction (flow loss) should be readily visible. It is sometimes difficult to deter-

84

mine if a valve is bypassing when it shouldn't, especially if the valve is built into the pump. It is probably best to remove the valve, do a partial valve disassembly and examine the mating valve seat surfaces or seat seals for wear or damage. Check any spring to be sure it is not broken. Work the valve mechanism manually if possible to detect any binding or galling. If the problem has still not been identified, be sure the pump driver speed is being achieved and that the pump shaft is actually rotating at its correct speed. These conditions must be met, especially in a new system startup. See "Loss Of Suction" for common causes of pump suction (inlet flow) problems.

EXCESSIVE NOISE OR VIBRATION As mentioned, excessive noise and/or vibration can be a symptom of cavitation, suction starvation or excessive gas in the liquid. This is especially true if the discharge pressure is fluctuating or pulsating. Mechanical causes of noise and vibration include shaft misalignment, loose couplings, loose pump and/or driver mounting hardware, and worn or damaged driver or pump bearings. Pump valves can also vibrate noisily. Especially on the discharge side of the pump, valves can sometimes go into a hydraulic

The Pump Handbook Series

Excessive power consumption can be caused by either mechanical or hydraulic problems. For rotary pumps, the pump power requirements are directly proportional to pressure and speed. If either has increased, the required input power will also increase. Power required will also increase if the fluid viscosity has increased. This can happen if the liquid has been changed to something new or the liquid operating temperature has been reduced. Some liquids (grease, for example) are shear sensitive and can become more or less viscous with shear (pumping action) as well as undergo permanent viscosity change from shear over time. Mechanical causes of high power usage include bearing wear out, pumping elements rubbing (a situation that can lead to pump failure), very bad shaft alignments and poor pulley alignments for belt drive arrangements.

RAPID PUMP WEAR Rapid pump wear is caused either by abrasives in the liquid or operation under conditions for which the pump is not suitable, such as excessively low viscosity or excessively high pressure or high temperature. If abrasives are a normal condition of the pumping application, as in slurry pumping, then pump wear will be a fact of life, and the best that can be done will include pump and drive speed selection that provides the best economic evaluation over the pump life cycle. While requiring bigger displacement and more expensive pumps, slower operation on abrasive service often pays back far beyond the initial purchase cost differential. Wear due to abrasives in the liquid is a function of speed raised to a power, usually between 2 and 3. If the abrasives are deliberately introduced, as when

fuel oil additives intended to reduce boiler corrosion are brought into a system, they should be injected downstream of any liquid recirculation to insure that they do not go through the pump. Obviously, if abrasive foreign material is not supposed to be present, strainers or filters should be employed wherever possible and practical. Rapid wear is sometimes not wear in the sense of a non-durable pump, but rather a catastrophic pump failure that occurs very quickly. Looking at the pump internal parts alone can frequently not provide much help in setting a direction to search. So it is important to know what was occurring in the time peri-

od immediately preceding the detection of a problem.

normally yield faster and more positive results. ■

CONCLUSIONS

A frequent contributor to Pumps and Systems, James R. Brennan is currently group manager for three pump divisions of Imo Industries specializing in crude oil transport pumps. He is a 1973 M.I.E. graduate of Drexel University in Philadelphia. He has more than 25 years experience with screw pumps at Imo Industries, is a member of the Society of Petroleum Engineers, and was engineering manager of a pump division for five years. He has authored many papers and articles as well as spoken at a number of industry conferences.

Careful and systematic examination of all aspects of a pumping system, sometimes with the assistance of the pump manufacturer, can solve most pump and system problems. System contamination with hard and/or abrasive foreign material is a leading cause of pump problems, followed by inadequate inlet pressure. There is a very strong tendency to place immediate blame on the pump. This usually causes other areas of the system to be overlooked, and the real problem remains undetected. An open and inquisitive approach will

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POSITIVE DISPLACEMENT PUMPS HANDBOOK

Selection Guide to Metering Pumps Understanding principles of operation and knowing the advantages of different designs is critical to proper specification of metering pumps. By Scott McKay ver the last 60 years, metering pump technology has been employed in some form or another in practically every industry. This technology is used from the actual manufacturing of a chemical to the end-use application of the chemical. Metering pumps have penetrated into ever widening areas of application, and they are becoming increasingly sophisticated. Thus, understanding how they work and the advantages of the different designs is critical to proper specification of a metering pump.

O

TYPICAL APPLICATIONS Metering pumps are used whenever it is necessary to pump accurate quantities of liquid. This is in contrast to the typical liquid transfer applications performed by centrifugal and many rotary pumps. Metering pumps can handle a wide range of fluids, ranging from acids, caustics and corrosive fluids to polymers and slurries in both continuous and batch control processes. The largest area of application by far is the water treatment market, but metering pumps are used anywhere there is a requirement for one of the following: • • • •

High accuracy High pressure Low flow Pump flow independent of system pressure • Leak free design • Handling of toxic, hazardous or high temperature fluids Some common applications in major markets include: Water & Wastewater Treatment- chlorination and fluoridation

86

of drinking water; coagulation with aluminum and iron salts, odor control; dechlorination and pH control of wastewater; softening, corrosion inhibition, and scale prevention for boiler feed water with chemicals such as hydrazine, ammonia and phosphate. Chemical Processing- formulation and introduction of additives, pH adjustment, foam control, solvent feed Petroleum & Petrochemical Industries- odorant injection in natural gas, addition of inhibitors to flash towers Pulp & Paper Manufacturingpaper stock preparation, bleaching of wood pulp, de-inking, coating, addition of fillers and process materials as well as coloring agents for corrugated cardboard Electric Power Generationbiocide and anti-scalant injection in boiler water treatment, fuel oil additive injection Food & Beverage Productionsyrup feed; stabilization of chemicals; flavor and additive ratioing and blending, controlled introduction of coloring agents, spices and vitamins; sterilization and process water treatment

the plunger, which displaces process fluid directly or actuates a diaphragm that displaces the fluid. The reciprocating motion creates a sinusoidal flow pattern that is different from the steady flow pattern of centrifugal and many rotary pumps. Knowledge of this sinusoidal flow pattern is crucial in view of the peak flows created and the acceleration head created. Peak flows of reciprocating pumps are 3.14 times the average flow. Since reciprocating pumps must start and stop the liquid on each suction and discharge stroke, acceleration head is created due to the changing liquid velocities (Figure 1). An understanding of both of these is important for proper piping and system design. Capacity of a controlled-volume pump per unit of time is a function of the plunger diameter, plunger stroke length and stroke frequency. Unlike centrifugal pumps, metering pumps have a capacity that remains relatively steady regardless of the pressure differential. Acceleration Velocity

PRINCIPLES OF OPERATION A metering or controlled-volume pump is a positive displacement reciprocating pump. It can accurately feed a pre-determined volume of liquid in a specific period of time. The primary feature that separates a metering pump from other reciprocating pumps is a stroke length adjustment mechanism that allows for capacity adjustment. Metering pump operation depends on the reciprocating motion of The Pump Handbook Series

Discharge

Figure 1. Non-uniform flow

V=A * h *n where V= theoretical output A = plunger area h = stroke length n = stroke frequency

0% 10

Flow

Since the plunger diameter is constant for a given pump, stroke length is most commonly used to adjust output capacity. Stroke frequency also can be used to increase the capacity adjustment range or turndown. Lost motion (hydraulic bypass and mechanical) and non-lost motion/amplitude modulation are the two categories of stroke length adjustment for capacity control (Figures 2 and 3). Lost motion designs are less costly and less complex and do not transmit forces back to the adjustment mechanism. Their disadvantage is that liquid is accelerated to the maximum velocity instantaneously, and this results in shock waves. Amplitude modulation designs are more complex and more costly, but they eliminate the instantaneous accelerations and liquid

velocities and resulting shock waves. The disadvantage of this design is the transmission of forces back through the stroke adjustment mechanism. Lost motion, the most commonly used stroke length adjustment mechanism, is typically used for low flows. It utilizes a constant eccentricity with a diversion of part of each stroke to idle motion. With hydraulic bypass, the stroke length of the plunger is fixed, but the effective stroke length is varied by bypassing hydraulic fluid during part of the cycle. In mechanical lost motion designs, the spring actuated plunger follows the complete cycle of the cam at a 100% capacity setting. However, a mechanical stop prevents the plunger from following the cam's complete cycle at lower capacity settings. Non-lost motion designs vary the eccentricity of the plunger drive mechanism. A cam or crank mechanism sets the stroke length, so a change in the radius affects the pump output. One design involves a rotating disc and connecting rod, and

50%

25%

Time

Figure 2. Typical lost motion characteristics

Flow

100% 50%

the plane of the disc is adjustable. Inclining the rotating disc adjusts the eccentricity and creates the plunger reciprocation. Another design involves an adjustable crank/rocker box to vary the eccentricity. A springloaded plunger design that eliminates crossheads, connecting rods and the shock loads created by those components is also available. A plot of metering pump flow rate versus stroke length shows a linear flow characteristic. By plotting two flow measurements at two stroke length settings and drawing a straight line between the points, other flow rates versus stroke lengths can be accurately determined. The accuracy of most metering pumps is +/- 1% of rated capacity. These pumps can be adjusted between 0-100% of rated capacity. However, pump accuracy is maintained over a range only. This is called the turndown ratio. Most metering pumps are rated to maintain accuracy over a 10:1 ratio. The use of stroke frequency control in combination with stroke length increases this ratio. Additionally, new metering pumps are hitting the market that offer increased accuracy and turndown. Metering pump designs have common operating characteristics. Flow is produced through operation of the reciprocating plunger. As the plunger retracts, the vacuum created draws fluid through a one-way suction valve. This retraction, combined with the system pressure, holds the discharge check valve closed during this portion of the stroke. As the plunger reverses direction and moves forward, liquid is forced through a one-way discharge valve. In this part of the stroke, the suction valve is forced closed. There is a valve-less design that utilizes a rotating and reciprocating plunger for very low flow applications.

DESIGNS 25% Time

Figure 3. Amplitude modulation flow characteristics The Pump Handbook Series

Metering pumps consist of four basic components: pump head/liquid end, flow/stroke adjustment, drive mechanism and powered driver. Pump Head/ Liquid End- The pump head is the portion of the pump in contact with the process fluid. There are three main liquid end designs: packed plunger, mechanical-

87

ly actuated diaphragm and hydraulically actuated diaphragm. To choose the proper pump head for a particular application, flow rate, pressure range and individual characteristics of the pumped liquid must be taken into consideration. Check valves are used on all pump heads and are available in several designs, each with different options. Ball check valves are the most common. Options include spring-loaded and double ball. Spring-loaded designs decrease slippage and are used for high specific gravity and viscous fluids to assure proper seating. Spring-loaded designs decrease a pump's suction lift capability and increase wear on the ball. Double ball check valves also decrease valve slippage and are used for high pressures (usually > 1000 psig) and for increased accuracy. Disk-type check valves are another common design and are usually used for higher capacities (>400 gph) and higher stroking rates. Disk valves have limitations with solids and viscous fluids due to their wide seating area. Check valves are critical to efficient metering pump operation, so care should be taken to select the proper design. Packed plunger designs have the plunger in direct contact with the process fluid. The plunger requires packing as a seal. This has disadvantages, including controlled leakage to lubricate the packing and plunger. Considerations for packed plunger designs include: Advantages • low cost • high flow, pressure capability • high temperature • high accuracy • low NPSH requirements Disadvantages • packing leakage • packing requires adjustment • unsuitable for abrasive slurries • unsuitable for fluids that crystallize • no built-in relief valve Mechanically actuated diaphragm designs have a diaphragm physically attached to the plunger. The diaphragm separates the process fluid from the plunger. This design is leak free and good for toxic or hazardous chemicals. However, it has a pressure limitation of approximately 150 psig because the non-process

88

side of the diaphragm is exposed to atmospheric pressure, and there is a resulting differential pressure across the diaphragm. Some advantages and disadvantages include: Advantages • low cost • no leakage • low maintenance • no hydraulic system • good solids and viscosity handling capability • no potential contamination of process fluid by hydraulic oil • good suction lift • extended dry run capability Disadvantages • low flow, low pressure capability • no built-in relief valve • accuracy varies more than other designs Hydraulic diaphragm designs are actuated by the plunger working against a hydraulic fluid that actuates the diaphragm. There is no mechanical attachment between the diaphragm and the plunger. Since the diaphragm is pressure balanced to limit the differential pressure, this design can handle higher pressures than the mechanically actuated diaphragm design. To maintain accuracy in a hydraulically actuated diaphragm pump, it is necessary to keep a consistent volume of hydraulic fluid, free of entrained air, in the hydraulic reservoir and to protect against overpressurization of the hydraulic fluid. A valve system consisting of a refill/compensator valve, air bleed valve and pressure relief valve is usually required to maintain this sound hydraulic system. However, a vertical piston design, eliminating the need for a refill/compensator and air bleed valve, is available. Advantages • no leakage • built-in relief valve • relatively high pressure capability • high capacities • better accuracy than mechanical design Disadvantages • higher cost • requires valve system for hydraulics Both hydraulically and mechaniThe Pump Handbook Series

cally actuated designs have various diaphragm options. The flat diaphragm is the most commonly used design for both hydraulic and mechanical pumps. This diaphragm is available in a wide variety of materials. Most hydraulically actuated flat diaphragm designs have a process side contour plate to limit diaphragm movement. This limits slurry and viscosity handling capability, decreases suction lift and increases the NPSH requirement. There is a mechanical refill design available for this design that overcomes these problems, however. Double diaphragm options are also available. These allow for diaphragm leak detection devices and eliminate the possibility of hydraulic fluid contamination of the process fluid. The tubular diaphragm design is unique to hydraulic diaphragm designs. Most tubular designs use a flat primary diaphragm acting on an intermediate fluid that compresses the tubular diaphragm. This is good for slurries and viscous fluids and also has the double diaphragm advantages. The disadvantage of this design is the requirement to synchronize the diaphragms. There is a single diaphragm tube design available that has all the tubular design advantages plus increased suction lift, lower NPSH requirements and no synchronization requirement. Additional diaphragm design options include bellows, forced return and remote mounting. Drive Mechanism- This usually consists of a high speed worm gear reduction unit that reduces the motor input speed to a slower plunger stroking frequency output. This is combined with a mechanism that translates the rotary motion of the motor/driver into the reciprocating motion of the plunger. Powered Driver- Metering pumps are generally driven by constant speed AC motors. Variable speed DC motors, variable frequency motors and solenoids are becoming more common, though, since they offer automatic and increased process control. Pneumatic and hydraulic drivers are occasionally used as well.

SUMMARY Metering or controlled-volume

pumps can handle an extremely wide range of fluids, including acids, bases, solvents, slurries and other corrosive and hazardous chemicals. Understanding the principles of operation and knowing the advantages of the different designs is critical. This understanding, combined with knowledge of the physical characteristics of the fluid, allows for proper pump selection and ultimately, a dependable and versatile system. Recent and on-going innovations in metering pump technology are

continually expanding the application base and increasing process control and automation. ■ Scott McKay is the Director, Sales & Marketing for JESCO America Corporation. He has a B.S. in Mechanical Engineering from Clarkson University and an MBA from the University of Rochester. He has seven years of experience in the metering pump field.

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POSITIVE DISPLACEMENT PUMPS HANDBOOK

Reciprocating Metering Pumps in Leak-Free Design Diaphragm designs are proving their worth in critical fluid handling processes as well as everyday applications. By Gerald Harting

I

90

pharmaceutical, and power, as well as the subset of water and waste treatment applications within each of these industries. Due to their exceptionally long up-times, high efficiency, zero emissions, and their simple, quick maintenance, the trend toward use of diaphragm pumps is extending even into the handling of liquids formerly considered not critical. Water and waste treatment applications for metering pumps represent 65% of the market, and are well known and straightforward services. Therefore, the focus of this article will be on the other 35% of the market for metering pumps, industrial and process applications. The article will explain how modern diaphragm designs function, and it will detail their performance characteristics, installation requirements, and

safety advantages.

RECIPROCATING DRIVE ELEMENTS (FIGURE 1) Diaphragm pumps employ the same reciprocating mechanisms as plunger pumps. Simple applications utilize a simplex drive element, usually with adjustable stroke length to vary the flow from the pump by changing the displaced volume (Figure 1a). Such drive mechanisms must be stroke adjusted manually, electrically, or pneumatically under full and pulsating load, so even modern designs with excellent internal load management are generally limited to about 50 hp. Drive elements without stroke adjustment are capable of much higher hp inputs. Modular drive element designs can be multiplexed together to achieve process

PHOTO COURTESY OF AMERICAN LEWA, INC.

t is hard to pick up a technical journal nowadays without seeing evidence of the increasing trend towards zero-leakage pumps. Although driven in part by environmental and safety requirements, the justification is often primarily economic – that is, plant managers want to improve reliability and process integrity while minimizing downtime, scrap losses and maintenance. Leak-tight pumps often have a much higher initial cost than equivalent sealed designs, but the payback can be short if they are properly applied. The most common rotating pump is the centrifugal, with the sealless designs incorporating canned motors or magnetic drives to deliver torque to the impeller without the need for moving seals. In reciprocating pumps a mechanically or hydraulically actuated diaphragm is used to replace sliding plunger seals. Although these pumps are called “sealless," they do have static seals at component interfaces, such as the can in the centrifugal or the diaphragm clamping area in the diaphragm pump. Reciprocating diaphragm pumps in single and multihead configurations are available for metering and controlled volume services over the remarkable power range of 0.1–1000 hp and they are well proven. They offer special advantages in the pumping of liquids that are corrosive, flammable, reactive, non-lubricating, abrasive, toxic, shear-sensitive, high purity, and sterile (medical or pharmaceutical). And they may be appropriate as well for fluids that are costly to produce. The application areas are very broad and include the industries producing chemicals, oil & gas, personal care products, detergents, paper, plastics, food and beverages,

Photo 1. Off-shore oil and gas skid for condensate injection. High pressure triplex monoblock diaphragm pump delivers 8,750 gph at 2,200 psi with 150 hp input. The Pump Handbook Series

size flow ranges, with each drive element equally phased from the other to deliver smoother discharge flows. The trend is to horizontally multiplexed designs (as opposed to vertical) for mechanical, lubrication, NPSHR and space efficiency reasons. As process flow rates become larger, a compact monoblock triple-crank drive becomes more economical, does not require a baseplate, and is mechanically more reliable. Flow is adjusted over the required range by a variable frequency drive. Power ranges from 10 to more than 750 hp are common (Photo 1).

MECHANICALLY ACTUATED DIAPHRAGM DESIGNS These pumps are well proven and reliable. If care is taken in their selection and application, they will perform extremely well to specifications. The diaphragm forms a flexible barrier, totally containing the process fluid and separating it from the reciprocating drive element, which is attached to the diaphragm at the center with a clamping disk. A hermetic seal is made at the circumference of the diaphragm and at the clamping disk. Because the diaphragm is simultaneously loaded by deflection as well as by the pressure generated by the pump, these designs are limited to pressures below 300 psi. The force acting on the plunger rod is proportional to the fluid pressure and approximately to the area of the pusher disk in contact with the diaphragm, so relatively large drive thrusts are required. Conservative designs limit powers to 2 hp or less. Cam-and-spring (lost motion) drive mechanisms are often used in the smaller sizes (Photo 2), with variable eccentric (amplitude modulated) drives on larger sizes. In designs that use elastomeric diaphragms, the process pressure can deform the diaphragm and reduce the metering performance of the pump. Recent developments to overcome this problem include a multi-layer pure PTFE design (Photo 3) that offers broad chemical resistance, condition monitoring, a built-in backup diaphragm, and a stiffness against process pressure and service lifetime approaching that of a hydraulically actuated design. The major advantages of mechanically actuated diaphragm pumps are their simplicity (no hydraulic system required), high suction lift capability, and low first cost. Modern designs should be consid-

ered for low pressure processes where their advantages can be decisive.

of use as well as the design of the diaphragm support.

HYDRAULIC DIAPHRAGM DESIGN BASICS

HOW A PUMP HEAD WITH PTFE DIAPHRAGM WORKS

The diaphragm totally contains the process fluid, and the separate reciprocating plunger operates in hydraulic fluid. The volume of confined hydraulic fluid displaced by the moving plunger causes flexure of the diaphragm and the equivalent displacement of process fluid. Therefore, the freely moving diaphragm is stressed by deflection only, and there is equilibrium of pressures on both sides of the diaphragm and pressure uniformity across the full face of the diaphragm. The only moving parts in the process stream are one-direction automatic check valves to create an upward flow through the pump head, a fluid motion that follows plunger oscillation. For such a design to work reliably for extended periods, a number of conditions must be met: 1. Constant volume of hydraulic fluid in the pressurized circuit by automatic leakage replenishment. 2. Continuous venting (degassing) of any vapor in the hydraulic fluid. 3. Relief of excess pressure to avoid overloading. 4. Protection of the diaphragm against cavitation, or excessive stresses from high suction pressure or improper operation of discharge valve. Additional important features of this kind of pump preferred by users include: 1. Multi-layer diaphragm for secondary containment. 2. Continuous monitoring of diaphragm condition. 3. Continued operation on one diaphragm layer for controlled shutdown or switch over to a standby pump without process interruption. 4. Fast and simple maintenance for minimum downtime. Modern diaphragm pumps provide all of these features using either PTFE diaphragms for pressures up to 5,000 psi or metal diaphragms for pressures to 17,000 psi and higher. While there are many diaphragm shapes and designs, the most commonly used are plain flat diaphragms which are considered optimal with respect to quality, economy and ease

Pumping Stroke (Figure 2) The process fluid (3) is totally contained by the rigid diaphragm cover (1) and the flexing pumping chamber wall formed by the hydraulically balanced PTFE diaphragm (2), which is clamped with a hermetic seal at its circumference. This prevents any moving or sliding seals from being wetted by the process fluid. The displacer is a packless plunger (4), which operates in clean hydraulic fluid with good lubricating properties and relies on piston rings to eliminate seal maintenance and adjustment. The volume of confined hydraulic fluid (5) displaced by the plunger on the forward stroke causes forward flexure of the diaphragm and the equivalent displacement of process fluid. One directional check valves (6, 7) create an upward flow through the pump head. On every forward stroke of the plunger, the vent valve (9) volumetrically "bleeds" a metered amount of pressurized hydraulic fluid back to the reservoir to remove any vapor bubbles that may form due to pressure and temperature fluctuations. This assures a stiff hydraulic system. At the same time a small, consistent slippage at the piston rings lubricates and cools this critical area for long service life. The built-in pressure relief valve (8) protects the pump against overload due to dead-heading. Also, since it functions in clean hydraulic fluid, it provides a reliable backup to a process side PRV, protecting the system from pump-generated over pressure. Suction Stroke (Figure 3) As the plunger moves rearward, the hydraulically coupled diaphragm follows rearward, dropping the pressure in the pumping chamber (3) to suction pressure and refilling this controlled volume with process fluid. On the hydraulic side of the diaphragm, the hydraulic fluid volume –being slightly reduced due to the metered venting and controlled slip at the piston rings – allows the diaphragm to reach and rest against the smooth, solid backup surface of the pump body (Photo 4) before the plunger reaches its

The Pump Handbook Series

91

Without Stroke Adjustment

b

Boxer Design

a

Photo 2. Cross-section of a heavy duty mechanically actuated diaphragm metering pump using a cam-and-spring drive element.

d

e Figure 1. Drive element configurations: a, b, c, and d (side view) can be multiplexed up to twelve. Monoblock triplex e (top view).

rearmost position. In this position the diaphragm is fully supported, and even pressures as great as 5,000 psi (resulting, for example, from high suction pressure or failure of the discharge valve) will not cause perforation or damage. The diaphragm mechanically opens the centrally located gate valve (10), permitting the refill valve (11) to replenish the stroke volume of hydraulic fluid precisely as the plunger continues its travel back to its rearmost point.

DIAPHRAGM POSITION CONTROL Diaphragm Position Control (DPC)

92

allows a freely oscillating diaphragm to be safely used, eliminating the need for a front perforated support plate. The arrangement of gate and refill valves shown controls the position of the diaphragm by continuously maintaining the correct hydraulic volume without overfilling. DPC always forces the diaphragm to start from the full backup support position, thus keeping it safely away from the inside of the diaphragm cover plate. In this way the diaphragm is also protected from any damage that might result from particles in the process stream. During periods of partially starved suction or transient cavitaThe Pump Handbook Series

PHOTO COURTESY OF AMERICAN LEWA, INC.

Triplex Drive Element

c

PHOTO COURTESY OF AMERICAN LEWA, INC.

Single Drive Element

With Stroke Adjustment

Photo 3. Multi-layer PTFE diaphragm with two working layers, monitoring layer, and backup diaphragm (U.S. patent 5,074,757).

tion (underpressure), the hydraulic system cannot overfill, since the diaphragm cannot travel far enough back to actuate the gate valve (10), and thus the refill action is blocked. Once the cavitation condition is cleared, the automatic venting action quickly restores normal performance and efficiency. If the pump is exposed to long-term underpressure periods – during shutdown, for example – hydraulic fluid may weep through the necessary clearances of the hydraulic operating components and gradually move the freely oscillating diaphragm out of position. It is important to understand that such conditions require ade-

12

7

8

1

9

2

4

PHOTO COURTESY OF AMERICAN LEWA, INC.

13

10 5 3 11 6 FIGURE 2.

1. 2. 3. 4. 5. 6. 7.

Diaphragm Cover “Sandwich Diaphragm” Process Fluid Chamber Packless Hydraulic Plunger Pressurized Hydraulic Circuit Suction Valve Discharge Valve

8. 9. 10. 11. 12. 13.

Pressure Relief Valve Vent Valve Gate Valve Refill Valve Breather Diaphragm Monitor Switch

13

12

7 1 2 10 5 3 11 6

8 9 4

FIGURE 3.

quate preventive measures in the installation and operating technique, especially at startup.

METAL DIAPHRAGM DESIGNS The diaphragm has multiple responsibilities: deflection, displacement, control of the hydraulic system, support of suction pressure (or even of discharge pressure if a discharge check valve does not close properly) and static sealing by the restricted compression principle. At pressures above 5,000 psi, it is neces-

sary to use metal diaphragms for secure, leak-tight clamping at the periphery. Also, metal diaphragms are stiffer and stronger. For its good toughness coldrolled chromium nickel stainless is normally used. Where special corrosion resistance is needed, higher alloys come into play. Metal diaphragms are designed for free deflection from the neutral position forward to avoid buckling stresses. To stay safely below the elastic limit and to maximize fatigue strength, the allowable deflection is about 10% of The Pump Handbook Series

Photo 4. Inside surface of pump head hydraulic side shows flush fit of gate valve at center to provide full support for PTFE diaphragm under high suction pressures or upset conditions.

that of PTFE diaphragms. Thus, pump heads become much larger and more costly for an equivalent flow rate. Unlike PTFE diaphragms, metal diaphragms are more sensitive to slight surface damage (scratches, grooves, dents), which become stress risers and shorten the operating lifetime of the diaphragm. A front (process side) support plate and the Diaphragm Underpressure Control (DUC) method is used for most of the installed base of hydraulically actuated metal diaphragm pumps. This is due to the high sensitivity of metal diaphragms to excess stress (beyond the elastic limit) and the fact that only very small deflections are permissible under any circumstances. In normal operation the volume between the front and rear support plate is slightly larger than the maximum stroke volume of the plunger, so the diaphragm stays safely away from the front support plate. This allows particles in the process stream to pass without damage to the diaphragm. DUC operates in a fashion similar to the DPC method, reliably refilling the slight reduction in hydraulic volume that might come through “underpressure." With a DUC system, however, no gate valve is present and excess hydraulic replenishment can occur in an upset condition. In this case, the diaphragm tries to move too far forward but is safely “caught” by the front support plate before any damage can

93

Figure 4. Hydraulically actuated pump head with “sandwich” metal diaphragms

occur. Development in this field continues, and one manufacturer has recently introduced a DPC system for metal diaphragms. For years single metal diaphragm designs have been available for pressures up to 17,000 psi. Dual metal diaphragms (Figure 4) are desirable for critical services such as the injection of highly flammable methanol on offshore oil and gas drilling platforms. The long term reliability and safety of these is well proven in installations involving pressures up to 11,000 psi.

LEAK FREE SAFEGUARDS Of major importance is a multi-layer diaphragm with condition monitor, as leakage or cross-contamination of the process fluid with hydraulic oil is highly likely in the case of the perforation of a single diaphragm. Two identical diaphragms coupled to act as one with essentially no space in between them is a widely proven design. Should one diaphragm rupture, the other is the backup, and the pump continues to operate at full performance without leaks while signaling the need for maintenance. A pressure rise between the two diaphragms gives notice of the rupture and is immediately sensed with simple, reliable devices that can trip alarms locally, remotely or both. Failure of both diaphragms simultaneously is rare and usually occurs only when there is an

94

extreme mechanical difficulty, such as solidification of the process fluid. “Sticking” the two diaphragms together by hydrostatic adhesion is a proven method of assuring that the diaphragms act as one even under the most difficult suction conditions. In this design, often called a "sandwich," any clear fluid compatible with the process is introduced between the two diaphragms at the time of pump head assembly, and all but a very thin film is squeezed out through a nonreturn valve upon the first few strokes of the pump. (The volume of fluid film in Figures 6 and 7 is overstated for clarity.) The film's important duties include holding the diaphragms together to act as one, enabling them to flex without abrading each other, and the rupture detector is prefilled for response with a few strokes. Vacuum coupling of two diaphragms is also widely used, but the normal drop in vacuum during operation can reduce the effectiveness of the coupling and pump efficiency. Once diaphragm perforation is sensed and indicated, nothing else happens—the pump continues to run at full performance and leakage free. Maintenance can be scheduled when convenient, switch-over procedures can be manually or automatically implemented to an on-line spare pump, and the pump to be shut down can be flushed with a neutral fluid, if appropriate. Even if both diaphragms of the sandwich are run to failure, the process is still contained. Process fluid can enter the hydraulic system but is prevented from leaking at a rate any more than a seepage from the breather (12) as it is stopped by the vent valve (9), packless plunger (4) , and gate valve (10). In this case, a pressure rise detector and shutoff valve at the breather is a reliable "fail-safe" solution.

DIAPHRAGM ENDURANCE The operating life of a diaphragm is influenced by many factors, the most important being correct selection of the pump and design for the service, proper installation in a suitable piping system, and operation in accordance with manufacturer's recommendations. Studies of up-times for a large range of pumps show that PTFE dia-phragms last to 20,000 hours or more in continuous, high pressure operation at pump speeds of 200 strokes per minute. Metal diaphragm designs can operate 8,000-10,000 hrs or more continuously, even at their highest operating pressures. Naturally, intermitThe Pump Handbook Series

tent operation will lengthen diaphragm lifetimes to many years, provided proper startup procedures are followed. Shorter lifetimes can normally be attributed to mechanical damage by the process fluid (solidification), induced mis-operation of the hydraulic system (e.g., wrong or dirty hydraulic fluid), or piping system problems such as excessive stresses from cavitation, pulsation or overload.

PUMP PERFORMANCE CONSIDERATIONS The volumetric efficiency of diaphragm pumps at high pressures is lower than for plunger pumps due to the compressibility and elasticity effects in the necessarily larger dead volumes in the discharge and hydraulic fluid working spaces. Energy efficiency, however, is typically better for diaphragm pumps as the high friction of plunger sealing is eliminated. This also eliminates the added heat in the process fluid, and it prevents particles of the seal from contaminating the process stream. Another energy advantage of diaphragm pumps is that a "start-up device" can be installed to recirculate the high pressure hydraulic fluid during start-up until the drive motor comes up to speed. This allows the pump to start without experiencing the process pressure and without use of the oversized motor and V/S frequency inverter that would otherwise be required. By-passing on the hydraulic side is also more reliable than on the process side, especially with slurries. The lower volumetric efficiency of diaphragm pumps at high pressures (i.e., 70-90% for diaphragm pumps vs. 9096% of plunger pumps) requires that more attention be given to the piping system and dampening devices, as pulsation and vibration excitation, even with multiplex pumps, can be significantly higher. The technologies of piping analysis and system interface devices (directional dampeners for flow, pressure and noise attenuation) has advanced to the point where degrees of smoothness approaching that of centrifugal pumps can be reliably predicted and achieved. Suction conditions with diaphragm pumps also must be thoroughly considered. NPSHR for diaphragm pumps is not much different from that of plunger pumps, but internal hydraulic losses must be added. Pump designs with perforated support plates will naturally have

higher losses. For continuous safe operation, diaphragm pumps with DPC should not be subjected to suction conditions below 8.5 psia.

CONCLUSION Diaphragm pumps for metering and process applications have developed to a high degree of reliability, safety and economy and they are available in a wide performance range. The successful application of diaphragm pumps requires a knowledge of their

function, limitations and characteristics. Partnering with a supplier that can provide a full range of products, references and services such as piping system analysis and recommendations (including system interface devices such as dampeners) as well as in-depth technical support is an important first step to achievement of design goals and overall lasting satisfaction. ■

LEWA, INC., Holliston, MA. He holds degrees in Mechanical Engineering from Wentworth Institute of Technology and Northeastern University, and has held positions in design, product development, sales, and marketing management. He has 33 years' experience with metering and process pumps and packaged systems using reciprocating metering pumps and rotary gear pumps.

Gerald Harting is Vice President and General Manager of AMERICAN

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95

POSITIVE DISPLACEMENT PUMPS HANDBOOK

Chemical Injection: Simplex or Complex? System requirements dictate type of installation, pump and accessories. By Jon Crowley hemical injection is a straightforward concept. It involves introducing relatively small amounts of dry or liquid chemical into a larger volume of liquid at a set ratio. Beyond that, the type of applications, variety of chemicals and sophistication of injection systems are virtually infinite. We will look at the reasons for chemical injection, its methods and equipment, and available options and guidelines for selection and specification.

C

CHEMICAL INJECTION APPLICATIONS Originally, chemical additives were used to facilitate health and safety in water and sewage treatment. This was followed by the need for corrosion control in steam power plants, and then generally for treating nearly every conceivable type of process flow in industry. Present-day industrial applications include the following: • Potable water chlorination and treatment. This consists of chlorine gas or liquid hypochlorite injection for mandatory disinfection. It also includes use of various chemicals for softening, coagulation, fluoridation, pre-filter flocculation and corrosion control to meet EPA and associated standards. • Wastewater treatment. Similar to potable water with more emphasis on flocculation and settling additives, the treatment of wastewater involves a variety of polymers and similar chemicals. • Boiler feedwater treatment. The goal here is to avoid boiler and turbine damage by adding phosphates for scale control and sulfites for oxygen elimination.

96

• Cooling tower treatment. This consists of adding acid and/or caustic to makeup water to adjust pH for scale and algae control. • Fuel/petroleum additives. Gasolines and lube oils require various additives such as dyes, TEL, gum inhibitors and pour depressants. The production process involves wideranging viscosities. • Corrosion inhibition. The field of corrosion control covers virtually any system involving piping or vessels carrying water, steam or waste. Elevated temperatures and pressures affect the type and amount of chemical used. • General industry process additions. Almost every industry requires specific chemical additives. Food products, plating plants, photo labs, chemical companies, and auto manufacturers all require chemical injection systems of some type for either specific products or associated streams of water, waste or steam involved in the manufacturing process.

DEVELOPMENT OF CHEMICAL FEED/INJECTION The concept of chemical injection dates back to the first uses of chemical additives in process flows. The early methods of proportional injection, however, were rudimentary. Static batch mixing of liquids and manual addition of solids to liquids sufficed for open-to-atmosphere systems. When pressurized flow arrived, the first answer was the pot "by-pass" feeder. This involved diverting a "by-pass" liquid flow around an orifice plate, then through a pot feeder containing dry chemicals. Given limited and known solubility rates, in addition to proportional differential pressure changes, rough proThe Pump Handbook Series

Photo 1. Skid-mounted industrial chemical injection system includes metering pump, tank, valving and accessories.

portional feed was achieved. As industry grew and the knowledge and use of chemical additives accelerated, the need for a specific type of low capacity, high pressure, adjustable rate injection pump became evident. Thus was born the proportioning, or metering, pump that for the past 60 years has accommodated 95% of the liquid chemical injection needs.

TODAY'S CHEMICAL METERING PUMPS Today's typical industrial quality metering pump is a positive displacement, motor driven, variable strokelength pump with a 10:1 capacity range from maximum to minimum setting. The general metering pump groupings are: • packed plunger, mechanically driven, adjustable stroke length • flat diaphragm, mechanically

driven, adjustable stroke length • flat diaphragm, hydraulically driven, adjustable stroke length • tubular diaphragm, hydraulically balanced, adjustable stroke length Most types of metering pumps are available with variable stroke length adjustments (while in operation), constant or variable speed drives, explosion-proof drives, stroke or speed controllers to respond to electronic or pneumatic signals and a multitude of reagent-end materials of construction to handle whichever chemicals might be involved. These materials include stainless steel, Hastelloy C, Carpenter 20 stainless steel, PVC, Viton and others. Capacities range from 0.1 gallons per hour (gph) to 2500 gph per pump head. Many models are supplied with multiplexed heads to expand the capacity range. Discharge pressure capabilities range from 20 - 4000 psi, but always in inverse ratio to the gph capacity (i.e., the higher the pressure capability, the lower the volume capacity, and vice-versa). Some metering pumps feature turndown ratios (maximum rated capacity vs. minimum repeatable delivery rate) of 100:1, using speed controls only. Achieving a turndown ratio of 100:1 without the use of stroke control has been a recent breakthrough in metering pump technology. This development has allowed for precision feed rates in a variety of processes. Various solenoid-driven, mechanically activated metering pumps have been developed during the last 20 years to meet certain low pressure applications. These pumps often employ a plastic body and are frequently tank-mounted. Pumps of this design are frequently used to vary feed rates in concert with 4-20 mA signals from other instruments. While maximum pressures from these pumps seldom exceed 150 psig, they are quite effective in metering liquids for specific purposes. Complementing metering pumps, dry feeders are available to inject chemicals directly into a process by gravity, in dry form. These feeders are of the bucket, disc, belt-gravimetric or loss-in-weight type and can be connected to and interfaced with metering pumps in various ways.

Figure 1. Chemical injection diagram showing typical pump installation

In addition to the primary chemical injection pump, failure of any of the associated accessories can impede the injection process. These accessories include the chemical storage/dissolving tank, mixer, piping, valving, level/flow control and various other automated electrical or SCADA (supervisory control and data acquisition)-type controls to assist the operator in adjusting feed rates from a central or remote location. These considerations emphasize the importance of overall system design and review.

DESIGN CONSIDERATIONS The subject of chemical feed systems must, by necessity, embrace a wide field of mechanical and electrical schematic considerations because of the variables encountered in the type of plant, chemicals used, quantities to be handled and conditions under which treatment is applied. Chemical state and storage conditions are major factors. Many decisions depend upon whether the chemicals to be injected are dry or in solution, packaged in barrels, bags, drums, cylinders or carboys, and, where large quantities are to be used, if bulk shipments are made by boxcar, gondola, tank car, tank wagon or truck. The design of a chemical feed system, therefore, may start with a materials handling problem that includes the unloading, storage and decision of whether to use the dry chemical and put it in solution at or near the point of application, or to purchase it as a concentrated liquid and handle it as a solution or slurry of unknown concentration. The economThe Pump Handbook Series

ic factors will affect the method selected, as will the type of feeding operation to be performed. Other design considerations include: • Degree of importance of the chemical and required accuracy. This will determine pump selection in terms of quality, feed range, mode of control, failure alarms and possible installation of a standby unit. • Pressure considerations. Both the discharge pressure into which the chemical is injected and the available suction pressure to the additive pump are critical. Proper relief valves to protect the pump against blockage and a back-pressure (anti-siphon) valve in the event of system depressurization are both vital. • Integrity of the related system components. While the chemical injection pump is critical, its successful operation is contingent mainly upon the system accessories. Consideration should be given to the idea of packaging the chemical tank, pump, mixer, piping and controls as an engineered sole-sourced packaged system. • Geographic and operating personnel considerations. An installation in a foreign country or any remote area prompts greater consideration of a packaged system with built-in backup components. Also, remember that even local installations have great disparities in availability and competence of operating personnel. A chemical injection system installed in a North American power plant most likely receives better maintenance attention than average overseas installations with personnel of unknown qualifications.

97

• Availability of replacement parts, local service and engineering assistance. This relates to the foregoing and is often a serious consideration relative to the extent of packaging and the selection of individual components (A key replacement part available only in an obscure country could present a problem.). • Cost considerations. All of the above factors are influenced by budget. Given the funds available, the design engineer must weigh the importance of the chemical injection system against other project priorities.

EQUIPMENT REQUIREMENTS In any chemical injection system for applications in the industrial or water and waste fields, certain basic equipment components must be identified, selected, purchased and brought together to operate cohesively. These basic components, in order of their presence relative to flow in a typical chemical injection system, include: • Chemical storage tank. Tank capacity must meet chemical delivery requirements. Materials of construction must be compatible with the chemical additive. The tank elevation and lateral distance from the metering pump are critical in assuring adequate suction pressure. • Level controls and gauges. Jobsite conditions determine the necessity for remote level gauges and alarms. In all cases, however, a visually calibrated and valved gauge glass should be provided. • Dissolving basket. Because chemical selections sometimes change and may be provided in pellet or granular form, a removable plastic dissolving basket of 1 or 2 cubic ft capacity should be supplied. The basket is placed inside the chemical storage tank. • Mechanical mixer. Many, if not most, chemicals require mixing, either initially or continuously. A motor-driven, slow-speed mixer with shaft of proper construction should be provided in many cases. • Floating seal. Some chemicals degrade upon exposure to the atmosphere. A plastic "donut" floating seal (1" thick) inside the tank may help. Coordinating the floating seal, mixer and dissolving basket in one tank may present a challenge.

98

Photo 2. Chemical injection systems in a municipal wastewater treatment facility

• Suction piping sizing considerations. The suction piping is critical, both with regard to materials of construction and sizing. Obviously, materials must be compatible with the chemical used, and pipe sizing must be adequate for free-flow to the pump suction intake. Distance is also a critical factor, as metering pumps have an entirely different flow curve from other pumps and require much larger suction piping. NPSH (net positive suction head) is a significant design parameter that is a function of piping size, suction piping length and viscosity of the liquid to be pumped. NPSH is considered a staple of systems design, and responsible metering pump manufacturers and representatives can assist designers with these considerations. • Suction strainer. The strainer is required to intercept dirt and sludge prior to its entry into the metering pump check valve system. Regular cleaning of the suction strainer is essential to prevent suction line blockage that might result in diaphragm rupture in some types of metering pumps. • Standpipe. An open-ended pipe located in the suction piping near the metering pump is used to help overcome friction losses in piping from the chemical tank. Adequate pipe height is critical. • Calibration column. Although not an absolute necessity, a calibration column is quite useful to the operator when attempting to verify the output of a chemical injection system (It has settled many jobsite The Pump Handbook Series

disagreements over the years!). • Chemical metering pump and drive. The pump must be selected for capacity, metering rangeability, pressure, materials of construction and type of control and drive. • Pulsation dampener. This device is commonly used to "flatten out" the pulsing delivery inherent in metering pumps, to safeguard against "hammering" effects in piping, and to produce steady, laminar flow. The need for a dampener depends on distances involved, the pump "stroking" rate and the sensitivity of the process performance to "pulses" in chemical delivery. • Pressure relief valves. A relief valve, relieving back to suction, on the discharge of each metering pump is mandatory to protect the pump in the event of line blockage. • Back pressure valves. Adjustable back pressure valves in the discharge piping are necessary to prevent overfeeding of chemical if a vacuum develops in the process system. • Injection devices. Varying from corporation cocks to open flow, the nozzle (or other method of chemical introduction to the process) must be carefully considered for maximum effectiveness and to avoid blockages due to crystallized buildup of chemical residue. • Timers and controls. This most dynamic feature of the chemical feed system will vary from system to system. Chemical injection systems range from constant feed, manual "on/off" controls to sophisticated flowresponsive systems with overriding

automatic adjustments for fluctuations in pH, chlorine demand and temperature. Careful study and integration of components is required.

CHEMICAL FEED SYSTEM PACKAGING The merits of specifying a completely pre-packaged system for shipping intact to the jobsite versus the accumulation of individual equipment items for installation by the end user or contractor are often discussed. Frequently, the assumption is made that end user control over the coordination responsibility results in cost savings, largely attributable to the apparently lower "upfront" equipment costs. A closer look may reveal quite the contrary due to these considerations: • The costs of preparing cohesive specifications addressing pressures, capacities, accuracies, chemical compatibilities, piping and accessories, and the cost of preparing

comprehensive, detailed CAD-quality drawings are borne by the end user or design engineer. • After completing the basic design requirements, the review of equipment submittal drawings, approval and release of equipment, expedition of delivery, assembly and coordination of components, integration of mechanical, electrical and structural subsuppliers and jobsite startup and training are all associated costs that must be calculated and included in any economic analysis by potential customers. • By nature, the infrequency with which most end users or nonspecialists engage in systems packaging inherently limits their familiarity with cost-efficient suppliers, and relative inexperience with the systems coordination process usually results in the expenditure of additional administrative time. Use of a systems integrator elim-

The Pump Handbook Series

inates the jobsite coordination of mismatched components, disputes among suppliers and general field startup problems. Overall costs are almost always reduced by significant savings on contract labor, overtime and efficiency in correlating components. Most major metering pump manufacturers make systems coordination available (either directly or through third parties) upon inquiry, and there are a select number of integrators nationwide that specialize in chemical feed systems packaging. ■ Jon Crowley is President of the Charles P. Crowley Co., a manufacturers' representative firm based in San Dimas, CA. He is the third-generation head of the firm, founded in 1932, which also specializes in packaged chemical feed systems for worldwide industrial and municipal use.

99

POSITIVE DISPLACEMENT PUMPS HANDBOOK

Built-in Relief Valves: The Case For and Against By James R. Brennan ositive displacement pumps need a pressure relief valve to provide a path for the pumped liquid to take if the discharge system is blocked, as when a discharge valve is closed or some other obstruction develops (Figure 1). Unlike centrifugal pumps, positive displacement pumps loose little flow with increasing differential pressure. Their "shutoff" head is essentially infinite. If discharge line blockage occurs and there is no relief valve in place, dangerously high pressure can result, placing the pump, driver and system components at risk of damage. The whole system can breakdown and property and people's lives can be in jeopardy. Many rotary positive displacement pumps are manufactured with a discharge pressure relief valve as an integral component. This valve is usually of the direct acting type, which diverts flow from the pump discharge to the pump inlet at some predetermined set pressure and overpressure.

P

the pump is suitably protected. At first glance built-in relief valves seem to offer a nice set of features to the pump user. There are, however, some prices to be paid for such designs. Nothing comes for free, and built-in relief valves are no exception.

VALVE CAPACITY The first issue in using built-in relief valves is valve size or capacity. Rotary, positive displacement pumps are used for moving liquids with large viscosity variations as well as at a variety of speeds. The size and shape of a built-in relief valve is dictated largely by the pump body configuration and relative location of the pump's inlet and discharge cavities. Valve performance will always be subservient to the physical constraints of the pump design. Usually, only one "size" valve is available to build into a given size pump. Frequently, the result is a valve design with many compromises and very

limited range of desirable operating capabilities. The most conservative of pump users – API members, for example – require valves to have set pressures above the maximum expected system operating pressure. The rule of thumb is 10 -15% above maximum system pressure. As a relief valve opens, there is an additional pressure rise from cracking pressure (the pressure that begins to open the valve) to full flow bypass pressure, where the valve is allowing bypass of the full pump capacity. Electric motor or other drivers will typically be sized for this maximum overpressure condition while operating at the lowest liquid temperature, which is the highest liquid viscosity condition. If the relief valve is undersized for the flow rate, the required driver power rating will be unnecessarily high. Also, the driver might be more expensive than necessary, and its operating efficiency at normal (low-

BUILT-IN RELIEF VALVES Figure 2 illustrates the internal construction of a typical built-in relief valve. A variety of styles and designs are available depending upon the manufacturer. As shown in Figure 3, however, there are common characteristics of operation. The advantages of having the pressure relief valve built into the pump are threefold. First, the cost of the valve is usually lower than that of a separately purchased valve. Second, there are no valve installation costs (labor, pipe and fittings). Third, the pump supplier is closely involved in the selection and operation of the valve – hopefully in such a way that

100

RELIEF VALVE

PUMP Figure 1. Typical relief valve configuration The Pump Handbook Series

surrounding environment. The pump body will not be able to dissipate any significant portion of this heat energy until the temperature is far higher than any reasonable allowable limit. This high heat can distort close-clearance pump parts as well as reduce the liquid viscosity below minimum levels needed to keep the pump in a reliable operating range. The liquid may flash to a gas as its vapor pressure is reduced at high temperature, and pump failure is one of several very likely outcomes. Therefore, full bypass of built-in relief valves must be limited to intermittent, very brief periods to avoid the substantial risks to pump and system.

OTHER ISSUES An array of other considerations must be addressed when using builtin relief valves. For example, most if not all built-in relief valves do not meet any recognized codes (ASME, for instance) or regulations concerning pressure protection. API rotary

Figure 2. Internal construction of a typical built-in relief valve

er) operating pressure unnecessarily low. The second consideration, perhaps the more serious one, is what happens to the liquid in the pump when a built-in relief valve bypasses. A rotary pump has an internal liquid volume 1/25 to 1/100 of its flow rate per minute. For example, a 100 gpm (378 l/m) pump, depending on type and design, might have an internal liquid volume of 1 - 2 1/2 gallons (4 10 liters). When a built-in relief valve bypasses 100% of the pump flow, the most likely operating point, the entire power rating of the pump is converted to heat input to the volume of liquid being recirculated. With built-in relief valves, the oil volume being recirculated is very small, and the power draw can be low or quite high depending on the size (flow) of the pump and valve overpressure rating. The end result is a nearly instantaneous temperature rise in the liquid to very high levels. Figure 4 presents a generalized picture of temperature rise versus percentage recirculation for a range of pressure levels assuming the liquid has a specific heat of 0.51BTU/pound/degree F, the approximate value for petroleum oil. The curve neglects heat radiation to the

INCREASING OVER PRESSURE INCREASING OVER PRESSURE

TEMPERATURE RISE

INCREASING INCREASING VALVE SIZE VALVE SIZE

INCREASING VALVE FLOW INCREASING VALVE FLOW Figure 3. Common characteristics of built-in relief valve operation The Pump Handbook Series

101

200 190 180 170

TEMPERATURE RISE, F

160 150

_ 2000 PSI

_

110

_ _

100

_ _

90

_ _

80

130 _ 120 _

70

140

110 _ 100 _ 90 80

60

1000 PSI

_

50

_ 40

70 _ 60 _

30

50 _ 40 _

20

30 _ 20 _

500 PSI 10

10 _ 40

TEMPERATURE RISE, C

210

150 PSI 50

60

70

80

90

100

PERCENT RECIRCULATION Figure 4. Generalized picture of temperature rise versus percentage recirculation for a range of pressure level assuming the liquid has a specific heat of 0.51 BTU/pound/degree F, the approximate value for petroleum oil.

pump Standard Number 676 specifically prohibits the use of built-in

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pump relief valves. Built-in relief valves usually do not have replace-

The Pump Handbook Series

able valve seats, so wear in this area will require remachining of the valve or valve replacement. Troubleshooting a low flow problem with a pump having a built-in relief valve is also a bit difficult as one cannot readily determine if the valve is bypassing or the pump is worn. Some manufacturers offer a built-on relief valve which, while having many of the same limitations discussed above, allows the user to pipe the outlet side of the valve either back to the source of the liquid (storage tank, for example) or at least well upstream of the pump in the inlet piping system so that the volume of trapped liquid is increased. These practices will help provide a larger volume of liquid to act as a heat sink for bypass flow power-to-heat conversion. Ultimately, the pump user community will decide which arrangement best meets needs on a case by case basis. This article is intended to provide users with some information about the trade-offs involved with built-in or built-on relief valves. ■ James R. Brennan, currently group manager for three pump divisions of Imo Industries specializing in crude oil transport pumps, is a 1973 MIE graduate of Drexel University in Philadelphia. He has more than 25 years experience with screw pumps at Imo Industries, is a member of the Society of Petroleum Engineers, and was engineering manager of a pump division for five years. He has authored many papers and articles as well as spoken at a number of industry conferences. He is also a frequent contributor to Pumps and Systems.

POSITIVE DISPLACEMENT PUMPS HANDBOOK

Applying the NPSHR Standard to Progressing Cavity Pumps Take NPSHR values into consideration to ensure proper continuous operation. By Michael Dillon and Klaus Vu¨ llings rogressing cavity pumps have characteristics unique to their class. These include high suction lift capabilities and the ability to transfer multiphase fluids. Their ability to handle vapors or gaseous materials, however, complicates the application of traditional NPSH values. To make full use of any pump's capabilities, it is crucial to take into account NPSH - Net Positive Suction Head. This acronym, which originated in the United States, is used internationally in pump engineering as a standard term for designating the net energy level in the pump inlet crosssection. Even though, due to their construction, progressing cavity pumps do not suffer immediate damage in the area where cavitation occurs, there will still be negative long-term effects caused by the reduction in flow and by pressure variations. Possible product damage, as well as a degradation of a pump's continuous operating capability, are inevitable if NPSHR values are not taken into consideration.

P NPSH

In simple terms, this acronym signifies: ■ the internal pressure loss of a pump, designated as the NPSHR or NPSH required of the pump, and ■ the inlet pressure available to the pump – that is, the pressure at the end of the suction line reduced by the fluid vapor pressure, designated as the NPSHA or NPSH available from the system. With this definition of the NPSHR value, cavitation will be within accept-

able limits if pump inlet pressure exceeds the fluid vapor pressure and the internal pressure loss (the NPSHR). The following formula must hold: NPSHA > NPSHR When planning pump systems, a safety margin, of 0.5 m (1.6') should be taken into account, to accommodate variances within individual parameters such as pump speed, fluid temperature, vapor pressure or inlet pressure. Moreover, a safety factor should be allowed for system changes over time, such as solids settling in the inlet pipe or erosion of the pumping components and a resulting degradation of the pump's NPSHR capabilities. These changes will not immediately cause an operational fault but can certainly cause problems later. To define the NPSHR as a pure pump characteristic, NPSHR is generally stated in the "meters" or "feet" as the level of fluid (water) head in absolute terms. The existing NPSHA is defined by the pump installation and calculated by NPSHA= (pi+pb-pD /d*g)+(vi2/2g)+H1.geo-Hj

where pressure in the inlet cross-section of the system, fluid level pb atmospheric pressure pD vapor pressure d density g acceleration due to gravity vi velocity in the inlet cross-section of the system H1.geo geodetic level Hj loss level

pi

As reference level, the horizontal level (Z1) will be defined as – in deviation from the definition for rotary The Pump Handbook Series

pumps (e.g., ISO 2548 or DIN 24260) – passing through the center of the pump inlet cross-section. This avoids the influence of an additional reference height.

PROGRESSING CAVITY PUMPS As is generally known, the progressing cavity pump was invented by Dr. Ren´e Moineau, following World War I, as a supercharger for an airplane engine. Due to the multiphase and high suction capability of the pump and its ability to convey large quantities of air, vapor or gas in a fluid, that may also contain solids, this pump is not usually used in the same way or considered to be the same as other rotary positive displacement pumps. The progressing cavity pump was designed to be a rotary and piston pump combination. It integrates the advantages of both types of pump constructions, such as high pumping flow rates, high pressure capabilities, minimal pulsation, valveless operation and excellent pressure stability. The progressing cavity pump distinguishes itself from other rotary pump types, in particular, by the fact that the external housing – in addition to its sealing function – is also the pumping element. This static pumping element, an elastomeric stator, is preferably designed with a compression fit for containing the rotating pumping element, the metallic rotor. In this way, fluids with very low viscosity can be conveyed with very little internal fluid slippage. Most other rotary positive displacement pumps require pumping media with a much higher viscosity than that of water in order to produce slip-free

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cross-section

v1 velocity in the measurement z1

cross-section height level of measurement point in relation to the entry cross-section

NPSHR STANDARDS

Figure 1. Cross-sectional view: progressing cavity pump of a modern modular type.

performance between the rotating and stationary elements. Since tolerance fits are integral to their designs, the resultant gaps and the inherent slippage in other pumps produce volumetric efficiencies below those of progressing cavity pumps. This is particularly the case for applications with higher differential pressures and lower viscosity. In addition, the higher internal leakage will induce a cavitation "gap" condition. As in all rotary positive displacement pumps, the progressing cavity pump volume is directly proportional to the speed and shows only a slight degradation due to the pumping pressure. In contrast to reciprocating pumps, as well as to gear pumps and peristaltic pumps, the progressing cavity pump has almost no pulsations. In this regard, NPSHR values in piston pumps show a particularly strong dependence on the mass forces of the pumping medium, originating from the pulsating flow. A periodic movement law, which is dependent on the displacement kinematics and the elastic characteristics of the pumping fluid, will be imparted to the liquid inside the pipes. The pressure pulsations thus generated in the pipe system must be taken into account. If low frequency pressure variations exceed the vapor pressure, cavitation will occur.

application parameters, define operational and fluid condition limits that will not be exceeded at a predefined NPSH value. In general, practical engineering – looking for an economical and operationally safe system – is interested in solving problems such as the deterioration of pump performance by cavitation and the consequences of such cavitation – i.e., noise emission, pulsation, vibration and wear. Generally, the most common indication of cavitation is reduction of the pumping height or the pumping flow (X%, ∆H or ∆Q). When such conditions exist, the amount of cavitation must already be significant since there must already be a significant volume of bubbles forming for the resultant effect to be noticeable. The standard procedures for measuring the cavitation sequence are designed to decrease the systemside NPSH value (NPSHA). Depending on the regulatory standard to be met, the pump manufacturer will chose from one of the following methods: ■

■ ■

CAVITATION STANDARD If the pressure in a flowing liquid decreases at any point below the level of the liquid vapor pressure, bubbles will form, and these bubbles will implode intermittently when they contact areas of the pumped fluid that are under higher pressures. Manufacturers can experimentally define a series of cavitation scenarios and, depending on the customer's

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The inlet pressure of the fluid entering the pump will be controlled by means of a throttling valve. The pump will be fed by a vessel with an adjustable fluid level. In a closed circuit, system pressure or temperature (and thus vapor pressure) will be changed.

Taking into account the real measurement cross-section, the NPSHA value here is calculated in the same manner as NPSH: NPSHA = (p1+pb-pD/d*g)+(vl2/2g)-Hj+-z1

p1

where pressure in the measurement The Pump Handbook Series

Varying standards among international regulatory institutions regarding the definition and testing of NPSH values can cause confusion about the correct way to operate progressing cavity pumps in relation to their suction capability. The Hydraulic Institute recognizes the difference between the general classification of rotary positive displacement pumps and reciprocating pumps or centrifugal pumps. For this reason, a separate standard as well as a test procedure for the required pressure at the pump inlet (NPIP=Net Positive Inlet Pressure) has been provided. However, discrepancies between the institutions and even pump manufacturers arise when the percentage of the reduction is to be determined. The German VDMA (Verein Deutscher Maschinenbau – Anstalten-Association of German Engineering Institutions) states in VDMA 24284 "Testing of Positive Displacement Pumps" that there is to be a decrease of pumping flow by no more than 2%. This is the same standard set by the Hydraulic Institute for centrifugal and positive displacement pumps. Yet API (American Petroleum Institute) describes in API 676 "Positive Displacement Pumps Rotary" a reduction by 3%. The Hydraulic Institute states additionally that the NPIPR for rotary positive displacement pumps is to be measured at a 5% reduction of the pumping flow. Users need to know which standard is being used by the pump manufacturer to calculate NPSHR.

NPSHR CHARACTERISTICS As described in the section on NPSH, the NPSHR value can be understood as the internal pressure loss of the pump, and this pressure loss can be read off directly from the Q-Ps characteristic curve. This criterion is frequently used in connection with reciprocating pumps. The pressure loss will be determined experimentally, by measuring the pressure inside the pump-

30 —

28 —

30 psi 60 psi 90 psi

26 —

24 —

22 —

20 —

18 —

16 —

14 —

90 psi

Capacity US GPM

10 — —6 60 psi

8—

30 psi

6—

—4

4— —2 2—

0— l 0

Starting Torque 66 lb. ft. l 100

l 200

l 300

l 400

l 500

l 600

l 700

l 800

l 900

—0 l 1000

Absorbed Horsepower

—8

12 —

damping acting of the elastomeric stator. If the bubbles remain intact while the pumping chamber is being created, the chamber will be formed and progress through the pump in a partially filled state, and this will decrease the flow. When the chamber opens on the pressure side of the pump, the bubbles will condense instantaneously. The consequent back-flow causes pulsations the strength of which is determined by the volume of vapor. The Q-Ps characteristic curve of a progressing cavity pump is dependent on both speed and fluid viscosity. In the lower speed range there will be a relatively immediate transition from partial cavitation to full cavitation for highly fluid media. At higher speeds this transition is more gradual. For high viscosity fluids the transition will be progressive across the entire speed range, and will frequently begin as early as at the point of entry into the vacuum range. The onset of cavitation will also be influenced by the type and characteristics of the fluid. For instance, with mixtures such as hydraulic oils, cavitation will start slowly and then become amplified across a wide boiling or vapor phase range. The Q-Ps characteristic curves of progressing cavity pump sizes have been recorded in comprehensive test runs. From the results it is possible to derive pump-specific coefficients that take into account the positive displacement geometry and the individual manufacturer's design 10 —

Figure 2. Characteristic curve of a progressing cavity pump

NPSH = ∆ p/ d*g The procedure described here can be used for progressing cavity pumps as well; however, the pressure measurement drilling that needs to be carried out does not permit any non-destructive testing of production pumps. By recording the pressure curves in the pumping chambers of a progressing cavity pump, any resulting cavitation phenomena can be analyzed. When a new pumping chamber is created, rotation of the rotor within the stator will cause a sudden opening of the chamber on the suction side, accompanied by a strong decrease in

the cavity pressure. Depending on the actual level of inlet pressure, speed or fluid viscosity, this pressure decrease may fall below the vapor pressure level for a defined time and thus determine the volume of the cavity filled with bubbles. If any bubbles implode in the area of higher pressure as early as this suction action, they will not affect the Q-Ps characteristic curve. However, with low viscosity fluids this situation leads to pressure variations. The resulting oscillating behavior and increase in noise can be considered as a first indication of cavitation, even though this phenomenon – when compared to other rotating positive displacement pumps – frequently is not as marked due to the The Pump Handbook Series

9— 8.5 — 8— 7.5 — 7— 6.5 —

NPSH (m)

ing cavity. The NPSH value will then be calculated as follows:

9.5 —

6— 5.5 — 5— 4.4 — 4— 3.5 — 3— 2.5 — 2— 1.5 — 1— 0.5 — 0— 0

1

2

3

4

5

6

7

8

9

10

Factor

Figure 3. NPSH curve

105

of the suction components. For the normal direction of rotation, construction of a hydraulically favorable joint seal and the length and design of the suction casing are both important. Within the turbulent flow range, the NPSH value will be independent of the pumped fluid, and the pressure loss will be in proportion to the density. For each pump size, using the flow rate and this coefficient, it is possible to calculate an NPSHR value, as is shown on the curve in Figure 3.

ue will be improved by the undisturbed suction conditions. This mode of operation eliminates the influence of the 90˚ angle inside the suction casing as well as the restrictions from the moving rotor joint on the inlet side of the pumping elements. The influence of the moving joint, in particular, represents the primary restriction, as frequently a 90˚ angle must be installed in the pipeline to make it possible to run the pump in reverse.

REVERSE DIRECTION OF ROTATION

CAVITATION DAMAGE

For applications in which the NPSHA value is very low, the performance of progressing cavity pumps rotating in reverse should be considered. The pumping medium flows into what is normally the high pres-

In long term studies we have made, progressing cavity pumps were subjected to many different cavitation conditions, and the pumps have proved themselves to be relatively insensitive to cavitation damage. In addition to the pump-specific cavitation phenomenon, minimization of damage is essentially due to the combination of materials in the pumping elements and the effect of the elastomeric stator. The stator acts as a shock absorber. Pressure surges caused by bubble implosion are better absorbed in this design than by a pump made from an inelastic metal or synthetic materials. Moreover, the smooth injection-molded stator surface provides a poor working surface for the currents that occur during cavitation. If casing walls are rough, these "microjets" shoot into fine grooves and cracks. By selecting an appropriately low strength elastomer quality, damage as a consequence of cavitation could be confined to the first third of the stator pitch - when viewed from the suction side. Cracks in the elastomer will appear, and these, in connection with the mechanical load due to squeezing, will cause – in part – material to peel off across a wide surface area. Photo 1 shows this cavitation damage across three phases. A total blocking of the suction line will cause damage in the entire area, and this will be similar to the damage generated by running the pump dry. In contrast to dry run following overpressure, however, the stator end will keep the original form. In addition to a temperature rise in the entire pump area, extreme cavitation will increase the power consumption of the drive due to temperature-dependent material expansions. Photo 2 shows a typical dry run which - in

Photo 1. Cavitation damage sequence

Photo 2. Dry run damage sequence

sure side of the pump and is discharged from what is normally the low pressure side. The advantage of this is that shaft sealing will be on the pressure side, excluding any possibility of losing prime through the packing or a mechanical seal. A further advantage is that the NPSH val-

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The Pump Handbook Series

contrast to cavitation damage - starts from the center. During cavitation tests no damage to transmission components was found within the full cavitation range. Strong variations in pressure and the resulting thrust load shocks would normally damage universal joint connections. The design of the joint to absorb these loads is an important consideration if low NPSHA operation is likely. The use of high quality materials with sufficient cross section designs to avoid fracturing is essential. Designs that include thin locking keys, brass or bronze or cast iron components or flexible shafts represent increased risk. Earlier investigations have shown that progressing cavity pumps can continue to operate without damage to the elastomer when only 10% of the flow through the pump is really a fluid. In relation to the NPSHR, this capability – when converted into a test standard – would allow a reduction in the pumping flow of around 90%. This standard could not be used for most other pump types, because of the destructive effect of cavitation on the internal components.

SUMMARY AND CONCLUSION Progressing cavity pumps are a unique type of rotary positive displacement pump. In addition to their ability to pump very thin fluids, they can pump multiphase fluids with an extremely high vapor or gas content. The typical standard of 2% or 3% reduction in pumping flow to determine the NPSHR may not be a prudent criterion for a decision against the use of a progressing cavity pump – if you consider that this type of pump can operate well below the listed NPSHR without suffering physical damage. Aside from the low speed operation, a progressing cavity pump incorporates specific design features and operational characteristics that minimize the damaging effects of cavitation. While the consequences of cavitation may be acceptable far below the 2% or 3% criterion, it may be better to use criteria such as noise level or vibration, particularly for larger pumps, even when there is no visible reduction in flow or pressure. A general application of a criterion that allows for a 50% reduction in flow would mean that dramatic lev-

els of pulsation, vibration, and noise must be accepted. For operators, this is not a good practice. The extent to which cavitation can be allowed is mainly dependent on the oscillation behavior and the noise emission of the pump and plant system. However, it also depends on fluid characteristics, the loss of performance that is deemed acceptable, and whether this condition is intermittent or continuous.

LOOKING AHEAD Currently, our lab is proceeding on systematic investigations for nonNewtonian fluids, whereby a criterion that is directly based on the inlet pressure loss of the pumps (as described earlier) is used. In the course of these investigations, new flow coefficients will be experimentally determined and optimized in relation to existing as well as newly developed pumping element geometries. ■ Michael Dillon is Vice President & General Manager of seepex, Inc. (Dayton, OH). A graduate of Ohio University, he has more than 23 years of experience in research, marketing, sales and general management of progressing cavity pumps. Klaus Vu¨llings is Director of Research and Standards, Seeberger GmbH+Co. (Bottrop,Germany). He has worked in the applications, research and development of progressing cavity pumps for 13 years.

The Pump Handbook Series

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POSITIVE DISPLACEMENT PUMPS HANDBOOK

The Canned Rotary Pump - Circa 1997 A new canned positive displacement 3 screw pump shows promise in handling corrosive liquids not suitable for a wet stator design. By James R. Brennan n the continuing effort to produce more reliable, leak-free pumping capability, the concept of a “canned” motor/pump has now been extended to the rotary, positive displacement three screw pump. “Canned” is the term applied to the motor/pump package in which the pump, electric motor and pumped liquid are completely contained within a closed vessel (can) such that no shaft seals are required. Canned centrifugal pumps have existed for many years and are used to pump toxic liquids, radioactive waste water and other liquids that pose serious risks should a shaft seal failure occur. Many fluids, however, are more viscous than can reasonably be handled using centrifugal pumps. Additionally, many applications lend themselves better to the constant flow, self priming, predictable performance characteristics of rotary, positive displacement pump technology. In 1995 the Imo Industries Pump Group completed its initial development of what is believed to be the world’s first production

I

INLET

canned rotary motor/screw pump package. Some of the development work was done in close cooperation with a long term customer seeking leak-free operation for a difficult, high inlet pressure pumping application. Three screw pumps were chosen not only for their versatility but because they can be driven at direct connected electric motor speeds in the 1500 to 3600 rpm range and do not normally require speed reduction devices such as belts or gear reducers. Figure 1 illustrates a cross section of a canned three screw pump and motor. All joints exposed to the pumped liquid are o-ring sealed; there is no shaft seal. Note that the pump shaft is drilled through, as is the electric motor shaft. A small flow of the pumped liquid escapes past the balance piston, as well as through an orifice between pump discharge and the motor, and circulates through the motor. This cooling flow returns to the non-drive end of the motor shaft before reaching the pump inlet chamber via the drilled passage. The

OUTLET

BALANCE PISTON Figure 1. Cross section of a canned three screw pump and motor

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The Pump Handbook Series

flow, in the order of 0.5 gpm (2 l/m) depending on liquid viscosity and pump differential pressure, provides the normal cooling for the electric motor. The motor casing contains the pumped liquid essentially at pump inlet pressure. The pump balancing piston also acts as a sleeve bearing for the drive end of the electric motor. An open ball bearing locates the rotor at the non-drive end of the electric motor. All cooling flow goes through this bearing. Extensive design and application experience with submerged hydraulic elevator screw pumps and motors led naturally to the canned design for use external to the liquid supply tank or vessel. The design shown in Figure 1 is of the wet stator type. The stator winding insulation is exposed to the liquid being pumped and must, therefore, be chemically compatible with it. The current design is suitable for reasonably clean petroleum products such as lubricating oil at liquid pumping temperatures up to 225°F (107°C), depending upon elastomers, and a pumping viscosity range of 32-1500 SSU (2-325 Cst). The motors are currently available in 2, 3 and 5 hp configurations (1.5, 2.2 and 3.7 Kw) continuous duty rated, and can drive screw pumps to 50 gpm (190 l/m) at discharge pressures to 150 psig (10 BARg) or to 20 gpm (75 l/m) at pressures to 400 psig (28 BARg). Noise levels are as low as 60 dBA or less at a distance of 3 feet (1 meter), a very low level. The motors can be supplied in 2-pole and 4-pole configurations for 50 hz or 60 hz operation at

(photo courtesy of IMO Industries) (photo courtesy of IMO Industries)

Photo 1. Standard motor/pump package using tie rod construction

Photo 2. Rotor mounted to its bearing

(body) and mounting foot and cast iron for the end covers or end bells. Photo 2 shows the rotor mounted to its bearing and the stator/winding assembly. Steel cased pumps can also be provided for API type services; ports are usually SAE 4-bolt flange pads. The steel cased pumps have replaceable liners in which the pumping screws rotate. For API services the motor end bells are cast steel or machined from steel plate. A proprietary lead pass-through arrangement is used to permit the lead wires to exit the motor while retaining pressure tight integrity for the motor. The pump can be mounted to the motor in 900 increments about its drive axis allowing some versatility in porting location. Some of the pumps used with these canned arrangements also have independently rotatable inlet heads such that the pump inlet and outlet ports can be independently directed relative to each other in 900 increments as well. These motor/pump packages can be mounted with the drive axis in any orientation, including vertical and motor up or down. They are suitable for outdoor, unprotected locations and require no guards, shields, screens or other physical protection from dirt, rain, snow/ice or creature ingress. The use of 100 mesh (nominal 0.005 inch, 150 micron) or better inlet strainer, with a 0.5 psi (0.03 bar) maximum pressure drop when clean, is recommended to protect the package from excessive abrasive wear or damage. Since the pump is manufactured for only one direction of rotation, correctly wiring the motor to suit this direction is important. Unfortunately, no shafting is visible to show direction of rotation when jogging a motor on initial startup. Neither is there a method for “disconnecting” the pump and motor shafts so that direction of rotation can be determined before operating the pump. For these reasons, it is recommended that a phase and motor rotation tester be used to determine correct lead wiring before energizing the motor. Alternatively, inlet and discharge pressure gages can be used to determine direction of rotation The Pump Handbook Series

(photo courtesy of IMO Industries)

standard three phase dual voltages. Synchronous speeds are 1800 or 3600 rpm for 60 hz service and 1500 or 3000 rpm for 50 hz operation. The motors should probably be classified as TENV (Totally Enclosed, NonVentilated) although they are not designed to NEMA or other conventional motor standards. While no review has taken place, as long as they are kept purged of air or gas these units may meet explosion resistance criteria. Motors use an epoxy class F insulation, a system that can operate indefinitely at rated temperature. The motors carry an ambient temperature rating of 104°F (40°C). Normal industrial air cooled electric motors require power derating at altitudes above 3300 feet (1000 meters) due to the lower air density and poorer cooling capability. This derating is unnecessary for the canned motor because it is process liquid cooled. The standard motor/pump package uses the tie rod construction design shown in Photo 1. The pump shown is a cast iron cased unit using SAE 4-bolt flange pad inlet and outlet ports. The standard canned motor uses a steel plate fabrication for the frame

Photo 3. Canned three screw motor pumps on a skid supply high dielectric strength fluid to a high power, electric cable cooling system.

when the motor is jogged, provided the pump is filled with liquid and there is no foot valve in the inlet piping to obstruct reverse flow and thus risk over pressurization. Finally, the discharge piping can be temporarily disconnected, the motor jogged and the portion of the shaft visible through the discharge port watched to determine if rotation is correct. Observe the same foot valve caution. Applications to date include refrigeration compressor lubrication and power cable cooling. Both services are closed loop system designs that maintain high inlet pressure on the pumps and are thus susceptible to leaks from shaft seals when conventional pumps are used. Photo 3 shows three of six canned three screw motor/pumps on a skid supplying high dielectric strength fluid to a high power, electric cable cooling system. The pump inlet port (green piping) has been rotated 90 degrees from the discharge port (red piping) for ease of installation. The current canned motor/ pump package costs compare very favorably with conventional shaftsealed pumps flexibly coupled to standard industrial 3-phase AC motors. Further development is expected to result in additional, larger size motors as well as a broader range of both pump and motor materials of construction to suit a broader range of liquids. Dry stator canned motor/pumps may also be included in the upcoming phases of development. The dry stator design imposes a non-magnetic barrier between the stator and rotor, effec-

109

tively isolating the stator windings from exposure to the pumped liquid. With proper metallurgy, such a canned motor/pump can be used to handle corrosive liquids that are not suitable for a wet stator design.■ James R. Brennan is Manager, Crude Oil Pumps, for three operating units of Imo Industries Inc. (Monroe, NC). His responsibilities include worldwide marketing and technical support for crude oil pumping applications. Brennan is a 1973 graduate of Drexel University in Philadelphia, a member of Society of Petroleum Engineers (SPE) and has 27 years of service with Imo Industries. He is a frequent contributor to Pumps and Systems.

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The Pump Handbook Series

POSITIVE DISPLACEMENT PUMPS HANDBOOK

Metering System Design Requirements Understanding fluid properties as well as installation requirements for metering devices is needed to specify or design the right system. By Daniel J. Guthrie and Jeffrey G. Ives o specify and design a metering system correctly, engineers need to understand the requirements for a successful installation. Along with the properties of the fluid to be metered (flow range, viscosity and temperature) and installation requirements (pressure, power and space constraints), they need to be familiar with the various metering devices and the effect they have on flow characteristics and the process used. Metering systems can range from complex pumping systems with fully automated closed loop flow control to simple manual additions of a certain ingredient during a time period. For the purpose of this article, emphasis will be placed on metering devices claiming steady state accuracy of 1% or better. This would encompass piston or plunger, diaphragm and precision gear type devices. It should be pointed out, however, that other pumping devices in conjunction with control systems are being applied with success, especially in higher flow rates (above 1,200 gph) and reduced turndown ratios.

T

Reasons for Meter Flows The requirement for a metering system versus a standard pumping system is usually the result of two important considerations. The first and foremost is the amount of fluid being added to a process is critical to ensuring the quality control of the end product. The second factor is the fluid being pumped. Does it have an associated cost that would justify a metering device over a non-

metering device because overdosing the product would result in higher usage of raw materials and higher operating costs? While the control example still remains the major justification for a metering system, the age of continuous improvement (cost reduction and inventory control) has shifted batch processes to continuous processes, furthering the argument for lower rate and higher accuracy flow control. When asked to find a metering system for a fluid, develop an understanding of whether the need is for flow control or to conserve raw material costs.

Accuracy and Repeatability All manufacturers have a flow rate accuracy specification over a turndown ratio. Accuracy, sometimes referred to as steady state accuracy, is the ability of the metering system to maintain and repeat the desired flow rate over the flow range or turndown ratio. Repeatability is the ability of a device to reproduce a flow rate when returned to a set point. In continuous processing where production rates can vary, a metering system’s ability to repeat the flow rate accurately for a given condition is critical to product quality. Some standards allow a metering device to have stated accuracy based on two tests at rated or maximum capacity. These tests may not be acceptable at the lower end of the operating range. A +/- 1% unit rated at 100 gph would allow +/- 1 gph in variation. This rated flow meets the requirements, however, when operated at a lower flow rate a variation of +/- 1 gph would be greater than The Pump Handbook Series

+/- 1% of the lower flow rate. When designing a metering system, be sure to consider accuracy and repeatability over the entire turndown range. It is very important to specify accuracy over the flow range desired to be assured that your conditions are met.

Flow Characteristics in Various Processes Three metering system devices are used and have general acceptance in industry today. They are piston or plunger, diaphragm and precision gear. A piston or plunger metering system consists of a movable plunger inside a cylinder, as seen in Figure 1. The plunger is connected to a shaft that is operated in a reciprocating motion, allowing the cavity to open and close, displacing a controlled volume on each stroke. Such devices are well suited for lower viscosity fluids because fluids of a viscous nature can require higher inlet pressures and sometimes booster pumps to overcome flow losses through the inlet check valve. In batch processes requiring agitation, these units can be coupled with

Figure 1. Piston or plunger type reciprocating device

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a stroke counter to fill the required amount. In continuous processes, they will require a pulsation dampener or accumulator and an inline mixer to provide a more homogeneous output of the metered fluid as the flow characteristics are of a slug or shot feeding nature (Figure 2). Hydraulic or piston actuated diaphragm metering devices are reciprocating pumps that incorporate a flexible diaphragm as the pumping element. The diaphragm can be directly coupled to a piston or flexed by a hydraulic fluid to increase and decrease its volume depending on the stroke position (Figure 3). While the operating principles of a plunger and diaphragm pump are very similar, the diaphragm serves as the pumping chamber and is a flexible member. Diaphragm pumps operate well in dispensing with stroke counters and similar processes when operation is timed with the process. In continuous processes, a diaphragm unit is also well suited for lower viscosity as its output remains a controlled volume. Flow characteristics are still of a slug or shot feed nature, and some

AVERAGE RATE

FLOW RATE

TIME PLUNGER BACK DEAD MID FORWARD CENTER DEAD CENTER ONE ONE DISCHARGE SUCTION STROKE STROKE

NEXT DISCHARGE STROKE

Figure 2. Flow characteristics of single head reciprocating pumps Inlet Manifold

FLOW RATE vs TIME

SIMPLEX

DUPLEX

TRIPLEX

4. Flow characteristics of multiple head reciprocating pumps

processes could require a pulsation dampener or accumulator and an inline mixer. To reduce the effects of slug feeding, many diaphragm units offer multiple heads, and the flow characteristics are dampened by synchronized displacements (Figure 4). In cases where the fluid to be metered has a containment concern, the diaphragm type reciprocating pump has a distinct advantage since it is sealless and the fluid is contained within the pumping and piping chamber. A precision gear metering device operates in a similar respect to all external gear pumps. Unlike normal gear pumps, however, it is precisely ground to operate with running clearances of .00015 inches or less and surface finishes less than 4 rms to negate the effects of slip or internal leakage back to the inlet. In lieu of using inlet and outlet check valves to seal the liquid in a controlled volume, the device seals the fluid between the close clearances of the tips of the gear teeth and center housing and in tooth-to-tooth contact, as shown in Figure 5. These

Diaphragm Inlet Valve

Delivery Valve

Low Pressure Inlet

High Pressure Outlet

Vent Delivery Manifold Piston Crankcase

Figure 3. Diaphragm reciprocating type device

112

Figure 5. Precision gear type device The Pump Handbook Series

types of devices are well suited for all viscosity ranges and pressures since they have no inlet check valves to restrict high viscosity fluids from entering the pump or to keep the valves from seating. For very low viscosity fluids with high pressures, a review by the manufacturer should be conducted to assure that gear contact will be lubricated to avoid excessive wear. Precision gear devices have continuous motion, and each gear mesh displaces an amount of fluid equal to its rated capacity divided by the number of gear teeth. Some sets of teeth are meshing while other sets are opening and collapsing (Figure 6). Therefore, the gear teeth spread the controlled volume and discharging action over a much greater number of pumping actions. As a result, the flow characteristic depicted in Figure 7 produces a virtually continuous flow because of the rotary versus reciprocation motion of the controlled volume. In intermittent dispensing applications, the precision gear device can be coupled to a stepper motor to produce a shot flow characteristic. In batch processes, this device must be coupled with a simple counting control to fill the amount required. In continuous processes, the flow characteristics are of a continuous feeding nature. Therefore, pulsation dampers or accumulators are not required. Because the precision gear device produces continuous feed versus a shot or slug at its discharge, it produces a more homogenous mixture.

Effects of Devices on Pumping Fluids As with all pumping devices, metering devices put work or energy into the fluid to create a pumping action. This energy can cause some shearing action as well as a temperature rise depending on stroke or speed ranges. It is important to understand the fluid’s reaction to these conditions before considering various types of metering devices. With fluids that exhibit dilatent properties (increase in viscosity when exposed to shear), one should have a rheology curve developed before attempting to specify a metering device because the vis-

Flow Characteristics

Figure 6. Simultaneous pumping action of gear teeth

Typical Reciprocating Pump

Typical Precision Gear Pump

Precision gear pumps are normally provided with a closed loop speed control that constantly monitors input signals of desired flow rate with actual pump speeds to hold a set rpm. While automated speed control is the norm, these pumps are also available with manual operation. In all metering devices, simple manual controls up to and including automated PLC control systems are available to meet the user’s requirements.

Calibrating and Reliability

Each pump is delivering the same flow rate

Figure 7. Flow characteristics of metering pumps

cous properties could exceed the hydraulic properties in power requirements. Fluids exhibiting Newtonian properties (viscosity remains constant when exposed to shear) should not pose a power requirement problem so long as the viscosity is known at the operating temperature. Fluids exhibiting thixotropic properties (viscosity decreases when exposed to shear) should also have a rheology curve developed so as to not oversize the power required. This can result in increased capital cost in a system. In all cases it is important to understand the viscosity of a fluid at its intended operating temperature.

As with all process equipment, a calibration and check off should be performed under actual process conditions to assure that the equipment specified produces the desired result. A periodic verification schedule should also be put in place to verify expected conditions and detect wear. Many processes have adapted sophisticated process controls such as flow meters, rheometers and other devices to verify product quality. Depending on your process, these additional systems may be redundant to the metering device specified and would do nothing more than check a product that is as accurate. In depth discussions with manufacturers on proper scheduled maintenance and inspection or calibration methods could reduce your capital costs.

Summary An understanding of fluid properties as well as installation requirements for metering devices is needed for correct specification or design of equipment. Various metering devices produce different flow characteristics – a fact that provides advantages for intermittent dispensing, batch and continuous processes. The points discussed in this article should assist you in designing the correct metering device for your process.■ Diagrams reprinted with permission from ZeTech Manual, Zenith Pumps Division, Parker Hannifin Corporation. Dan Guthrie is the Business Manager for the Chemical Market for Parker Hannifin Corporation, Zenith Pumps Division. He is a graduate of Western New England College and has more than 12 years of experience in the application, engineering and marketing of precision gear pumps and systems. Dan has recently joined Roper Pumps. Jeff Ives is the Domestic Chemical Marketing Manager at the same company. A graduate of Michigan State University, he has spent the last 13 years involved in the application and marketing of metering pumps and systems.

Piston and diaphragm metering devices are normally provided with two types of flow control adjustment. The first is a stroke adjustment designed to change the length in which the reciprocating action occurs, therefore adjusting the volume of liquid displaced or acted on by the diaphragm. The second is a variable speed control for motor rpm. The combination of stroke and motor speed adjustment provides a large turndown ratio. Both devices can be provided with manual or automated speed control, where an input signal can be provided to operate at a desired flow rate or change with a mainstream process as product rates change.

(Photo Courtesy of Zenith Pump Division, Parker-Hannifin Corp)

System Control Considerations

Today’s metering pump systems feature digital operating technology, such as that found in the Zenith B-9000 Series. The Pump Handbook Series

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POSITIVE DISPLACEMENT PUMPS HANDBOOK

Sealing Technology for VOC Control The most effective approach is the simplest system that meets your application criteria and emissions control requirements. By Ken Lavelle and Bill Key n recent years mechanical seal manufacturers have developed extremely low emission seals in response to government regulations limiting emissions of volatile organic compounds from pumps. This benefits users by reducing product loss and increasing service life. Low emission mechanical seal designs generally adhere to API Standard 682: Shaft Sealing Systems for Centrifugal and Rotary Pumps. Low emission single seals and liquid/liquid dual seals are routinely achieving seal life of 5+ years while continuing to comply with emission limits. The Chemical Manufacturers Association (CMA) and the Society of Tribologists (STLE) conducted a study of leak rates from pumps equipped with mechanical seals. Their survey found typical leak rates of 0.0023 lb/hr (25 gm/day) for single mechanical seals with good face materials and good gasket materials. The CMA/STLE study concludes that: “Analysis of the data indicates that single mechanical seals can perform to meet the requirements set forth by the Environmental Protection Agency’s current and proposed future standards.” Pumps equipped with dry gas secondary seals are also showing low emission readings. Lift off dry gas secondary seals can theoretically last many years unless the lift augmentation devices become clogged with debris. Dual gas seals with a pressurized barrier gas show near zero emissions. This technology is especially suited for toxic fluids or when the process

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fluid provides poor lubricity.

Low Emission Single Seal Design Figure 1 shows a typical API 682 Type “A” Arrangement 1 single seal. Single mechanical seals provide reliable sealing for most VOC services when the following conditions are satisfied: • Fluid specific gravity > 0.45. • Vapor pressure margin in the seal chamber > 25 psi. • Flush fluid provides good lubrication of the faces.

seals often operate with seal chamber pressure close to vapor pressure. Seal face surfaces may be hotter than the fluid boiling point. In these cases the entire fluid film between the faces is a vapor. Well designed seals can run successfully with vapor between the faces provided that careful attention is paid to the following: • face deflections • seal balance ratio • face width • materials • flush • vapor pressure margin • recoverability from an upset A few of these considerations are covered in detail here.

Figure 1. Typical API 682 type “A”

A dual seal with a barrier fluid is recommended if these conditions are not met. Low emission single seals are designed to run with contacting faces. To minimize heat generation and wear, contact loading must be light. Face contact promotes longer seal life by minimizing the possibility of abrasive particles getting between the faces. Contact occurs on face high spots (asperities). A small amount of fluid migrates across the sealing dam in the spaces between asperities. Typically, this leakage is on the order of 1 gm/hr for low emission single seals. In flashing hydrocarbon services The Pump Handbook Series

Face Width Selection of face width involves contrasting criteria. A narrow face generates less heat, but a wider face is more resistant to pressure deflection and provides a longer leak pathway. A narrow face is more sensitive to O-ring drag effects. For narrow faces, O-ring drag is a larger fraction of net closing force. Net closing force consists of hydraulic closing force, spring loading and O-ring drag. O-ring drag can either add to or subtract from the axial closing force. Since hydraulic closing force is proportional to face area, O-ring drag has a larger relative effect on closing force for narrow faces. Seals with very narrow face widths on the order of 0.100” (2.5 mm) generally have a short life when used in flashing hydrocarbon services.

PRODUCT: PRESSURE: TEMP: SPEED:

0.218"

BUTANE 290PSI 190°F 3550 RPM

0.100"

CASA ∆T+178°F

Figure 2. Wide vs. narrow face analysis

An example is depicted in Figure 2. The service is butane at 290 psi. Narrow faces of 0.100” (2.5 mm) were found to last less than three days, with the carbon nose worn off. Replacement seals with the nominal face width of 0.218” ran for more than two years. Computer Aided Seal Analysis (CASA) predicts a large amount of pressure deflection for the narrow face, causing OD contact. The divergent gap creates low opening forces within the fluid film and, thus, high contact loading. The model predicts a face delta T (face temperature rise above bulk fluid temperature) of 58°F for the standard width face, and a face delta T of 178°F for the narrow face. Optimum face width to achieve both low emissions and long service life is a function of both seal design and service conditions. Properly designed and operated seals with face widths ranging from 0.140” to 0.280” can comply with strict emission limits and attain long life. Narrow faces can perform satisfactory at pressures below 10 atmospheres (150 psi), but should not be used in higher pressure applications. Recoverability A mechanical seal should be able to absorb short term upsets and recover to provide long life and low emissions. For example, during startup, the seal chamber may be filled with vapor only. This is referred to as “running the pump dry.” A laboratory test was devised to evaluate face materials under dry running conditions. Each test started with a 3-4 hour run-in on liquid propane at 225 psig (15.5 barg). Then pressure was decreased to about 160 psig (11 barg) to vaporize the seal chamber fluid. After two minutes of “dry running” at 3600 rpm, pressure was increased to 225 psig for a final hour on liquid propane. More than two dozen grades of carbon-graphite recommended by

vendors were run against a high quality silicon carbide mating face. The silicon carbide face survived this two minute dry running test with no visible wear. Of the carbon grades evaluated, only high quality antimony filled faces survived without blistering or other signs of damage. Two high quality resin impregnated faces experienced minor blistering. Other grades showed heavy damage. Flush Arrangement Seal flush is used to cool the faces and maintain bulk fluid temperature below the vapor point. In flashing hydrocarbon services it is common for pump suction pressure to be near the fluid vaporization pressure. Experience shows that satisfactory seal performance requires that seal chamber pressure be 25 - 50 psi (1.7-3.4 bar) above the vapor pressure at bulk fluid temperature. Special flush arrangements are used to provide an adequate vapor pressure margin in the seal chamber.

∆T (°F)

∆T (°C)

160

88

140

77

120

66

100

55 44 33 22 11

80 60 40 20 0

1

2

3

4

5

Test Time (hrs)

Figure 4. Seal face temperature increase during simulated dry running SINGLE PORT INJECTION

Emissions (PPM)

CASA ∆T+58°F

1000 800 600 400 200 0

108 PPM AVG

0

1

1000 800 600 400 200 0

2 3 MULTI PORT INJECTION

4

5

74 PPM AVG

0

1

2

3

4

5

Test Time (hrs)

Figure 5. Emissions readings single vs. multiport injection

port seal required about 20 minutes to return to a stable face temperature. Apparently, the single port flush is not very effective in eliminating the vapor pocket that forms during the dry running period. Emission rates for the two seal arrangements are shown in Figure 5. Average emissions were 74 ppm for the multiport injection and 108 ppm for the single port flush. The multiport arrangement ran considerably more stable, as evidenced by reduced spiking in emission levels.

Dual Seal Arrangements Figure 3. Multiport injection baffle

Multiport (Figure 3) versus single port flush systems were evaluated on a laboratory seal tester using propane as the process fluid. The seals were run on the recoverability test previously described. After a 3 to 4 hour run-in, pressure was dropped from 225 psig to about 160 psig to vaporize the chamber fluid. After two minutes of “dry running,” pressure was increased and a final hour run on liquid propane. Face temperature rise above bulk fluid temperature is shown in Figure 4. The multiport arrangement resulted in cooler and more stable face temperatures. After the two minutes of dry running, the single

The Pump Handbook Series

Dual seals consist of two single mechanical seals per seal chamber. A barrier (or buffer) fluid or gas usually fills the space between the primary and secondary seals. This fluid can be either pressurized or unpressurized. Dual seals with a buffer fluid nearly eliminate product leakage to the environment, achieving emission levels less than 10 ppm. Unpressurized dual seals are simpler to operate and generally more reliable than pressurized systems. Barrier/Buffer Fluids The choice of barrier/buffer fluid can have a significant effect on secondary seal performance. Figure 6 shows face temperature rise above seal chamber fluid for different barrier fluids. Data are for a 2.625” seal

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vapor handling system is provided, then emissions are solely determined by the effectiveness of the primary seal. There are two basic types of dry gas secondary mechanical seals: • Seals with flat faces (contacting) • Seals with lift augmentation (lift off or non-contacting)

Figure 7. API 682 type “A” Arrangement 3 standard seal Figure 6. Face temperature rise with different barrier fluids

at 3600 rpm, approximately 3 feet fluid pressure, carbon vs. SiC faces. The secondary seal was designed for no visible leakage. A 50/50 mixture of water/ethylene glycol ran with the coolest faces. Diesel and kerosene resulted in relatively low face delta T and a low wear rate. Diesel #2 has proved to be an excellent barrier fluid in refinery services. Automatic Transmission Fluid (ATF), not shown, resulted in very high temperatures and rapid wear. Optimum fluid viscosity is in the range of 1-5 cSt (at bulk fluid temperature). Higher viscosities generate more heat and may cause blistering of carbon faces. Viscosity should be at least 1cSt to provide adequate lubrication. Diesel fuel #2 works very well in field applications. One Southern California refinery installed 75 unpressurized dual seals to meet the area’s stringent emission requirements. The refinery is mostly using 5 gallon water cooled reservoirs with diesel buffer fluid. Both primary and secondary seals are low emission designs. Almost all of the installations have run more than 5 years and continue to be very low VOC emitters. Degraded diesel is disposed of by feeding it either to the refinery flare or into the pump process fluid. Dual Seals with a Pressurized Barrier Fluid Pressurized systems are used to minimize migration of process fluid into the environment. They are also used in applications where the process fluid provides poor lubricity for the seal faces or in which the process fluid may change frequently as in pipeline services. Both primary and secondary seals are lubricated by the pressurized barrier fluid. The

116

standard arrangement is for both seals to be in series (Figure 7), per API Standard 682. Circulation of the barrier fluid is by means of an internal pumping ring (API plan 53), or by an external pump or pressure system (API plan 54). Barrier fluid pressure is usually set to be at least 10% higher than the pump process pressure at the seal chamber throat bushing. (It should be at least 25 psi greater). Too low a barrier fluid pressure may result in process fluid migration into the barrier fluid. Too high a pressure results in heavier face loading and more rapid wear.

Dry Gas Seals Seal manufacturers have developed dry running secondary seals to minimize emissions to the environment and act as a safety backup if the primary seal fails. A dual seal with a dry running secondary seal is a simpler system (Figure 8) compared to liquid/liquid dual seals. Eliminated is the support system: barrier fluid, reservoir and provision for cooling water. Near zero emissions to the environment are attainable if the secondary seal faces are contacting and the space between the primary and secondary seals is vented to a vapor disposal system, such as a flare. If no Orifice .125” typical

Pressure or flow indicator or switch

Normally open To flare or vapor recovery

Check Valve 1-2 psi opening maximum P.I.

To drain or suitable enclosed recovery system

Sump w/liquid level indicator or switch .250” min

Normally Closed

Figure 8. Dry running backup seal piping (drain not required if product is a vapor under ambient conditions) The Pump Handbook Series

Figure 9. Dry running contacting seal face with hydropad pattern

Contacting Dry Gas Secondary Seals BW/IP has developed a contacting gas seal that incorporates OD and ID hydropads on the carbon rotating face (Figure 9). The hydropads promote cooling of the faces and provide a means for wear particles to migrate from between the mating surfaces. The carbon is a special grade that yields exceptionally low wear on dry gas operation. The stationary face is silicon carbide. Carbon wear is less than 0.0001” after 500 hours of testing on 5 psi propane vapor. Projected service life exceeds 3 years, provided flare back pressure is not above 5 psi for extended periods. Emissions are about 5 ppm on 5 psi propane vapor, and 10 ppm on 10 psi propane. Tests were also conducted in which the seal was suddenly subjected to 300 psi liquid propane. The seal holds, with no visible condensed vapor. Emissions, however, are on the order of 10,000 ppm when the pump is run on 300 psi liquid propane. The faces experience negligible wear on high pressure liquid propane, and the seal can be run in this mode for a period well in excess of two weeks. Several hundred of

these seals have been running well in the field for more than two years. Noncontacting Dry Running Secondary Seals Noncontacting seals ride on a film of gas and thus should have very low wear rates and run for many years. Lift augmentation is provided by incorporating features such as spiral grooves, T-slots, Rayleigh pads or circumferential waviness on one of the seal faces. Lift forces are produced when there is relative sliding between the two faces. Film riding backup seals are not intended to achieve near zero ppm emissions unless provided with a neutral gas purge, with both leakage and purge gas vented to a vapor disposal system. Some noncontacting seals can, however, achieve less than 1,000 ppm on low pressure vapors. SEAL DAM

WAVINESS AMPLITUDE

TILT

Figure 10. Wavy face type lift off dry gas seal

A seal employing a silicon carbide face with a circumferential wave pattern (Figure 10) has been developed by BW/IP International. Wave amplitude is 300 to 400 µin (7.5-10.0 µm). Rotation of the mating carbon face produces hydrodynamic lift. Laboratory tests show that the wavy seal operates in a noncontacting mode for gas pressures from 0 to 600 psi. Gas leak rates are one fourth to one half that reported for noncontacting seals using grooved faces. Tests to 500 hours show no measurable wear. Even though this is a noncontacting seal, emissions are less than 1000 ppm for secondary seal chamber pressures less than 25 psi. Lab testing shows that the seal successfully contains the sudden introduction of 300 psi liquid propane. More than 500 wavy seals are in the field in a variety of applications, mostly flashing hydrocarbons. In several installations the wavy sec-

ondary seal successfully contained the pumpage upon failure of the primary seal. In each of these cases there was no damage to the secondary seal faces; they were cleaned and reinstalled. Besides being used as a secondary seal, the wavy seal can serve as the primary seal in low pressure blowers to achieve emission compliance.

Dual Gas Seals for Pumps Several seal vendors market dual dry gas seals (Figure 11) for pumps that use a barrier fluid of pressurized gas, typically nitrogen. The barrier gas is maintained at a higher pressure than that for the process fluid. A small amount of barrier gas leaks across the primary seal into the pumped product; slightly more barrier gas leaks across the secondary seal to atmosphere. This arrangement assures that essentially no VOC’s leak to the environment. This technology is especially recommended for toxic fluids or for process fluids having poor lubricity for seal faces. The support system is more complex than that required for nonpressurized dual seals since a control panel is used and barrier gas pressure must always be maintained at least 25 psi above process pressure. Barrier gas must be a reliable source, such as plant nitrogen or instrument air. Gas consumption is on the order of 10-100 standard cubic feet per day, depending on seal size, pressure and shaft speed. This usage will deplete a standard nitrogen bottle (approximately 230 scf) within a few days. Thus this technology may not be suitable for remote sites, such as tank farms, that do not have an inhouse supply of inert gas.

Conclusion Figure 12 shows a relative cost comparison between VOC control sealing systems. This comparison is based only on initial purchase cost of the system. Factors such as the cost of plant utilities and mean time between failure must by analyzed to determine the actual system cost to the user. High

Dual Dry Gas Seal w/ Pressure Amplifier Dual Seal w/ Plan 54 Dual Seal w/ Plan 53

Relative Cost

Dual Dry Gas Seal w/ Control Panel Dual Seal w/ Plan 52 Dual Seal w/ Noncontacting Dry Gas Seconday Dual Seal w/ Contacting Dry Gas Secondary Single Seal w/ Plan 11 Flush

Figure 12. Relative cost comparison between VOC control sealing systems

Seal manufacturers today have developed a wide-ranging technology to meet VOC control requirements. The most effective choice is the simplest system that meets the application criteria and emissions control requirements. This will provide the highest reliability and ultimately the lowest operating cost.■ Ken Lavelle is Manager of Engineering and Product Development at BW/IP International Seal Division (Temecula, CA). He has more than 20 years of experience in the mechanical sealing industry and is a member of Pumps and Systems’ Editorial Advisory Board. William E. (Bill) Key is Manager of Research at the Seal Division of BW/IP International, Inc. He is a member of STLE Seals Technical Committee and the Advanced Projects Subcommittee, and is a former Chairman of the STLE Seals Course. He is currently responsible for development of new sealing technology and the generation of mathematical models of seal performance.

Figure 11. Dual dry gas seal for pumps

The Pump Handbook Series

117

POSITIVE DISPLACEMENT PUMPS HANDBOOK

A Users Guide to Rotary Pumps By Robert A. Platt, P.E.

otary pumps account for about 10% of the $15 billion world pump market. About $1.16 billion of this is for rotary pump completes – that is, pumps, drivers and auxiliary equipment purchased at the time the pump is purchased. Of this amount, gear pumps are the single most common type of rotary pump, and by themselves, account for over one third of the total. Aftermarket sales add another $400 million to the total. In this series we will be looking at rotary pumps from a user perspective, discussing the general attributes of these pumps and, when appropriate, comparing them to the more familiar centrifugal pumps (Part 1). We will then consider the features of rotary pumps and give guidelines for determining typical applications and how to identify the best pump for a

R

given situation.

What is a Rotary Pump? The Hydraulic Institute defines a rotary pump as a positive displacement mechanism consisting of a casing with closely fitted gears, cams, screws or vanes to provide a means for conveying a fluid. Its principal motion is rotating rather than reciprocating, and it displaces a finite volume of fluid with each shaft revolution. As the shaft rotates, the pumping elements form cavities that capture incoming fluid. The cavities open and close in a continual progression from suction to discharge in a smooth continuum of flow without pulsations or pressure spikes. As fluid fills the cavities, it also fills the clearances between them and the pump casing. This creates a liquid

Reciprocating 8%

Centrifugal, Turbine 82%

The $15 billion world pump market

seal that helps hold pressure and lubricates the pumping elements as they rotate.

Comparing Performance Curves Pumps move fluid by either centrifugal force or positive displacement. A rotary pump uses mechanical and hydraulic forces to create flow at a pressure equal to the system back pressure. Centrifugal pumps, on the other hand, create pressure. Unlike rotary pumps, they impart a velocity to the fluid and convert the velocity energy to a pressure energy as the fluid flows around the casing and out the discharge nozzle. The conditions of service will usually determine the best pump for an application. For instance, for constant pressure at varying flow rates, a centrifugal pump would be the best choice. An example of this is a municipal water system, where consistent pressure must be maintained over a wide range in usage levels.

A typical 6” flanged gear pump

118

Rotary 10%

The Pump Handbook Series

two screw

9,000 gpm/ 1,500 psi

vane

1,000 gpm/ 150 psi

other

1,200 gpm/ 150 psi

Standard industrial rotary pump performance

A rotary pump would be better for constant flow in the presence of varying back pressures. An example of this is an oil pipeline, for which system economics dictate constant flow rates regardless of any system pressure changes due to varying viscosity or pipe diameter. There are other differences between centrifugal and rotary pumps. The performance curves, affinity laws and the terminology used to describe rotary pumps are different (Fig.1). And, since rotary pumps are primarily for viscous fluids, the applications and markets served by these two pump types are also different. 100

Rotary

Head (feet)

80

Centrifugal

60 40 20

System

0 0

10

20

30

40 50 60 Flow (gpm)

Figure 1. Typical performance curves

The Pump Handbook Series

100,000

1,000 gpm/ 500 psi

50,000

three screw

10,000

1,000 gpm/ 1,000 psi

5,000

progressing cavity

1,000

1,200 gpm/ 500 psi

500

lobe

To develop its 100,000 pumping action, a centrifugal pump assumes the absence 10,000 of viscous drag across the impeller shroud 1,000 and vane surfaces. However, viscous drag forces developed 100 as the fluid passes Rotary Recip'g Centrifugal across these surfaces Figure 2. Viscosity handling ability of the major pump can be considerable. groupings With increasing 100% viscosity an increasing amount of energy must be expended to Rotary 80% overcome these forces and produce 60% the same amount of Crossover point hydraulic work. At 40% some point viscosity Centrifugal simply overtakes the 20% centrifugal pump – it can no longer over0% come the inertia and viscous drag losses of the fluid (Fig. 2). Since a rotary pump Viscosity (SSU) does not depend on Figure 3. Typical efficiencies as a function of viscosity centrifugal force for its pumping action, it from pump to pump and depends on is well suited for viscous fluids up to many factors. Generally it is 100,000 SSU. Some rotary pumps, between 700 and 1,000 SSU. In fact, such as twin screw pumps, can go efficiencies actually increase with even higher. increasing viscosity because the With rotary pumps each shaft higher viscosities are more effective revolution displaces a theoretical in sealing the clearances between the amount of fluid. In actual operation, pumping elements and casing. clearances must exist between the Rotary Pump Affinity Laws pumping elements to lubricate the rotating parts. Thus, a small percentThe affinity laws for theoretical age of fluid, known rotary pump performance with varyas “slip,” passes ing speeds are shown in Table 1. through the clearIt should be noted that for rotary ances from the pumps there are no considerations higher discharge for radial thrust loads as a function of pressure to lower position on the pump curve relative suction pressure. to BEP. Similarly, there is no rotary Compared to pump parameter for designers to use their centrifugal the way specific speed (NS) is used counterparts, with centrifugal pumps. Instead, rotary pumps are rotary pumps generally use flow per less efficient at revolution to make comparisons. lower viscosities, Other considerations such as flat but they are more versus steep curves or matching a efficient at higher system curve to a pump curve also 70 80 90 100 viscosities (Fig.3). can not be applied to rotary pumps. This crossover Instead, a rotary pump is selected for point will vary 100

1,200 gpm/ 500 psi

1,000,000

Viscosity (SSU)

gear

The Influence of Viscosity

Efficiency (%)

INDUSTRIAL HYDRAULIC RANGE

ROTARY PUMP

119

CORRESPONDING EFFECT

{rpm2 / rpm1} {gpm2 / gpm1} {rpm2 / rpm1} no direct effect on differential pressure {rpm2 / rpm1} {bhp2 / bhp1} {rpm2 / rpm1} {NPSH2 /NPSH 1}X where X varies from 1.5 to 2.5 Table 1. Rotary pump affinity laws for speed

VISCOSITY CORRESPONDING CHANGE EFFECT

V1 > V2

gpm1 > gpm2

V1 > V2

no direct effect on differential pressure

V1 > V2

bhp1 > bhp2

V1 > V2

NPSHR1 > NPSHR2

Table 2. Rotary pump affinity laws for viscosity

a given differential pressure and viscosity at the nearest commercially acceptable speed. For lower viscosities under this means that synchronous motor speeds, with the flow rate falling at the design point (hopefully) or slightly above it. Rotary pumps operate over a wide viscosity range, and simple generalizations are often difficult to make. Nonetheless, for changing viscosity with all else held constant, refer to the relationships shown in Table 2. One of the direct technical comparisons that can be made between centrifugal and rotary pumps is with single versus multistage designs. Even here, however, the analogy is not a perfect one, and certain rotary pumps such as progressing cavity and screw pumps fit the analogy better than others. Additional considerations – such as vertical versus horizontal mounting, metal versus non-metallic materials, sealless magnetically driven versus dual containment mechanical seals and conventional drivers – are pretty much the same whether you are considering a centrifugal or rotary pump.

120

General Rotary Pump Equations

Eoverall =

Rotary pumps are measured by their volumetric, mechanical and overall efficiencies. This often requires certain trade-offs when selecting a pump for a given set of conditions. It can also be confusing because a rotary pump does not have a single best efficiency point (BEP) the way a centrifugal pump does. Volumetric efficiency compares the actual to theoretical (geometrically calculated) output flow. It is an indication of both internal wear and a pump’s ability to handle a given differential pressure and viscosity. The more flow a pump can deliver under these conditions, the higher its volumetric efficiency will be. The volumetric efficiency Evol is defined as: Ptheo Emech = Ptheo + Plosses Ptheo = Pactual

{ {

}

}

{

Qactual Qtheo

} { x

Ptheo Pactual

}

=Evol x Emech

Suction Considerations and NPSH Required It is often said that a rotary pump can pump out anything that can get into it. Getting the fluid to the pump is another story. The same precautions regarding viscosity, suction line losses, fluid vapor pressure and NPSH Available apply to rotary pumps as they do with other pump types. (Fig.4) 100 Percent of Rated Flow

SPEED CHANGE

80 Low viscosity 60 Medium viscosity 40 High viscosity 20 0 14.7

12

10 8 6 4 2 Net Inlet Pressure Available (PSIA)

0

Figure 4. The effects of speed and

Volumetric efficiencies will be viscosity on NPSHR in the range of 70% to 98% for rotary There are several ways to state pumps. A changing volumetric effisuction conditions. Net Positive Succiency over time usually indicates a tion Head (NPSH) is the most comchanging viscosity or accumulated mon, although strictly speaking it pump wear. applies only to centrifugal pumps. Mechanical efficiency compares With rotary pumps the term Net Posthe actual to theoretical power input itive Inlet Pressure (NPIP) is prerequired. The less power input a ferred. However, to avoid confusion pump requires to produce a given in an area that already has more than amount of hydraulic work, the highits share of it, many manufacturers er its mechanical efficiency. Mechanjust use the units of Net Positive Sucical efficiency E mech of a rotary tion Head (NPSH) regardless of the pump is given by the equation: specific pump type. Qtheo - Qslip Vapor pressure is particularly Evol = Qtheo important when handling hydrocarbons and petrochemicals, which can Qactual have very high vapor pressures. For = Qtheo instance, low-sulfur crude oil vapor pressures can be as high as 100 psia under summer ambient temperaOverall efficiency is the most tures. Unless this is known when important efficiency because it alone selecting a pump, the results could determines the overall effectiveness be a pump that works in cold weathof a pump for an application. That is, er but cavitates in warm weather. knowing only the volumetric or The absolute pressure above mechanical efficiencies can result in vapor pressure available at the pump misleading conclusions. Only with a inlet must always exceed the satisfactory overall efficiency one absolute pressure above vapor prescan be reasonably assured of overall sure required by the pump. For satisfactory performance. Overall rotary pumps this pressure is deterefficiency Eoverall is defined as: mined by Hydraulic Institute stan-

{ {

} }

The Pump Handbook Series

dards similar to those used for centrifugal pumps. With the pump initially operating with a 0 PSIG inlet pressure and constant ∆P, temperature, speed and viscosity, a valve in the inlet line is gradually closed. This action continues until cavitation noise is clearly audible, there is a sudden drop-off in capacity, or there is a 5% overall reduction in output flow. The value at which any of these developments occurs is defined as that pump’s Net Positive Suction Head Required (NPSHR) for the stated conditions. Since flow is directly proportional to speed, it can be tempting to increase the pump speed above its maximum speed to obtain more flow. However, besides the obvious concerns about safety and component life expectancy, the effects of cavitation must be considered as well. As the cavities in a rotary pump rotate, they present a void(s) for the incoming fluid to fill. This void is available for a fixed amount of time. Incoming fluid drawn into the suction chamber must accelerate to fill this void in the available time. The higher the fluid viscosity, the more energy, in the form of an acceleration head, is required to accelerate the fluid to fill the void. If the fluid cannot fill it in the time available, partial vaporization and cavitation will occur. The best way to reduce NPSHR is to select a larger pump and run it slower. This will give the fluid more residence time to fill the voids on the suction side. And, with larger internal passages and ports, pump entry

losses will be reduced as well.

Other Rotary Pump Considerations

Feature Flow is independent of pressure

Benefit Predictable pump performance over varying system conditions

Pump Speed (% of Max)

Centrifugal pumps typically run at synchro- Large hydraulic coverage Almost impossible to find an application a nous motor speeds. Flow - Flows to 10,000 gpm, and head are fine tuned by differential pressures to rotary pump can’t handle such procedures as trim- 5,000 psi ming the impeller or underfiling the impeller Efficiently handles high Less system operating viscosity fluids – over cost as efficiencies actuvanes. 100,000 SSU in some ally increase with visRotary pumps, on the cosity (up to a point) other hand, operate with cases their internal dimensions fixed. Since they depend Smooth, pulse-free flow Less system cost – no need for vibration isolaon close running tolertors or vibration dampances for their output and eners efficiency, rotary pump gears, screws or lobes can- Self priming and will not Less maintenance intennot be trimmed the way vapor lock sive – no need to prime centrifugal pump impeller or re-prime vanes can. Instead, performance is usually fine Non-shearing pump Will not degrade shear tuned externally by speed action sensitive polymers and adjustment. petrochemicals There is usually no minimum speed for a Table 3. General features and benefits rotary pump. Certain vane pumps, which depend on with integral relief valves. While centrifugal force to draw the vanes these safety devices are a wise out of their slots, do have minimum investment, they are not a substitute speed requirements. Another excepfor external system relief valves. tion is with thin fluids at higher presLike the parking brake on an autosures, where at slow enough speed mobile, they are a secondary safety the slip can equal the theoretical disdevice of last resort and not intended placement, resulting in no net output for continuous duty. flow. To avoid this situation and Summary assure that the pump is running at a suitable volumetric efficiency, the We have discussed rotary pumps pump should always be run within in a general sense and covered the its recommended speed range. technical attributes that give them Because it is a their unique technical characterispositive displacement tics. The advantages that these 120% pump, a rotary pump pumps have over other pump types will work against any can be summarized in Table 3. 100% back pressure imAs we have seen so far, in any posed upon it. Operapplication involving viscous fluids, 80% ating against a closed high pressure requirements, selfdischarge valve, for priming needs, or flows rates virtual60% instance, it will conly independent of discharge pressure tinue building presone should consider a rotary pump. 40% sure until it damages But which one? itself or overloads the 20% Part 2 motor. Therefore, a pressure relief valve 0% I n this second part of our series on rotary pumps, we will focus on 500 1,000 5,000 10,000 50,000 100,000 500,000 must be in the system. gear, screw, vane, lobe and progressViscosity (SSU) Most rotary ing cavity designs – the most comFigure 5. Typical speed reduction as a function of pumps are available mon rotary pumps in use today. We viscosity The Pump Handbook Series`

121

(Photo Courtesy of IDP)

A 1 1/2” threaded gear pump

will examine the main features and benefits of each and the markets they serve, and we will provide basic guidelines for matching rotary pumps to given applications. The driver on most rotary pumps turns a shaft, which in turn physically meshes in some way with another shaft to form the cavities that move the fluid. This is known as an untimed arrangement. However, in some applications problems can occur with the gears, lobes or screws meshing in this way. For example, stainless steel gears will gall and seize if rubbed against each other. High wear rates will also occur if any dirt is trapped between the meshing lobes of a lobe pump, regardless of the material, and a pump with meshing screws will seize if allowed to run dry. The timed rotary pump was developed to overcome these problems. A timed pump uses timing gears positioned outside of the pumping chamber to transmit torque between the pump shafts and synchronize the pumping elements relative to each other. Preventing the elements from touching each other eliminates most of the problems caused by pumping dirty fluids,

Top Ten Features 1. High flow (up to 1000 gpm) 2. High pressure (over 300 psi) 3. High viscosity (over 100,000 SSU) 4. High suction lift (over 25 ft) 5. Dirty or abrasive fluid capability 6. Low fluid shearing 7. Quiet running 8. Direct drive throughout range 9. Integral relief valve available 10. Dry running for short periods

Figure 7. Typical internal gear pump Figure 6. Typical external gear pump

material compatibility and dry running. Most lobe pumps are configured this way. Gear and screw pumps can be timed or untimed.

Gear Pumps With drawings dating back to the 16th century, the gear pump is one of the oldest designs around – and one of the most popular because of its ability to handle a wide variety of applications. Gear pumps fall into two classifications: external and internal. External designs can be either timed or untimed. Their gear teeth cut on their external, or outside, diameter and mesh about the outside diameter (Figure 6). Bearings support the shaft at both ends with the gears located between the bearings. This resists deflection of the shafts, prevents contact between the gears and casing wall, and allows the pump to operate at higher pressure and with less overall wear. External gear pumps are capable

Gear Gear PC Three Vane Lobe Ext’l Int’ Screw ✔ ✔ ✔ ✔ ✔ ✔ ✔ ✔ ✔

✔ ✔ ✔ ✔

✔ ✔

✔ ✔ ✔ ✔ ✔ ✔ ✔

✔ ✔ ✔ ✔

✔ ✔

✔

✔

Three Screw Pumps Three screw pumps are untimed with a central driven screw (power rotor) driving two sealing screws (idler rotors). As the screws rotate, they capture the fluid between their

✔ ✔ ✔ ✔

✔ ✔ ✔

Table 4. Top ten features

122

✔

of flows to 1500 gpm, pressures to 500 psi and viscosities to 1,000,000 SSU. They are found in both clean and dirty fluid handling services serving OEM, refinery, tank farm, marine and API-related industries. The design of an internal gear pump incorporates one larger gear (rotor) with internally cut gear teeth that mesh with and drive a smaller externally cut gear (idler). Pumps of this type can be designed with or without a crescent shaped partition (Figure 7), and they have an overhung shaft arrangement. The internal gear pump is basically a low speed design (400 to 1200 rpm) with lower pressure limits than its external gear counterpart. Internal gear pumps are capable of flows to 1100 gpm, pressures to 225 psi and viscosities of up to 1,000,000 SSU. Primary applications include low pressure transfer of fuel oils, paints and various chemicals in the CPI and OEM industries.

Figure 8. Typical vane pump The Pump Handbook Series

threads and convey it along the length of the pump (Figure 8). These pumps can run at speeds as high as 5,000 rpm. However, because the screws are in virtual contact as they “float” on a viscous fluid film during operation, they cannot tolerate any dirt or abrasives. One of the most important markets for three screw pumps is supplying lube oil to large rotating machinery, where they typically provide flows to 400 gpm and pressure to 500 psi. Some pumps can operate at even higher (to 4000 gpm and 4500 psi) capacities, but they are beyond the industrial market we are discussing here.

Progressing Cavity Pumps The progressing cavity pump is one of the most unique of all rotary designs. Developed in the early 20th century for moving abrasive fluids, it can be characterized as more than a pump for dirty fluids, but not quite a slurry handling pump. Progressing cavity pumps are used in applications in which dirt and abrasives would wearout other types of pumps. All PC pumps consist of a rotor and stator, suction, discharge and bearing housing, and a mechanism to accommodate the eccentric motion of the rotor within the stator. An interference fit between a single helix (spiral) metal rotor and double helix elastomeric stator forms the chambers that create the pumping action (Figure 9).

turers are often difficult to make. The most common rotor material is chrome plated tool steel (at least 4 mils thick). Stainless steel offers better chemical resistance but is not as good for abrasion resistance. Stators are usually made of Buna N because of its low cost and excellent tear and abrasion resistance. Other stator materials used are Viton®, Teflon®, hypalon and natural rubber. Similar to a multistage centrifugal design, the number of stages in a progressing cavity pump determines its pressure capability. PC pumps usually have up to four stages, with each stage good for 85-90 psi. Pump life can be extended by limiting the pressure per stage and increasing the total number of stages. PC pumps for general industry can handle pressures to 350 psi, flows to 1,000 gpm and viscosities as high as 1,000,000 SSU. Operating speed is also important, and most PC pumps run in the 200-800 rpm range. They are used in the wastewater treatment, food, pulp and paper, chemical processing, mining and general OEM industries.

Vane Pumps Vane pumps offer flows to 1000 gpm and pressures to 125 psi. They use external sliding vanes rotating about a non-concentric arm. This arrangement draws fluid into the pumping chamber, moves it around the casing, and expels it out the discharge port (Figure 10). There are several types of vane

kerosene and similar light hydrocarbons. One of their biggest application areas is as truck mounted unloading pumps.

Lobe Pumps Similar to vane pumps, lobe designs can meet flow requirements to about 1000 gpm and pressure to 125 psi. Each rotor has one or more lobes, and the pump is similar in appearance to a gear pump with two or three teeth per gear. A lobe pump carries fluid between the rotor lobes in much the

Figure 11. Typical lobe pump

same way a gear or vane pump does (Figure 11). However, since the lobes cannot drive each other, they require timing gears. Lobe pumps are used to pump sludge in wastewater treatment plants. They are also employed in stainless steel execution for handling product in the food and beverage industries.

Conclusion

Figure 9. Typical progressing cavity pump

Rotor and stator geometry varies from one manufacturer to another, and with no standards to define connections or component sizes the way, for instance, ANSI B73.1 applies to end suction centrifugal pumps, comparisons between different manufac-

As we have seen in this introductory series, rotary pumps are among the simplest and most versatile of all major pump designs. Hopefully, you’ve become more familiar with their basic attributes and the applications where their best features can be exploited.■ Figure 10. Typical three screw pump

pumps. The most common is the sliding metal vane design. Another design has flexible elastomeric vanes and is used for dirty or chemically aggressive fluids. Vane pumps work best for low pressure transfer of gasoline,

The Pump Handbook Series`

Bob Platt is Manager of the Standard Products Group for IngersollDresser Pumps in Chesapeake, VA. He is a member of Pumps and Systems’ Editorial Advisory Board and a frequent contributor to this magazine.

123

POSITIVE DISPLACEMENT PUMPS HANDBOOK

Rotary Pump Inlet Pressure Requirements Shedding light on one of the least understood yet most important aspects of a successful pump installation By James R. Brennan, Imo Industries Inc.

ncorrectly specifying the required inlet pressure for a pump will result in either poor performance, noise, premature wear, high operating and maintenance expenses and failures, or a seemingly excellent installation that costs a good deal more than it should. Rotary pumps handle the broadest range of liquids of any generic pump classification – from molten metal, food, liquefied petroleum gas and sewage to asphalt, fuels, chemical slurries and plastics, polymers and pharmaceuticals. Capabilities and user expectations for rotary pumps are significantly different from those of other pump classifications. In the United States, the Hydraulic Institute is the major controlling organization for pumping definitions, and each of the centrifugal, rotary and reciprocating pump manufacturers have its own set of similar but not necessarily identical standards. Inlet pressure requirements for rotary, positive displacement pumps are similar to NPSHr (Net Positive Suction Head Required) for centrifugal pumps. For rotary pumps, pressure units are normally in force per unit area (psi, bar, MPa) rather than elevation difference (feet or meters). The variety of labels used for this parameter, as well as an astonishingly long list of units of measure and reference scales, perpetuates misunderstanding of required inlet pressure. The purpose of this article is to provide a basic physical understanding such that the various scales

I

124

and units do not, at least initially, intrude on our ability to grasp the principles. Every pump has a minimum required inlet pressure. What that minimum pressure is depends on pump type, size, speed and the viscosity of the fluid pumped. If the minimum required inlet pressure is not made available to the inlet side of the pump, cavitation will result. Cavitation is the incomplete filling or feeding of the pumping elements with liquid. This results in a reduction of flow and, if the condition is severe, noise, vibration, instability, internal erosion and catastrophic failure can result. Cavitation must therefore be avoided. Pure cavitation is the partial vaporization of the pumped liquid caused by allowing the fluid pressure to fall below its vapor pressure at the pumping temperature. Pseudo-cavitation can occur if the liquid contains dissolved gas or air – a not uncommon condition. The dissolved gas will expand as the fluid pressure is reduced and cause exactly the same symptoms as pure cavitation. Entrained gas or air in the fluid, such as can be found in some restricted or poorly designed lubrication systems, will also cause pumps to exhibit cavitation symptoms, as will an air leak in a pump inlet line below atmospheric pressure. The Hydraulic Institute defines minimum inlet pressure as that pressure, for a specified pump and set of operating conditions, that will result in a flow loss of 3% for reciprocating pumps, and 5% for rotary pumps and The Pump Handbook Series

5% FLOW REDUCTION (ROTARY)

PUMP FLOW OR HEAD

3% HEAD REDUCTION (CENTRIFUGAL) 3% FLOW REDUCTION (RECIPROCATING)

MINIMUM REQUIRED INLET PRESSURE

INLET PRESSURE

Figure 1. Hydraulic Institute cavitation definitions

a 3% head loss for centrifugal pumps while all other operating conditions are held constant. Most pump manufacturers accept these fairly arbitrary definitions as a condition that their pumps will tolerate indefinitely (Figure 1). It is, however, operation in a very mildly cavitated condition. Figure 2 illustrates what is happening to the pump above and below this minimum inlet pressure. The lower case letters in the diagram correspond to the horizontal axis locations in the graph. Pump manufacturers have conducted extensive tests and determined the empirical equations used to calculate the required minimum inlet pressures for their products. So from where does the required minimum inlet pressure come? It comes from either an upstream pump or atmospheric pressure pushing on the free surface of the fluid upstream of a pump in question. Atmospheric pressure can be the natural pressure exerted by the column height of air above the pump, or it can be the artificially maintained pressure above the fluid surface, such as a deliberately maintained

ATMOSPHERIC PRESSURE ON SITE

ABSOLUTE SYSTEM INLET SIDE PRESSURE

STRAINER PRESSURE DROP - DIRTY PUMP INLET PRESSURE NET INLET PRESSURE AVAILABLE MAXIMUM LIQUID VAPOR PRESSURE 0 a

b

c

d

e

f

g

h

i

j

LOCATION RELIEF

VALVE AND FITTING LOSSES VENT

BLOCK VALVE

e

STRAINER PUMP

g

f

i

h

j GRADE

ATMOSPHERIC PRESSURE AT JOBSITE ALTITUDE

STRAIGHT PIPE FRICTION LOSS d

MAXIMUM LIQUID VISCOSITY AND VAPOR PRESSURE

ELEVATION DIFFERENCE

ACCELERATION LOSS

a

ENTRANCE LOSS c

MINIMUM LIQUID LEVEL

FOOT VALVE

b

Figure 2. Factors impacting net inlet pressure available

vacuum or pressure in a process vessel. If natural atmospheric pressure is used, then the job site altitude above sea level is an important factor. Figure 3 illustrates the reduction in atmospheric pressure with altitude. Higher elevations have less pressure available for use in pushing fluids into a pump, and this often overlooked factor can make or break

an application. The idea that a pump can “suck,” while seemingly obvious, is in fact incorrect. The pressure reduction at the inlet of the pump is simply the result of frictional pressure loss due to the flow of fluid from its source to the pump and into VAPOR PRESSURE- BARA 0.5 1.0 1.5 2.0 2.5 130

20

15 FEET ABOVE SEA LEVEL10 X1000

6 5 4 METERS ABOVE SEA 3 LEVEL X1000 2

50 120 TEMP. °F 110

TEMP. °C 40

100

5 1

90 40 50 60 70 80 90 100 ATMOSPHERIC PRESSURE, %

Figure 3. Effects of altitude on atmospheric pressure

30 10 15 20 25 30 VAPOR PRESSURE- PSIA

Figure 4. Effect of temperature on vapor pressure The Pump Handbook Series

the pumping element(s). If fluids always remained in their liquid state, establishing the minimum required inlet pressure would be somewhat simpler. However, many liquids exhibit a vapor pressure of sufficient magnitude at pumping temperature – a factor that must be taken into consideration for proper pump operation (Figure 4). Vapor pressure is the inverse of boiling temperature. As we all learned long ago, water boils at 100ºC (212ºF). This boiling temperature is only correct when the water is at a pressure of one atmosphere (one international atmosphere is equal to 101,325 Pascals, 1.01325 bar, 1.03323 kg/cm 2, 14.696 psi). At an elevation of 3000 meters (9842 feet), the atmospheric pressure is only 69% of what it is at sea level. At this reduced pressure, water will boil at about 90ºC (195ºF). The inverse way of looking at this is to say that the vapor pressure of water is 1 atmosphere at 100ºC (212ºF). If you wish to pump water in its liquid state and the water happens to be at a temperature of 100ºC (212ºF), the inlet side of the pump must not be exposed to a pressure below 1 atmosphere or the liquid will begin to convert to a gas (steam) and the pump will enter a cavitating region of operation, a very undesirable condition. If liquid water is to be pumped at 160ºC (320ºF), then the inlet side of the pump must be maintained at or above the 6.1 atmospheres that represent the vapor pressure of water at this temperature. Many liquids handled at their normal pumping temperature exhibit such a low vapor pressure that this factor can be ignored for all practical purposes. Refined lubricating oils, for example, at normal operating temperatures up to 82ºC (180ºF) have vapor pressures in the range of 0.01 atmospheres. On the other hand, volatile liquids such as gasoline and alcohol will readily evaporate (boil) at ambient temperatures. Propane is kept liquid at ambient temperature only because it is stored in a pressure vessel. The vessel must contain the propane’s vapor pressure at the vessel’s temperature. Vapor pressure invariably increases with temperature. It is this very fact that

125

REQUIRED INLET PRESSURE

is put to use in refining petroleum and in many other petrochemical and chemical processes. While it is almost always important to know a fluid’s minimum vapor pressure at the maximum pump suction temperature, the pumping of high temperature fluids should involve a careful analysis of the possible impacts of vapor pressure. Some fluids will exhibit multiple vapor pressures. Raw crude oil is an example. This fluid is composed of many different complex molecules. It is a mixture and, as such, its component fluids will each have its own vapor pressure. The lowest discernible component vapor pressure is the one to use for net inlet pressure calculations if pumping this fluid. Alcohol mixed with water will exhibit two vapor pressures: that of alcohol and a different one for water. The way to separate these two liquids is to apply heat. The alcohol will boil first at a lower temperature than water’s boiling point. The alcohol can be collected as a gas, then cooled to its liquid state. This process, called distillation, is a good example of vapor pressure at work. Cavitation causes its damage by the abrupt, violent compression of the vapor (gas) back into liquid at the pump discharge. This compression occurs very rapidly as an implosion. There is enough energy to erode minute metal particles from the rotating and stationary pumping elements. Such erosion is frequently visible on outboard marine engine propellers in which the propeller velocity exceeds the water velocity, thus cavitating the blades. Given enough time, a blade failure is inevitable. Photo 1 shows cavitation

V4 V3 V2 V1

VISCOSITY

Figure 5. Effect of inlet velocity and viscosity on required inlet pressure VISCOSITY - CST 200

300

400

130

50 120

TEMP. 110

TEMP.

°F

40

°C

100

90 30 1000

1250

1500 1750

VISCOSITY - SSU

Figure 6. Effect of temperature on viscosity

damage to a pump. Most rotary positive displacement pumps use incremental pressure staging within the pumping elements to withstand differential pressure. Examples of this staging include multiple wraps of screw pump thread, multiple teeth on gear pumps and multiple vanes in vane pumps. This staging is only effective if the fluid pumped is nearly incompressible, i.e., a liquid. Introduction of gas, air or vapor causes the fluid’s compressibility to increase, and this compressibility defeats the staging benefits. Most of the pressure rise across a pump handling compressible fluids occurs at the last stage, overloading the unit. The minimum required inlet pressure of a pump also depends upon its size and speed. The product of size and rotational speed is velocity. Fluids moving at high Photo 1. High pressure gear pump destroyed velocities entering the pumpby severe cavitation

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The Pump Handbook Series

ing elements will consume more energy (pressure) than slower moving fluids. Consequently, large and/or high speed pumps will require a higher minimum inlet pressure than smaller and/or low speed pumps. Fluid viscosity (fluid resistance to shear) will also adversely affect minimum required inlet pressure. Friction losses within the pump suction side casing (minimal) and friction losses entering the first pumping chamber increase with increasing viscosity. Thus, pumps will require higher minimum inlet pressures when handling higher viscosity fluids. Figure 5 shows the effect. Velocities are labeled V 1 increasing in magnitude to V4. One solution to the high required minimum inlet pressure is to use a larger pump operating at a lower speed to reduce the internal velocity. The price paid is, of course, a larger, more expensive pump and a more expensive driver. Getting the inlet side of the pump correctly specified, and providing as much net inlet pressure as possible, will result in an optimally sized, minimum cost pump selection that can be expected to operate well for a long time. Excessively conservative inlet pressure specifications will result in larger, slower and more expensive – and perhaps even less efficient – pumps.■ James R. Brennan is Market Services Manager for Imo Pump (Monroe, NC). His responsibilities include worldwide marketing and technical support for pumping applications. He is a 1973 graduate of Drexel University in Philadelphia and a member of the Society of Petroleum Engineers.

POSITIVE DISPLACEMENT PUMPS HANDBOOK

Rotary Pump Overhaul Guide By James R. Brennan, Imo Industries Inc. rotary pump overhaul can be handled on a professional basis with clearly understood expectations and communications as to whether the repair is to be done at the pump manufacturer’s facility, in house, or by a third party (on or off site). Using the pump manufacturer has some obvious advantages, not the least of which is access to original component design/dimensional/tolerance data and experience with the pump as well as usually complete testing facilities. New pump warranties are sometimes available from the original manufacturer as well as upgrades and modernizations. Third party repair is frequently less costly, at least on a first cost basis, but it rarely includes a warranty, almost never includes any meaningful testing and can sacrifice reliability for expediency. In-house overhaul – facilities and personnel permitting – is a good alternative if some in-house expertise or experience is available. With the many “right-sizing” initiatives taking place, maintenance is frequently the first to go, resulting in the loss of in-house expertise as well as historical perspectives that can frequently lead to very quick analyses and fixes. If there are company or industry-specific or unique requirements to be imposed on the overhaul, such as sanitary specifications or testing liquids compatibility, be sure they are spelled out in writing before proceeding. If troubleshooting, failure analysis or other services are expected or needed, this, too, should be made clear before the overhaul commences. Regardless of where pumps are overhauled, basic steps to be followed are similar. Pumps for overhaul should be delivered to the facility as clean as possible and include a material safety data sheet

A

(MSDS) for any liquid residue in the pump(s). Basic steps include (as applicable): 1. Received condition report (external damage, missing parts, extraneous parts, etc.). 2. Review of MSDS to ensure proper handling of residue liquid in pump. 3. Disassembly of the pumps. 4. Thorough cleaning of pump component parts for inspection. 5. Inspection of each part to determine if it should be: Scrapped (always for elastomers and gaskets, usually for antifriction bearings and frequently for mechanical seals unless sufficiently valuable to have them overhauled themselves). Reworked (remachined, plated, scraped, polished, weld build-up, etc.) Reused as-is. 6. A written inspection report relating the above component condition and disposition recommendations should be produced. The report should include pricing for the recommended overhaul, noting what is or is not included (painting, testing, special preservation.) A comparison to the cost of a new pump may be appropriate if such is still manufactured. Many “off-the-shelf” pumps are manufactured in large lots and sell for about the same price as an overhaul handled on a one-at-a-time basis. The party paying for the overhaul should review these recommendations and, depending on cost, time, and how critical the pump in service is, either challenge the observations if any appear questionable or authorize the overhaul. Frequently, time will determine the re-use of questionable or worn parts. Overhaul should specifically state that new-pump perThe Pump Handbook Series

formance will be achieved or something less due to economics or time. A loading or unloading pump flow rate may not be very critical to dayto-day operations while pump capacity may be critical for gas sealing, machinery lubrication or process applications. If something less than new-pump performance is expected, then the minimum flow rate should be agreed upon before the overhaul proceeds. If weld build-up or repair is to be conducted, be sure that the weld procedures and welders have been adequately qualified. Plating and spray build-up processes are frequently technique sensitive and may need some non-destructive testing after application to confirm suitability to the objective. Not all pumps are tested after repair. If this is a requirement, specify it in the ordering document. Testing ranges from turning the pump shaft by hand, hydrostatic test and spin test to full qualification testing as conducted on a new pump. If newpump flow rate is expected after the repair, again, so specify. If the overhauled pump will be installed and operated upon its return, no special preservation or packaging may be needed. On the other hand, if the pump will go into storage, internal preservation and sturdy boxing should be specified. Application of a paint should be included if the pump exterior is of cast iron, steel or other material susceptible to rusting.■ James R. Brennan is Market Services Manager with Imo Pump (Monroe, NC). He is responsible for worldwide service and support for the company’s rotary screw and gear pump product lines. He has more than 30 years of pump application and troubleshooting experience, has authored many articles and spoken at a variety of technical conferences and seminars

127

Rotary Pump Inspection Report Owner_____________________________________ Model No. ___________________________________ Serial No.___________________________________ MSDS_____________Yes______________No Received Condition______________________________________________________________________________ Part I.D.

Qty

Part No.

Part Name

Recommended Disposition- Check one Scrap Reuse Rework & Comments

By___________________________________________________ Date__________________________________________________ Comments____________________________________________ Figure 1. Example of rotary pump inspection report form

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The Pump Handbook Series

Price $

Subtotal $ Assembly/Test Other__________________ Other__________________ _______ $

POSITIVE DISPLACEMENT PUMPS HANDBOOK

A Guide to High Pressure Reciprocating Pumps When a number of high pressure designs meet your initial performance requirements, how do you choose the right one? Start here. By Dallas Simonette and Dave Semotink, General Pump, a member of the InterPump Group reciprocating pump is a mechanical device used to impart a pulsating, dynamic flow to a liquid and consisting of one or more single or doubleacting positive - displacement elements (pistons or plungers). The elements in the liquid end are driven in a more or less harmonic motion by a rotating crank and connecting rod. The liquid flow is directed from the pump inlet (suction) to the pump outlet (discharge) by the selective operation of selfacting check valves located at the inlet and outlet of each displacement member. This definition puts us in the ball park for the subject at hand — high pressure reciprocating pumps. The first distinction between normal reciprocating pumps and high pressure pumps is that most high pressure pumps are single-acting. The number of displacement elements on commercially available pumps ranges from 2 (duplex) to 6 (sixplex). The high pressure designation refers to the mechanical design of the pump, which enables it to handle high operating pressures

A

Piston Pumps Piston pumps have a dynamic primary seal. The seal is affixed to a moving piston and creates a seal by its interaction with the cylinder wall. The most popular high pressure piston pumps are a flowthrough design, incorporating a floating piston contained by an inlet valve, spacer and retainer, which are fixed to the piston rod. Each cylinder has a spring-energized

check valve positioned at its outlet. In operation, the inlet valve seals on the backside of the piston as it drives it forward through the displacement cycle. At the same time, the piston pulls the next charge of fluid into the cylinder. As the piston rod reverses direction, the spacer allows the inlet valve to open before the retainer starts to pull the piston back. The piston then moves through the fluid drawn into the cylinder on the displacement stroke. The fluid flows between the piston and inlet valve into the discharge side of the piston, and the cycle repeats. The advantage of this design is its suction capabilities. Inlet pressure ratings as low as –9 psi are possible, depending on fluid type and temperature. Limitations of piston pumps in high pressure applications are numerous when compared to plunger pumps: 1. Discharge pressure: The most common range is 800–1500 psi maximum. 2. Slow operating speeds: 800– 1400 rpm. This limits applications to belt drive systems. 3. Sound level: Very noisy. A lot like the old high performance V-8 engines with solid lifters.

Plunger Pumps Plunger pumps have a static primary seal. The high pressure seal or packing is captured in a gland within the manifold and retained with a preload by a seal retainer. The inside diameter of this packing seals on the dynamic plunger surface, which is an extremely hard material like ceramic. The Pump Handbook Series

Figure 1. Cutaway view of an industrial high pressure triplex plunger pump

Spring-energized check valves are positioned on the end of the plunger chamber in the manifold (Figure 1). In operation, the plunger moves forward to displace the fluid in the plunger chamber. The inlet valve checks the water flow and prevents it from going back into the inlet port. The outlet valve opens and allows the fluid to enter the discharge collector bore. When the plunger reverses its direction, the outlet valve closes and the reduction in plunger chamber pressure draws fluid through the inlet check valve. The cycle then repeats. There are several reasons why this style of pump has become the dominant force in the reciprocating pump world: 1. Discharge pressure:500–50,000 psi models are readily available. 2. Wide range of operating speeds: up to 3600 rpm. Most manufacturers have bore and stroke combinations to deliver the desired flow at speeds to match electric motors and internal combustion engines for direct drive mounting.

129

3. Sound level: very quiet for the amount of power throughput. I’ve had to touch pumps many times to make sure they were running. 4. High temperature and chemical compatibility: pumps are available in a wide variety of materials and elastomers. 5. Low maintenance 6. Ease of service

Diaphragm Pumps Diaphragm pumps are another option for high pressure applications. These specialized pumps drive diaphragms with oil pressure to create a uniform balanced load. This allows high pressure operation with long diaphragm life. Tough solutions are the mark for these pumps, such as reclaim water, fluids containing abrasives, machine tool cutting fluid, maple syrup, melted chocolate or glue.

Pump Selection Criteria Fluid Identification Many factors influence pump choice for specific applications. Fluid identification is a major consideration in selecting the proper pump. First, the viscosity of the fluid to be pumped has a tremendous impact. The more viscous the fluid, the harder it will be to use a reciprocating pump. This is due to the rapid motion of starting and stopping the inlet fluid by the inlet check valves. The higher the rpm, the faster this cycle occurs. Second, one must bear in mind the temperature of the fluid being pumped. Reciprocating pumps start and stop the fluid during their pumping cycle, which may create a vacuum during intake. This can reduce the boiling point of the fluid to its actual temperature. If this happens, the pump will cavitate during a portion of the operational cycle, reducing the life of the manifold and its internal components. Fluid composition also impacts pump operation. Applications today vary greatly in the types of fluids used. The specific fluid being pumped, including the percentage of chemical used and pH of the solution, can have an adverse effect on wetted components. In a number of

130

applications the pump must be customized to ensure compatibility with the fluid. This could be as involved as changing the manifold to 316 stainless steel, or as easy as changing the composition of the seals and packings. Contaminants within the fluid stream also affect what materials are used and their life span. Reciprocating pumps have very tight tolerances between components in the manifold. The introduction of contaminants can cause premature wear by etching (eroding) key components. Or contaminants can become lodged in key areas, preventing proper function. Contaminants can be introduced into the fluid stream intentionally, by adding them through the use of a soap or other additive, or as a by-product of the application, such as the use of reclaim water. Another factor to consider is the use of hazardous or flammable liquids. Standard reciprocating pumps usually have a vented area between the manifold and crankcase. Fluid escapes through the vent to signal the beginning of seal or other type of failure. Typically, this is not acceptable when using hazardous or flammable liquids. Inlet Design Inlet conditions for the pump will not only affect its dynamic performance, but will determine the ultimate service life. Inlet requirements for a particular application can be divided into two basic categories: pressurized and non -pressurized. Most high pressure pumps are sensitive to contaminants larger than 10 microns. This mandates the use of filtration prior to the inlet of the pump. Use a 50–80 mesh screen strainer mounted in a convenient position for easy cleaning. When sizing inlet plumbing, a good rule of thumb is not to be smaller than the inlet port of the pump. Preferably, use the next larger size, especially on suction inlets. Mount the pump as close as possible to the inlet supply and keep plumbing to a minimum. Sweep elbows are preferred over hard 90s for a smoother flow path, and make sure to reduce down right at the pump. High flow requirements may require dual feed The Pump Handbook Series

lines to supply adequate flow. The goal is to prevent cavitation. Excessive pressure to the inlet may also be detrimental to the pump or other inlet components. An inlet regulator valve can be used if supply pressure is too high. Some pumps have low pressure seals with maximum ratings of only 40 psi; higher pressures will cause premature seal failure. Pumps that use an unloader valve mounted on the discharge port typically route the bypass flow back to the inlet. This recirculation loop will generate enough heat to damage the pump seals, given enough time. Always use a PTP (pump thermal protector) valve, which will open and dump fluid at 145°F. Outlet Performance Outlet pressure is probably one of the most misunderstood processes that takes place in positive displacement reciprocating pumps. Pressure is actually created by the amount of restriction applied to a specific amount of water in a positive displacement condition. The easiest example of this process is your average garden hose. When the water is not restricted in any way, it flows at a very slow speed and pressure. When you restrict the water flow with your hand or a nozzle, it travels at a very rapid rate (velocity), creating pressure. Pressure can be fixed by using a constant or varied resistance, whichever fits the needs of the application, and can be adjusted in a number of ways. One way is to vary the rpm of the motor, which will change the pump flow rate. Another is to use an adjustable orifice. Or you can change fixed orifices for each application. The last way is to use some type of diverter or regulating valve to divert a portion of the flow, decreasing the flow through the restriction, thus lowering pressure. A number of complications can be created when outlet pressure is adjusted through the use of a regulating or diverter valve. Excessive bypass can heat the water to a point where the pump cannot adequately handle the temperature. Another complication, caused by setting the pressure regulator improperly, is

Inlet effects on efficiency, 1 inch bore plunger pump As pump speed is increased, performance decreases with inlet pressure. 14

13 1750 rpm @ 1800 psi

Outlet Flow (gpm)

12

11

Photo 1. Cutaway view of a hollow shaft, direct drive plunger pump

1450 rpm @ 2100 psi

10

9

8

1150 rpm @ 2500 psi

7

6 40

20

10

5 Inlet Pressure (psi)

0

-2.5

-5

Figure 2. Inlet effects on efficiency, 1 inch bore plunger pump

pressure spiking when the system changes from a discharge mode (valve open) to a bypass mode (valve closed.) A pressure spike is the rapid increase in pressure on the system during the transition from discharge to bypass. Although pressure spikes are only a fraction of a second long, they are cumulative and affect the life of the pump and other system components. Like all moving parts, there is always a chance that the regulator or diverter valve will fail in the discharge mode (especially with the use of reclaim water.) As discussed earlier, positive displacement reciprocating pumps will pump a given amount of flow for a certain rpm. If this flow is not diverted, a rapid increase in pressure occurs until system failure takes place. A secondary safety relief device should always be installed to prevent damage to equipment or personnel. Another key factor in the life of the pump is the ratio of operating to rated pressure. The closer the operation of the pump comes to its rated pressure, the shorter the life of that pump. This will be evident in the increase in number and frequency of routine maintenance operations performed on the pump during its life (Figure 2). Flow Flow of the pump is directly

related to shaft speed. Although rpm is the main factor in deciding flow, a number of other variables are also influential. They affect the efficiency of the pump, which can be calculated by figuring the theoretical flow rate at a given rpm versus the actual flow rate at the same rpm. One of these variables, shaft speed, can be a major consideration in pump efficiency. A pump running at 93% efficiency at 1000 rpm may only run 87% efficient at 1750 rpm. In this case, the main reason for lower efficiency may be the design of the check valves and the speed at which they operate. At 1000 rpm, the check valves will be opening and closing approximately 16.5 times per second. At 1750 rpm, this rate increases to about 29 times per second. Another factor contributing to pump efficiency is the condition of the inlet. If the pump cannot draw fluid through the check valves, due to an adverse inlet condition, such as high vacuum or inadequate flow, pump efficiency will decrease. This will correlate directly to outlet flow. A side effect of the pump not receiving enough flow is cavitation. Component life is directly related to rpm and the number of cycles that take place over a given amount of time. The faster the rpm, the more cycles the components will see, reducing pump life. The Pump Handbook Series

Photo 2. Duplex and triplex plunger pumps with gear box

Drive Selection Internal Combustion Engines Calculate the horsepower requirements: Pressure x Flow = hp 1100 (Efficiency) Loaded speeds for gasoline engines range from 3200–3600 rpm, depending on manufacturer, governor type and no load setting. Solid shaft pumps are available for belt drive or flexible coupling direct drive. Direct mount gear reduction drives are another popular option. Hollow shafted pumps with direct mount flanges are the most economical and provide the smallest overall envelope size for the unit, (Photos 1 and 2). Electric Motors Calculate the horsepower requirements: Pressure x Flow = hp 1460 (Efficiency) Hollow shafted pumps with direct mount flanges for C–face motors are available for units requiring up to 10 hp. Solid shaft pumps with a bell housing and flexible coupling can be used on motors up to 15 hp. Pump mounting rails are provided for belt drive or flexible coupling direct drive (Photo 3).

Troubleshooting Troubleshooting a system can be a very simple process if taken step by step. The initial goal is to identify and isolate the problem. When a problem arises, the usual assumption

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Photo 3. Solid shaft pump utilizing a bell housing and flexible couplings for direct mount to an electric motor

is that something is wrong with the pump. Although the pump is the heart of the system, 80% of problems are either upstream or downstream. Like most systems, there are number Problem

of components that need to be checked for wear or cleaned on a frequent basis in order to perform well. Filters need to be cleaned, tanks and water supplies need to be checked and nozzles need to be inspected for wear. All these can contribute to system malfunction. Once you have checked the system for problems, check the pump. Refer to Table 1 for information on identifying the causes of problems and the steps to fix them. In closing, there are a number of pumps that may meet the initial performance requirements for a particular application. But when all factors are accounted for, the choices narrow and the final decision should be

made based upon the duty cycle of the system. This will insure that the pump chosen will provide maximum performance over the life of the system.■

References 1. Miller, John E. The Reciprocating Pump: Theory, Design and Use. Krieger Publishing Company. Malabar, Florida, 1995. Dallas Simonette received his B.S. Engineering from St. Cloud State University and has 18 years of experience in the high pressure pump industry. Dave Semotink has 15 years of applications experience in the industry.

Cause Valve stuck open Worn nozzle Belt slippage Air leak in inlet plumbing Relief valve stuck, partially plugged or improperly adjusted, valve seat worn Inlet suction strainer clogged or improperly sized Worn packing, abrasives in pumped fluids or severe cavitation; inadequate water Fouled or dirty inlet or discharge valves Worn inlet, discharge valve blocked or dirty Leaky discharge hose

Remedy Check all valves, remove foreign matter Replace nozzle, use proper size Tighten or replace; use correct belt Disassemble, reseal and reassemble Clean, adjust relief valve; check for worn or dirty valve seats. Kit available Clean; use adequate size and check more frequently

Pump runs extremely rough, pressure low

Restricted inlet or air entering the pump Inlet restriction and/or air leaks Stuck inlet or discharge valve

Proper size inlet plumbing; check for air-tight seal plumbing Replace worn cup or cups, clean out foreign material, replace worn valves

Water leakage from under manifold. Slight leakage

Worn packing Cracked plunger

Install new packing Replace plunger(s)

Water in crankcase

May be caused by humid air condensing into water inside the crankcase Worn packing and /or piston rod sleeve O-ring on plunger retainer worn Cracked plunger

Change oil intervals; use nondetergent oil

Pulley loose on crankshaft Broken or worn bearing or rod(s) Valve stuck open or shut, or not opening enough Inadequate water supply to pump inlet

Check key and tighten set screw Replace bearing or rod(s) Replace bad valve

Scored, damaged or worn plunger Overpressure to the inlet manifold Abrasive material in the fluid being pumped Excessive pressure and/or temperature of fluid being pumped Overpressure of pumps Running pump dry Upstream chemical injection

Replace plunger(s) Reduce inlet pressure Install proper filtration on pump inlet plumbing Check pressure and fluid inlet temperature; be sure they are within specified range. Reduce pressure DO NOT run pump without water Use downstream chemical injection

Pulsation Low Pressure

Loud knocking noise in pump

Frequent or premature failure of the packing

Install proper filter; suction at inlet manifold must be limited to lifting less than 6 feet of water or 3.5 psi Clean discharge and valve assembly Replace worn valves, valve seats and/or discharge hose

Replace packing; replace O-ring Replace plunger(s)

Check inlet feed conditions and adjust accordingly

Table 1. Troubleshooting reciprocating pumps

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POSITIVE DISPLACEMENT PUMPS HANDBOOK

Sealing Hazardous Fluids with Dry Seal Technology The most exciting seal technology in decades, and the applications you can use it for. by Pete Barnhart, John Crane Inc.

T

he past few years have seen an almost endless stream of articles lauding dry gas seal technology. Admittedly, many readers may be groaning at the idea of another such article. If apologies are in order, so be it. The truth is that dry gas seals are the most exciting and far-reaching advancement in sealing technology in several decades. Their ability to seal efficiently and, in most cases, absolutely is unsurpassed in the industry. The technology is relatively new (six years), and there are countless new applications just waiting to be thought of and tried. This is the reason for this article, to review the technology and a wide range of applications that seal hazardous fluids in an attempt to stimulate understanding and interest. “Hazardous” can have several meanings. For this discussion, it will be defined very broadly to include VHAPS (Volatile Hazardous Air Pollutants), VOCS (Volatile Organic Compounds) and products that are toxic, extremely corrosive or even explosive. This includes any product that poses a safety threat to people. In most cases, severely limiting leakage to meet emission regulations or providing zero leakage of these products is desirable, if not required. Dry seal technology also has more than one definition. Dry running seals can have either contacting or non-contacting faces. Dry running contacting seals definitely have their place, but their application is severely limited by pressure and speed. The dry seal technology addressed in this article refers to seals with noncontacting faces. Here, application ranges for both pressure and speed are very wide, encompassing most

rotating equipment in use today. Historically, the sealing of hazardous fluids has been accomplished by using various forms of wet, contacting seals. Conventional single seals were, and continue to be, very effective in ”controlling” leakage and emissions of hazardous fluids. However, in certain services, dual mechanical seals are desirable or required because they provide an added measure of safety by greatly reducing or eliminating hazardous process leakage to atmosphere. Dual seals are either pressurized or non-pressurized. Pressurized dual mechanical seals require the careful selection of a barrier fluid and a costly support system. The liquid barrier is maintained at a pressure within the seal that is higher than the process pressure acting on the seal. Inherent in the operation of this seal

arrangement are frictional torque and heat generation associated with dynamic seal face contact, the potential for process fluid contamination, support system reliability problems and costly horsepower consumption. Non-pressurized dual mechanical seals also require a barrier fluid (in this case, a ”buffer” fluid because it is not pressurized) and a costly support system. The arrangement requires the non-pressurized outboard or secondary containment seal to be in line with a vapor recovery or closed loop system to contain the buffer fluid as it becomes contaminated with the process fluid. Frictional torque and heat generation, support system reliability problems, costly horsepower consumption and a certain amount of process leakage must be taken into account in using this arrangement.

Counter-Clockwise Shaft Rotation

Spiral Groove Inner Groove Diameter

Sealing Dam

Opposing Face Inner Diameter

Oute Diamet

Figure 1. A gas seal with a spiral groove design The Pump Handbook Series

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consumption and dramatically increased seal life. The only heat generated at the seal faces comes from the shearing action of the gas. This heat is dissipated through the materials of construction and by virtue of the small amount of gas being pumped across the faces. Horsepower consumption, when compared to a wet seal, is reduced by a factor of 20 to 50 depending on size, speed and differential pressure. As an example, when comparing the energy costs of a double wet seal and a double gas seal (Figure 2) on a 2.000 inch shaft, the savings can be more than $500 per year! Most notably, no contact friction means no face wear. This translates into considerably longer seal life. Dry gas seals can be applied in single or dual arrangements, much the same as contacting wet seals. As single seals they are essentially limited to fan, blower and steam turbine applications and to pumps in which the process fluid is in gaseous form at the seal interface. Single dry seals can limit leakage of hazardous fluids. In many cases limiting leakage may be all that is required. Here are a few applications:

Gas Seal Technology Satisfactory life of any mechanical seal depends on its ability to minimize the effects of contact friction. Without the proper design and materials, the seal will fail due to friction generated heat. The non-contacting design of gas seals eliminates contact friction, making it possible to use the seal where energy levels are too high to use dry running contacting seals. This design also gives gas seals certain advantages over contacting wet seals. These advantages will be discussed shortly. There are several designs on the market that achieve the non-contacting feature of a mechanical seal. These include spiral grooves, T-slots, V-grooves and wavy faces. Because it is by far the most prevalent design in use today, this discussion will focus on the spiral groove design. Gas seals ride on a gas film generated by the spiral grooves as the shaft rotates (Figure 1). The grooves are recessed into a silicon carbide or tungsten carbide mating ring that comprises one of the two seal faces. A carbon primary ring that generally rotates with the shaft opposes the spiral groove face. Gas is drawn into the tip of the groove at the OD of the face and is compressed at the sealing dam (the ungrooved portion of the mating ring). The gas expands at the dam, and the resulting pressure separates the faces, creating a non-contacting, gas lubricated seal. The sealing dam plays other important roles. It restricts the amount of gas flow across the faces, and it provides contact sealing when the shaft is idle. Eliminating contact friction on mechanical seals has many advantages, such as reduced face operating temperature, reduced horsepower

Hot Air Blower This unit’s (250-300°F, 3600 rpm, 1.500” shaft diameter) lip seal arrang-ement was failing in a matter of days. The leakage was impinging on the motor bearings, causing premature failure. The hot air was also a hazard to the workers in the area. An outside mounted single dry seal (Figure 3) was installed, greatly reducing the leakage, the bearing failure and the safety hazard. Steam Turbine This application (150 psig back pressure, 750°F, 6675 rpm, 2.937”

shaft diameter) had recurring bearing failures due to moisture in the oil caused by excessive wear on and leakage from the carbon bushing seals. A single dry seal has drastically reduced steam leakage and improved the unit MTBR (Mean Time Between Repair). Repeller Pump This arrangement (Figure 3) in nitric acid service (99.5%, 150°F, 1850 rpm) has been running successfully for more than nine years. As the shaft rotates, the repeller generates a vacuum in the seal chamber, removing all liquid. The dry seal operates on atmospheric pressure. At shutdown, the seal chamber is flooded and pressurized with product. The sealing dam provides a positive seal for the duration of this condition. Cryogenic Pump This service, which involves pumping liquid nitrogen (87 psi, -320°F) from a tank truck, generally had an MTBR of several weeks. Converting to a dry seal and using welded metal bellows and polymer secondary seals to handle the temperature, (Figure 4), has increased seal life to years! Dry seals are revolutionizing the way cryogenic fluids like liquid oxygen, nitrogen and argon are being sealed. (See sidebar.) The dual arrangement of dry seals has become far more prevalent than the single because it can be more readily applied to pumps. In this arrangement (Figure 5) a barrier gas, usually nitrogen, is pressurized between the dual seals at 20-30 psi above the process pressure in the seal chamber. A small amount of the gas is pumped across each seal (into the process and to atmosphere). This gas consumption is extremely low and is related to the size, speed and

ANNUAL ENERGY COSTS ($)

3.500 3.000 2.500 2.000

DOUBLE WET SEAL

1.000 500

(1535) (155)

(2750)

SAVINGS PER PUMP

1.500

ATMOSPHERE OUTBOARD

DOUBLE GAS SEAL (532)

0 1.000

2.000

Double Wet Seal Operating Cost Typical 2.000" D. seal consumes 1.40 HP Where: HP x .745 = KW (1.40 HP) x (.745) = 1.043 KW Based on $ .06/KWH Energy Cost (1.043 KW) x ($ .06/KWH) (24) (365) = $548/year

SIZE (INCHES)

SIZE 1.000 2.000 3.000 4.000

3.000 COST/YR WET $158.00 548.00 1,575.00 2,890.00

4.000 COST/YR GAS $2.35 15.70 39.15 136.60

PRODUCT SIDE INBOARD

SAVINGS/YR W/GAS $155.65 532.30 1,535.85 2,753.40

Energy costs comparison between a double wet seal and a double gas seal.

Outside mounted gas seal for a blower or repeller pump.

Figure 2. Energy costs comparison between a double wet seal and a double gas seal

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Figure 3. Outside mounted gas seal for a blower or repeller pump

The Pump Handbook Series

differential pressure of the application. With this arrangement, all the previously mentioned advantages of dry seals are enhanced by one additional benefit—zero product-to-atmosphere leakage! Dual dry seals are very versatile and user friendly. The seals come in cartridge form, making installation easy. The cartridges are designed to fit both standard bore ANSI-type pumps and oversized bore seal chambers, making them applicable to most process pumps. The present application range for these seals is vacuum to 600 psig, -40 to 500°F, and speeds from 10 to 5000 fpm. However, these boundaries have been and will continue to be expanded in specifically engineered applications. The technology has no limiting factor due to the pressure-velocity relationship or wear at the seal faces. The limiting factor appears to be the materials of construction. Pressure is limited due to the effect it has on deflection of the seal faces. Temperature is limited due to the limitations of the secondary seals (O-rings, polymers, etc.). New materials are being tested, on an ongoing basis, that will continue to expand these ranges. Also, unlike dual wet seals, the support system for dual dry seals is neither costly nor complicated. A simple, dependable gas panel is shown in Figure 6. So, what are the applications? There may be many candidates in your facility that you have not considered. The use of dual dry seals in pumps is only in its infancy. Many possible applications are yet to be tried. Here are a sampling of applications to awaken you to the possibilities.

Phenol-cumene This product (475°F, 28 psig, 3000 rpm, 1.187” shaft diameter) solidifies when the temperature drops. Dual wet seals had problems with an adequate barrier liquid (due to the temperature) and tended to hang up the inner seal if the temperature dropped. The dual dry seal has the following advantages: 1. The nitrogen barrier gas is not affected by temperature, so there is no degradation of the lubricant. 2. Nitrogen does not circulate, so there is no cooling of the inner seal. Hot Thermonol C Heat transfer fluids are a safety hazard because of their elevated temperatures. One facility was failing single wet seals in 10-15 days due to a low suction pressure that caused severe cavitation in the pump. Correcting the suction problem would have been a major expense and was not feasible at the time. Dual dry seals have been running successfully now for two years. The solution is in the seals’ ability to handle upset conditions. The pressurized barrier gas and non-contacting faces create a sealing system that is nearly unaffected by the process fluid. As long as the barrier gas pressure is maintained 20-30 psi above the process pressure, changes in process temperature or pressure, or the onset of dry running or cavitation, will not adversely affect the seal. Toluene A facility was having problems with mag drive pumps in this service (30 psi, ambient temperature, 3600 rpm, 1.750” shaft diameter). The

pumps had been selected to meet emission regulations. Several had been down due to dirty pumpage and dry running. Repairs, rebuilds and reportable incidents were costly. The dual dry seal eliminated the dry running and abrasives concerns and provided zero emissions. This seal has been running without failure since late 1995. Sofasets This is a generic term for a very troublesome acid soap solution. This application (30 psig, 137°F, 1750 rpm, 1.750” shaft diameter) was failing wet seals in short order because even minute leakage of the barrier liquid into the process resulted in crystallization of the product at the seal faces. The dual dry seal, using nitrogen as a barrier gas, prevented crystals from forming. The root cause for failure was thus eliminated, and the seal continues to operate as designed. Wet Chlorine Gas This centrifugal fan application (510 psig, 180°F, 2900 rpm, 1.312” shaft diameter) had to provide zero emissions, with a high degree of reliability, of this very hazardous fluid. The nitrogen barrier, set at 29 psi above the process gas pressure, has run leak free since mid-1995. Methylene Chloride, Acetic Acid, Acetone These products are regularly pumped at a leading pharmaceutical company. In this particular application (95 psig, 110°F, 1450 rpm), all three chemical compounds are pumped at different times. Besides the obvious concern about leakage of these products, the dual wet seal Inject clean dry barrier gas into seal chamber at 20-30 psig above process pressure.

PROCESS FLUID

ATMOSPHERE

Stationary spiral grooves pump higher pressure barrier gas across inner faces into Stationary spiral grooves pump higher pressure barrier lower pressure process and across outer faces to atmosphere gas across faces into lower pressure process

Figure 4. A dry seal using welded metal bellows and polymer secondary seals to handle cryogenic fluids

Figure 5. A dual dry seal for hazardous applications.

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Thus, leakage from a dual wet seal, even a drop, could mean a product loss of more than $100,000. The solution: dual dry seals with a nitrogen barrier. Nitrogen is a preferred blanket for many of these batch operations. There is also no wear debris from the faces, since the faces are non-contacting. Whether the mixer speed is high or very low (as low as 10 fpm) dual dry seals are a viable solution.

Figure 6. The support system here provides a dependable flow of barrier gas to the double seal shown in Figure 5.

approach caused contamination of the process. This required costly cleaning. The dual dry seal eliminated leakage and enhanced reliability, and the inert nitrogen barrier gas did not contaminate the process. Butadiene This product polymerizes readily, particularly if oxygen is present. Using dual dry seals with a nitrogen barrier eliminates the possibility of oxygen contamination. Many other products act similarly. Pharmaceutical Mixers In most cases the products being mixed are not hazardous. But any barrier liquid that gets into the product will be fatally hazardous to it.

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High Speed Hydrocarbon Service There are many applications in refinery and chemical plant service where a high head, low volume pump is preferred. Many of these applications involve high shaft speeds (10,000-35,000 rpm). In this particular service (hydrocarbon, 0.57 specific gravity, 136 psig, 72°F, 12,425 rpm, 1.249” shaft diameter), a single wet seal was failing every 4-6 weeks. The root cause was the temperature rise at the seal faces due to the high speed contact friction. This caused the product to vaporize and led to subsequent blistering, wear and failure of the seal faces. A dual dry seal (specifically engineered, non-cartridge design) was installed in early 1996 and has run without failure since then. Non-contacting technology removed the pressure-velocity limits of the wet seal. Applications with speeds greater than 34,000 rpm have been sealed successfully with dual dry seals.

Conclusion In summary, dry gas seals are the most significant advancement in

The Pump Handbook Series

sealing technology in decades. There are countless new applications just waiting to be conceived and tried. I hope this brief article will promote understanding of this exciting new technology and stimulate interest in new applications. n

REFERENCES 1. Wasser, J. R., “Dry Seal Technology for Rotating Equipment,” 48th Annual Meeting of STLE (1993). 2. Wasser, J. R., Sailer, R., and Warner, G., “Design and Development of Gas Lubricated Seals for Pumps,” Proceedings of Eleventh International Pump Users Symposium, Turbomachinery Laboratory, Texas A&M University, College Station, Texas (1994). 3. O’Brien, A. and Wasser, J. R., ”Design and Application of Dual Gas Seals for Small Bore Seal Chambers,” Proceedings of Fourteenth International Pump Users Symposium, Turbomachinery Laboratory, Texas A&M University, College Station, Texas (1997). Pete Barnhart is the Branch Manager at John Crane Inc.’s Concord, CA facility. He has worked for JCI since 1978, gaining 15 years of field sales and service experience before entering management. Pete received his B.S. degree (mechanical engineering) from the U.S. Merchant Marine Academy, Kings Point, NY (1971).

POSITIVE DISPLACEMENT PUMPS HANDBOOK

Well Pump Applications for Mine Dewatering Choosing the right pumps means knowing drainage requirements, dewatering schedules and well construction as well as system and fluid conditions. by Mark List, Miller Sales and Engineering

nce the economics of a mineral deposit have been determined as favorable and a decision made to proceed with mine development, significant financial commitment is placed at risk in expectation of a calculated return on investment. Many aspects of mining carry relatively high levels of uncertainty that contribute to the overall degree of risk. One important consideration is groundwater control, should it be a factor, during mining operations. Where mining must take place below the water table in highly permeable geologic material, operations would not be possible without effective control of groundwater. Several mines in the western United States have developed large well fields, up to 70,000 gpm capacity, to intercept inflow and divert groundwater from the workings.

O

Mine Dewatering Objectives There are two general objectives to most mine dewatering programs: Keeping the working conditions relatively dry and maintaining the stability of the excavation or opening. Operating costs and production rates are directly influenced by working conditions. Wet floors and working faces create poor ground conditions for heavy loading and hauling equipment. They also increase tire wear, reduce cycle times and impact tonnage factors. Safety becomes a direct issue if electric powered machinery is used. In the worst case, a submerged work-

ing level becomes inaccessible and production is halted. The stability of open pit walls and underground openings is of vital concern for safety and economic reasons. Adequate drainage must occur in order to keep pore pressure at acceptable levels based on geotechnical stability analysis. General Types of Mines and Groundwater Control Methods Mines are either open pit excavations or underground excavations, or both. In some cases large scale open pit mines have succeeded prior episodes of underground mining in the same area. In other situations, underground mines are developed adjacent to or from within existing open pit mines. Driven by metal prices and advancing technology, companies have continued to explore deeper and/or more challenging geologic territory for mineable orebodies. Mine dewatering methods have evolved out of necessity in response to the increasing groundwater control requirements of contemporary mining. Where conditions permit, open pumping from collection sumps is a standard practice. This method is commonly used in open pit mines to control surface water drainage and in both open pit and underground mines where groundwater inflow rates are small enough to be managed in this manner. Booster stations might be required employing positive displacement pumps or horizontal centrifugal pumps The Pump Handbook Series

Photo 1. In-pit well in service with loading operation on left, blasted material on right, and high wall in background

designed for high head dirty water, abrasive solids or slurry service. Depending on the hydrogeologic setting, however, some mines cannot be effectively or safely dewatered using this method alone. Well Field Systems for Dewatering Several mines in northern Nevada require the use of wells to intercept and lower groundwater levels around open pit and underground excavations. The orebodies associated with these mines are hosted in fractured bedrock forma-

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depths exceeding 2000 feet. Pit perimeter wells typically discharge to gathering mains at relatively low well head pressures. Inpit well pumps can be staged for the additional head required to discharge to a surface location outside the pit, or to a booster station in the pit. Vertical turbine can pumps and horizontal centrifugal pumps are in service for this application.

Understanding the Application

Photo 2. Twin 800 hp, 5000 gpm vertical turbine can boosters with 42” discharge main in background

tions along mountain ranges that contact alluvial basins high in groundwater storage. Mining companies drill large diameter deep wells in bedrock fracture zones tested for favorable production yield. Other wells are completed so as to intercept shallow recharge or promote drainage of less permeable zones. Wells are typically located outside the open pit, but drilling sites in the pit are often unavoidable due to local hydraulic compartmentalization. A well field can be comprised of 30 or more individual wells with completion depths up to 2000 feet or greater and production casing diameters up to 24 inches. Well-specific capacities can exceed 60 gpm per foot of drawdown. Vertical turbine line shaft pumps (400 hp) are in service at setting depths of 1020 feet, and 1500 hp vertical turbine line shaft pumps are in service at setting depths of 800 feet. Single 2200 hp submersible units work at setting depths of more than 1800 feet, and at more than 2000 feet deep single 1500 hp submersible units are applied. Submersibles are also installed in series using tandem 1000 hp (2000 hp) units, tandem 1170 hp (2340 hp) units and combination 1500 hp/800 hp (2300 hp) units with the lower setting at

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The uncertainties involved in developing an efficient mine dewatering program become much better understood as operations progress. Groundwater flow information available at the onset of large scale dewatering can be very complete and supported by sophisticated model simulations, but such information is usually based on field test data that cannot be conducted at a scale proportional to what will actually be undertaken. Granted, pump applications engineers are most comfortable when customers assume all risk by specifying the necessary conditions for pump equipment selection. The outcome is likely to be better for all parties involved if knowledge is shared prior to establishing the conditions for equipment selection. Getting the Right Concept Well field dewatering involves lowering the water pressure level in the mine area to permit safe, efficient excavation. Over the life of the mine, the change in pumping lift from the initial static water level to some future pumping water level can be large—on the order of 1000 feet or more. Individual well production capacities can decrease dramatically depending upon aquifer system characteristics. From a pump applications perspective, this means selecting equipment with initial operating points that best match starting conditions and which can be made to fit, if possible, conditions that are expected to occur as formation dewatering progresses. Dewatering Schedule and Pumping Rates The rate at which a mine is expected to be deepened below the static water level is an important planning factor. It is used to estabThe Pump Handbook Series

lish a schedule for lowering groundwater heads before excavation begins, and it is a major consideration in predicting the required overall pumping rate. The change in pumping lift over time, indicated by the dewatering schedule, is the variable component required to evaluate intermediate and final TDH conditions for pump selection. Pump capacity range can be estimated assuming that sufficient test or operating data are available to be confident in doing so. The most reliable values for individual well production capacity and efficiency (well drawdown) are not available until the well is constructed and test pumping has been completed. However, the overall mine development schedule might not allow for the long delivery times that may be needed for special pump engineering or construction. If this potential problem is not addressed during project planning, pump equipment orders can be placed with results that are not cost-effective over the long term. Well Construction Well dimensions limit the size and type of pump equipment that can be installed. Although very costly to construct, large diameter deep wells can accommodate the installation of the large diameter four-pole (1800 rpm) 1500 hp and 2200 hp submersible electric motors that are required for high volume deep set applications beyond the practical setting depth limitations of line shaft pumps. Similar deep set applications with smaller casing diameters can require the use of two-pole (3600 rpm) submersible motors with an overall length of 100 feet or longer. In both situations a well must be drilled deep enough to achieve the design pump intake elevation and to accommodate motor equipment length and standard clearances. In addition to well diameter, well alignment is of critical importance for deep set line shaft pumps. Even the closest attention to construction alignment standards, however, cannot prevent ground movement from adversely affecting well alignment as dewatering progresses. Depending on the severity of

Photo 3. 400 hp vertical turbine line shaft (1020' setting) with mineral processing plant in background

ground movement and resulting deflection, shaft vibration can result in shaft and motor bearing failure. System Considerations System conditions vary in response to production well field changes and discharge method modifications. A well head pressure condition can usually be determined for use in staging the well pump, but a conservative approach is often taken to ensure that the desired pumping capacity can be maintained. If required, throttling is used to impose pressure temporarily until system conditions are within pump operating conditions. Fluid Conditions Water temperature and corrosivity are major factors influencing the selection of dewatering well pump equipment. Water temperature can influence the type of construction and materials used in a line shaft pump, but elevated water temperature adds significantly to the cost of submersible electrical equipment and thus can be a limiting factor in selection. At one particular dewatering operation, line shaft pumps are not an option, and submersible motors rated at up to 2200 hp are operating in water temperatures of 140°F. These are oil filled motors of specialized construction sometimes fitted with heat exchangers. Because the motors are located in the lower reaches of the wells, below the pump intake, shrouds designed for adequate water flow past the motors are usually required for cooling purposes. Unforeseen corrosion damage to pump cases, impellers and column pipe joints can ruin the best efforts in hydraulic applications engineering. Corrosion potential can some-

times be estimated up front by water quality analysis, and should be taken into account, if possible, in the specification of pump equipment materials and coatings. Unfortunately, geologic formation water conditions can and do change during dewatering. Partial aeration can occur with the rapid displacement of groundwater, and this can lead to unanticipated corrosive damage. Bronze and bronze alloys should be considered if conservatism is justifiable. If standard materials are selected, the first pump tear-down will reveal what doesn’t work.

Equipment Selection Making a reasonable attempt to understand the factors that dictate initial conditions and influence future conditions is key to selecting dewatering well pumping equipment that will remain effective under actual operating conditions. Nevertheless, there are limitations, and well yields can eventually decline to the point that pumps must either be operated intermittently or replaced with lower capacity units. Electric Submersible vs. Line Shaft Pumps Vertical turbine oil lubricated line shaft pumps are operating successfully at setting depths of more than 1000 feet. The slower pump and driver speeds (1800 rpm or less) of these units are favored by many operators. Pump damage from abrasive particles or partially aerated formation water is significantly less intense at slower speeds. Electrical problems are much simpler to troubleshoot and correct because motors are located at the surface above the discharge head. On the downside, slower speeds require larger bowls and well casing diameters. Line shaft equipment is mechanically complex and requires special engineering and manufacturing for bowl tolerances to accommodate the effects of relative shaft elongation under high thrust loading. Well alignment problems can adversely affect shaft and bearing life or even preclude the use of line shaft equipment. The Pump Handbook Series

Photo 4. Pump rig installing deep set line shaft pump

High capacity submersible equipment is available in both 1800 rpm and 3500 rpm classes. Small diameter high yield wells or setting depths greater than 1000 feet generally restrict pump equipment to the submersible type. The limitation here is motor power output. Four-pole 20inch diameter motors (1800 rpm) are available to 2200 hp for 24-inch casing applications. Slim line two-pole motors (3500 rpm) can be coupled in tandem to produce more than 1000 hp. These two-pole motor assemblies are more than 100 feet long, requiring additional well depth. Because the motors are installed below the pumps, electrical faults that occur in the down hole power cable or motor system require retrieval of the equipment string from the well for testing and repair to take place. The Pump Curve For a desired initial performance and estimated final performance, there is a simple rule of thumb for dewatering pump selection: start on the right side of the H-Q curve, run back to the left through the Best Efficiency Point, and plan to refit the pump end with additional stages if necessary to conform with estimated future conditions.

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Another method is to throttle the pump during initial operations if the range of expected conditions indicates that this will eliminate the need to pull and refit the pump end. Throttling is most common in deep set submersible applications involving relative certainty in the drawdown rate and final conditions. Pump Mechanical Considerations Line shaft applications involving deep settings and high thrust require special consideration for relative shaft stretch and bowl endplay requirements to establish adequate lateral impeller clearance under running conditions. Enclosing tube tension design as well as manufacturing tolerances for tube, shaft and column pipe lengths also need to be considered. Surface Equipment Power transformers and switching equipment are available in modular form on skid mounted platforms. Also, individual components can be custom assembled on a common skid or placed on individual pads if preferred. Mine power distribution systems often are plagued with swing loads and transients depending on the variety of electric machinery in service. Power factor correction, system protection and motor control

140

requirements vary with the application. Vertical hollow shaft type motors used with line shaft pumps are usually 460 volt or 4160 volt. Submersible motors are typically 460 volt, 2400 volt or 4160 volt. Installation Considerations Line shaft pump installation can be more mechanically involved than submersible pump installation. The oil tube and shaft are usually shipped assembled in lengths of 20 feet and must be individually placed in each piece of column pipe before installation in the well. Because the column, tube and shaft assembly are run in the well casing, three threaded connections must be properly made at each 20-foot interval. The projection dimensions of the tube and shaft, which start at the pump discharge case, must be maintained over the length of the column assembly for proper fit at the discharge head and motor coupling. Depending on the manufacturer of the submersible pump equipment, motor system and pump assembly during installation can be more or less complicated, generally requiring manufacturer’s field service in addition to the installation crew and equipment. However, once the pump and motor equipment are assembled in the well and

The Pump Handbook Series

tested for continuity, column installation is a straightforward process of making one threaded joint per pipe length and securing the power cable to the column. This goes relatively quickly, especially if the pump rig can handle pipe lengths of 40 feet. Operating Considerations It is important to follow up on the performance of a pump after it has been placed in service. The operator will no doubt inform someone associated with the sale of the equipment if a failure has occurred or performance is not as represented; conversely, the operator will be concerned with other matters if equipment performance is acceptable. Either situation involves information that can assist the applications engineer in selecting proper equipment and recommending the most effective modifications. n Mark List has more than 25 years of experience in the construction and mining industries, with 10 years of practice in groundwater investigation and mine dewatering equipment design, construction and operation as Chief Dewatering Engineer for a western gold mine. He is a registered engineer and state water rights surveyor in Nevada.

POSITIVE DISPLACEMENT PUMPS HANDBOOK

Sealing Positive Displacement Pumps: What Are Your Choices? Match the correct sealing option to your pump and your application. by John W. Wood Jr., Garlock Sealing Technologies here are two kinds of positive displacement (PD) pumps, rotating and reciprocating. The types are similar in that both move product through a system by displacing it with a dynamic member, which varies for the different styles of pump. Reciprocating pumps use a piston action to transport fluid, and rotating pumps, as their name implies, use a rotating element such as a screw, vane, lobe, gear or other element. Either way, product moves positively from one point to another in measured amounts. Sealing devices between the assembled (static) parts of the pump housing are necessary to prevent fluid leaks, whether the fluid is being contained within the pump or is being transferred by it. These areas are normally sealed with gaskets or O-rings. However, the most difficult area to seal in any pump is between the dynamic and static members, such as between the housing and the piston in reciprocating pumps and between the casing and the shaft in rotary pumps. Sealing these areas of the pump requires knowing which options exist for each type of pump and how to choose which one is right for your application.

(Figure 1). The higher the pressure in the system, the tighter the packing engages the shaft, resulting in higher temperatures. Heat thus generated causes thermal growth of the metal parts and the packing set. In the last step of this chain reaction the increased interference causes wear damage to the sealing members and piston shaft.

T

Piston PD Pumps (Figure 1) Sealing devices for these reciprocating pumps are normally located between the casing and the piston shaft with some type of hydraulic sealing arrangement or packing.

Photo 1. A typical rotary positive displacement pump with an installed mechanical seal (Viking universal bracket)

Such seals are available in various material compounds. Which one you use depends on the pumped product, as well as the speed (fpm) and temperature of the application. Seals in piston pumps are subject to hydraulic loads while the pump is in operation, and they must also maintain sealing integrity when little or no pressure is acting on the packing, such as when the pump is shut off. Typical seals or sealing sets for these pumps are compressed or retained in the seal chamber or stuffing box by some type of gland follower or cover plate. These devices prevent the sealing set from moving axially along the piston shaft. As reciprocation speed increases, system temperature and pressure rise, and more careful attention must be given to the types of sealing materials used. Heat must be transferred away from the dynamic surface. This requirement is complicated by the fact that packing sets are affected by hydraulic pressure The Pump Handbook Series

How to Choose the Right Packing Set Knowing that the wrong packing set can trigger a set of unwanted events like the one just mentioned, how does one choose the right packing? First, remember that the best sealing product is always one that is compatible with the product being pumped, is appropriate for the application (rotating or reciprocating), can handle the temperature, speed and pressure, and does minimal damage to the pump sleeve. At this point, many in the industry also consider cost a deciding factor. But if cost were the best way to choose packing, we could seal everything with tallow-lubricated flax! A more important consideration than straight purchase cost is service life. In other words, how much will it cost the company if this packing fails, rendering the equipment useless? Even the most

Figure 1. A packing set

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expensive sealing product available costs only a fraction of the cost of downtime and rebuilding. Or, as they say in the packing industry, “Packing systems are only expensive if they fail.” There are a few tips for success. First, don’t continue to use packing sets that have a short life. Second, get the best training available for your personnel to ensure proper installation and break-in techniques. Premature failure is often caused by improper installation. Matching Packing Sets to Applications: Important Considerations Most packing sets require some small amount of leakage along the shaft (weeping) to help cool and lubricate the area between the dynamic and stationary parts. Sound choices must be made concerning the type of sealing system that best suits the application. 1. If it is not practical or logical to disassemble the equipment to replace seals because the unit it too large and/or time is short, use split rings. Keep in mind that split rings normally weep or leak a small amount. 2. If Environmental Protection Agency (EPA) regulations dictate containment, zero leakage of the product, or leakage on the order of a few parts per million, use solid rings. Disassembly will be required to replace the sealing system, but solid rings provide a much tighter sealing system and less leakage to the atmosphere than split rings. 3. If the application is sanitary service and/or resistance to chemical attack is required, it might be necessary to use Polytetrafluoroethylene (PTFE) seal components. However, because PTFE is an insulator, heat cannot escape from the dynamic areas. Heat can still increase as the equipment runs even though PTFE is very slick. 4. If high system temperature is damaging or cooking the seals, consider changing to a sealing system that dissipates heat more freely, such as one whose parts are made of carbon or graphite materials. Remember that black materials will almost always run

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faster and cooler than white ones. Heat is a packing and pump sleeve killer. Carbon and graphite are normally more expensive, but they often pay dividends with the resulting longer life.

Diaphragm Pumps Another type of reciprocating PD pump is the diaphragm pump. Like the piston pump, the diaphragm pump drives product in a pulsating flow. The diaphragm flexes and forms a void that is then filled with fluid as the pump cycles. Either mechanical force or air is used to move the diaphragm in one direction then the other, drawing and pushing the fluid along. Diaphragm pumps can push almost any type of fluid, and they are often used for products that cause problems for centrifugal and rotary PD pumps. They are especially useful for thick, viscous fluids such as slurries and polymers, and for more aggressive paints, solvents and chemical mixtures. The diaphragm materials of construction are selected to match the service. Elastomers of various types are used for general purpose and abrasive applications. For chemicals and solvents, however, PTFE and GYLON® are the best choices. These two materials are resistant to chemical attack and are used for gaskets as well as diaphragms. Although diaphragm pumps do not have typical shafts to seal, as do the reciprocating and rotary pumps, they are not without sealing challenges. The diaphragms are vulnerable to abrasives, chemical attack, fatigue and, in the case of PTFE pumps, cold flow. Abrasives can wear a diaphragm thin and eventually cause a rupture at the flex point. The wrong chemical can be fatal to a diaphragm. As in all applications, chemical compatibility is one of the first considerations when specifying a diaphragm material. Even trace amounts of incompatible substances can cause premature failure of the diaphragm. Cold flow is a real problem when using a PTFE diaphragm or housing gaskets. When a pump is first assembled, all the PTFE joints will seal properly. However, after a The Pump Handbook Series

Photos 2 and 3. Before and after pictures of a starch-based glue leaking from a typical inface mechanical seal in a typical PD pump application. The after picture shows zero leakage using a P/S®-II seal.

short time cold flow will result and the joints will leak. Retightening the bolting seldom lasts. The PTFE will again extrude and cause leaks. Cold flow can also cause diaphragms to stretch and rupture.

Rotating PD Pumps There are many types of rotating positive displacement pumps, including single or multiple rotor, fixed member, screw, vane, gear lobe and circumferential piston designs. They are used most often to move difficult-to-seal products such as asphalt, thick oils, polymers, latex, isocyanate, TDI, glues, salts and sugars. Products like these pose some of the most bewildering and challenging sealing problems in the industry because of the way they react with seals and with the atmosphere (Photos 2 and 3). Typically these applications are sealed in one of the following ways. From the most senior to the most modern they are: packing, single end face seals, double face seals and high performance sealing elements.

Photo 4. An end face mechanical seal. The ceramic (white) face remains stationary while the carbon face shrouded with stainless steel rotates with the shaft. The two faces are pressed together by spring and hydraulic pressure.

2. Expensive labor and supplies for cleanup 3. Damage to equipment, particularly bearings and casings 4. Violation of EPA emissions regulations leading to shutdowns and possible fines 5. Recovery cost for lost product or waste disposal 6. Safety problems including slippery floors and injuries from inhalation and contact The question is, can a company afford these costs associated with traditional packing?

Single End Face Mechanical Seals The second way to seal rotating PD pumps is to use single end face Packing seals (Photo 4). Various versions of Pump packing has been with us these have been around for the last since rags were pressed in around 60 or 70 years. They have dramatithe first dynamic pump shaft. cally less leakage than packing and Packing continues to serve us well in serve well in many applications. many applications. The newer packThey work best in clean, lubricating ing materials seal better, last longer, fluids where they are less likely to require less or no flush, and seal at clog or stick together. Two extremehigher speeds and pressures. Also, ly flat faces form the primary seal. they are generally less trouble in Most manufacturers try to make the terms of maintenance. However, all seal faces flat within approximately packing leaks —even if only a small one helium light band—a mere amount. .0000116”. This degree of flatness When the kinds of difficult-tocan be measured only by looking at seal products listed above leak out the face through an optic flat in of the pumping system, there are monochromatic light. several consequences, none of them The product being sealed serves good: as a lubricating film between the faces, enabling them to hydroplane 1. Product loss and the related cost as the shaft rotates. This keeps the faces cool and results in Atmosphere longer operating life. Even when precisely Product

yyy ,,, ,,, yyy

machined and working properly, face seals leak (Figure 2). As pressure and friction act on the lubricating fluid, the product vaporizes and escapes. The vapor is not visible to the naked eye but is measured in parts per million by the EPA with its electronic sniffing devices. Single face end seals do have drawbacks. Imagine something as simple as sugar water passing between the faces. When the pump is turned off for a while, the sugar will “glue” the faces together and crystallize on the atmospheric side of the faces. When the pump is restarted, several problems could occur, including damage to the face and drive lugs, entry of debris between the faces, and ultimately seal failure. This has been the plight of mechanical seals when used in PD pumps. Double End Face Mechanical Seals The industry eventually developed a double face end face seal (Figure 3). Over time these double seals have taken on many design configurations, depending on the manufacturer and the intended service. However, the basic elements of all double seals are the same. A double face seal arrangement is merely two end face seals fitted on a rotating shaft or within a single seal chamber. In PD pumps there are three basic situations that call for the use of double seals. 1. To hydraulically load a fluid between the faces to prevent the migration of a problem product across the faces. (Problem products are often in slurry form, salt-

Vaporized products to atmosphere

Figure 2. Simple seal face combination showing the natural migration of product (red) between the faces. The product vaporizes into the atmosphere.

Figure 3. Double end face mechanical seal with two sets of faces installed on a single seal chamber

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there should be two ports in the gland. Always introduce the Product Atmosphere barrier fluid into the bottom of the gland and have it exit the top. This will expel all air from the barrier area and, after all, heat rises. The barrier fluid should be maintained at 15-25 psig greater than the pressure behind the rotor or in the seal chamber (Figure 5). The higher pressure forces the barrier fluid to migrate or weep Barrier/Buffer Fluid In between the inboard Figure 4. Always introduce a barrier/buffer fluid into faces, an action that the lowest gland inlet. It will exit the top and expel air. hydraulically blocks the sealed product, preventing or sensitive to the atmosing it from migrating outboard. The phere.) faces will not stick together when 2. To capture emissions that would the pump is shut off. otherwise escape into the atmosA good barrier fluid is clean, phere, such as dangerous, lubricating, compatible with the volatile, carcinogenic or ozonesealed product, non-hazardous and depleting products. non-foaming. The barrier system 3. As a built-in spare for added procan control the product migration tection if the inboard seal fails. from the faces, but there is always This is common for vital systems the risk of contamination when the or expensive products where inboard seal fails. The failed inboard interrupting the batch could be seal will make it possible for the costly and problematic. barrier fluid to flow into the One of the most common errors process. Often small tanks of barrier in installing double face seals is failfluid, called convection tanks, are ing to purge air from the barrier used to help alleviate this problem. area (Figure 4). The seal should be The tanks contain only two to five mounted with two barrier fluid gallons of barrier fluid and have a ports, one near the bottom in the sight glass that enables an operator gland and an exit port above the to monitor changes in fluid level— seal. In the case of cartridge seals indicating even the slightest leak. Purge Air Out

,yy,y,y,y, ,y,y,y,y,y,y , y y, y, ,y ,y Product

Atmosphere

Buffer Fluid Pressurized 15 to 25 PSI greater than the product (red)

Figure 5. Double seal with a barrier fluid (green) applied at 15-25 psig greater than the pressure behind the rotor. This setup hydraulically blocks the product (red) from migrating between the faces. The barrier also cools and lubricates the outboard faces.

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Product

Photo 5. A cross sectional view of the standard P/S®-II seal showing an elliptical gland configuration

The convection tank approach is a good option, and it avoids excessive dilution. Often an outside source for barrier fluid such as plant water can leak into the product and go undetected until a serious problem develops. Double seals are often used to contain dangerous products or EPA regulated products that cannot be diluted or contaminated. In these situations, install the double seal as normal but don’t pressurize the buffer fluid (Figure 6). The product will migrate across the faces as with the single seal, and the EPA-regulated emission will migrate into the barrier fluid. It can then be routed to a collection area or recycled. This method is used when handling regulated fluids, but it does nothing to solve the sticking face problem.

Positive Sealing Elements A new sealing option from Garlock Sealing Technologies is called Positive Sealing (P/S®) technology. The heart of this method is the GYLON® thermal elastic sealing

y,y,y,y,y, y , y, y,

,yy,y,y,y, , y y, y, Atmosphere

Buffer Fluid Not pressurized

Figure 6. Double seal with a buffer fluid (green) not pressurized. This fluid can capture emissions and/or act as a built-in spare seal for extra safety. This method does allow product (red) between the faces, which can be a problem.

The Pump Handbook Series

Figure 7. P/S®-III seal with one ring of packing, split spacer bushing and lantern ring

element. GYLON® is a special compound of PTFE base, which was first formulated for use as a gasket material. It resists cold flow and the shrinkage that is typically experienced after a hot-to-cool thermal cycle. Positive Sealing technology incorporates a special lip seal formulation for high performance in some of the most difficult applications. Unlike a face seal or packing set, the P/S® element positively prevents the product from weeping along the dynamic surface to the atmosphere. When product migration is avoided, the product is less apt to change state and become debris between the sealing surfaces. Also, the P/S® seal has only one moving part (the shaft), so there is very little opportunity for clogging. Because the P/S® element can run dry up to 700 fpm, it can handle most PD pump speeds. However, if the temperature is elevated, it may be best to incorporate some cooling/lubrication such as steam, water or a standpipe of oil. As an example,

when the P/S® system is used to seal asphalt at 300°F, the injection of 2-4 psig of steam between the second and third sealing elements neutralizes some of the heat from the sleeve. This cools the sealing elements and reduces coking. Often products that salt or solidify in atmosphere require a buffer fluid to isolate them from the world outside the pump. The P/S® system is available in three configurations. The P/S®-I is a single lip seal that can be installed in the bottom of a seal chamber to augment other sealing systems such as packing or face seals. While most lip seals can seal a maximum of about 10 psig, the P/S®-I can seal up to 75 psig when pressed into a proper size bore. However, if it is held in position with positive backing such as a snap ring or gland follower, it can seal up to 150 psig, unique in the rotating shaft seal industry. Ideally, positive sealing elements should rest on a dynamic surface that has a finish of 4-8 RMS and a hardness in excess of 50 on the Rockwell “C” scale. The GYLON® element is filled with graphite for lubrication and requires no fluid weeping under the lip, as most elastomeric seals do. The P/S®-II cartridge seal is the most commonly used positive sealing system for PD pumps (Photo 5). Its standard three-element, two-port (inlet/outlet) configuration affords the widest possible range of envi-

The Pump Handbook Series

ronmental controls. Often this system can be operated without additional environmental control elements. The P/S®-III is the newest offering in this positive sealing family (Figure 7). It is a system that sandwiches a spiral GYLON® sealing element between two .250” gaskets. Because the three pieces are split, the set can be installed without disassembling the pump. The P/S®-III can be used outboard of packing for further reduction in packing leakage, as a stand alone outboard seal or in the bottom of the seal chamber with bushing segments to apply loading pressure applied by the gland. Only the imagination and sound sealing sense limit the variations.

Conclusion All sealing options mentioned in this article are doing a good job in various applications. Evaluate your PD pump sealing systems and ask yourself if you are happy with your solution. There may be another option. n John Wood is Manager of Technical Services for Garlock Sealing Technologies’ Mechanical Seal Division. He has more than 20 years of experience in the sealing industry. GYLON® is a registered trademark of Garlock Sealing Technologies.

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POSITIVE DISPLACEMENT PUMPS HANDBOOK

Flow Control With Metering Pumps These precise pumps are the simplest way of controlling flow. Learn how to match the right one to your application. By Claude LeFrapper, Zenith Pumps What Is a Metering Pump? he flow produced by a metering pump is independent of the differential pressure it generates (Figure 1). Everyone in the rotating equipment industry is familiar with centrifugal pump “curves,” which show the flow at a given differential pressure, assuming a constant speed. It is easy to see on these curves that any change in the pressure can drastically change pump output. If one were to draw a similar curve for a positive displacement pump on the same graph, with the same scale on the axis, the curve would be very close to a vertical line because the flow rate fluctuates little with pressure changes. When the scale is expanded, the change in flow rate becomes visible and can be significant. Different pumps will have curves that depart more or less from the vertical. The degree of offset from the vertical will depend on the amount of “slippage” or leak

T

between the high pressure side of the pump and the low pressure side. As with any leak, that slippage will increase with the differential pressure, decrease with the viscosity of the liquid, and depend on the tightness of the sealing mechanism between the high and low pressure. A packing seal around a “packed” plunger provides an extremely good barrier between the two sides. Hence, packed plunger pumps have a curve that is the closest to the vertical, with flow that is practically independent of pressure. Practically speaking, the leak can never be zero—and besides, the compressibility of the liquid is not zero, either, a fact that cannot be neglected at high pressures. However, packed plunger pumps are not suitable for all applications, and many other positive displacement pumps can be used as metering pumps, provided that steps are taken to minimize slippage. Many plunger pumps have diaphragm heads,

where a flexible membrane separates the liquid from a more viscous hydraulic fluid that will not leak much past a close-clearance plunger with no seal. Rotary pumps, such as some gear pumps, can also feature very close clearances, allowing very little slippage between high and low pressure sides.

Flow Rate Is Adjustable The flow out of a metering pump is proportional to the displacement of the pump, in cubic inches or cubic centimeters per revolution, and to the speed of the pump. (We must assume here that the liquid is flowing into the pump unimpeded, i.e. that the NPSH is sufficient to overcome the friction losses and, when applicable, acceleration losses in the suction line.) There are therefore two ways to adjust the flow: by changing the displacement, or by changing the speed. A metering pump can be a variable displacement pump running at constant +/-1%=+/-0.8 cc/min

Q

P

Centrifugal Pumps 80 cc/min

x

x

x

x x

x

Metering Pumps

x

+/-1%=+/-0.04 cc/min

Most PD Pumps

4 cc/min x

x

x

x

x

x

x

Time

Q = Mean Delivered Flow

Q

Figure 1. Flow vs. pressure

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Figure 2. Steady State Accuracy The Pump Handbook Series

measures are taken to keep that speed under control. This can take the form +/-0.8 cc/min +/- 1% of simply oversizing the motor, or of employing Actual Error sophisticated speed regulation systems. The overall stability achieved is called “Steady State Accuracy.” It is stated in % deviation Q 8 cc/min 80 cc/min Actual Error = Flow Dependent (e.g. motor speed variations) + Flow Independent (e.g. check over time from the “mean valve closing time presence of bubbles in liquid or hydraulic fluid). delivered flow” at any setFigure 3. Accuracy and turndown ratio ting within a stated range or “turndown ratio.” For a more complete definition speed, or a fixed displacement of the above terms, see API 675, the pump driven at a variable speed. only standard published so far on Conceivably, both the displacement the subject. and the speed of a given pump could be adjusted, and this is someMetering Pump Range times done to extend the pump’s (Figure 3) range. More on this later. Whether the output is adjusted Steady State Accuracy by changing the displacement or by (Figure 2) changing the speed, there will be a minimum flow setting below which The purpose of a metering pump the “steady state accuracy” will no is to deliver a liquid flow that will longer be maintained. This is the remain constant at a predetermined lower limit of the range of flow setting of displacement or speed, achievable that can still be called regardless of minor fluctuations in “metered” flow. Without an accurapressure, temperature, viscosity, cy constraint, ± x% of set flow rate, etc. The ability of a pump to many metering pumps can achieve achieve this will depend on the stazero flow and therefore have an bility of the leak between high and infinite turndown ratio. Conversely, low pressure, and on the evenness by limiting the usable flow rate to a of the input speed. The former narrow range, very high “accuradepends mostly on mechanical cies” can be attained. It is meaningdimensions that are inherently staless to specify one without the ble (with the exception of reciproother. cating pump check valves and the oil refilling system of diaphragm Methods of Flow Control pumps), and the latter depends on Closed Loop on Flow Measurement the ability of the driver to maintain (Figures 4 and 5) a constant speed, which can vary Flow control, like most process greatly from one type of motor to controls, is most frequently of the an other. “closed loop” type. The process Most electrical motors do not variable, flow (generated by any maintain a constant speed under type of pump), is measured by a variable load conditions unless some Absolute Error

Example: 10:1 Turndown Ratio

measuring element, a flowmeter, which sends a signal to a process controller that compares it to a predetermined set point. The controller then signals the final control element, typically a valve, which will open or close in order to bring the flow rate to the desired value. The controlling variable is sometimes the pump speed, a situation requiring a variable speed motor rather than a control valve. The ability of the pump to generate a constant output is irrelevant because whatever “error” that may result from process or environmental fluctuations will be compensated for by the controlling variable, the position of the valve stem or the speed of the motor. Such a flow control system requires four elements: pump, flowmeter, controller, and control valve or variable speed motor. The accuracy of the flowmeter and the response time of the control valve or variable speed drive will determine the “accuracy” of the flow rate. Open Loop (Figure 6) A single metering pump can generate, measure and control the flow at the same time. In most cases, the set point will be established manually, and the pump will reliably deliver a predictable flow rate – within the limits of its mechanical components and of the speed stability of its driver, as discussed in the previous section. Whenever a metering pump is available and will meet the process conditions, it is definitely the simplest way of controlling flow. Closed Loop on Speed (Figure 6) Since the principal source of flow instability is usually motor speed fluctuations, an improvement in the

Variable Speed Controller

Control Valve

FC

FC FM

FM

Figure 4. Centrifugal pump + control valve + flow controller + flowmeter

Motor

PD Pump

Figure 5. Variable speed PD pump + flow controller + flowmeter

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147

Manual Stroke Adjustment

Speed Adjustment

Figure 6. Flow control with metering pump Variable Speed Controller

FC

FM Motor

Metering Pump

Figure 7. Ultimate flow reliability—the pump and flowmeter validate each other

open loop system is to “close the loop” on the pump speed, i.e. measuring the speed, feeding the measurement back to the speed controller, and adjusting the voltage (for DC motors) or the frequency (for AC motors) to maintain the desired speed. This is standard procedure on “servo-motors,” stepping motors and the brushless DC motors commonly used in automation. The approach adds one element to the system, but it is still simpler than closing the loop on flow. For fixed displacement metering pumps, it is the system of choice since a variable speed motor is always necessary. Metering Pump and Flowmeter (Figure 7) For the “belt and suspenders” crowd, a metering pump and a flowmeter in series will provide a constant check of one by the other. This is a true redundancy of the flow measurement and is sometimes required on very critical processes. Both mechanisms will eventually wear out or go out of calibration for many reasons—not to mention process upsets beyond the range of either that may occur. Possible reasons for an upset could

148

be a clogged filter, a pulsation dampener loosing its charge, the flowmeter being driven too slow, etc. Any discrepancy between the flow measurement and the pump flow setting will point to a malfunction of one or the other, or to an unacceptable process condition, and thus guarantee that no incorrect flow will go unnoticed.

What to Choose

$

Metering Pumps

7

40

Q (GPM)

Easy application: water, low pressure Difficult application: viscous, corrosive

Figure 8. Flow vs. cost

For very large flows, there is little choice since only centrifugal pumps can produce thousands of gallons per minute. Controlling such a flow means a closed loop system unless the process calls for “the maximum we can get.” For very low flows, the choice is also simple because only positive displacement pumps can be used, and with flows in the same order of magnitude as the slippage between the high and low pressure sides of the pump, metering pumps are the obvious choice. The tight control of that slippage even allows what is called “differential pumping,” whereby the controlled flow is the difference between the two flows of the two smallest pumps one can find. Controlled flow rates in the cc per hour range can be obtained in this way. There is, however, a wide range of flow and pressure combinations for which it is not easy to select the best solution. Of course, the ultimate choice will be based on economic factors, the cost of the installation and the subsequent maintenance costs (Figure 8). The major factor to consider is that the cost of centrifugal pumps with a closed loop system does not increase as fast as the cost of positive displacement metering systems when the hyd- raulic horsepower, QP (flow x pressure) increases. Whereas the price of a pump will be proportional to its size, the price of the controls is independent of the flow being controlled. There is a break-even point, beyond which centrifugal pumps with closed loop control will be The Pump Handbook Series

Closed Loop Control

more economical, and below which metering pumps will be more cost effective. For “easy” applications, such as a low viscosity non-corrosive liquid at low pressure and ambient temperature, that break-even point will be around 10 gpm (2400 l/h). For more difficult applications, the point shifts upward. High pressure centrifugal pumps, whether high speed or multistage, are expensive, whereas metering pumps can all produce higher pressure even though it is not always required. Because of their relatively small size, metering pumps are more likely to be made of corrosion resistant materials as standard. Centrifugal pumps are also very limited in the fluid viscosity they can handle. For these “difficult” applications, the break-even point will be in the 40 gpm (10 m3/h) range, sometimes higher. Another consideration is whether the process calls for remote or automatic control of the flow rate. Controlling the set point of a flow controller or a speed controller from a remote location is a simple matter. Controlling the displacement of a metering pump from a remote location, or automatically following a process signal, requires additional equipment, namely actuators. This shifts the break-even point downwards, to around 7 gpm (1800 l/h) for an “easy” application, or 30 gpm (7.5 m3/h) for a “difficult” application. The closed loop speed control used on rotary metering pumps falls somewhere in between the variable displacement metering pumps and the closed loop flow control used with centrifugal pumps, since it does not use a

flowmeter but is easy to control automatically.

Conclusion For controlling low flows, metering pumps should always be the first choice. There is a point beyond which positive displacement pumps are not practical, and this is why few are available for very large flows – and even these are appropriate only for very special applications. In the 7 to 50 gpm range (1.8 to 12 m3/h), metering pumps should be considered, but the choice will depend on the many factors listed above. n Claude LeFrapper is International Sales Manager for the Zenith Pumps Division of Parker Hannifin Corporation. He holds a degree in Mechanical Engineering from the University of Strasbourg, France as well as an MBA from the University of Chicago. He has spent 26 years in the field of metering pumps in various engineering and marketing capacities. He lives in Chapel Hill, North Carolina.

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POSITIVE DISPLACEMENT PUMPS HANDBOOK

Want to Go From 0 to 100 in 3 Seconds? Methods for controlling the pumps that control your flow. By Katherine Quarfordt, Milton Roy Company

S

uccess in today’s world requires fast, accurate responses to change. This is true for industrial process plants as much as for any other endeavor or pursuit. How can we ensure that a process is performing as we designed it? By ensuring the chemical dosages of all components are exactly as required. One of the best tools for the job is a metering pump, especially when it is coupled with instrumentation providing process feedback. Metering pumps inject various chemicals at the required rates with accuracy as precise as +/-0.5%. They can change output as much as 100:1 while still maintaining 0.5% accuracy, and they can do it in less than three seconds. Applying the correct control devices, either manually or remotely, can maintain the exact process specifications required. Metering pump injection rates are a function of three parameters: plunger diameter, stroke length and strokes per minute. The formula for determining the maximum theoretical gallons per hour is: 2

(3.14) x (D ) 60 xLxNx 4 231 Where D=plunger diameter in inches L=stroke length in inches N=strokes per minute Adjusting these parameters singly, or in combination, enables the user to alter pump output. The usual methods of changing stroke length are manual adjustment (typically called micrometer adjustment) and the use of pneumatic and electric or electronic actuators. The number of strokes per minute can

150

be changed using variable frequency or variable speed drives. The selection of an adjustment method affects the pump’s accuracy, cost, response time and turndown ratio.

Method 1: Manual Stroke Adjustment Manual stroke length adjustment (micrometer) has been the dominant method of control for metering pumps since their inception. The process requires the operator to know how much adjustment is required to bring about the desired change in the injection rate. To make the change, the operator goes to the pump and adjusts the micrometer either to increase or decrease the setting between 0%100% in relation to the process requirements. The amount of change will be determined by test results particular to the process. The change in rate is the result of the actual change in stroke length for the piston’s movement in nonlost motion pumps or the change in effective stroke length for mechanical lost motion or hydraulic lost motion designs. The rate of change is linear, but not proportional. It is affected by process conditions (e.g., compressibility of fluid and discharge pressure among other factors). A linearity test of the pump in the system under normal operating conditions can help the operator determine the amount of adjustment. This method enables the pump, once set to a given output, to maintain accuracy within +/-2.0% to +/- 0.5% of the manufacturer’s design. (For the purposes of this article, accuracy statements are assumed to be the steady state accuThe Pump Handbook Series

Photo 1. Micrometer adjustment

racy of the set point of the pump (Ref. 1.) There is no increase in price or additional maintenance associated with this feature. This type of control is slow to tune in when a process requires chemical injection changes. It is most effective in processes that vary little over time and when remote signals are not available.

Method 2: Pneumatic Actuator The pneumatic actuator is a control element added to the pump in place of the micrometer knob. It is typically used to adjust the stroke length remotely. An operator can adjust the output of the pump by using a 3-15 psi instrument air signal, where 3 psi sets the actuator to 0% output and 15 psi sets the output to 100%. This method enables the

pump, once set to a given output, to maintain it to +/- 1% accuracy at best. This control can be integrated in an automated process control loop through the use of an I/P converter. The pneumatic actuator is especially well suited for remote control in highly volatile environments where a standard electric powered unit would pose a fire hazard. In some cases, newer technology explosion proof electronic actuators, with their increased responsiveness of control, are replacing pneumatics. Because they use air power, the operating cost of pneumatic controls is fairly low. Maintenance costs are small due to the minimal wear of the internal seals and the use of air power. However, a pneumatic control device tends to be sluggish and slow to settle. Losses in accuracy due to hysteresis are compounded due to the losses in the I/P controller and the mechanical losses in the actuator itself.

Method Three: Electric and Electronic Actuators Electric and electronic actuators can also be added to the pump in place of the micrometer knob. Like pneumatic actuators, they are typically used to adjust the stroke length remotely. Electric actuators are adjusted by turning the power to the unit on and off. Feedback is provided by an internal potentiometer (an instrument that measures electromotive forces or differences in force). The stroke length limitation is determined with microswitches set at 0% and 100%. Some electric actuators utilize clutches to engage the screws of the capacity adjustment assembly. Clutches convert linear motion to rotary motion. Some electric actuators utilize motors to turn the shaft of the adjustment assembly. The off/on pulsing of the actuator to change the pump output responds to the 420 mA or 1-5 volt DC signal. The repeatability of the actuator’s position is from +/- 3.0% to +/- 1.0%. (This is separate from the repeatable accuracy of the pump’s performance as defined by API 675.) The electric actuator is not 100% duty cycle and in some units is as low as 25% duty cycle. This is important when the process changes frequently, especially when parts are worn causing capacity control to bind. The many parts required to make

the electric actuator function are wear components requiring maintenance and replacement on a regular basis. Due to unreliable performance compared to today’s modern electronics, metering pump manufacturers are now replacing these units with electronic actuators. Electronic actuators are best utilized with process control loops. Control inputs and outputs can be 4-20 milliamp (mA) signals, 1-5 volt DC signals, or more recently, digital interfaces (Fieldbus, Modbus). These controllers are comprised of electronic circuit boards and motors. The electronic circuit board is calibrated with potentiometers to set the upper and lower limits of travel and the 0% and 100% set points relative to the DC signal. The servo motor turns a shaft connected to the capacity control assembly to adjust the pump’s stroke length. The circuit board maintains its position relative to the input signal. The adjustment of the pump maintains a +/- 1% accuracy or better. The accuracy limitation is now determined by the pump, not the controller. Electronic actuators are 100% duty cycle, lending to a long life. Upkeep on them is minimal as they require no routine maintenance. The circuit board and motors are replaced at the end of their life (usually 2-5 years for circuit boards and approximately 10 years for motors). These devices can offer a control system that is upgradeable and flexible enough to adapt to a changing process environment. The electronic actuators require more initial investment, but little maintenance due to simplicity of design and long motor life. Thus, there is only a moderate economic impact on the process.

Method Four: Variable Speed Drives Another way of varying pump output is use of variable speed motor drives. These will vary a pump’s strokes per minute. There are four basic types: standard DC drives using motors with brushes, standard AC variable speed drives using inverter duty motors, the new vector AC drives using vector duty AC motors and DC drives using high efficiency motors. When applying any variable speed drive, consideration needs to be given to The Pump Handbook Series

Photo 2. Pneumatic actuator

the significant change in a metering pump’s torque level that takes place between the suction and discharge strokes. One must also factor in the increase in torque required as the speed drops due to the increased friction of the worm gear drive. For more demanding applications, drives with feedback should be employed. The following control methods will affect the output accuracy of the pump based on the accuracy of the control equipment. Standard DC Drives (Motors with Brushes) Standard DC motors are best applied to applications requiring smooth, accurate speed adjustments. They are also suited to equipment with varying torque requirements. These are economical choices for applications up to 3/4 hp. They are typically used in clean, dry, non-hazardous environments. They can be permanent magnet or shunt wound. Permanent magnet designs are the norm in smaller sizes. Turndown is 5:1 without feedback and up to 100:1 with feedback depending on the manufacturer’s design. Standard DC motors utilize linear speed/ torque characteristics and usually operate with SCR controls. The SCR (thyristor based) drive controls the rotating mechanical commutator and carbon brushes to change speed. These drives require more maintenance than the other types. In addition to the standard motor maintenance of bearing lubrication, the brushes need to be replaced periodically. Due to its limited accuracy, this method of control has a low economic impact on the process. Standard AC Drives (With Inverter Duty Motors) Typical performance of these dri-

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Their simplicity adds little cost to the process.

Hybrid Controls Photo 3. Electronic actuator

ves is limited 4:1 to 10:1 turndown. This method allows the pump, once set to a given output, to maintain the accuracy of the motor controller at between +/-1% and +/-5%, depending on the manufacturer. When selecting this control method, consideration should be given to the possible increase in motor horsepower associated with worm gear driven pumps. This increase is the result of the pump’s higher torque requirements at slower speeds. Bearings need to be lubricated. High Performance AC Vector Drives These drives offer performances up to 5000:1 turndown. The various “gearing” options such as worm, cam, helical, spur and pinwheel are affected differently when driven at faster speeds, thus enabling a larger turndown ratio. When selecting this control method, consideration should be given to pump design so as to not jeopardize the efficiency of the process. High Performance Brushless DC Drives Brushless DC motors are selected when precise speed and torque control is required. The motors must be paired with the appropriate solid state controls due the electronic commutation process. These drives use a feedback loop to monitor and maintain motor speed. They offer up to 4000:1 turndown with analog control and circuitry, and 16,000:1 with digital systems. The electronic design and precise control features make them more costly, however. Another type of DC drive is the stepper motor. Stepper motors allow accurate speed control and impressive turndown ratios even when applied without a feedback loop. Stepper motors are common in instrumentation such as the electronic actuators discussed previously. They can operate directly from a 4-20 mA or 1-5 vDC signal or from a state-of-the-art digital communications system. There is no maintenance required for these motors.

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Combinations of the above control methods can also be utilized. A hybrid control uses a variable speed drive in addition to electric, electronic or pneumatic actuators. This type of control is practical when two process control signals are required. When used together, automatic control from two different signals can adjust the pump’s output in response to changes in two varying process conditions. An example of this application is pH control in a water plant. When the flowrate through the plant changes—for instance, from 300 gallons per minute (gpm) to 150 gpm— the flow meter sends a signal to the motor controller, which adjusts the stroke speed of the pump, thereby changing the injection rate of the base and/or acid proportionately. However, the pH of the water may also change. If this happens, the pH controller sends a signal to the actuator, telling it to adjust the stroke length accordingly. The use of multiple control devices offers flexibility in the process. This option will require additional wiring, power resources and the associated maintenance costs, which must be taken into consideration when figuring overall cost. In terms of accuracy of injection, varying the speed provides higher accuracy than varying the stroke length. Hydraulic losses occur on each stroke in metering pumps. These losses are due to the slight compressibility of the fluids and hydraulic leakage. The hydraulic performance of a pump can be characterized by the following formula: Pump Capacity = 60 x SPM x % stroke x theoretical volume - fluid compressibility - hydraulic leakage As can be seen from the above formula, the capacity losses from fluid compressibility and hydraulic leakage become larger in proportion to the theoretical volume as the stroke length decreases. In variable speed control, because the stroke length is a constant, the losses remain the same nominal percentage. When stroke length and variable speed control are used together, the accuracy of the stroke length control becomes the limiting factor in the repeatability of the pump. The Pump Handbook Series

We can apply the root mean squared error equation to aid in evaluating the effect of hybrid control. Total error=√error 12+error 22+…+error N2

Therefore, the turndown ratio of the hybrid control is not a simple multiplication of each control method’s turndown. For example, an electronic actuator with a turndown of 10:1, when used with a variable speed drive with a turndown of 50:1, will not yield a useful 500:1 chemical injection turndown rate. In practice, the use of multiple control devices limits the turndown to approximately 30:1 to 50:1 depending on the pump “gearing,” stroke speed and design. How can a process be changed with high accuracy and cost-effectively in three seconds? This can be accomplished by capitalizing on the advantages offered by variable speed control and those pumps in the industry that have been specifically designed for these high performance drives. This has been achieved with high performance brushless DC motors and the high performance AC Vector motors coupled with their respective drive systems. Success is achieved when a process is maximized with equipment that provides the most value for your investment dollar. The economic impacts are different with each control method. When selecting the appropriate control method, initial investment, maintenance and expandability needs should be considered. In the future there will be requirements to do everything faster and more efficiently. What will be the next generation of equipment offering solutions to our processes? The author wishes to thank Bob Bean, Don Weidemann and Craig Nagler, all of Milton Roy Company, and Bronson Clay of Oliver Equipment Company for their contributions to the article. n

Reference 1. American Petroleum Institute, Positive Displacement Pumps— Controlled Volume, API Standard 675, Second Edition, October 1994 Katherine Quarfordt is part of the Industrial Process Customer Focus Unit of Milton Roy Company, Flow Control Division. Ms. Quarfordt received her B.A. degree from Texas A&M University.

POSITIVE DISPLACEMENT PUMPS HANDBOOK

Maximize the Performance of Progressive Cavity Pumps If you choose the right geometry, protect the parts, consider the application carefully, control temperature and pressure your PC pump could last a long, long time. By Michael L. Dillon, seepex, Inc.

s positive displacement pumps, progressive cavity (PC) pumps have the same benefits as other PD pumps. They can handle high viscosity fluids; they can produce accurate repeatable flow; the output capacity is relatively independent of head; and they can operate with fairly high efficiency at high heads. One of the major progressive cavity pump benefits is that they have no valves. A PC pump works like a piston pump, but with the piston operating in a cylinder of infinite length. Without valves, the pump will not air lock, clog, foul or leak, and flow is predictable and repeatable. The absence of any valves on the suction side of the PC pump contributes to its low NPSHR. The PC pump is also excellent for abrasive slurries, since in its most common form, it is rubber lined. It can be repaired on-site with only a few tolerance fits. It can handle solids up to several inches in diameter, is good for low shear pumping, is self-priming, has a reversible flow direction, and has only a single shaft that requires sealing (Figure 1).

A

Pump Theory Renee Moineau, an aircraft designer who was trying to invent an engine supercharger, designed the PC pump in 1936. The range of Dr. Moineau’s patents is truly amazing. Many of his designs have only just recently become commercially available due to their manufacturing complexity.

Figure 1. A progressing cavity (PC) pump fitted for industrial and municipal service

The most familiar of the PC pump designs uses a single metal rotor, machined as an external helix, that revolves eccentrically within a rubber injection molded double internal helix that is twice the pitch length. The rotor is usually metal, the stator elastomeric with a compression fit. However, where a tolerance fit is used, rigid materials such as metals and plastics can be used for these components. With rigid materials, PC pumps behave similarly to other rotary PD pumps such as gear, screw or lobe pumps. These designs are a special consideration and will not be covered in this discussion. Discrete cavities are created when the rotor and stator are combined. The cavities spiral or progress through the pumping elements as the rotor turns. As one cavity diminishes, the following increases. The fluid cross section is unchanged, regardless of rotor position, so it functions like a piston in a cylinder of infinite length. One of the unique and defining properties of the PC pump is that more than 90% of the loading is axial instead of radial The Pump Handbook Series

Figure 2. Conventional (top) and long (bottom) geometry single helix rotor and double helix stator designs

because the liquid is moving along the same axis as the rotating parts (Figure 2).

Lengthened Pitch Geometry Lengthened pitch geometry was developed in early 1970s as a result of the availability of whirling head lathes—cutting machines in which the bar stock is held stationary and several cutting tools are placed in a head that rotates or “whirls” around the bar stock. Because the cutting forces were balanced it became possible to manufacture very thin and long rotors. The previous method of cam operated single point machines required that rotors be fairly short and thick to minimize deflection

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Figure 4. Failure of chrome plating due to corrosion of rotor base material (carbon steel)

Figure 3. Double helix rotor and triple helix stator design delivers 50% more flow but with 100% increase in internal velocities

caused by the cutting tool. In the new lengthened pitch geometry, cavity volume could remain unchanged while pitch length increased and cavity and rotor diameter decreased. Reduced rotor diameter and circumference resulted in reduced surface velocity at the same rpm as the conventional geometry. The reduced diameter and rotor cross sectional resulted in thrust loads that were similarly lowered. This is because thrust load (lb.ft.) is a function of the cross sectional area of the rotor (in2) times the differential pressure (psi). The longer geometry also increased the sealing line cross section while maintaining the same compression. This yields less slippage and higher volumetric efficiencies than the conventional geometry. Longer pitch geometry is useful for abrasive applications because of its lowered velocity between the rotating and stationary components. It can also accomplish low pressure (<60 psi) metering because it has a flatter performance curve. It should not be used on high viscosity liquids, large solids or very low NPSHA applications, however, because the entrance into the cavity is considerably smaller than on the conventional design.

Multiple Helix Geometry Dr. Moineau also designed the newest development in progressive cavity pumps. Multiple helix rotors and stators became popular and have been used for drilling mud motors since about 1980. They became affordable for pump designs in 1993 when new whirling machines that could form double helix rotors became available (Figure 3). Multiple helix designs can include

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an unlimited number of helices on the rotor and stator, as long as there is one more helix on one member than its mate. The most affordable and practical of these designs is a double helix rotor and a triple helix stator. In this design, the fluid crosssection decreases by 25%, but the internal velocities double. Flows increase by 50% in the same physical space. Additionally, in the 2⁄3 multiple helix geometry starting torque decreases with equivalent running torque. This design is good for thin liquids and abrasives, and it is excellent for variable frequency drive (VFD) applications. It should not, however, be used on high viscosity liquids, to pump large solids or on very low NPSHA applications. This is due to the reduced size of the opening into the cavity in relation to the linear velocity of the liquid. Comparisons to determine which pump is best should be between similar pump styles and models from different manufacturers. Internal linear velocities and the velocity of the surface of the rotating parts must be compared, as well as the NPSHR, to determine which pump is best-suited for a particular application. Certainly the new 2⁄3 geometry pumps will provide the lowest cost per unit pumped, but they are severely restricted on other parameters.

Mechanical Difficulties Unfortunately, PC pumps are prone to a variety of mechanical difficulties. These can be classified into five major areas: pumping elements, universal joints, shaft seals, bearings and drives. Rotor Problems Erosion is the most common problem for rotors. Operators should try to use the hardest material that is chemically compatible with the pumped liquid. Rotor coatings, such as chrome plating, ceramic and other special hard coatings (which are harder than the rotor base material) will all increase the longevity of the rotor. Unfortunately, some care must be taken with the coatings. If the The Pump Handbook Series

coating cracks, peels or flakes, the rotor will destroy the stator. Remember that the rotor is loaded with thrust by the discharge pressure, and the drive train of the pump is trying to keep the pressure from expelling all of these components out of the rear of the pump. The rotor bends as it rotates, sometimes more than slightly, and brittle materials can easily crack. Sometimes it is better to use hardened base materials with no coating than to use a very hard but easily fractured coating. Another common way to increase rotor life is to derate the pressure capability of the pump as the differential pressure increases. While this increases the initial cost of the pump, and decreases its mechanical and over-all efficiency, lowering the “pressure-per-stage” can dramatically increase rotor and stator life. Since the discharge end of the stator is pushed away from the rotor by the pressure of the liquid, the stator becomes conical on the discharge side of the pump. This results in a reduced contact area on the suction side of the rotor and stator. Using more stages (or rotor and stator helix lengths) in the pump creates more contact area between the rotor and the stator. It’s like adding tread depth on a tire; there is simply more material to wear away and part life is extended, sometimes dramatically. While using pumps for higher pressures can increase the purchase price by 20% or more and increase power consumption by 50% or more, this selection technique can extend the mean time between maintenance requirement by a factor of four or six or even more. Corrosion is also a problem with rotors, and selection of the correct base metal is extremely important. Luckily, there are a variety of material compatibility guides available. If there is any question as to compatibility, the pump supplier should be willing to provide test coupons of any material used in the pump, whether it is metallic, elastomeric, plastic or ceramic. Coating failures,

Figure 5. Rubber to metal bond failure on a “cut-to-size” rotor

as previously mentioned, are common. Hard chrome is subject to porosity problems, and the cracking of ceramics is also common. If the base metal starts to corrode, the coatings can easily “lift.” It is critical that there is no possibility of base metal corrosion. In many cases, where some corrosion of the base metal is inevitable, rotors without any coating are the best choice. In fact, very few Hastelloy® or titanium rotors are sold with coatings because the fluids being pumped will have some corrosive effect on the base material over time (Figure 4). There are some special rotor problems related to material failure. Generally, chrome plating cannot be used on hardened steels because of the hydrogen embrittlement of the base metals during the chrome plating process. The base metals can crack and, unless X-ray inspected, the rotors can actually break in operation. Special coatings can be used with hardened rotors; however, some of these are proprietary. Corrosion of hardened steels in intermittent use can also be a problem. Hardened steels are not good for applications where the pump is routinely drained, as corrosion rates are amplified with the increased exposure to oxygen. It is also important to use separate materials in the u-joint to prevent galling. This is particularly true for designs such as food grade pumps, where the rotor is an integral part of the joint. Stator Problems Elastomer compatibility, along with dry running, is one of the most common problems with PC pumps. While the use of an elastomer stator increases the utility of the pump for abrasive or solids laden applications, chemical compatibility can cause a whole new set of difficulties. Again, there are a variety of reference

sources that can help with elastomer selection. If there is any doubt as to the acceptability of a material, conduct an immersion test that lasts at least two weeks. Make sure to conduct the test at the temperature at which the pump will operate, as compatibility is dramatically affected by this variable. In general, the pump can tolerate a 10% change in the hardness of the elastomer and a 5% change in the volume of the elastomer. Undersizing the rotor diameter can easily accommodate volume change due to temperature. However, a volume change from chemical absorption can not be similarly tolerated. If the elastomer swells due to chemical attack, it will usually continue to increase in size, and its physical strength will deteriorate, eventually causing premature failure. Abrasive resistance varies widely with elastomers. While Buna N (NBR) is the most commonly used and least expensive base polymer for stators, a relatively new compound, hydrogenated nitrile buna rubber (HNBR), is far superior. Of course, HNBR is considerably more expensive than NBR. Just because a material is chemically compatible, there is no guarantee that the material will provide the longest life. Fluoropolymers (FPM) are notoriously poor for the physical properties needed by PC pumps. There are compounds that are both less expensive (HNBR) and more expensive (Aflas®) that might be better choices. The choice ultimately depends on the material pumped. Certain compounds, buna among them, and especially white compounds that are filled with kaolin clay as opposed to carbon black, are hydrophilic. Pure water, deionized and distilled water applications require special elastomers. Another problem with some elastomers is the high mineral oil content (This is one reason why old truck tires burn so well.) Some hydrocarbons can break down these oils over time and cause the elastomer to shrink and harden. A final concern of stator compatibility is the presence of gases in solution in the pumped liquid. Some gases, like CO2, can be absorbed by the stator in their compressed or liquid state. This has no effect on the performance of the pump or the elastomer, but if pressure is relieved from The Pump Handbook Series

the system and the material changes phase from liquid to gas, it will completely destroy the elastomer. Bonding adhesive failures are also common problems, especially with stators that are “cut-to-size.” Some manufacturers injection mold their stators in long, multiple-stage sticks and then cut off the number of stages needed. This exposes the adhesive that holds the elastomer inside the metal tube to the material being pumped. This may not be a problem for many applications. However, it is a common problem with EPDM elastomers on applications with ketones. EPDM is resistant to a wide range of chemicals, including many adhesives, and most adhesives are easily dissolved by ketones. Separating the adhesive from the pumped liquid by molding an integral rubber lip at the end of the stator is the best way to prevent adhesive failure. Gaskets tend to not be very effective because the rubber is usually pulled away from the metal when the stator is cut. This can also be a problem if carbon steel stator tubes are used for applications where all of the wetted parts are made of stainless steel or other corrosion-resistant metals. Bonding adhesive failures are also common on jacketed pumps, as heat reduces the strength of the adhesives holding the rubber inside the metal tube (Figure 5). Controlling the temperature at a point below the manufacturer’s recommendation for the elastomer is important. Thermostats for hot oil, water or resistance heaters, as well as pressure regulating valves for steam jacketed pumps, are an absolute requirement when heating the pump, as heat destroys the bonding chemicals. Another stator failure problem is named after the hysteresis viscosity curve used to measure the cure rate of the elastomer during vulcanization. Excessive flexing of the rubber due to high temperatures and/or pressures, which activate the vulcanizing enhancing additives in the elastomer and can cause additional hardening, causes hysteresis failures. An easier understood explanation is the “exploding truck tire” phenomenon. No one ever sees big chunks of truck tires inside town. They are always on the highway. If a truck is overloaded, with too much

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Figure 6. Thermistor imbedded in stator to protect against run-dry damage

compression on the elastomer tire, it can run at low speeds without a big problem. If this same truck runs at higher speeds on the highway, the elastomer is compressed and relieved at a higher frequency. Heat builds up that can cause the elastomer to further cure and become hard, like plastic. Without resiliency the elastomer starts to break apart and big chunks of rubber come off the tires, just like they will come out of the end of the pump. To avoid this problem, the rotor must always be sized for the correct temperature and the pump should never be exposed to higher than rated pressures. Hysteresis failure can be caused by repeated “dead head” operations that take place for short periods of time but at high frequencies, which is why it is always important to cycle valves when the pump is not operating. Run Dry Damage The most common problem with stators is probably run dry operation. The compression fit used with elastomer stators needs lubrication to carry away heat. Heat will build up if there is no fluid, or if there is fluid but it is not moving through the pump. In this situation, it is heating up to above the temperature limit of the elastomer. There are various devices available to prevent the problem, but all of them have limitations. Thermistors can be imbedded in the stator to measure the operating temperature, and a set point controller can be used to shut off the pump at a specific temperature. These devices are inherently very reliable, but are only available as part of a new stator; they cannot be used with adjustable stators; they

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are not allowed in dairy (3A) applications, and they cannot be used on very small PC metering pumps (Figure 6). Paddle-type flow switches can be used, and they are very inexpensive. However, they are not recommended on viscous or solids-laden materials, which are a common application area for progressive cavity pumps. Heat transfer flow switches are reliable on a great number of applications, but they are generally more expensive than the thermistor devices and are not suitable for most explosion proof applications. There are a number of fluid detectors that can be used to protect the pump. Keep in mind that these devices generally only measure the presence or absence of a fluid and not heat, which is the real cause of stator failure. These devices should always be installed with a pressure switch to prevent either dead head or closed suction conditions. They also have to be installed in a vertical or self-draining line to be effective. Capacitance-type detectors are about the same price as a thermistor ($650), and are available in a food grade design at a somewhat higher price. Some customers have used tuning fork fluid sensors, but they need to be inspected periodically because they can foul and become unreliable. They are, however, less expensive than the capacitance type fluid detector. It is also possible to integrate flowmeters into process instrumentation to prevent run dry operation. This is the least expensive way to protect the pump, provided you can get the equipment installed, programmed and operating before start-up. Pressure switches can also be The Pump Handbook Series

used, with some precautions, to protect the pump from run dry damage. A switch, used on the suction side with a minimum pressure set point, can prevent damage to both starved suction and emptied suction tanks (if the static fluid level is high enough to detect these conditions). Many customers also like to draw down tanks to empty, but there is not enough sensitivity in a standard pressure switch to draw a tank down to zero static head without running the pump dry. A pressure switch, rupture disc or pressure relief valve should always be used on the discharge side of the pump to protect from pumping against a dead head. These pumps are positive displacement pumps, and if they are operated against a dead head, they will try to build pressure until one of the components or the piping fails. Pressure switches can protect against running the pump dry by using a dual set point switch that protects against both over and under pressure conditions. Discharge pressure is a combination of both static head and friction loss. If a suction line is closed or the feed source empties, the friction component of the total head will disappear and the recorded pressure will drop. This system is not very reliable when the friction loss component of the total head is very low. Universal Joint Problems While some manufacturers try to make universal joint design a major difference between progressive cavity and other pumps, they generally do it to give themselves an advantage in written specifications. The truth is that for most applications, it is only important to ensure that the universal joints are positively sealed from the pumpage and are properly lubricated. Sealed u-joints are imperative, as any auto mechanic working on front wheel drive cars can tell you. There is also no real use in having a universal joint that will not need rebuilding before the rotor needs replacing, since you have to disassemble the joint to replace the rotor. Generally, it is also advisable to have sacrificial or low cost wearing parts in the joint. This will minimize the cost of repair and makes it fairly easy to work on. Gear type u-joints are notorious for their high replacement cost and repair difficulty, but

of the pumping elements, and the drive train is mounted in a resilient mounting (the elastomer stator). Cavitation, however, will have a negative effect on your pump.

Figure 7. Gear type and sealed pin-type universal joints in comparable sizes

they are very good for high thrust load applications. Remembering that thrust load is a function of pressure and rotor diameter, high thrust loads are only seen in very high pressure situations (>1000 psi) or in very large pumps (rotor > 6”). Cardan joints can also be used in these applications, but they have to be oil filled, the same is true for gear type universal joints. For 95% of PC pump applications, simple sealed and grease lubricated pin joints, which are made by almost all of the major PC pump manufacturers, are acceptable (Figure 7). Excessive thrust loads caused by high differential pressures (ÆP), as mentioned above, can cause problems with the universal joints. On abrasive applications the stator and rotor will normally fail before the universal joint. The opposite is true for non-abrasive, ambient or lower temperature applications—the joint will fail first. The weakest component in these cases is usually the joint lubricant. The high thrust load combined with friction between the joint components can produce enough heat to vaporize the lubricant, which is why some manufacturers in their larger or higher pressure pumps use oil-filled joints. It is always wise to use only the manufacturer’s recommended lubricant in the joint and check the operating pressures, especially on the second and third shift operations where plant personnel are always more willing to use throttling valves than variable speed drives to control flow. Suction pressure can cause joint problems as well. If the seal that protects the joint from the pumpage is damaged or displaced, the joint will fail. Many gear joints are limited to 25 psi in the suction casing. There are some double seal designs that can handle up to 50 psi. Hydraulically balanced sealed pin

joints are available that can operate up to 175 psi. This is something that you really need to consider if you are operating the pump in reverse, such as with a suction lift application or when pulling against a vacuum, i.e., a crystallizer or concentration unit. The discharge pressures will affect the universal joints. Universal joint angularity, again, is a point pushed by some manufacturers to try to exclude competition. The truth is that the correct angularity varies with the type of universal joint. The angularity must match the u-joint design. For gear joints it should be < 2°, because they generate a lot of heat due to the thrust plates, which enables them to absorb the higher thrust loads. Pin joints should have an angularity of around 3°. Cardan joints need to have an angularity of > 5° to ensure that the needle bearings inside of each bearing cup rotate to promote even wear. Shattered joint parts are a sign of cavitation, and this happens in any type of universal joint in a PC pump. The implosion of vapor bubbles on the discharge of the pump causes high momentary thrust loads on the drive train. Most manufacturers use some hardened steel or cast iron components in their joint, and the shock loads caused by cavitation can fracture these parts. In gear joints with soft alloy thrust plates, galling and deformation leading to failure of the joint seal is common with cavitation. Pumps with flexible shafts, instead of u-joints, will break these shafts. Cavitation should be avoided in any pump, and while Dr. Moineau originally designed PC pumps as compressors, cavitation will have a damaging effect over time. PC pumps can withstand some cavitation because they operate at slow speeds. There is a pressure gradient as the fluid travels the entire length The Pump Handbook Series

Sealing Problems PC pumps are famous for not having sealing problems. Compared to many other rotary PD pumps, they have an advantage because they only have one shaft to be sealed. This seal is usually only exposed to the suction pressure condition and 95%+ of the total load is axial rather than radial. In cases with low NPSHA or suction lift applications, as mentioned previously, it is advisable to run the pump in reverse. This prevents air from being sucked in through the packing or the mechanical seal. A flush or a quench can help alleviate this problem. Failure to prevent this loss of prime through the sealing area can lead to run dry damage of the stator. When operating with packing, it is important to flush abrasives with a lantern ring spacer in the packing and an API plan 32 or 52 flush. These flushing systems create a pressure barrier in the packing that prevents abrasives from entering. They also keep the packing cool and lubricated, and this prevents damage to the pump shaft. There is a myriad of packings available, and they can be fitted to any PC pump. Your PC pump manufacturer or local distributor can install your favorite packing for you. Like pump vendors, if you have a packing supplier that you trust and provides good service, use his or her recommendation. Your trusted pump supplier will respond in kind. For packing, though, it is always recommended that you use a hard coated shaft or shaft sleeve. Packing rubs against the shaft and where there’s rubbing, there’s wear. The harder the shaft, the longer it will last on packing. Single seals are gaining in popularity. For seepex, more than half of all the pumps we sell have single seals. The prices of hard faced, silicon carbide or tungsten carbide seals, especially rubber bellows seals, have gone down dramatically over the last several years. Generally, the price for a single seal is equal to the cost of a good packing with a

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casing, require expensive modifications to accept this seal. Double seals, of course, are still widely used. However, most double seals in use now come in the form of a cartridge seal. Double seals are still widely used on fine abrasives (like pigments) and crystallizing materials (like latex), which can foul single seals with a quench. Double seals must have a flush system with the seal flush liquid at a pressure that is usually 5 psi higher than the pressure in the suction housing. Again, an API plan 52 flush system, which can cost as much as a small PC pump, needs to be installed with a double seal. If a double seal is not flushed, it can be destroyed within minutes. One last word regarding double seals: Split cartridge seals may not fit into some PC pumps. Please be careful when using these seals, which have very large diameter glands. Some manufacturers are making units that will fit. Others may fit but may require modification to the suction casing or the bearing housing of the pump. The added cost may not be worth the added convenience of the split seal.

bearing cover plate is important, and while bolted plates are easier to use than snap ring fitted plates with adjustment shims, the tolerances are the same. Once properly set and lubricated, you can expect tapered roller bearings to last a long time. Close coupled PC pumps are becoming more popular. In Europe, the majority of PC pumps sold are close coupled or block configuration. These arrangements have made it possible for PC pump suppliers to reduce the costs to users by as much as 40%. It has also helped to reduce lead times. Additionally, it is safer than “V” belts and pulleys, and more reliable. It’s a great idea. But like a lot of great ideas, it can be abused. These units are typically oil lubricated. If a manufacturer proposes a close coupled pump, make sure that the gear box is rated with, at the minimum, a 1.5 service factor based upon the motor input power. This is the minimum factor according to AGMA for class II gearing with a positive displacement pump. Secondly, ensure that the pump manufacturer provides you with the actual maximum thrust load calculation and the maximum thrust load rating of the gear box. The gear box should be rated for more load than the actual calculated load. Otherwise, use a larger gear box or a pump with dedicated bearings (Figure 8). Protection of the bearings from contamination is always a major concern. Use IP65 double lip seal protection or labyrinth seals for the bearings on fine slurry or coating type fluids to prevent contamination. Because of the extremely slow speeds used with PC pumps on highly abrasive applications, shaft slinger rings can be useless. High quality fluid-end and bearing seals will solve most bearing problems associated with PC pumps.

Bearing Problems Bearing problems, because so much of the load is axial rather than radial, are not a severe or common problem in PC pumps. First, new bearings always run hot (up to 160°F), and they will take several days to run in. Because of the high axial loads, avoid ball bearings except on very small pumps. Tapered roller bearings can handle higher thrust loads and are a better investment. Proper fitting of the

Drive Problems Until about a decade ago, variable pitch pulley belt drives were the most popular drives for PC pumps. They are still one of the least expensive ways to achieve variable flow in a rotary pump. Unfortunately, there are some sizing and reliability concerns with these drives. It is very important to ensure that the proper hp rating is used. Some manufacturers rate these units on their input power, and some rate

Figure 8. Integral or “block” pump construction (top) where gear reducer bearings absorb pump thrust and radial loads

hard coated drive shaft and a lantern ring seal flush. In some PC pump designs, it is actually easier to replace a seal than it is to replace packing. There are a few caveats with single seals. You should use a quench if there is more than a five foot suction lift when running the pump in standard rotation. A quench is just an area behind the seal faces, sealed by a rubber lip seal, where a clean liquid (water or machine oil, generally) is present at atmospheric pressure. This forms a viscous barrier that prevents air from entering through the seal when the suction pressure in the pump is at less than atmospheric pressure. The quench can also be used to prevent crystallization of the product on the seal faces. Single seals with quench have proven to be effective on paper coatings, paint, sugar solutions, honey, salt solutions and a variety of other crystallizing materials that are known for destroying packing and single seals without a quench. On certain applications it is important to use slurry seal housings for abrasives. It is advisable to keep these hard faced seals cool by placing the seal inside the suction casing of the pump. Ceramic or carbide seal faces can fracture or chip if they are allowed to get too hot. The pumpage keeps the seal cool and prevents the build up of solids around the seal. This cannot be achieved by placing the seal inside a stuffing box designed for packing. Some PC pumps can easily accommodate this seal mounting arrangement. Others, which have the stuffing box cast as part of the suction

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The Pump Handbook Series

them on their output power. This can be a big difference, due to the mechanical losses associated with both the variable pitch belt and the gear reduction unit. Make sure that you have enough power at the drive output to power the pump. Since these units are invariably integral with a gear reduction unit, make sure that the service factor of the gear box at low rpm is sufficient. Because they are a mechanical drive, the torque increases as the speed is reduced. Some units may have very low service factors on the gears at slow speeds—some may be less than 1.0. It is recommended that customers buy the hardened or hard coating pulley option on these drives. The belt has a tendency to run a groove in conical pulleys. Most manufacturers recommend that users run the drive all the way to its maximum speed and all the way down to its minimum speed once a week to prevent grooving. I’ve never known a user to have this step in their formal PM procedures, so the hard coating is a worthwhile safety precaution. One of the most common problems with gear boxes and mechanical drives, which are integral with gear boxes, is venting. If the box is not vented, vapors will build up and the pressure will blow out the oil seals. Be sure gear box breathers, which are shipped separately to prevent loss of lubricant during shipping and installation, are installed. Failure to install breathers is probably the most common cause of gear box failure. Mechanical friction drives can be used on PC pumps, but it is important to ensure that the drive is rated for maximum pump starting torque and not the maximum running torque. The shock load associated with PC pumps can shatter the phenolic friction ring in these drives if they are undersized. Again, be sure that the gear box breathers are installed. Electronic drives have become extremely popular for all variable speed applications in the last decade, including PC pumps. While the electronics have eliminated a lot of the problems associated with mechanical components, they have given rise to other problems. Progressive cavity pumps are constant torque devices,

if the differential pressure is constant. Therefore, if a variable frequency drive (VFD) is used, it must be of the constant torque type. VFDs must be sized considering the drive’s starting torque capability as well as its operating torque capability. VFDs cannot generate as much starting torque as a mechanical variable speed drive, and PC pumps, due to the compression fit between the rotor and the stator, require a lot of starting torque. Depending on the operating pressure, starting torque is usually 50% higher than the running torque and can be as high as two or three times the running torque if the pump pressure capability has been severely derated to improve rotor and stator life. Use the following formula to calculate the proper drive size, given the pump maximum speed and starting torque in lb. ft.: (lb. ft. x rpm)/(5250 x starting torque current boost)=VFD horsepower VFDs are basically computers, and they can be difficult to program. Of course, each VFD is different; but there are a few guidelines to remember: • Set the current boost for starting to the maximum setting. • Minimize the ramp up and soft start capabilities. This increases the amount of starting torque. • Locate the VFD as close as possible to the motor. Problems start to arise when the VFD is more than 100 feet from the motor. • For more turn down (>6:1), set the maximum pump speed @ 90 Hz with four- or six-pole motors. This will enable you to use a higher reduction on the gear box, which provides more torque for starting. It will also have the motor running at higher speed to allow for improved cooling of the motor. Just about any type of prime mover including, but not limited to, air motors, hydraulic motors, DC motors, gasoline and diesel engines can drive PC pumps. There have even been hand-operated PC pumps. Remember to match both the starting and running torque requirements, and size the unit over its entire operating speed range.

The Pump Handbook Series

Application Problems Some very specific application problems need to be mentioned to finalize this topic. High viscosity applications usually require open hopper pumps. It is important to size the hopper so that material will not “bridge” in the unit. These pumps are available in a variety of designs depending on the particular application conditions. They are available with and without extension tube “induction” zones to improve the volumetric efficiency of the pump, and they are available with internal cutters for chopping up potatoes or beets or fruit. Special augers, some with mixing capabilities, can be installed as part of the drive train, and separately driven “bridge breakers” can be installed in the suction casing of the pump. These not only protect against bridging of the product above the open hopper pump’s auger, but will impart additional shear to lower the apparent viscosity of thixotropic or pseudo-plastic fluids. This makes them easier to pump (Figure 9). Excessive speed is another common cause of premature PC pump failure. Again, rpm is not a useful measure of speed. What is important is the surface velocity of the rotor against the stator. Necessarily, as higher flows are needed, the cavity in the pump increases in both length and diameter, and the rotor summarily increases in diameter and circumference. Some broad guidelines for maximum speed relative to the flow rate required are: > 500 gpm =< 250 > 50 gpm =< 350 > 5 gpm =< 500 > 0.5 gpm =<1000

rpm rpm rpm rpm

Cavitation, as mentioned previously, can cause problems with the universal joints, and it can damage the rotor and stator. Don’t ever install a pump without comparing NPSHA to NPSHR. Even though the suction may be “flooded,” high viscosity applications commonly have enough friction loss associated with valves and piping to reduce the NPSHA to a level below the NPSHR. It is also important to know the standard being used by the pump manufacturer to measure NPSHR. There are differences between the standards used

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and prevent settling. Using a long or multiple helix geometry will also increase the linear velocity in the pump while still keeping the surface velocity of the rotor low.

Conclusion

Figure 9. Open hopper type pump used for high viscosity liquids and sludges. Note auger feed device.

by the Hydraulic Institute, API and other organizations. The pump may not produce according to the published performance curve, even if the NPSHA is above the manufacturer’s listed NPSHR, because of the standard used to measure NPSHR. It is important to install gauges on both the suction and discharge sides of the pump. Some manufacturers do not have gauge connections on their pumps or they may only offer them at an extra charge. It is impossible to diagnose a pump problem in the field without pressure gauges. Pressure is, after all, one of only two components that defines work in a pump. If no provision is made for gauge installation, your piping will have to be changed. It’s like flying in a snow storm without an altimeter. Excessive pressure and/or temperature in an application will cause several of the earlier listed conditions: stator hysteresis, u-joint or bearing failure. It is imperative that you know the temperatures and pressures for your PC pump installations. One of the things that PC pumps are not good for are applica-

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tions with wide temperature fluctuations. This is due to the expansion and contraction of the elastomer. High temperatures will cause either increased erosion rates or hysteresis failure of the elastomer. Low temperatures will cause reduction of the compression in the pumping element, excessive slip and necessitate premature replacement of the stator and/or rotor. Proper selection of the elastomer material will help to minimize this problem, but applications with temperature fluctuations of more than 150°F are generally not recommended for PC pumps. While PC pump manufacturers like to promote slow speeds to increase the life of the pumping element, this can backfire on some heavy and hard-solid laden slurries where there is insufficient internal velocity, and the solids settle in the pump cavities. In these instances, rotors will wear as fast or faster than stators, u-joint covers can get holes worn in them and fail, and stators will quickly erode. The best solution is to install an auger or propeller type coupling rod to add turbulence

The Pump Handbook Series

Progressive cavity pumps are versatile and adaptable for a wide range of applications. Unfortunately, they are somewhat more “sensitive” than other more commonly used pumps. If care is taken with proper selection and installation, PC pumps can be a superior choice. Choosing the proper materials of construction, speed, geometry, seals and drive are only part of the job for a good pump installation. Proper installation requires the inclusion of a reliable and appropriate device to prevent dry running and offer pressure protection. In addition, process temperature controls and properly programmed electronic drives will ensure that your PC pump is dependable for a long time. " Michael L. Dillon is Vice President and General Manager of seepex, Inc., a manufacturer of progressive cavity pumps in Enon, Ohio. A graduate of Wright State University with an MBA, Mr. Dillon is the author of numerous articles on positive displacement and progressing cavity pumps.

POSITIVE DISPLACEMENT PUMPS HANDBOOK

Minimizing Pressure & Flow Pulsations from Piston/ Diaphragm Metering Pumps Dealing with the pulsating flow of these precise pumps requires knowledge, a little know-how and maybe even a trick or two. By Ed Warwick, Milton Roy Company aterhammer? Pipes jumping around? Pressure relief valves opening unexpectedly? Pressure gauges wearing out in a week? Flow rate is lower than it should be? If you’ve got these symptoms, your pumping system is experiencing the adverse effects of pulsating flow! They are caused by pressure surges in suction and/or discharge pipe systems due to either unexpected high frictional pressure drop or acceleration pressure drop. This article contains information about how to cure those headaches. Before we get into the details, however, let’s lay the groundwork for some common understanding of metering pumps.

W

Metering Pumps What is a metering pump? It’s a pump used to deliver liquids at a predictable, stable flow rate, relatively independent of system pressure changes. Metering pumps are also known as dosing pumps, controlled volume pumps and chemical feed

Figure 1. Elementary liquid end of a metering pump

pumps. Metering pumps are positive displacement—as opposed to kinetic or centrifugal. Some other examples of positive displacement pumps are gear, vane, peristaltic (hose), progressing cavity and reciprocating piston. This discussion relates only to reciprocating piston metering pumps. Metering pumps find use in many industries and applications. Oil field, petroleum refining, textile, public and industrial water treatment, electric power generation and food processing are just a few industries served. Applications include injecting acids and alkalis to control pH, adding corrosion inhibitors to boiler feedwater and downhole oil and gas wells, adding chemicals to purify and clarify water systems and online blending of “chemical recipes.” The applications are widespread and varied. Reciprocating plunger (or piston) metering pumps operate by moving a plunger back and forth in a pump head, and using one-way directional check valves (usually ball-style valves). Diaphragm pumps are simply a variation of plunger pumps. This discussion applies equally to pure plunger (usually called “packed plunger”) and hydraulically actuated diaphragm pumps. Both designs use a reciprocating plunger to generate the pumping action. A mechanism, usually powered by an electric The Pump Handbook Series

motor or air, is attached to the plunger and moves the plunger back and forth (Figure 1). As the plunger moves to the right on a suction stroke, liquid tries to follow it from all directions. Liquid from the discharge pipe, trying to flow in, seats and seals the discharge check valve. No flow can enter to follow the plunger from that source. On the suction side, however, liquid can move up and lift the suction check valve ball, which admits flow from that direction. During the suction stroke, there is flow in the suction pipe and no flow in the discharge pipe. When the plunger reaches the end of its suction stroke, the pump mechanism pushes it to the left on a discharge stroke. As the plunger moves, it displaces liquid that again tries to flow out in all directions. This time, the liquid trying to move back out the suction seats and seals the suction check valve, and it lifts and flows around the discharge check valve ball. So during the discharge stroke, there is flow in the discharge pipe, and no flow in the suction pipe. The instantaneous pipeline flow rate diagram (Figure 2) depicts this alternating, or pulsating flow. Notice that when the plunger moves, it doesn’t just snap forward and back. It accelerates and changes speed (velocity) smoothly throughout the entire distance of each stroke. Most

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results in a momentary reduced pressure at the pump inlet at the beginning of each and every suction stroke. When considering the liquid in the discharge piping, this same acceleration pressure concept applies, except that it results in a momentary pressure increase at the pump outlet at the beginning of every discharge stroke. Frictional and acceleration pressures cause pulsating pressures in the piping attached to reciprocatingplunger metering pumps. The pulsations occur on every stroke of the pump. If the pump is stroking at 120 strokes per minute, pressure in the pipe pulses at 120 times per minute.

How to Predict Pulse Pressure in Pipelines

Figure 2. Instantaneous pipeline flowrate

metering pump stroking mechanisms transmit a nearly sinusoidal velocity to the plunger stroke. This motion is transferred to the liquid in the suction and discharge pipes.

Effects of Pulsating Flow Notice the value of peak flow rate generated during each stroke (in the pipelines) compared to the value of the average flow in the pipelines. The peak flow rate is significantly greater than the average flow rate. For example, if 100 gallons per hour (or 100 liters per hour,) are being delivered by the metering pump (average flow) to the process, the peak flow rate would be 314 gallons per hour. That’s right, peak flow in the suction and discharge pipelines from a plunger type metering pump is π (3.14) multiplied by the average flow rate. This has a serious effect on the friction pressure drop encountered in the pipes. If there is a peak flow rate in the pipeline when only average flow rate was expected (based on centrifugal pump experience), there will be more friction pressure drop than anticipated. Widely accepted equations used to predict frictional pressure drop in pipelines indicate that the pressure drop caused by friction does change proportionally with the flow rate. It fol-

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lows, then, that if the peak flow rate from a metering pump is 3.14 times greater than average flow rate, then the peak pressure drop will be 3.14 times greater than the average pressure drop. This is a significant increase over standard, centrifugal frictional pressure drops and must be accounted for.

Acceleration Pressure Drop Pulsating flow from metering pumps introduces a second type of pressure drop in the system piping— acceleration pressure drop, sometimes referred to as inertia pressure drop. To understand what this drop is, visualize the liquid motion in the suction pipe during the pump discharge stroke. It’s stationary, right? The pump must accelerate (get moving) all the liquid in the suction system at the beginning of the suction stroke. Remember Newton’s first law of motion: “A body at rest will remain at rest until acted upon by an outside force?” The outside force in this case is applied to the liquid by the pump. It accelerates the liquid in the suction pipe from its state of rest to flow into the pump head. This force is applied to the liquid at the beginning of each and every pump suction stroke. When applied to the suction liquid, the force The Pump Handbook Series

Here are two equations that can be used to predict pulsing pressures. Remember that the equations are valid for both the suction and discharge sides of the piping system. The only difference is that the calculated pressure in the suction pipe subtracts from the suction pressure at the pump; the calculated pressure in the discharge pipe adds to the discharge pressure at the pump. Pipeline Pulse Friction Pressure .0000143QµLp 4 dp ∆p=friction pressure, psi Q =pump flow rate, gallons per hour µ =liquid viscosity, centipoise Lp=pipeline equivalent length, feet dp=pipeline inside diameter, inches ∆p =

Use suction pipe length and diameter for friction pressure in suction pipe. Use discharge pipe length and diameter for friction pressure in discharge pipe. Pipeline Pulse Acceleration Pressure ∆p =

NQ (SG)Lp 27,700 d2p

∆p =acceleration pressure, psi N =pump strokes per minute, SPM Q =pump flow rate, gallons per hour SG=liquid specific gravity, no units Lp =actual physical pipe length, feet dp =pipeline inside diameter, inches Use suction pipe length and diameter for acceleration pressure in suction pipe. Use discharge pipe length and

Vent to Headspace of tank

Top of Standpipe higher than highest supply tank level

Chemical Supply Tank

Minimum Diameter 1.5X Pump Piston Diameter Standpipe

Full size Tee

Metering Pump

Length less than 25 pipe diameters

Figure 3. Standpipe configuration

diameter for acceleration pressure in discharge pipe. To illustrate a typical suction side problem, let’s check frictional pressure drop from a metering pump delivering 40 gph (2⁄3 gpm) of Caustic (NaOH), with a viscosity of 50 centipoise. The suction pipe is 3⁄8” schedule 80 (.42” inside diameter) with pipe, valves and fittings adding up to 15’ of equivalent pipe length. If the conventional centrifugal pressure drop equation were used to predict the pressure, a drop of only 4.4 psi would be predicted, and a successful installation would be expected. The equation for metering pumps, however, predicts a suction pipe pressure drop of 13.8 psi. This is enough of a pressure loss to create inadequate NPSH and disable most metering pumps for this service. Pumping acids, in particular, often causes unexpectedly high acceleration pressure because of the chemicals’ high specific gravity. This can cause problems in both the suction and discharge systems. One example of this occurred at a petroleum refinery. The installation was complete, but the pump produced less than 1⁄4 of its rated flow output at start-up. The problem turned out to be overpressure pulses in the discharge line that caused the pipeline pressure relief valve to open and divert flow back to the supply tank. The discharge piping was ordinary for a centrifugal pump system, but created adverse conditions for pulsating flow. The pump was piped to inject sulfuric acid (SG = 1.85) into a 200 psi process, and the relief valve was set to relieve at 250 psig. Pump flow rate was 500 gph, the pipe was 2” schedule 80 with 300 feet from pump to injection point.

The pump’s stroking speed was 142 strokes per minute. The acceleration loss equation yields a pulse pressure of 378 psi. Because this is in the discharge line, the pulse pressure increases the pressure at the pump (and relief valve) to 578 psig (200 psig system pressure plus 378 psi acceleration pulse pressure). With the relief valve set to 250 psig, but exposed to 578 psig, it did its job and opened to bypass excess flow, resulting in low flow to the process.

How to Reduce Flow Pulsations and Their Effects Reducing pressure drop in the discharge piping is usually a fairly straightforward effort. Look back at the pressure drop equations cited earlier. They show each of the elements of the system that can affect pressure. For both friction and acceleration, the equations show that a great reduction in pressure results from increasing pipeline diameter. Increased cost is a factor with this solution. To a lesser extent, but still frequently very effective, shortening the pipeline reduces the pulse pressure. Although pump stroking speed plays no role in generating frictional pressure, stroke speed plays a direct role in acceleration pressure. In the design and specification stages of a project, selecting a pump that strokes slower for the required flow rate will yield a lower acceleration pressure. Probably the most convenient and cost-effective way to reduce flow pulsations (and therefore pressure pulsations) in the discharge pipeline is to install a properly sized pulsation dampener. For most chemical and water services, dampeners that use a flexible diaphragm or bladder to separate the chemical from the compressed gas pad work very well. A pulsation dampener should be installed at a tee intersection in the pump discharge pipe-line, reasonably close (within approximately 40 pipe diameters) to the pump connection. The pipe between the tee and the dampener should be straight, the size of the dampener connection or larger, and no longer than 15 pipe diameters. As a rule of thumb, the dampener should be sized to be a minimum of 15 pump stroke volumes. Of course, The Pump Handbook Series

it must be made of corrosion resistant materials and be rated for pressure and temperature service consistent with process conditions. The dampener will eliminate enough of the flow pulsations to reduce acceleration pressure to practically zero. It will reduce frictional pressure drop to the same as a steady, nonpulsing flow. Use conventional steady-state flow equations to calculate the dampened flow frictional pressure drop. Using a pulsation dampener in the discharge piping is the most frequent method of reducing frictional and acceleration pressure pulses in the discharge pipe system (A pulsation dampener was the solution to the refinery problem stated earlier.) Manipulating pipe diameter and length or using a standpipe as a dampener are usually the most effective ways to reduce pressure drop. Most frequently, the goal is to obtain good NPSH. Usually the task is to eliminate all but 3 or 4 psi of suction pipe pressure loss. This 3-4 psi is a very small pressure, but is significantly large when compared to the pressure usually available in the suction system. For whatever reasons—laws of physics, operating efficiency of the dampeners, etc.— experience has shown that compressible-gas/membrane dampeners do not provide adequate suction dampening for most metering pump applications. Further complicating the issue is the fact that pressure in the pipeline frequently dips below atmospheric pressure on every stroke. The procedure to precharge and maintain a dampener operating in a vacuum is not practical in most installations. Gas/membrane dampeners are excellent for use in discharge systems, but should be applied in suction systems only after thorough analysis and direct communication with the dampener manufacturer. Sizing suction pipe diameter and locating the supply tank to provide elevation head and minimum suction pipe length are by far the best solutions for achieving minimum pressure drop and adequate NPSH. If flow pulse dampening is needed to control suction pressure loss, consider using a simple standpipe (Figure 3). A successful standpipe installation will eliminate all flow

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pulsations, and therefore pressure pulsations, from a suction line. It will do this with no adjustments or calibration and with practically no maintenance. It almost sounds too good to be true, but it’s not. A standpipe can be used in many situations, but it does have some limitations. It is best applied in a system that provides static “flooded” suction. This means that the lowest liquid level in the supply tank is always higher than the pump inlet level. A standpipe can never be used in a suction lift situation (where tank level is lower than the pump) because it will cause the pump to “suck air.” A standpipe is difficult to apply in frictional pressure drop situations. It is best applied in static flooded suction conditions with chemicals with a viscosity less than 50 centipoise. The following are some useful standpipe design parameters: 1. Diameter should be 1.5 times the pump piston diameter or larger. 2. Top of standpipe must be higher than highest liquid level in the supply tank. 3. Top of standpipe must be capped and vented (usually vented back to the headspace of the supply tank) to a safe location to control possible overflow and spillage of chemicals.

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4. Install the standpipe with a full size tee. If there is a 2” standpipe installed in a 1” suction pipeline, use a 2” tee. 5. Locate the tee no more than 25 suction pipe diameters from the pump inlet connection. A standpipe is a simple, efficient flow pulse dampener that reduces suction piping pulsations to near zero levels while improving NPSH.

Conclusion Most reciprocating plunger type metering pumps produce significant flow rate pulsations in suction and discharge piping systems. Unlike flow from centrifugal pumps, the peak flow, which occurs in the pipeline with every pump stroke, reaches more than three times the average flow in the pipe. This high instantaneous flow causes higher frictional pressure drop than usually expected. The stop-and-start, or pulsating flow, also introduces an additional pressure into the piping system, acceleration pressure. Both of these pressures can be predicted by simple calculation. When they occur in discharge piping, they increase the pressure in the discharge line, and can cause pipeline shaking, pressure gauge bouncing, pump pressure overload and inadvertent pressure safety relief valve openings. When it

The Pump Handbook Series

occurs in suction lines, inlet pressure to the pump is reduced, resulting in pipe shaking and reduced pump flow rate from inadequate NPSH (starved, or restricted inlet). Ways to reduce these undesirable pressure pulsations can be seen in the elements of the equations given. Shorter pipeline length and larger pipeline diameter both reduce pressure pulses. Acceleration pressure can be reduced by selecting a pump to provide the flow rate required, but at slower stroking speed. Accessories can be added to the pipe system to reduce pressure pulsations; membrane-type pulsation dampeners are very effective in the discharge piping, and a standpipe is generally most effective in suction lines. Metering pumps are essential, reliable and effective elements in many liquid flow processes. Unlike centrifugal pumps, they do have pulsating flow characteristics, which must be planned for to achieve a reliable, trouble-free pumping system. " Ed Warwick is a Technical Support Engineer for the Flow Control Department of Milton Roy Company in Ivylan, Pennsylvania. He is a graduate mechanical engineer from the University of Florida and has worked in various design application engineering positions for the company since 1965.

POSITIVE DISPLACEMENT PUMPS HANDBOOK

Installation, Start-up and Operation of a Reciprocating Pump A seemingly simple switch from centrifugal to reciprocating pumps becomes a learning experience involving troubleshooting and problem-solving on these offshore oil platforms. By James W. (Tray) Geeslin, Unocal Alaska Resources nocal Alaska Resources is the primary oil and gas producer and operator in Cook Inlet, Alaska. The company operates 10 of the 15 existing offshore oil and gas production platforms. Most of the facilities were installed in the 1960s. Peak daily production reached 225,000 Barrels of Oil Per Day (BOPD). Today the average daily production is near 30,000 BOPD. The lower rates require lean staffing, creative engineering and cost-effective operations. Withdrawal of fluids from oil and gas producing formations causes the reservoir pressure to drop. This pressure is what drives the oil and gas into the well bore. The loss of pressure causes lower recoveries and premature abandonment. As a way to compensate for fluid withdrawal, water is sometimes

U

pumped into the formations. This technique, called ”water flood,” maintains pressure and ”steers” the formation fluids to the producing wells. The process can extend the life of a field by years and is a vital part of production operations. This article reviews recent experiences concerning the selection, installation, start-up, operation, shutdown and troubleshooting of water flood pumps. Throughout this article several pumps will be discussed; each was purchased from a different manufacturer. The names of the companies will not be revealed. The pump installation is more important than who manufactured the equipment.

Water Flood Water flooding is a method that pumps water back into the reservoir through old production wells or new

Vent Drain

Vaccum Pumps

Vaccum Pumps

Reciprocating PD Injection Pumps

RP Filters

M Injection Charge Pump

RP Charge Pump

Injection Well

M Injection Charge Tank

Raw Water Tank Sand Filters Raw Water Pumps

Surface Reservoir

Cook Inlet Alaska Source Pump

wells drilled primarily for injection. The injected water replaces the produced fluids by replenishing the reservoir and increasing the pressure within the formation. The water flood process can extend the life of a field several years by increasing the recovery of oil (Figure 1). Injection water, drawn from the Cook Inlet, must be cleaned and processed before it can be injected into the reservoir. To prevent damage to the reservoir, such as plugging, all of the particulates larger than five microns in diameter are removed from the water by a filtration process. Additionally, the water is de-oxygenated to prevent scale corrosion damage to the well. The de-oxygenation process also eliminates microorganisms that could cause the formation to ”plug up.” Most of the equipment on the platform included in the water flood operation is used to filter and deoxygenate the water. Until recently, most of the water flood pumping systems consisted of turbine-driven centrifugal pumps. These pumps continue to operate well on the facilities that still inject high volumes of water (30,000 Barrels of Water Per Day (BWPD)) at a relatively low pressure (3,500 psig). Over time, it has become much more difficult to inject water into the Granite Point Field reservoir, compared to the other fields in

Figure 1. Layout of water flood injection system The Pump Handbook Series

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Cook Inlet. Consequently, injection rates are considerably lower and injection pressures are much higher today than they were when the equipment was originally installed. To inject water in Granite Point Field, surface pressures ranging from 5,000 to 6,500 psig must be reached with flow rates ranging from 5,500 to 6,800 BWPD. Two of the three platforms in the Granite Point Field continue to operate with a water flood process, the Granite Point and Anna Platforms. Both platforms used turbine-driven centrifugal pumps for many years, but in the past several years have changed to electric-driven reciprocating pumps. Anna Platform had been injecting close to 4,500 BWPD at 4,500 psig using a gas turbine engine and a centrifugal pump. The gas turbine had approximately 3,830 hp available. However, the application required only 500 hp. The centrifugal pump was pumping at a rate of 14,000 BWPD, but a majority of the water was being recirculated back to the suction. Using the pump in this manner caused premature wear. The operating and maintenance costs of running the turbine and centrifugal pump were compared to operating two electric-driven reciprocating pumps. The analysis illustrated that switching to two electricdriven reciprocating pumps would yield cost savings of more than $150,000 per year. The cost of the new equipment (in addition to the $150,000 savings) was much less than the cost of overhauling and reworking the existing turbine and pump. An additional financial benefit was that we could salvage and sell the old gas turbine package. Based on the economic analysis, we decided to replace the gas turbine and centrifugal pump. Granite Point Platform had a similar water flood operation, but it utilized more injection equipment. The platform was injecting approximately 6,000 BWPD with a surface pressure of 6,500 psig. Two 1,500 hp gas turbine engines were driving two centrifugal pumps in series to inject the water. Again, the available 3,000 turbine hp was much more than the 800 hp required for the electric motors and reciprocating pumps. The cost of operating and maintaining the two

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turbines and centrifugal pumps was compared to that of two electric-driven reciprocating pumps. The analysis illustrated annual savings of more than $400,000 per year. Again, it was decided to swap the turbines and centrifugal pumps for electric motors and reciprocating pumps. Although we had made the decision to change the injection pumps and drives to electric motors and positive displacement reciprocating pumps, the change was not as simple as buying a set and plugging it in. The selection, installation, startup, operation, shutdown and troubleshooting of the positive displacement (PD) pumps has required a considerable amount of time and capital. Research, communication and experience all played a part in our learning process. The result is that over the past three years, Unocal has installed four electric-driven PD pumping systems. The first installation on the Anna Platform failed to work due to mechanical problems and pump material. The pumps were replaced with another manufacturer’s pump of proper material and mechanical design. The new installation is in operation today and running well. The Granite Point platform has two systems that utilize PD injection pumps: water flood and produced water injection. The produced water injection system is the most recent installation.

Selection When Unocal began researching the use of PD pumps, we relied on several vendors to recommend and guarantee a feasible application. We learned the hard way (after numerous installations and one major failure) what to look for in a pumping system. We have found that a PD pumping system is more difficult to design and build than the typical centrifugal setup. The design of a feasible PD pumping system begins with the specification of the pump. There are many crucial decisions and things to know when selecting a pump, including the following: • required flow rate(s) • required discharge & maximum pressure • facility limitations (electrical, instrumentation, utilities) The Pump Handbook Series

Photo 1. Reciprocating water flood pumps at Anna Platform, Cook Inlet, Alaska

Photo 2. Working on the reciprocating water flood pumps at Granite Point Platform

• • • • • •

type of driver type of driveline materials of construction pressure protection vibration/pulsation protection subsystems (lube oil pumps, grease pump, coolers, etc.) • history of similar equipment (reputation) • vendor support (location) When evaluating a pump, the first concern is the sizing (i.e., available flow rates and pressures). Review your facility’s operating history and consult reservoir engineers to obtain the design flow rates and pressures of the injection pumping system, both current and future. At Anna Platform, injection rates of 5,500 BWPD at pressures up to 5,500 psig were required. The Granite Point Platform would need 6,800 BWPD at pressures up to 6,500 psig. Familiarity with facility limitations is key when retrofitting pumping systems within existing operations. For example, the platforms generate power on board. When adding new electrical equipment, the entire electrical load on the platform must be evaluated to ensure that there will be no overload. On both Granite Point Platform and Anna Platform, power generation was in surplus. When we considered adding new electric motors to drive the PD pumps, we

Q = 200 gpm (~6,800 BWPD) Ptd = 6,500 psig ME = 90%

Photo 3. Discharge piping in produced water injection pump. Note that there are no sharp bends from the piping to the well (section goes straight up).

evaluated the electrical surplus to ensure that the additional load would be feasible. The following equations were used to do a preliminary check on the motor size and load on the platforms: bhp= Q x Ptd 1714 x ME Where: bhp = brake horsepower Q = desired flow rate [gpm] Ptd = developed pressure [psi] ME = mechanical efficiency [1] The results for Anna Platform were: Q = 160 gpm (~5,500 BWPD) Ptd = 5,000 psig ME = 90% At Anna Platform, we found that a little more than 500 hp would be required. The platform has a generator capacity of 1600-kw; however, only 800-kw was being used. The 500 hp equates to approximately 375-kw, which would not present a problem when added to the existing power grid. The 500 hp could be split into two pump and motor systems. The split would additionally create a redundancy that would be beneficial in case a pump was down for maintenance. This would also lower the starting load on the power grid to only 250 hp at a time. Although the power grid could handle the additional 500 hp, the grid was not capable of starting a 250 hp motor across the line. To start the motors, soft motor starters (SMCs) would have to be incorporated. An SMC regulates the power to an electric motor during start-up by ramping it up slowly. The motor is allowed to come up to full speed without draining a large surge of power from the generators. The Granite Point Platform had similar results:

Granite Point Platform would need approximately 850 hp (634kw). Granite Point also has a 1600kw power grid, which was carrying an 800-kw load. Similar to Anna Platform, the additional electrical load was handled using SMCs. For both Anna and Granite Point Platforms, we only had two requirements for the motors. They were to be Totally Enclosed Fan Cooled (TEFC) and capable of being placed in a Class I, Division II location. Speed A PD pump rotates much slower than an electrical motor; therefore, a driveline between the pump and motor must be incorporated. On Anna Platform, Unocal specified a floating shaft coupling because of the limited installation space available for pump skids. For the Granite Point Platform water flood pumps, we allowed the pump manufacturer to specify a V-belt driveline. Based on our experience, either driveline would have been acceptable. However, the floating shaft coupling was found to be easier to handle during alignment. Materials Our injection water source is the Cook Inlet; therefore, material resistant to corrosion must be used to protect the pump from the salt water. In the treatment process we use a variety of chemicals to prepare the water for injection. Each chemical is designed for a specific purpose. Not all are consistently required, but all are used at one time or another. Because all of the chemicals could be used at any time, their reaction with various materials must be analyzed. The chemicals involved in our processes include corrosion inhibitors, oxygen scavengers, flocculent and biocide. We relied on the pump manufacturer to specify a material that would stand up to these. The vendor was provided an analysis of the water and a list of the chemicals we proposed to use. With this analysis, the vendor was asked to select a material that would hold up to the water and chemicals in any combination. When the first set of pumps we installed on Anna Platform failed due to corrosion, Unocal took the The Pump Handbook Series

Photo 4. Installation phase of water flood injection pumps on Granite Point Platform

initiative and specified the material for the replacement pumps. After analyzing the service conditions, we selected stainless steel and specified it as the material to be used in all future PD pump applications. We have also learned that stainless steel works well in a continuous operation, but for an intermittent operation, Duplex Stainless Steel is better. The most recent pump installed on the produced water system at the Granite Point Platform was specified and designed with Duplex Stainless Steel fluid ends because it operates intermittently. Pressure To protect the pump and facility piping from over-pressure damage, steps must be taken to control the pressure. Unocal uses a pressure control valve (PCV) linked to the platform computer (PLC) and a mechanical relief valve or rupture disk to protect from potential overpressure in the piping system. The PCV can be adjusted to keep the discharge pressure constant at any specified value. The mechanical relief valve cannot be adjusted and is set to relieve at the maximum system design pressure. Pressure control is achieved by routing the excess water out of the injection system through the PCV. Vibration A reciprocating PD pump can cause two types of vibration: mechanical and hydraulic. Both of these can create serious problems on an offshore platform. The PD pump generates mechanical vibration through the physical piping or baseplate. Hydraulic vibration and pulsation occurs through the pumped fluid. Both vibrations can be disruptive to personnel and damaging to the piping and structure. In some instances, resonance can occur. The vibration match between two pieces

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Photo 5. Water flood injection pump installed on Granite Point Platform (99 gpm, 6,500 psig discharge pump with 500 hp electric motor)

of equipment causes this phenomenon. Resonance can cause major damage to the facility structure and/or surrounding equipment. To avoid these problems, we specified that each pump skid be delivered with vibration isolation mounts. In addition, the piping to and from the skid is made with flexible hoses. Pulsation dampeners were supplied for both the suction and discharge sides of the pump to minimize the hydraulic vibration (pulsation). With each new pump installation a variety of sub-systems were tied into the main pumps. For example, the pumps on Anna Platform require a supply of cold water to circulate through the lube oil pan; this keeps the pumps from overheating. The pumps installed in the water flood system on Granite Point Platform require a forced lube oil circulation pump and a grease pump for the plunger packing. The lube oil and grease pumps are electric-driven; therefore, another power supply was added to the pump skid. When we chose a pump supplier, we verified that all of the sub-systems were understood. This helped us ensure that our vendors were making the appropriate decisions to accommodate each new pump installation. Upon receiving quotes for each pump installation, careful attention was given to any history of operating the same pumps using the same design we had developed. In the request for proposal, the vendor was required to supply a list of customers who were using the same pumps in a similar application. Interviewing the contacts on the list gave us a greater insight into the maintenance, operation, troubleshooting and economics associated with each pump. Another important factor in the

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pump selection was vendor support for the purchased products. Alaska’s location makes part replacement and vendor accessibility difficult when compared to the “lower 48.” We have purchased pumps previously from vendors located in California, Ohio and here in Alaska. Based on the company’s available inventory, spare or replacement parts can be received within days. However, if the parent company is located overseas, difficulty can arise when broken parts need to be analyzed to determine the mode of failure. The delivery of results and solutions can be delayed by several weeks. Such logistics can considerably hinder the repair of a pump. When selecting equipment, we look for companies that can supply parts quickly and have a representative on site promptly when problems arise.

Installation Installing new equipment on platforms is not a simple task. Each facility has very limited available space; equipment installation can take weeks to complete. Designing the piping, electrical connections, instrumentation and mechanical layout must be both functionally and ergonomically reasonable. The layout of the pumping systems on both the Anna and Granite Point Platforms was designed to enable smooth operations for the pumps and accessible working conditions for platform personnel (Photo 3). The first task in platform pump installation is choosing a physical location for the pump. We wanted the pumps to be near the filtration and de-oxygenation equipment to minimize piping runs. The pumps on Anna Platform were not installed in the location of the previous turbine and centrifugal pump. Instead, the equipment was installed in a smaller area near the existing water flood system. The area where the turbine and pump were located was then converted to a workshop and storage area. The pumps on Granite Point Platform were installed in the area vacated by the two turbines and centrifugal pumps. This kept open the possibility of adding another pump, if needed (Photo 4). Another important consideration was the pumps’ required suction. Charging a PD pump is critical to its operation. The flow to the pump The Pump Handbook Series

Photo 6. Two vertical 80 gpm, 5,000 psig discharge water flood injection pumps on Anna Platform

needs to be adequate to eliminate the possibility of cavitation. To prevent cavitation, the pump vendor was asked to provide minimum and maximum suction pressures. The first installation on Anna Platform was designed with a flooded suction, per the vendor’s instructions. Within a short time of operation, the pumps, suction valves and springs were breaking, resulting in long periods of downtime. The problems were caused by cavitation. The solution was to charge the suction side of the pumps with 40 psig of pressure. Based on experience, our charge pumps are now designed at a suction pressure close to the maximum. The installation currently on Anna Platform operates with a suction pressure of close to 90 psig, while the water flood pumps on Granite Point Platform operate with a suction pressure of 130 psig. Even with adequate suction pressure to the pump, cavitation can still occur if pressure and flow are not steady. The rapid movement of the plungers in and out of the fluid end creates pulsations that can temporarily stop the flow. Cavitation can result if there is an insufficient section of straight pipe to and from the pump suction and discharge. There is not enough room on the platform to allow for long runs of pipe to and from each pump, so we have used hoses with large sweeping turns. In addition, the pulsation dampeners that were specified for the discharge and suction of each pump were installed as close to the pump as possible and pre-charged with ample pressure to suppress the pulsation. The dampeners store a small volume of the process fluid during the plunger’s stroke out and redeliver the fluid on the inward stroke. Electrical supply to the pumps

Start-up

Photo 7. Creative solution for suction and discharge piping on Anna Platform’s injection pumps developed by Unocal’s Bob Baker

becomes difficult when dealing with large horsepower motors. The SMCs had to be within 50 feet of the pump motors because of the wire size. We were able to install each pump and motor on both facilities close to the electrical motor control center (MCC) room, which in both cases was under the required 50-foot limit. Timing the electrical hook-ups and the piping installation proved to be difficult. The limited workspace on the platforms could not accommodate simultaneous piping and electrical work crews. The operation and protection of our PD pumps rely heavily on instrumentation and the platform computer. When installing the pumps, we added alarms, meters, transmitters, indicators, switches and control valves so that the units could be monitored and controlled. An annunciator system on the platform alerts personnel when there are problems. Pump alarms include systems that alert us to low suction pressure, high discharge pressure, low discharge pressure, high vibration, motor overload, low lube oil pressure and high oil temperature. Every alarm has the ability to shut off power to the pump motors. Power shutoff reduces the risk of equipment damage. The flow through a PD pump is constant; however, the discharge is split to several wells. Meters that indicate where the water was injected were installed on each of the wells to monitor the reservoir. In addition, a meter on the suction side of the pump was installed to indicate the flow rates of the injected and overboard water. If too much water is being overboarded, one of the pumps will shut off to save the related operating costs. Local indicators throughout the pumping system make it possible for operators to monitor performance and locate potential problems.

When the PD pumps were installed we took considerable caution starting the equipment by working through a series of procedures that ”double-checked” the installation. In addition to our pre-start-up procedures, we required representatives of the pump vendors to visit the platform and evaluate the installation for errors. Upon acceptance from the pump vendor representative, the start-up procedures were completed (Photos 5 and 6). The first step was to ensure that all of the piping was flushed to prevent welding slag, rags, or other materials that could damage the pump, from entering. In addition, fittings, equipment and the rotation of the motors were checked for proper installation and make-up. The instrumentation was also tested to ensure proper function. We did this by setting off the alarms. We then determined if the platform computer took the correct action. After the equipment was checked out, the pumps were started. The pumps on both the Anna and Granite Point Platforms are controlled by PLCs. The operators initiated the start-up sequence by switching the control panel to the auto position. For safety reasons, the pumps cannot run in the manual position. The PLC verifies several parameters before the pumps start: • adequate suction pressure exists • lube oil pressure (if forced) is present • low discharge pressure is seen; this keeps the pump from starting against a load • no ESD triggered • the other pump is in the unloaded condition The PLC also makes sure that both generators are running; this confirms that there is enough power to run the pumps. Running the pumps without enough available power results in the generators shutting down, which in turn causes the platform to shut down. Once the above procedures have been satisfied, the PLC permits the SMC to begin starting the pump motor. While the motor is being started, the PLC evaluates the drawn current. If the value exceeds a specified limit, the breaker will drop and the power flow to the The Pump Handbook Series

motor will stop. The PLC also monitors the time required for the motor to ramp up to full speed; it stops it if it takes too long.

Operation Once the PD pumps are started, we begin injecting water into the reservoir. During the start-up sequence, all flow was directed overboard to eliminate pressure buildup. To begin water flooding, the overboard valves are closed and flow is directed to the injection wells. The continuous monitoring of the pumps and systems within the water flood process remains an important priority during the operation of the pumps. Our most frequent problem is cavitation. To protect the pump from cavitation, the PLC verifies that there is always an adequate amount of pressure and flow delivered. If the pressure drops below a specified set point, a “low suction pressure” alarm sounds. If the pressure drops further, the PLC stops the pump. Another indicator of cavitation on a PD pump is vibration. To detect high vibration, we have installed instrumentation to pick up the vibration levels of the pump and alert the PLC when the levels become excessive. If they do, the PLC shuts down the pumps. The discharge pressure of the pump is always monitored during operation. If for some reason a valve has been closed, the pressure in the pump can exceed the maximum design pressure. To protect the system, the instrumentation and platform PLC will shut the pump down before it builds up enough pressure to damage itself. Although the platform computer monitors the operation of the pumps, there are times when we cannot rely on the computer. We rely heavily on the personnel that live and work on the platforms. They watch the oil levels and pressures, and they visually inspect the pumps for anything that seems out of the ordinary. The mechanics do preventive and predictive maintenance to ensure that the pumps will run with a very low percentage of downtime. The operators complete hourly equipment checks, which enables them to develop operating parameter trends. If the parameters begin to vary from the established

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Photo 8. Packing grease mess caused by excessive flow rate

trend, they will shut the pumps down and begin troubleshooting.

Shutdown The PLC instrumentation, operators and mechanics can all shut down the pumps when operational problems are detected. All of the equipment installed on the platform has been designed and set up so that shutdowns occur safely. In the case of an emergency shutdown (danger, fire, etc.), electrical power to the entire pumping system must be shut off and the valves moved to their fail safe position. When one of the pumps is shut down under non-emergency conditions, it is done in a “controlled” manner. The PCV on the discharge side of the pump is opened to relieve the back pressure overboard, and the power to the pump motor is stopped. Once the pump has stopped, all of the sub-systems to the skid are shut off. In addition, the charge pump supplying water to the pumps is stopped. The remainder of the water flood equipment used in the filtration and de-oxygenation process continues operating.

Troubleshooting We encountered several very challenging troubleshooting problems with these pumping systems. Problems with improper material selection, cavitation, valve design, spring design, vibration isolation and lubrication all reared their heads. Each of the problems took time to decipher, but once the solution was executed, spotting problems in future pump installations became easier. Material Failure As previously stated, the first PD pump installation at Anna Platform experienced material failure. This surprised us because the pumps had a good reputation for operating with

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several fluids. The vendor, however, did not have much experience pumping salt water, and specified materials of construction that consisted only of carbon steel and an additional protective layer of nickel. Within a very short time, the nickel plating wore or flaked off, and the pumps began to corrode. As the corrosion worsened, small plugs on the head of the pump discharge began to blow out. We opted to return the pumps to the manufacturer and install different models. Cavitation This common problem has been the largest and most difficult one to detect and solve. Usually when cavitation was occurring or had already occurred, some of the components in the pump were destroyed. The primary indicator is high vibration or ”banging” noises coming from the pump. The suction side of the pump is usually damaged most severely when cavitation occurs. We have observed broken suction valves, valve seats peened beyond repair and broken springs wedged into the fluid end. Some of the items we check when cavitation is suspect are: • adequate suction pressure • piping to the pump is correct (Photo 7) • pulsation dampeners are charged properly • no air entrainment in the suction fluid • whether or not the problems seem to stem from bad design Cavitation became a problem with the first installation on Anna Platform. The pumps were installed based on the vendor’s recommendations, which included a flooded suction. Several replacement parts were purchased before we determined that the damage was from cavitation. To alleviate the problem, we added a charge pump to increase the suction pressure to the pump. The problems on the Granite Point Platform proved to be more difficult to troubleshoot. The pumps could not operate longer than two weeks before a breakdown. We asked the vendors to visit the platform and evaluate the problem more extensively. The piping to the pumps had a very sharp turn on the inlet, which we were instructed to remove. The tight corners were eliminated and The Pump Handbook Series

replaced with long sweeping turns using hoses. A technician from the pump manufacturer used an oscilloscope/pressure transducer connected to the suction of the pump to read the pressure pulsations; we could clearly see the pressure drop below the cavitation point. We were instructed to increase the size of our charge pump so that the pump could deliver twice the needed flow rate at a specified pressure. We discussed, in detail, why the vertical turbine pump used for charge pressure would not deliver twice the flow rate at a certain pressure. If we followed the recommendations, they said, the flow rate of the charge pump would be equal to that of the PD pump and the pressure would be on the curve of the charge pump. This resulted in our installing a larger pump that delivered more pressure at the flow rate of the PD pumps. Even after the larger charge pump was installed, the main pump still cavitated, as indicated by the oscilloscope. Rather than depending on the pump vendor any further, we researched every component on the pump skid ourselves. Our investigation led to the pulsation dampener on the suction side of the pump. When it was first installed, we were instructed by the pump vendor to charge the pulsation dampener with 30% of the pressure supplied to the pump. At one point during the troubleshooting phase, we were instructed to lower the charge pressure to 15% of the suction pressure. After talking to the manufacturer we found that the charge pressure needed to be 60% to 70% of the suction pressure. As we raised the charge pressure on the pulsation dampener, the oscilloscope showed the elimination of cavitation. Another factor related to cavitation was air entrainment in the suction piping. Because the pulsation of the fluid was caused by the pump operation, we were skeptical that cavitation was occurring as a result of air being introduced into the water. We had several valves that we used to drain the lines, and we had installed filters to fill with water during a start-up or shutdown. The valves are equipped with internal balls that float when filled with water. When the ball floats, it blocks the discharge of the valve. We assumed that the pulsations in the

water were causing the balls to temporarily unseat, making it possible for air to come backwards into the valve. To alleviate the problem we installed manual block valves upstream, which were closed after the pumps were started. Under the assumption that our cavitation problem was solved, we ran the system for several more weeks before the pumps came apart again. All of the broken pieces indicated cavitation. The valves were damaged, the springs were broken into several pieces, and a “banging” noise was heard before the pumps shut down. Surprised that the pump had cavitated again, we carefully inspected the broken pieces and investigated for another source of the problem. We found that the springs were becoming so compressed during the stroke of the valve that the coils flattened and broke. Once the spring broke, the valves did not have a guide. Without a guide, they began to rattle and hit the side of the seats until they broke. Because of our complaints, a pump representative investigated the situation and concluded the same thing we did—the pump failed due to mechanical design. Within days, we installed a new valve and spring design. We have been running the pumps for nearly nine months now without experiencing any cavitation damage. Vibration Vibration of PD pumps can be

very high at times. We have had several instrument tubing lines break and the isolation pads on two of the units have come loose from the floor. We switched the tubing from stainless steel to flexible plastic. The vibration isolation pads used on the pumps at Granite Point were attached to the skid by a bolt screwed into the floor. The vibration of the pumps caused the bolts to shear and the pumps moved about the floor. To solve this problem, we welded brackets to the floor on each side of the pump, holding it in place. The brackets have Teflon® on the inside, which enables the pump skid to move on the brackets without the two metal surfaces rubbing together. Lubrication The latest problem being investigated is plunger lubrication. Grease is dropped onto the plungers to lubricate them as they stroke in and out of the packing. We have been experiencing very short run times on our packing, sometimes as short as two weeks. We were burning out the grease pumps because the unit was unable to move the grease (Photo 8). Recently we switched to different grease and have not experienced any more problems with the grease pump. We have noticed, however, that soon after the grease change the plungers became scored and we had used a great amount of packing. To correct the problem, we decided to switch to a different grease pump

The Pump Handbook Series

manufacturer and return to the original grease. This is still in the works, so we do not know the results.

Conclusion Changing the types of pumps used for water flood within Cook Inlet has been a huge learning process. With the time and capital spent in the selection, installation, start-up, operation, shutdown and troubleshooting of the positive displacement pumps, Unocal has become very experienced. We have proven that the use of electric-driven reciprocating pumps is both feasible and economical in our water flood systems. Although we have been operating the pumps for several years, we expect that there is still much to learn. " James W. (Tray) Geeslin is Production Engineer for Unocal Alaska Resources in Anchorage, where he is assigned to the company’s Cook Inlet Oil and Gas Operations Team. His responsibilities include coordinating and implementing the Facility Engineering for three off shore platforms and one onshore processing site. Mr. Geeslin has a B.S. in Mechanical Engineering from Florida State University and is an associate member of the Anchorage Section of the American Society of Mechanical Engineers.

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POSITIVE DISPLACEMENT PUMPS HANDBOOK

Extending the Life of Positive Displacement Pumps Part 1: Gear Pumps By Dr. Lev Nelik, P.E., APICS Cert., Liquiflo Equipment Co. P -patience U -understanding M-maintenance P -prevention = PUMP se these and you will solve almost any difficult pump troubleshooting job. Skip any, and your headaches will never go away. In the pump business, there are no quick fixes. Those folks who have actually worked on pumps with their own hands, I’m sure, would agree. Good field service people rarely start at the pump itself when beginning a troubleshooting job. Much to the irritation of the plant management, who want it solved “right away,” a troubleshooter looks at the piping, supply tanks, listens to a noise from a neighboring valve, and asks a bunch of questions. On the surface, these actions have little or nothing to do with the problem. Just about when their escorts are convinced they are dealing with a “flunky,” the troubleshooter will get to the pump.

U

Patience Get the overall picture. Start from the far end, and zero in to the pump from all ends—suction, discharge, bypass lines and other auxiliary connections. Sketch it up. Sit down. Have a cup of coffee. Ask questions. You have gathered the facts. It looks messy, but it’s a start.

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Understanding Now thinking time starts. Do some calculations (as we will touch on later), see if anything obvious sticks out. By maintenance we imply proper maintenance, with records of wear parts replacement, and not just fixing it when it’s broken. “If it ain’t broke, don’t fix it” only applies to cases where the actual failure is not critical. For example, if the “ain’t broke” philosophy was applied to boiler feed pumps, we would all be sitting in the dark most of the time. Prevention is a culture. If maintenance can be thought of as activities directed to a pump itself, prevention covers the complete system that can affect pump operation. Prevention is a key to reliability and a low maintenance budget.

Applying the Formula to Gear Pumps Many aspects of gear pump troubles are common to other types of pumps as well, including centrifugal, screw, vane, etc. But some problems are unique to gear pumps. Below are some of the more popular issues. By understanding and addressing these, most gear pump reliability problems can be avoided or minimized. Cavitation In contrast to centrifugal pumps, cavitation is not nearly as “hot” an The Pump Handbook Series

issue for gear pump users. However, it is still serious enough not to be disregarded. Cavitation (liquid vaporization when pressure falls below the vapor pressure) can take place (usually in the suction region), regardless of the pump type. It starts from the incipient stage, when just a few bubbles begin to form. As suction pressure drops further, the vaporization becomes more vigorous, resulting in vapor-locking of the suction region. When this happens, centrifugal pumps begin to lose pressure (head drops); gear pumps lose flow. Pulsations and vibrations accompany this process, leading to pump damage—seals running dry, bearings failing due to shocks, loss of material (eroding of the impeller blades or gears) caused by the bubbles collapsing (implosions) against the internals. The intensity of these collapses depends on the liquid properties, operating speed, materials of construction and other factors. High-energy boiler feed centrifugal pumps, for example, are known to suffer particularly from cavitation. This is because cold water has high internal energy of vaporization. (The energy of bubble collapse is very high.) Double suction cooling water pumps are also plagued by cavitation headaches. Traditionally, gear pumps have been applied for much more viscous fluids, such as oils. Oil obeys physics just as diligently as

Figure 1a. Suction pressure vs. flow

Figure 1b. Gear pump cavitation

water, but the energy of collapsing bubbles is much less then that of cold water. The damage, therefore, is also much less. There is another type of cavitation, which is unique to gear pumps, as well as for many other positive displacement pump types, but not characteristic to the centrifugals. That is the ability of fluid to fill the gear teeth cavities, in order to be transferred (i.e. displaced) from suction to discharge. A gear pump will transfer the fluid only if the fluid gets to the pump, i.e. the pump does not “suck” the fluid into it’s suction port - the suction pressure does. If the suction pressure is too low, the gear pump will tend to pump out more fluid than it receives at the suction port (Figure 1).

Figure 2. Gears loaded by ∆p

To help the situation, you need to either increase the supply pressure or eliminate/reduce the reasons for its drop. You can increase the pipe size, minimize the number of restrictions (bends, orifices, valves) or reduce the viscosity of the pumped fluid by preheating it. All these measures reduce hydraulic losses in the suction line, causing the available suction pressure at the inlet flange of the pump to increase. People often assume that the pump inlet pressure is approximately equal to the supply tank pressure. This assumption could result in big mistakes. Don’t guess. Install a gauge as close to the pump inlet as possible and read it. You may need an absolute gauge, because a partial vacuum (below atmospheric) could be the case, and needs to be measured. If you did all you could to increase the available inlet pressure, the next step is to make sure that the pump’s required inlet pressure is as low as possible. For a gear pump, the minimum required suction pressure is approximately proportional to speed: A 1000 rpm running gear pump would require an inlet pressure twice that of a pump running at 500 rpm. This is another reason why we hear much less about cavitation in gear pumps—gear pumps run much more slowly, although there are exceptions to the rule. As recently as 10-15 years ago it was popular to select smaller pumps that ran faster to save on initial cost. However, as many maintenance people have realized since, with speed comes trouble. Today, a smaller 3600 rpm pump is no longer considered a wonderful idea. A larger 1800 rpm pump is considered a wiser selection

The Pump Handbook Series

from the reliability standpoint. Besides, a larger, slow-running pump will require less suction pressure, and have a better chance of avoiding cavitation. The bottom line is that available suction pressure at the inlet flange must be greater than that required by the pump. It is your responsibility to know (not guess) what is available See the suction gauge! It’s the pump manufacturer’s job to let you know what the required pressure is for the given viscosity and rpm. Discharge Pressure It is a good idea to have the discharge gauge as close to the pump as possible. You can then calculate the differential pressure using the following equation: ∆p = Pd - Ps Differential pressure should not exceed the pump’s rated differential pressure. But do not rely solely on the pump brochures. Manufacturer’s brochures often state the “envelope” parameters for pressure, flow, viscosity, etc. A brochure may claim that a particular pump model is a “400 psi” design, period. However, this may be true when moving fluids with a viscosity higher than 1000 cSt. Below that, the pressure is derated to 300 psi, 200 psi or even less. The two main reasons for differential pressure limitation of gear pumps are the capabilities of the product-lubricated bearings, and shaft deflection (Figure 2). The higher the differential pressure, the greater the shaft deflection. If deflection exceeds the available radial clearance, the gears will begin rubbing against the case. Many gear pumps for oil transfer are made from cast iron, which has a reasonable ability to withstand such rubbing and only a gradual wear will occur. The pump may not necessarily fail catastrophically, but it will “speak for itself” with noise and vibrations. However, for stainless steel construction (i.e. transfer of chemicals), this situation is definitely not acceptable: stainless steel tends to gall (transfer of metal from one rubbing piece onto another). Martensitic stainless steel resists galling somewhat, but austenitics (such as 316) can gall quickly, causing the pump to seize.

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Figure 3. Gear loaded, bearings P,V

Bearings and Lubrication Most gear pumps have productlubricated bearings. Depending on viscosity, rpm and clearances, these bearings operate either hydrodynamically or hydrostatically. (A mixed lubrication regime falls between these categories). Hydrodynamic lubrication means that there is no contact between the journal and the bearing ID—the fluid film is “strong” enough to support the rotor. The higher the viscosity, the further into the hydrodynamics area the pump tends to operate, i.e., the better the fluid film is supporting the shaft. In theory, the material from which the bearings are made is irrelevant in this case because the fluid film prevents the contact of parts. If viscosity is low, or the fluid film is thin or absent, it cannot support the rotor. Differential pressure would then succeed in displacing the rotor from the discharge side (higher pressure) to the suction side, and the parts would rub. In this case, bearing material is important. The ability of different materials to resist rubbing wear is characterized by the PV-value (psi x ft./min.), where P is the pro-

jected unit loading on bearings (not the same thing as pump pressure) and V is the circumferential velocity of the rotating journal. The operating PV-value must be compared with the allowables. As a rule of thumb, PV=60,000 for bronze, 30,000 for iron and 120,000 for carbon. Example: determine if the bearing material is acceptable for a gear pump running at 1800 rpm, with 6 psia suction pressure, 128 psig discharge. The pump rotor geometry is shown in Figure 3. The differential pressure pushing each gear is: ∆ p = (128+14.7) - 6 = 137 psi (Note: pay attention to consistency of units, do not mix psig with psia, but convert.)

The force on the gear is: 137 x (2 x 1.2) = 329 lb. Half of this force is transmitted at each bearing, i.e., 165 lb., with bearing unit loading of: 165 =110 psi (1.5 x 1) The journal velocity is: (rpm x d) =1800 x 1= 427 ft./min. 3.81 3.81 The PV value is then: 110 x 427 = 51,920 psi x ft./m. Therefore, bronze bearings (bushings) would be acceptable in this case, but the iron bushings’ PV-limit would be exceeded. Abrasives Gear pumps can tolerate very moderate amount of abrasives, usually in small concentrations and with small particle sizes (preferably

Figure 4. Casing life extended by 180° turnaround

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The Pump Handbook Series

under 5 microns). Since slip is sensitive to clearances approximately in proportion to cube of clearance, it does not take much to wear out clearances to the point that a significant amount of flow is lost (slipped). Also, product-lubricated bearings become less effective when they wear out, resulting in reduced rotor support, rotor deflection onto the casing and rubbing. For these reasons, it is important to filter out the abrasives before they enter the pump suction by using filters, centrifuges and settling tanks. Corrosion Carbon bearings provide better corrosion resistance than bronze; iron, obviously, is the worst. This is exactly opposite to the selection of bearings from the standpoint of abrasion resistance. Where corrosion is concerned, bronze turns out to be in the middle ground. This is why, historically, bronze is the most widely used bearing material in gear pumps. It’s a great compromise from abrasion, corrosion and lubrication standpoints. For corrosive applications, the rest of the pump is usually made from the austenitic or martensitic stainless steels, with care taken to size all clearances properly (wider) to prevent galling at low viscosity liquids.

Seals Failure of mechanical seals is a leading cause of pump failures. There are many different reasons for this. The cause can be a problem pump, a bad system, or both. For non-critical applications, packing is still a good choice. Technology has made good progress for both mechanical seals and packings. Packing requires lubrication in order to operate without burning, so a certain amount of leakage is required by necessity. Newer designs are capable of operating with very small leakage, but still require more than mechanical seals. Keep in mind, though, that mechanical seals also require some leakage to lubricate their seal faces. However, this amount is negligibly small, and the leakage essentially comes out in a vapor form and in low concentrations. Mag drive gear pump designs are very common nowadays, and are applied, just as their centrifugal cousins, for applications requiring no

leakage at all. Similar to other pump types, seals leak in cases of excessive piping loads, and proper alignment is required. Because of the slower speeds and more robust shafts, piping problems are not as prevalent for gear pumps. This is a bigger issue for centrifugal units. However, especially for larger sizes and horsepowers, gear pumps can be just as sensitive to piping misalignment.

Case Turnaround The radial gap between the casing and the gears is not uniform, due to the differential pressure. The gap is tighter at the suction side and wider at the discharge. The higher the differential pressure, the more the gears are deflected onto the case suction side, and the tighter the clearance (gap) becomes in that area. The casing is, therefore, wearing more on the suction side. Ironically, it is the differential pressure that helps keep the tight clearance at the suction and prevents the slip. This pressure, however, does

not always stay constant. For many reasons, it changes over the life of a pump. But even at constant pressure, the wear progresses with time, and the gap opens up, increasing the slip. If a worn casing is turned around 180°, so that the suction port becomes a discharge port, the radial clearance is restored at the “important” side, from the slip point of view. The tighter clearance (on the suction side) will now seal the slip from discharge to suction, similar to an “as-new” situation, and, at the minimum, this will buy time for the plant to purchase and install a replacement pump (Figure 4). A note of caution: If wear persists and continues to cause short equipment life, take a closer look at the application: Is the pressure too high? Is pumpage more abrasive than expected? Is it less viscous than you thought?

Documentation

gram the system. Have suction and discharge gauges installed and monitor them. Is the suction gauge steady or fluctuating? If it fluctuates, figure out why. Don’t just evaluate the magnitude of a parameter (e.g. vibration amplitude), but also its trend over time, and take action if the situation begins to deteriorate. It is much less expensive to set up a preventative program than it is to react to sudden breakdowns. At that point, the failure analysis may be difficult and masked by the severity of the damage. Use common sense, and, of course - may the “force be with you”! " Dr. Nelik is the Regional Sales Manager for Liquiflo Equipment Company. He is a full member of ASME, a P.E. and a certified APICS (CIRM). He received his Ph.D. in Mechanical Engineering from Lehigh University. Dr. Nelik can be contacted at: [email protected]

When troubleshooting a gear pump (or any pump), take time to dia-

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Extending the Life of Positive Displacement Pumps Part 2: Multiple Screw Pumps By James R. Brennan, Imo Pump

Three Screw Pumps Three screw pumps are the largest class of multiple screw pumps in service today. They are commonly used for machinery lubrication, hydraulic elevators, fuel oil transport and burner service, powering hydraulic machinery and in refinery processes for high temperature viscous products such as asphalt, vacuum tower bottoms and residual fuel oils. Three screw pumps also find extensive use in crude oil pipeline service as well as gathering, boosting and loading of barges and ships. They are found in the engine rooms of most of the world’s commercial marine vessels and many combat ships. Subject to material selection limitations, three screw pumps are also used for polymer pumping in the manufacture of synthetic fibers such as nylon and lycra. Three screw pumps will generally

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have only one mechanical shaft seal and one or two bearings that locate the shaft axially (Figure 1). Internal hydraulic balance is such that axial and radial hydraulic forces are

opposed and cancel each other out. Bearing loads are thus very low. Another common characteristic of three screw pumps is that all but the smallest, low pressure designs incor-

Figure 1. Three screw pump Low Pressure

Low Pressure

High Pressure

High Pressure

Pressure Rise

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xtending the life of a rotary multiple screw pump consists of optimizing the pump and its immediate environment as well as system considerations that can enhance longevity. Modern three screw high-performance pumps can deliver liquids to pressures of more than 4500 psi (300+ Bar) and flows to 3300 gpm (750 m3/h) with long term reliability and excellent efficiency. Twin screw pumps are available for pressures to 1450 psi (100 Bar), flow rates to 17,600 gpm (4000 m3/h) and can handle corrosive or easily-stained materials, again at good efficiencies.

Wraps or Stages Figure 2. The meshed outside screws, called idler rotors, cause each liquid-holding chamber to be separated from the adjacent one except for running clearances. This effectively allows staging of the pump pressure rise. The Pump Handbook Series

pumped liquid nor its cleanliness. Four mechanical shaft seals keep these bearings and timing gears isolated and operating in a controlled environment. As with three screw pumps, twin screw pumps use the staging effect to both minimize rotor deflection under pressure and to provide a longer leak path for internal slip to maintain good efficiencies.

The Pump Figure 3. The vast majority of twin screw pumps are of the double suction design. The opposed thread arrangement provides inherent axial hydraulic balance due to its symmetry.

porate replaceable liners in which the pumping screws rotate. Field repair is thus a simple matter. The center screw, called the power rotor, performs all the pumping. The meshed outside screws, called idler rotors, cause each liquidholding chamber to be separated from the adjacent one except for running clearances. This effectively enables staging of the pump pressure rise (Figure 2). High pressure pumps can have as many as 12 stages; low pressure pumps might have only two or three. Because the center screw is performing all the pumping work, the drive torque transferred to the idler rotors is only that necessary to overcome viscous drag of the cylindrical rotor spinning within its liner clearance. The theoretical flow rate of these pumps is a function of speed, screw set diameter and the lead angle of the threads. Essentially, flow rate is a function of the cube of the center screw diameter. Slip flow, the volumetric inefficiency due to clearances, differential pressure and viscosity are functions of the square of the power rotor diameter. This results in larger pumps being inherently more efficient than smaller pumps, a fact that applies to most rotating machinery.

Two Screw Pumps Generally, two screw or twin screw pumps are more costly to produce than three screw pumps and are not used as extensively. They can, however, handle applications that are well beyond many types of pumps, including the three screw designs. Twin screw pumps are especially well-suited for very low available inlet pressure applica-

tions—more so if the required flow rates are high. Services similar to three screw pumps include crude oil pipelining, refinery hot, viscous product processing, synthetic fiber processing, barge unloading, fuel oil burner and transfer as well as unique applications such as adhesive manufacture, nitrocellulose explosive processing, high water cut crude oil, multiphase (gas/oil mixtures) pumping, light oil flush of hot process pumping, cargo off-loading with ballast water as one of the fluids, tank stripping service where air content can be high and paper pulp production needing to pump over about 10% solids. The vast majority of twin screw pumps are of the double suction design (Figure 3). The opposed thread arrangement provides inherent axial hydraulic balance due to its symmetry. The pumping screws do not touch each other and lend themselves well to manufacture from corrosion resistant materials. The timing gears serve to both synchronize the screw mesh and transmit half the total power input from the drive shaft to the driven shaft. Each shaft effectively handles half the flow and thus half the power. Each end of each shaft has a support bearing to react the radial hydraulic loads that are not otherwise balanced. A few designs leave the bearings and timing gears operating in the pumped liquid. While this results in a significantly less expensive pump, it defeats much of the value that twin screw pumps bring to applications. The more common and better design keeps the timing gears and bearings external to the pumped liquid. They need not rely upon the lubricating qualities of the The Pump Handbook Series

Proper alignment of the pump to its driver is obvious but often it is often not achieved. Accurate alignment reduces driver and pump bearing wear and improves the pump shaft seal life. There are many alignment procedures available, all of which will achieve at least good alignment when followed. Don’t forget to do a hot alignment if the pump or pumped liquid is much hotter than about 150°F (65°C). During pump repair, do not mix used (worn) and new parts unless extensive experience has proven that practice works. In general, worn parts impose high loads and score the surfaces of new parts, resulting in very short life for the repaired pump. On difficult pumping services, premium parts may be available from the pump manufacturer. Features such as hard plating, coatings and case-hardened parts may be available to enhance pump life at a modest cost increase. This is especially true if the unit in question is an older pump. Bearings that have started to run hotter than normal or exhibit increased vibration should be changed immediately. Allowing a bearing to run to failure usually ensures that a great deal more damage is done to the pump. This will, of course, require a more costly repair. Leaking shaft seals should also be replaced immediately. Otherwise, leakage may reach the lubricant in a bearing and cause premature bearing failure with the corresponding downtime and repair expense. The use of carbon on silicon carbide mechanical seal faces has become more popular for extending seal life. Avoid the use of tungsten carbide unless running two hard faces together, e.g., silicon carbide on tungsten carbide. If running two hard faces, ensure that the seal chamber is maintained above atmos-

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pheric pressure as even momentary dry running of hard-on-hard faces will damage them. Pumps that require renewal of lubrication (oil or regreaseable) should be put on a rigid relubrication schedule. This is a very low cost way to avoid bearing failures. Do not over-lubricate the bearings or they will run hotter than necessary, reducing the useful life of both the lubricant and the bearing. If antifriction bearings are removed during pump servicing, replace them with new bearings rather than reusing the old ones. Bearings are generally not very expensive, have a finite life expectancy and can be easily damaged during the removal process, especially if pulled by the outer race. Multiple screw pumps that handle inherently “dirty” liquids such as residual fuel oils or relatively raw crude oil should be equipped with pumping elements that are enhanced to withstand this duty. Operating these pumps at reduced speed will result in larger, somewhat more expensive pumps, but the life expectancy will be greatly improved. Pump life is inversely proportional to at least the square of pump speed. You can achieve significant life improvement if you operate dirty liquid pumps at slower than their maximum speed. New or rebuilt/repaired pumps should be filled with liquid before

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start-up. The seal chamber should also be filled to avoid a dry start of the seal(s). Dry starting pumps or seals carries the risk of damaging the running surfaces before enough liquid reaches them and removes the frictional heat and prevents metal-to-metal contact.

The System The use of strainers and filters is highly recommended. These can minimize or prevent abrasive particles from reaching pump running clearances and scoring pumping elements or even causing a seizure. Properly operated in a clean liquid environment, rotary multiple screw pumps can provide astonishingly long operating life. Cases exceeding 100,000 hours operating time have been repeatedly reported. System contamination is probably the number one determinant to pump life. If the system includes stand-by pumps, rotate the use of the pumps between weekly and monthly. Spread the wear evenly over all the pumps. Stand-by units that have not been operated for prolonged periods are more vulnerable to sediment accumulation, air pocketing, rust and other difficulties that periodic operation will avoid. Some systems run a small bypass line around discharge check valves, which provide a small flow of liquid that runs through the non-operating pump(s).

The Pump Handbook Series

This keeps the stand-by pumps at operating temperature and prevents them from accumulating air pockets while they are at idle. Relief valve set pressure and proper valve operation should be verified several times per year. Over-pressure of a pump can result in reduced life or actual pump failure. Alarm systems for low liquid level, low pump inlet pressure, high liquid pump inlet temperature and excessive vibration will alert operators and/or shut down the pumps before the damaging effects of such out-ofrange operation can be experienced. Avoid changes in operating parameters (pressures, speeds, temperatures, liquid viscosity range, etc.) without first confirming that the pump is suitable for the new conditions. Consult the pump catalog or pump vendor. Small changes in operating conditions might take the pump beyond its capabilities, and thus shortens its life expectancy. " James R. Brennan is Market Services Manager for Imo Pump, Monroe, North Carolina. His responsibilities include worldwide marketing and technical support for pumping applications. Brennan is a 1973 graduate of Drexel University in Philadelphia and a member of the Society of Petroleum Engineers. He has 30 years of service with Imo Pump.

Extending the Life of Positive Displacement Pump Part 3: Progessing Cavity Pumps By Ed Wallace, Tarby

Progressing Cavity Pump Principles rogressing cavity pumps (also know as progressive cavity) are found in a wide range of applications because of their ability to handle difficult pumping applications. Classified as positive displacement pumps (flow rate directly proportional to the pump’s shaft speed), these pumps can move clean, thin fluids to fluids that are very viscous (up to 1,000,000 centipoise) with large amounts of solids (up to about 3 inches in diameter). The heart of the pump is the rotor and stator element. The rotor consists of a single helix external shape that rolls eccentrically inside a double internal helix stator. The stator is usually elastomeric and chemically bonded to a steel or stainless steel tube. The pitch of the double helix stator is twice that of the rotor. When the rotor and stator are assembled together they create what the inventor, Dr. Moineau, called “RM (René Moineau) capsulisms” (Figure 1). These capsulisms or cavities are of consistent size and shape and progress from pump inlet to discharge as the rotor turns in the stator.

Figure 2. Illustration of seal lines at the rotor and stator interface

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Figure 1. Shape of fluid in a RM capsulism

The rotor is manufactured slightly larger in diameter than the stator so that there is an interference fit between the two. This interference fit generates seal lines that seal the fluids in each capsulism or cavity (Figure 2). These lines enable the pump to move fluid against high discharge pressures. As the required discharge pressure increases, more seal lines are required to overcome this pressure. This requirement is met by providing an increased number of pitches or stages to the rotor and stator. Because these cavities do not change shape, solids contained in the pumped fluid will be transported through the stator element as the rotor rotates and pushed into the discharge line. The relatively short path around the rotor and the absence of any valves in the pump result in very low shear imparted to the fluid being pumped. For a more thorough description of the progressing cavity pumping principle see René Moineau’s U.S. patent (number 1,892,217 dated December 27, 1932). The Pump Handbook Series

With this rather unique set of pumping features, progressing cavity pumps can be found in service pumping fluids from clean but shear-sensitive fluids such as polymers to very abrasive solids-laden “fluids” like cements and everything in between. These pumps are heavily used in areas such as water/wastewater, pulp and paper, food and oil pumping, to name just a few.

Maintenance Considerations To help assure maximum possible progressing cavity pump life, the following maintenance and process related issues are discussed. Lubrication Lubrication is often one of those items that shows up in daily preventative maintenance programs. As with most mechanical components, progressing cavity pumps require lubrication for maximum service life. Follow your pump manufacturer’s lubrication guidelines closely. Over-lubrication of the pump’s bearings can actually reduce overall service life. On pumps that have lubricated drive components, many require little if any lubrication except at certain maintenance occasions. Again, consult the service manual that was supplied with your pump for lubrication type(s) and frequency recommendations. Packing Although progressing cavity pumps are well suited for mechanical seal use, most are manufactured

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Figure 3. Open throat progressing cavity pump for very viscous fluids

with packing. A wide spectrum of packing styles is available to suit nearly any application, but proper installation and maintenance is important. The typical packing installed by many manufacturers is not intended to totally prevent leakage. Most require a small amount of leakage to flush away any abrasives caught between the packing and the drive shaft. It is strongly recommended that you follow the packing manufacturer’s installation and maintenance recommendations. Excessively tight packing can result in scoring and abrading of the drive shaft, accelerated packing wear and higher horsepower demands from the pump. Pay particular attention to the packing manufacturer’s recommendation for start-up and break-in. When it is time to remove or replace the packing, use packing pullers to remove the packing and lantern rings. Use of screwdrivers or other objects can damage the chrome-plated (or other hard-surfaced) area of the drive shaft, leading to accelerated wear on replacement packing. Avoiding Run Dry Because of the interference fit between the rotor and elastomeric stator mentioned above, significant heat can build up and destroy the stator if the pump is run without fluid in the capsulisms or cavities. This is often referred to as “run dry” and can occur suprisingly quickly, especially with pumps turning at higher rpms. After any maintenance has been performed, and before start-up, check for fluid blockages and remove any you find. It is easy to forget to open an inlet or discharge valve somewhere in the piping system, but the results can be damage to and shortened stator service life. On initial installation and after service reinstallation, consideration should also be given to the length of time the pump must run before it can pull the fluid to be pumped from the source

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to the pump’s inlet. This is especially important in suction lift applications. After any maintenance shutdown, flooding the inlet piping can help assure the pump gets fluid before it runs long enough to experience a “run dry” condition. Process Changes Changes in process and operations over time seem inevitable, but progressing cavity pumps are quite flexible in handling these changes. However, changes are added to changes, and after a period of time the service the pump is currently performing often bears little resemblance to the original performance specification. If it has been a while since the current performance criteria have been reviewed, it may be time to do so. Increased Flow/Speed Changes Changes in pump speed to meet increasing flow demands, especially with fluids containing abrasives, can result in a significant reduction in service life. As a general rule in abrasive service, increasing the pump’s speed leads to an inverse square reduction in life. As an example, doubling the speed of the pump can result in the service life being reduced to 1⁄4 of the life achieved before the change was made. With the modularity of most progressing cavity pumps however, changes in the rotor and stator elements to a new size may be a simple and economical way to avoid reduced service life.

Piping Considerations Piping, although not part of the pump directly, can have a significant impact on pump longevity. It is usually recommended that the piping system have no abrupt diameter transitions, excessive elbows, valves or other flow obstructions. The greater the number of obstructions, the higher the pressure the pump will have to overcome to push the fluid at the desired flow rate. Piping The Pump Handbook Series

systems with short 90° elbows or pipe diameter reductions immediately following the pump’s discharge flange create significant turbulence and unnecessary backpressure, which can reduce the life of the rotor and stator. Maintaining the proper diameter of piping throughout the length of the piping system can also keep discharge pressures to a reasonable level and increase pump life. Piping must also be properly installed and aligned. Misaligned piping exerts stresses on the structural components of the pump that, if high enough, can cause deflections. Misaligned inlet or discharge piping tends to push or pull the pump’s suction housing out of alignment with the bearing housing. If this occurs, the packing that is supported in the suction housing can become misaligned with the drive shaft that is supported by the bearing housing. During the initial installation and subsequent piping maintenance, don’t forget the effects of thermal expansion and contraction and the stress effects this expansion and contraction can have on the pump. Provision must be made for the movement of the piping system, and expansion joints can help eliminate the problem. One final consideration for inlet side piping: Make sure that all flanges, gaskets and threads are in place and properly tightened. Suction leaks in the inlet can enable cavitation, creating noise and if severe enough result in the pump running dry.

Baseplate Factors Just as misaligned piping can create stresses in the pump reducing component life, failure to properly shim the pump on its supporting baseplate can induce stresses. Proper installation, shimming and alignment needs to be done with appropriate preventative maintenance checks to ensure that alignments have not been lost over time.

Fluid/Elastomer Compatibility Compatibility of the stator’s elastomer is a prime consideration. Any changes in the characteristics of the fluids being pumped can result in a

chemical incompatibility issue with the stator’s elastomer. If there have been any changes in the fluid’s chemical make-up, it’s probably a good idea to contact your pump’s manufacturer to see if there is a chemical incompatibility issue. Compatibility applies not only to the fluid being pumped on a continuous basis, but also to fluids introduced to the pump during flushing operations. Keep in mind that any fluid coming into contact with the pump at any time will be able to attack the stator’s elastomer if it is not chemically compatible. Because the rotor rolls eccentrically but parallel to the axis of the stator, some type of universal or flexible joint system is required to enable this rolling motion. Typically one and often both joints are exposed to the fluid being pumped, which requires some sort of flexible sealing method. Seals, washers, boots or gaskets protecting the pump’s pin, gear or flex joint drive mechanism can also be attacked by fluids that are not chemically compatible. This can result in loss of protection to the mechanical components of the joint system and severely reduce the pump’s service life. Any time the pump must be down for servicing, a thorough inspection of these sealing systems is recommended. Close attention should be paid to signs of chemical attack, nicks, gouges, punctures, abrasion or tears. If any of these conditions are found, the sealing components must be replaced, the internal components of the drive system inspected and a determination made as to suitability for continued service. Assembly of the rotor into the stator can be quite difficult if attempted without lubrication. Both rotor and stator should be liberally lubricated. Exercise caution in the selection of the lubricant. Chemical compatibility of the lubricant with the elastomer in the stator is an important consideration. Generally, soaps such as dishwashing liquids are not

recommended, as they tend to cause the rotor to “adhere” to the stator. It can be extremely difficult to break the rotor/stator assembly free if it has set for any length of time. A readily available lubricant that does not react with most elastomers is caster oil.

Elastomer Temperature Compatibility Along the same line as chemical compatibility is thermal compatibility. Each elastomer has a temperature limit. Check the manufacturer’s literature to determine your pump’s limit. Exposing that stator to temperatures greater than the published limit can result in significantly reduced stator life. This applies not only to the routine product(s) pumped, but also to any pump or piping flushing operations performed. While on the topic of temperature, as the temperature of the fluid being pumped increases, the rotor is subject to thermal expansion as well. The stator’s elastomer expands at a much greater rate than the metallic tube that it is bonded to; thus, expansion occurs inward toward the rotor. This combination of part expansion increases the interference fit between the rotor and stator, which can have a significant effect on starting torque. An increase in fluid temperature from 70°F to 100°F for some elastomers will result in a 10% greater starting torque, while a change from 70°F to 175°F increases starting torque by a staggering 100%. Therefore, if starting takes 10 hp at 70°F, and if no changes have been made to the rotor and stator, starting will take 20 hp at 175°F. If fluid process temperatures have been increased over time, consideration should be given to undersize rotors. Machining of the rotor to smaller dimensions (undersizing) compensates for the thermal expansion that occurs at higher pumping temperatures and brings the interference fit between rotor and stator back to standard values.

The Pump Handbook Series

Fluid Solids and Abrasives Considerations Process changes can result in the need to increase the percent of solids content or the abrasive characteristics of the pumped fluid. As the abrasiveness of the solids increases, changing to a “softer” elastomer and/or slowing down the pump can result in increased stator life. However, there may be some trade-offs in pressure capabilities. When evaluating the effects of abrasives, review the discharge pressure in terms of psi per stage. Progressing cavity pump manufacturers have recommended maximum psi per stage based on the characteristics of the fluid being pumped. The more abrasive the fluid, the lower the recommended pressure per stage. Thus, changing the number of stages of the rotor and stator may yield a significant improvement in pump longevity.

Conclusions Although progressing cavity pumps are selected for their ability to handle difficult fluids, it is not hard to maintain them or obtain maximum service from them. The pump’s design is relatively simple and awareness of proper maintenance and application considerations will help ensure lasting value. As service conditions change, and they frequently do, minor changes to the pump and a good preventative maintenance program can return significant cost reductions when compared to using the pump in changing circumstances without compensating for those changes. " Ed Wallace is Director of Technology at Tarby, a company that has been manufacturing progressing cavity pumps and parts for more than 20 years. He has been involved with various manufacturers in the progressing cavity industry for the past 20 years. Mr. Wallace is a graduate of Piedmont College and has an M.B.A. from Oklahoma City University.

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POSITIVE DISPLACEMENT PUMPS HANDBOOK

Peristaltic Pump Technology for Industrial Applications Once just lab pumps, these designs are making the move to more and more factory floors. By John A. Beahm, Watson-Marlow/Bredel Pumps eristaltic Pumps have been in use for decades, but have always been subject to a “lab pump” prejudice. True, the technology was first developed as a contamination-free method of fluid transfer in medical and laboratory applications, but the development of this technology has been driven by the desire to put its inherent benefits to work in industrial applications. Many features of peristaltic pumps are truly unique and have made them the pump of choice for many applications in industries including pharmaceutical processing, mining, paint and coating manufacture, and water and wastewater treatment.

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Peristaltic Technology Development The basic principle of a peristaltic pump is a mechanical application of “peristalsis,” a physiological term referring to the alternating contraction and relaxation of muscles around a tube (e.g., throat or intestines) to force the contents through it. It is not surprising, then, that the application of this basic peristaltic principle as a method of mechanical fluid transfer was developed for medical and biotechnology applications. A smooth wall, flexible tube (or hose) is rolled and squeezed along a pre-determined length, positively displacing the fluid contained within. The tube’s restitution after

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the squeeze produces a vacuum that draws more fluid into the tube. This creates a gentle pumping action that causes minimal damage to the media, particularly in comparison to other methods of mechanical transfer. Also, because the fluid is contained completely within the tube, the opportunity for contamination is minimal. These attributes have made peristaltic pump technology attractive for many industrial applications, and today’s designs are capable of handling flow rates and pressures unthinkable thirty years ago. From Lab to Industry The application of the peristaltic principle in numerous pump designs on the market is similar. The pump employs a drive-coupled rotor with rollers or “shoes” mounted on it (Photo 1). These continually compress and occlude some portion of the tube. This action moves the fluid

Photo 1. Close-up of a peristaltic pumphead on a tubing pump

The Pump Handbook Series

through the tube with a constant rate of displacement for each revolution of the rotor, enabling a precise measurement of the volume of fluid pumped through the tube. The amount of displacement is determined by the width of the rotor’s “sweep,” coupled with the tube’s inner diameter and a precise flow rate can be determined by counting or calculating the number of rotor revolutions over a given time period. The key component of the pump is its tube. The tube is the only part to undergo wear, and the interior surface of the tube is the only component that contacts the fluid. The tube’s restituitive ability (the number of squeezes it can withstand) can determine its viability for a given application. The development of long-life, chemically compatible tubing has been a primary factor in bringing peristaltic pumps to the industrial market. In addition to its precise metering capability, there are a number other characteristics that have made this type of pump an attractive method of fluid transfer. The construction is sealless, glandless and valveless. The tubing material itself is much more resistant to abrasive slurries and corrosive chemicals than metal or PVC. Peristaltic pumps’ simple maintenance requirements reduce downtime and repair costs. Changing the tubing is the only regular mainte-

nance needed. Despite all of these performance features, the use of peristaltic pumps had for decades been restricted to low flow, low-pressure applications, which limited its industrial uses. Even within its flow and pressure limitations, the peristaltic pump has moved out of the laboratory and onto the shop floor, providing chemical dosing, transferring and metering of pigments, dyes, and other additive-type applications, as well as transfer of emulsion-sensitive and shear-sensitive products. The ability to quickly change tubing and thus clean the pump is ideally suited for applications with numerous product changeovers. The gentle action of the pump, with the contents cushioned by the tubing material, helped fulfill a need for a pump that would not damage sensitive products. The wide range of “niche” applications served by peristaltic pumps encouraged manufacturers to continue its development toward heavier duty service.

An “Industrial” Peristaltic Pump Bredel B.V., a Dutch company that manufactured mixers and other industrial processing equipment, pioneered the development of a heavy-duty industrial peristaltic pump. In 1979 the company introduced a new pump option to U.S. processing industries: the Bredel heavy-duty “hose pump.” The development was in response to the fundamental problem in applying peristaltic technology to industrial applications, which required a hose element that could accommodate the higher pressures and flow rates demanded by the process industries while maintaining the capacity to handle highly abrasive and chemically aggressive fluids. The tubing traditionally associated with peristaltic pumps, and generally manufactured as a single extruded element, needed reinforcement. In the early 1970s Bredel developed a composite-laminated hose element with a patented reinforcement structure. The design is a triple-layered construction with an inner layer surrounded by a reinforcement cord matrix and a durable machined outer layer. (The original hose had steel cord reinforcements; later nylon cord was found to be

more suitable.) The hose was field tested in Europe along with a prototype pump assembly. Although the hose performed well, there were some concerns with the pump assembly. After re-designing the assembly, Bredel introduced the first heavy-duty peristaltic hose pump for industrial applications, bringing the inherent advantages of peristaltic technology more in line with the flow and pressure requirements of large-scale processing operations. The prototype pump assembly was a close-coupled design (the pump rotor was mounted on the output shaft of a motor/gearbox configuration). This arrangement made use of the gearbox bearings and sealing to handle the weight of the rotor. These prototypes revealed the inability of the bearings and seals to withstand the high loads that often exceeded their ratings and subsequently caused gearbox failures. This failing led to a new innovation, the bare-shaft pump assembly, which was introduced in 1979. This design incorporated Bredel’s patented hose construction along with a bare-shaft pumphead coupled to a gear reducer by a flexible coupling that utilizes two sets of bearings. With the development of heavyduty hose pumps arose an inherent design problem with the “rollers” traditionally associated with peristaltic pumps. Rollers need bearings, and under the high pressure experienced with these pumps, the roller bearings tend not to roll. The rollers actually slide, thus no longer functioning as “rollers.” The 1979 Bredel design replaced the rollers with two smooth-surfaced “pressing shoes” mounted to the rotor. In addition to being a simpler, lower maintenance, lower cost design, the shoes, with a larger radius of curvature, provide a gentler and more complete means (longer slip path) of hose occlusion. These shoes were designed with profiled leading and trailing edges to give a gradual lead-in and termination to each tube occlusion. This prevents an abrupt imposition and release of pressure, which could significantly reduce hose life. Additionally, the use of a specially developed lubricant contained in the pumphead reduces the hose wear that results from contact with the pressing shoes. The Bredel SP Series The Pump Handbook Series

hose pump was the first design capable of handling the demands of highflow and high-pressure industrial applications. Since the introduction of the Bredel hose pump, several other manufacturers have introduced their own hose pump designs that have proven to be viable variations on the Bredel theme. (The Bredel design has changed little since the late 70s.) Hose pumps enabled manufacturers to expand their “niches” because they employ the same inherent benefits of peristaltic pumping. The pumps have been successful in replacing other pump technologies that could not meet the demands of a difficult application. Two decades after the introduction of the Bredel design, peristaltic hose and tube pumps have become known as problem-solving pumps in metering and transfer applications as diverse as the process industries. For most pumping applications, many pump technologies are available that could “do the job,” and initial purchase price often dictates the pump decision. However, for applications where high maintenance, lengthy production downtime, or careful handling of product are critical issues, peristaltic pumps are very cost competitive. This is especially true when you compare their total cost of ownership to lower priced, but perhaps less effective technologies. Although still somewhat limited in flow rate capability, peristaltic pumps have answered the call in an increasing number of fluid processing applications.

Peristaltic Applications Abrasive Slurries One area that peristaltic pumps have been very successful with is in applications involving the handling of abrasive slurries. Most other pumps in this service will have the rotor and stator together in the product zone. Abrasive particles can wear these areas of the pump assembly. This can open up critical tolerances between the rotor and stator and cause a drop in pumping efficiency and eventual failure. “Traditional” positive displacement pump designs require this tight tolerance to function. A peristaltic pump rotor, however, stays out of the product zone. The hose is the only thing to touch the product, and

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this keeps abrasive wear to a minimum. The slurry particles are actually “cushioned” as they ride through the pump. In fact, a properly manufactured peristaltic pump hose does not fail due to abrasive wear; failure is ultimately a result of the accumulated number of compressions. In many applications, including those handling abrasive slurries, hose life can be counted in thousands of hours. Due to their reliability and abrasion resistance, hose pumps have been successfully replacing “traditional” pump types in a variety of slurry applications including mining, ceramics processing, and the transfer of magnesium hydroxide and other abrasive chemical compounds. Corrosive Fluids Corrosive fluids have also traditionally been troublesome to many pump types. Corrosive attack on wetted metal or PVC parts can lead to pump failure; chemical attack on pump seals can result in leakage, leading to product loss and/or hazardous materials containment problems. Remember, the peristaltic pump’s rotor is not in the product zone—the fluid is contained completely within the tube. Peristaltic pump designs are sealless, and the available hose or tube element materials meet a wide range of chemical compatibility. Products That Crystallize Products prone to crystallization are also well served by a peristaltic pump. There are no valves or glands in a peristaltic design, so clogging of those parts cannot occur through the buildup of crystals, a drawback with some other pump types. Pumps with valves or diaphragms can also be adversely affected by other chemical actions (such as off gassing in sodium hypochlorite applications) that do not affect peristaltic pumps. High-Viscosity and Shear-Sensitive Fluids Other problematic fluids for many pump types include shear-sensitive and high-viscosity fluids. Peristaltics are a good choice for transferring shear-sensitive fluids such as latex, live cells, or even live fish. The pump’s gentle, non-emulsifying action minimizes damage to the product or its consistency. High-viscosity products have always posed a pumping problem

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because they can clog and coat a pump’s wetted parts. This is in addition to the priming difficulty some systems have when moving high-viscosity product. Peristaltic pumps are inherently self-priming and can draw almost full vacuum, producing the pull necessary to draw viscous materials such as grout, ground beef or waste sludge into the pump. Some heavy-duty hose pumps have lasted for years in service with periodic hose and lubricant changes as their only maintenance requirement, even while handling the most aggressive and destructive fluids.

Peristaltic Pumps as Metering Units Peristaltic pumps have also performed well in applications where accurate volumetric control of the pumped fluid is a requirement. This “metering” capability has become one of the primary functions of the hose and tube pumps in industrial applications. Total flow is computed by the manufacturer’s specification of volumetric displacement per one revolution of the rotor multiplied by the number of revolutions. Flow rate is calculated by multiplying the displacement by the revolutions per minute. Because the hose always maintains a constant volumetric displacement, variable speed pumps offer the capability of changing the rate of volumetric output by changing the rotor speed. (There is a linear correlation between rotor speed and flow rate.) Typical hose pump metering applications often involve the controlled release of an adjunct product into a process stream. For example, effluent in a water treatment application involves the addition of a variety of chemicals to the process in regular, controlled doses. A hose pump can handle these often corrosive or abrasive compounds (without damage to the pump) and release them in precise quantities into the process stream. One of the keys to metering accuracy and repeatability is the restitution of the hose. With a properly constructed and maintained hose, a hose pump can deliver fluid with accuracies to 1% and a total drop-off over the life of the hose of about 3%. This kind of accuracy can also be achieved by tubing pumps down to The Pump Handbook Series

very low volumes and flow rates. Hose (and tube) life can vary greatly from application to application, so “a rule of thumb” for how many hours a hose should last is impractical. Hose life is mainly dependent upon rotor speed and process pressure. Hose pumps can maintain their metering accuracy even at the high end of their flow and pressure-handling capabilities. Running a pump faster, however, will decrease the hose life. For this reason, pumps are often sized so that the desired flow rate can be achieved by running the pump at a slower speed, toward the lower end of its range. This “slow running” will optimize hose life, and a hose life of thousands of hours is not uncommon in many industrial applications. Consistency in the manufacturing of the hose element makes it possible for a typical hose life to be established for a particular application. Consistent volumetric displacement from hose to hose can also be determined. A predictive maintenance schedule for hose changing ensures consistent metering accuracy. Chemical Metering Chemical metering is one area in which hose pumps have become an exceptional alternative to traditional pumping and metering methods. Many process streams controlled by a combination of pumps and flowmeters can be served just as well with a well-designed hose pump and tachometer—a considerable savings in equipment costs. High concentrations of caustic materials or particulate matter can attack the wetted parts of the flowmeter and affect the reliability of the measurement. Viscous media can coat flow sensors, affecting their measurement. Many slurries also have proven difficult to measure with a flowmeter. Utilizing a hose pump, the chemically compatible material can be checked and selected prior to installation, and fluid characteristics or process pressure will not affect the metering accuracy of the pump. Hose pumps have been successfully employed in a wide variety of highly automated metering applications with accurate control of both flow rate and total flow through the regulation of pump speed. Another feature of the peristaltic design, when it’s integrated into more complex systems, is that

Photo 2. A “duplex” hose pump configuration

Photo 3. A close-coupled tubing pump at work in a printing application (metering ink)

the pump can act as its own check valve. Some portion of the hose or tube is always occluded, so when the pump stops, that point stays squeezed shut. This feature removes the need for an additional valve to prevent over-dosing or backflow.

Expanding the Application Range Although the basic peristaltic pumphead design has not changed for many years, there have been several design innovations that have expanded the application range of both hose and tube pumps. Sanitary Applications One of these developments has been the introduction of CIP/SIP hose pumps for applications that require an unobstructed passage for cleaning. Using a hinged roller or shoe mechanism and mechanical or electrical actuator that enables the rollers to retract, the hose pump’s flow path can be “opened up” for unobstructed clean-in-place or stream-in-place processes. Within the flow passage there are no cavities, crevices or obstructions to collect residue or bacteria. These pumps are generally constructed with mounting frame, rollers, and hardware of stainless steel and a casing employing a bakedon epoxy powder finish. This combi-

nation makes them suitable for most sanitary processes and washdown environments. With the availability of food grade and pharmaceutical grade hose and tube materials, peristaltic pumps have found a number of applications in the food and pharmaceutical processing industries. These features, in addition to the fluid containment and easily changeable tube or hose inherent to peristaltic design, make these pumps a natural fit in applications where product contamination is an issue. Duplex Design Another innovation that has expanded the application range of peristaltic tube and hose pumps is the “duplex” design: two pumpheads are driven from a single motor/gearbox configuration (Photo 2). Duplex pumps offer substantially increased flow rate capacity with the added benefit of significantly reduced pulsation in the output flow. The higher flow, low pulsation is a result of incorporating two pumpheads with a single drive. The design uses four pressing shoes or rollers on the hose or tube instead of the two traditionally used in a single pumphead configuration. The shoes are staggered 90° apart, which not only produces a smoother flow profile, but also reduces the power requirements necessary to drive two pumpheads. Applications that would employ two hose pumps to increase flow capacity also benefit from the reduced space requirements of the duplex design. There are duplex tubing pump designs as well, although often pulsation reduction and flow capacity increase are achieved by incorporating additional rollers or multiple tubes in a single pumphead. Close-Coupled One tubing pump that has utilized duplexed pumpheads is a “close-coupled” design where the pumphead is coupled directly to a motor/gearbox configuration (Photo 3). Unlike larger hose pumps, the gearbox bearings in a close-coupled tube pump are capable of supporting the loads placed on them, providing fully assembled, highly compact pumps without the added cost of special couplings, flanges or adapters. One of the primary benefits of this innovation is that it creates a simpler and more economical pump without comproThe Pump Handbook Series

mising the benefits of the peristaltic design. The motor/gearboxes used in the close-coupled pumps can be washdown-rated and explosionproofed if the application demands these capabilities. The close-coupled design has also enabled peristaltic pumps to be more in line with the prices of other positive displacement pump technologies. Tubing Developments Another innovation expanding the application range of the tubing pump has been the development of more pressure-resistant and longerlife tubing materials. Pressure ratings to 100 psi are now available and tube life of up to 10,000 hours or more can be expected in some applications. These recent developments have significantly increased the number of industrial tubing pump applications because these parameters are more in line with the expected performance of industrial process equipment. The peristaltic tube pump can no longer be considered simply a fragile “lab pump.” Limitations Despite the numerous application benefits derived from peristaltic pump technology, there is no denying the limitations that have kept the design from becoming the pump of choice for many other applications. Considering pump size in relation to flow output, the peristaltic design cannot deliver the flow capacities that can be achieved with centrifugal or rotary lobe pumps. Even the largest duplex pumps can only deliver 350 gpm. Although chemical compatibility of the hose or tubing material makes peristaltic pumps ideal for some applications, it also limits their usage in applications where the elastomeric materials are not compatible, as with certain oils, fats and hydrocarbon-based fluids. Temperature is also a process parameter that limits the application range of peristaltic hose and tube pumps; the elastomers are not as temperature-resistant as metal or PVC components. Also, despite pressureresistant hose designs, some process pressures, or the combination of temperature and pressure, make the use of a hose pump impossible. If an acceptable hose or tube life cannot be obtained, peristaltic pumps begin to

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New Innovations

Photo 4. The new Bredel direct coupled hose pump

lose their lifetime cost advantage. Peristaltic pump manufacturers have recognized these limitations, however, and current research and development efforts are working toward finding solutions. The development of high endurance hose and tubing, without sacrificing restitution and accurate metering capability, is a primary goal of manufacturers.

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One of the most recent hose pump product enhancements has been introduced by Bredel—a new “directcoupled” hose pump design, which improves upon both a bare-shaft and a close-coupled design (Photo 4). At the core of the new design is a large diameter twin-bearing hub integrated into the pumphead. This unique feature ensures that the rotor is always fully and centrally supported by its own bearings, and it also eliminates the need for additional gearbox couplings or adapters. The gearbox plugs directly into the pumphead. An additional feature is a unique buffer zone that protects the motor and gearbox, and helps prevent lubricant seal failure. This new design reduces the weight and footprint of similarly sized hose pumps, and it is easier to install and maintain. Although this advancement does not expand the hose pump’s application range per se, it does make the pump considerably more

The Pump Handbook Series

attractive for those applications that can already utilize hose pumps.

Conclusion Peristaltic hose pumps offer precise positive displacement and minimal maintenance requirements, with the ability to handle abrasive and aggressive fluids. Although peristaltic technology is still a new concept to many process industries, the success of these pumps in handling a variety of difficult process fluids ensures that their utilization will continue to grow as a valuable asset in industrial processing applications. " John A. Beahm is the Marketing Communications Manager for WatsonMarlow/Bredel Pumps. He has been working for manufacturers of processing equipment for the past eight years, specializing in technical writing and market research. Mr. Beahm holds B.A. and M.A. degrees in communicationsrelated fields.

POSITIVE DISPLACEMENT PUMPS HANDBOOK

Pulsation Control for Reciprocating Pumps Suppress the surge in the suction and discharge systems of your positive displacement pumps. By V. Larry Beynart, Fluid Energy Controls, Inc.

his article reviews the specific requirements for the selection and sizing of pulsation dampeners to reduce the fluid pressure pulsations generated by positive displacement reciprocating pumps. It includes analyses of the nature of pressure pulsation and explains the advantages of using pulsation dampeners and suction stabilizers. Sizing procedures for dampeners illustrate how their application improves suction and discharge control in reciprocating pumps, reduces pressure pulsations, increases Net Positive Suction Head (NPSH), and makes the service life of pumps longer and more efficient.

T

Combating Pulsation The primary purpose of pulsation control for reciprocating pumps is to attenuate or filter out pump-generated pressures that create destructive forces, vibration and noise in the piping system. Every reciprocating pump design has inherent, built-in pressure surges. These surges are directly related to the crank-piston arrangement (Figure 1). Fluid passing through a reciprocating pump is subjected to the continuous change in piston velocity as the piston accelerates, decelerates, and stops with each crankshaft revolution. During the suction stroke, the

piston moves away from the pump’s head, thereby reducing the pressure in the cylinder. Atmospheric pressure, which exists on the surface of a liquid in a tank, forces liquid into the suction pipe and pump chamber. During the discharge stroke, the piston moves toward the pump’s head, which increases the pressure in the cylinder and forces the liquid into the discharge pipe. This cycle is repeated at a frequency related to the rotational speed of the crank. At every stroke of the pump, the inertia of the column of liquid in the discharge and suction lines must be overcome to accelerate the liquid col-umn to maximum velocity. At the end of each stroke, inertia must be overcome and the column decelerated and brought to rest. Figure 2 compares pump pulsation frequencies to pipe span natural frequencies of vibration. A key benefit of pulsation control is the reduction of fatigue in the pump’s liquid end and expendable parts. Reducing the pressure peaks experienced by the piston will in turn reduce the power-end peak loading.

charge pulsation control equipment. While multiple pumps may be arranged to run at slightly different speeds, it is impossible to prevent them from frequently reaching an “in-phase” condition where all the pump flow or acceleration disturbances occur simultaneously. Having more than one pump can multiply the extent of the pipe vibration caused by such disturbances because the energy involved is similarly increased. The reduction of the hydraulic pressure pulses achieved by using pulsation control equipment is usually reported in total pulsation pressure “swing” as a percentage of average pressure. This method is used widely within the industry and is often recommended as a standard. Referring to Figure 3, the percent of residual pulsation pressure can be calculated as follows:

The Source of Fluid Pressure Pulsations It is almost mandatory that multiple pump installations have well-des igned individual suction and disThe Pump Handbook Series

Figure 1. The rotary motion of the crank movement is transformed into the reciprocating motion of the piston.

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be difficult to measure with some conventional types of flowmeters.

Figure 2. (a) Typical pump and pulse frequencies (b) Typical piping natural frequencies

Case I: P/Ave x 100 = 460/1000 x 100 = 46% Case II: P/Ave x 100 = 70/100 x 100 = 7% It is very important that any reference to degree of pulsation should apply to the total excursion in terms of pressure and percentage. In Case I, the total pulsation of 460 psi (46%) implies that the excursion is from 680 psi (32% below the average) to 1,140 psi (14% above the average). Reciprocating pumps introduce into the suction and discharge systems three apparently unrelated pressure disturbances, which are illustrated in Figure 4. These include: • a low-frequency disturbance, based on the rate at maximum flow velocity pressures, occurs at A • a higher frequency due to maximum acceleration pressure at the beginning of each piston stroke at B • a pressure disturbance at the point of flow velocity change (valley) at C

Major Problems Caused by Pressure Pulsations Unsteady Flow The most obvious problem caused by pulsation is that flow is not constant, which can lead to problems in processes where a steady flow rate is required, such as applications involving spraying, mixing or metering. Flow rates can

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mechanical fatigue can lead to total failure of components and welded joints, particularly when the liquid being pumped is corrosive.

Noise and Vibration Many reciprocating pump installations suffer from problems that can lead to excessive maintenance costs and unreliable operation. Typical examples are noise and vibration in the piping and the pump. Vibration can lead to loss of performance and failure of valves, crossheads, crankshafts, piping and even pump barrels. High levels of pulsation can occur when the pulsation energy from the pump interacts with the natural acoustic frequencies of the piping. A reciprocating pump produces pulsation at multiples of the pump speed, and the magnified pulsation in the system is generally worse at multiples of the plunger frequency. Most systems have more than one natural frequency, so problems can occur at differing pump speeds.

Cavitation Under certain pumping conditions, areas of low system pressure can occur. If this pressure falls below the liquid’s vapor pressure, local boiling of the liquid will occur and bubbles of vapor will form. If the liquid has dissolved gases, they will come out of solution before the liquid boils.

Pressure Waves Summary The earlier analysis of the source and nature of fluid pressure pulsations enables us to build the idealized pressure wave pattern (Figure 5). Although identical forces are acting on the liquid in the suction and discharge sides of a pump, the effects of these forces vary greatly. The major difference is that acceleration effects in the suction at B tend to separate the liquid into vapor or “cavitation” pockets that collapse with a high-magnitude pressure pulse. Acceleration on the discharge side of the piston at B, however, tends to compress the liquid and create an impulse pressure pulse. This requires separate consideration of the suction and discharge systems when choosing and installing pulsation control and suction stabilizing equipment.

Shaking Shaking forces result from the pulsation and cause mechanical vibration of the piping system. These forces are a function of the amplitude of the pulsation and the crosssectional area of the piping. When the exciting frequency of the pulsation coincides with a natural frequency of the system, amplification and excessive vibration occur. Amplification factors can be as high as 40 for pulsation resonance and 20 for mechanical resonance. If the mechanical resonance coincides with the acoustic resonance, a combined amplification factor of 800 is possible. Wear and Fatigue Pulsation can lead to wearing of pump valves and bearings and, if coupled with mechanical vibration, will often lead to loosening of fittings and bolts and thus to leakage. In severe cases,

Pulsations in Suction Systems The friction losses in a suction system are usually low because of the relatively short length and large diameter of the piping involved. Accordingly, the flow-induced A-type

MAX 1140 PSI

MAX 1021 PSI AVE 1000 MIN 951

AVE 1000 MIN 680

 P 70 ZERO

P 460 ZERO (a)

(b)

Figure 3. Reporting the degree of pulsation and control (a) Non-dampened waveform from flow variation (Case I) (b) Dampened waveform from flow variation (Case II)

The Pump Handbook Series

Figure 4. Triplex single-acting pump—points of induced pressure disturbances

pressures are of low magnitude compared to the accepted percentage of change (Figure 4). For example, if the suction pressure is a static 20 psi, the 23% variation of a triplex single-acting pump would generate a theoretical Atype pressure variation of only 9.2 psi. This is hardly enough energy to set pipes in motion. The same flow variation at 1,400 psi working pressure, however, creates more than enough energy—a 644 psi A-type pressure variation in the discharge line. Even so, the forces of acceleration become overwhelming disturbances in the suction. Pressure pulses of more than 25 psi are often encountered in pumps—even in systems with short suction pipes. A small amount of dampening of the flowinduced pressure can reduce pulsation to a negligible amount, leaving the 25-psi C-type acceleration pulsations. These remaining pulsations can create damage by enabling cavitation (Figure 4). Cavitation In practice, cavitation occurs at a pressure slightly higher than the vapor pressure of the liquid. Because vapor pressure is a function of temperature, a system may run without cavitation during the winter months but be subject to cavitation problems during the summer when temperatures are higher. Cavitation is connected with the presence of microscopic gas nuclei, which are in the solid materials bonded in the liquid. It is because of these nuclei that liquids cannot withstand tension. Without them, it is estimated that

water could transmit a tension of approximately 150,000 psi. During the beginning stages of cavitation, the nuclei give rise to the formation of gas bubbles, which are swept along to an area where the pressure is higher. The bubbles then collapse, producing high intensity pressure waves. The formation of successive bubbles quickly follows. This repeating cycle lasts for only a few milliseconds, but during this time the localized pressures can be as high as 60,000 psi and temperatures can rise as much as 1500°F. Cavitation can occur at a pump’s inlets, inside the barrels, at the draft tubes of turbines and on propellers, mixing blades, Venturi meters and other places where changes in fluid velocity occur. The cavitation effects in rotary positive displacement pumps are similar to those in rotodynamic pumps. The usual effects are wear of surfaces in contact due to erosion, slow increase in clearances, reduction in performance, noise and vibration. Cavitation damage in reciprocating pumps can happen very quickly and be catastrophic. A cavitating 60 hp pump can break a major component every 350 hours. Suction Control Suction control is an essential step in the design of any pump system and is equally important for reciprocating and centrifugal pumps. This process requires the following major steps: 1. Compare Net Positive Suction Head available (NPSHa) to Net Positive Suction Head required (NPSHr). (NPSH is a feature of the system layout and liquid properties. It is often referred to simply as NPSHa or NPSH available. Pump manufacturers usually quote NPSHr. This is a function of the pump’s performance and is related to the pump’s susceptibility to cavitation. To calculate the NPSHa to the pump, it is necessary to take into account the fricThe Pump Handbook Series

tion losses of the suction piping. NPSHa is then the difference between the total pressure on the inlet side of the pump and the sum of the vapor pressure of the liquid and the friction losses of the suction pipes.) 2. NPSHa is greater than NPSHr, the general suction conditions are under control and their improvement (with respect to NPSHa) by installation of a suction pulsation dampener is not required. In this situation, the installation of a pulsation dampener would be required only for pressure fluctuation and noise reduction purposes. 3. If NPSHa is less than NPSHr, and the piping rearrangement of a suction system is impossible or insufficient, installation of a pulsation dampener before the pump’s inlet is essential. 4. Choose the appropriate dampener design based on steps 1-3. Using this process will help you achieve maximum reduction of pressure fluctuation.

Pulsations in Discharge Systems The same forces at work on the suction side also affect discharge systems, but the pressure at A due to flowinduced pulsations becomes overwhelming at 460 psi (Figure 4). The 25 psi contribution from acceleration at C is a small percentage (2.5%) of the total discharge pressure. One exception is when the pump is delivering into a low-friction, high-pressure system such as a short vertical discharge system (Figure 6 (a)). An example is a

Figure 5. Idealized triplex single-acting pump pressure waveform showing the effect of surgelike acceleration pressures

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noise in the workplace. One of the most practical and inexpensive ways to meet OSHA noise level requirements and achieve the overall performance improvement of reciprocating pumps is installation of pulsation dampeners on suction and discharge systems). Dampener Types After evaluating the causes and types of pressure pulsation effects in reciprocating pumps, we can review which available fluid pressure pulsa-tion control equipment is best for specific conditions. Hydro-pneumatic pulsation damp-eners are gasFigure 6. Dynamic differences in discharge systems (a) vertical discharge type energy-absorbing de(b) horizontal discharge sign that uses a gas-filled (c) pressurized systems bladder to absorb the “A” mine dewatering application, where flow on peaks and to give back that the pump is delivering to an already flow on the “lows,” thereby reducing pressurized system—a pressurized flow variations beyond the pump pipeline—through a short connecting and consequently reducing the flowpipe. Another exception is when the induced pressure pulsations. This is pump is delivering to already pressurthe most efficient, low cost and ized systems such as hydraulic press widely used type of pul-sation conaccumulators (Figure 6 (c)). In those trol equipment. two cases, the acceleration pressures This discussion is limited to the can become the overwhelming disturtwo major designs of bladder type bance, particularly if the piping syshydro-pneumatic pulsation dampentem is relatively long compared to a ers: appendage and flow-through. suction system (but considerably shorter than a “pipeline”). Appendage Dampeners Figure 7 shows a typical apControl of Pulsation with pendage design suction pulsation Dampeners dampener. It consists of a homogeSuccessful control of damaging pulnous shell and a nitrogen-filled flexisation requires careful selection of pulble rubber bladder. The fluid port sation dampeners and proper location can be threaded (NPT or SAE) or of those dampeners in a piping sysflanged. This design is recommended tem. The earlier pressure pulsation for installation on pumping systems analysis showed that dampeners will where suction is flooded. The dampreduce the level of pulsation and conener should be installed as close as sequently pressure fluctuations. With possible to the pump inlet, and the lower pipe vibration and noise, wear nitrogen in the bladder should be at the pump, risk of cavitation, and precharged to 60% of the pump’s metal fatigue will be less as well. inlet pressure. Figure 8 shows a large (There are several OSHA standards capacity version of this design. dedicated to the control of industrial The same type of suction pulsa-

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The Pump Handbook Series

tion dampener (150 psi) was installed on a 120 gpm and 30 psi Wheatley triplex pump with 3” piston bore, 3-1/2” stroke. Another two discharge 1,500 psi pulsation dampeners were used with a 263 gpm and 642 psi Wheatley quintuplex pump with 3-3/8” bore, 4-1/2” stroke. Flow-Through Dampeners This design should be installed on a pump’s suction system when inlet pressure (NPSH) is very low and the fluid can contain entrained gases such as air, carbon dioxide, sulphur dioxide, etc. Flow-through units have significant advantages over the appendage design for a few reasons. First, the inlet and outlet diameters of the dampener are equal to the pipe diameter. Second, the in-line installation enables reduction of the inlet fluid velocity into the pump. In addition, the bladder response time is faster than an appendage dampener, due to the minimum distance between the shell inlet and pipe centerline. The flow-through design can also be used to improve the NPSHa of systems with low NPSH because it acts as a suction stabilizer. Figure 9 shows an actual installation of a discharge flow-through 2.5 gallon capacity dampener. This carbon steel, 3,000 psi, 3” flanged unit

Figure 7. Appendage type 60 in3 1,500 psi stainless steel pulsation dampener with threaded fluid port

Figure 8. Typical large capacity appendage type flanged pulsation dampener

was installed on the discharge system of an existing triplex single-acting reciprocating pump with 2.88” piston bore diameter and 5” stroke— one of seven such units. The owner, Ormet Aluminum Mill Products Corporation, is a Hannibal Rolling Mill Division, Hannibal, Ohio. Currently, the company is scheduled to install another two pulsation dampeners of the same type. The installation of pulsation dampeners for the remaining pumps is scheduled for sometime in 2000.

The Abilities of Pulsation Dampeners Hydro-pneumatic bladder pulsation dampeners are able to prevent the generation of the most destructive low-frequency pulses such as those generated by the pump’s rotary motion and the combination of flow from each of the pump’s cylinders. (This is basically rpm multiplied by the number of cylinders.) Accordingly, based on a maximum of 500 rpm for most small pumps, the maximum frequency involved should not be more than about 50 Hz. Performance Comparisons Figure 10 compares the performance of the two types of pulsation dampeners on a typical pump. In this particular case, installation of hydro-pneumatic dampeners offers overall improved performance. An analysis of reciprocating pump pulsations is based on the premise that pressure or pressure variation (pulsation) due to flow velocity (flow- velocity variation) is a function of the square of such velocity. And the system must present enough resistance to flow—due to pipe friction, bends, fittings, chokes—to enable the generation of a pressure or pulsation in the first place. The magnitude of that pressure is directly related to the resistance. The analysis is also based on the premise that the acceleration of the liquid at the beginning of an increase in velocity (start of stroke) generates a pres-

sure pulse, and the magnitude is directly related to the length of pipe connected to the pump suction or discharge. It is also inversely related to the diameter of the pipe and directly related to the rate of change of velocity. Because the inherent flow variations of a reciprocating pump are of the same intensity (but out of phase) in both the suction and discharge, for many years it was believed that a pulsation dampener of the same size as the discharge should be used on the suction, and many such installations can be found. However, from the standpoint of flow-induced pulsations alone, it is logical that a smaller dampener can be used on the suction side. For example, a typical triplex single-acting pump has a flow variation of about 23% in both the suction and the discharge. If all of this flow variation were converted to pressure pulsations, the pulsations would be a maximum of 46% of the average pump pressure. In our example, the maximum change in pressure at the discharge would be about 460 psi at 1,000 psi average pressure. If there is a 50 psi suction pressure, the change in pressure there would be only 23 psi; that is, there would be about 20 times less energy available to set the piping in motion. In addition, the acceleration due to velocity variations is the same on the suction and discharge sides. In the suction system, however, such disturbances usually overwhelm the disturbances due to flow variation, and conversely, in the discharge system the flow-induced pressure pulses usually overwhelm the acceleration pressure pulses. There is evidence that only a flow disturbance or an acceleration disturbance (not both) can exist during any suction or discharge stroke of a pump. One will overwhelm and reduce the effect of the other. A plausible explanation is the theory that in pumps of greater than about 300 fpm piston speed, a disturbance of either type will momentarily reThe Pump Handbook Series

duce the volumetric efficiency to such an extent that it cannot recover fast enough to enable the flow rate to reach an adequate value to affect another disturbance in the short period of time (about 3 ms) between the disturbances. There is no such worry with slow speed pumps because few if any disturbances are generated. It is possible for a pump with high inherent flow variations to actually deliver an almost pressure-pulseless liquid into an open-ended (in effect), short, relatively large-diameter horizontal discharge pipe. Some typical cases are pumps with short connecting pipes feeding a large-diameter already-pressurized system such as an existing pipeline or hydraulic press supply. If the discharge pipe is vertical, as in a mine dewatering system, the acceleration of the liquid column due to gravity becomes the predominant generator of pressure pulses. This great dynamic and hydraulic difference is shown in Figure 6, where the pump “sees” the same discharge pressure under three different physical arrangements. Energy-absorbing or hydro-pneumatic pulsation dampeners may not solve pulsation problems in cases (a) and (c) of Figure 6. Application of a large capacity flange type of surge suppressors should be carefully considered (Figure 8). Advantages of Bladder Type Pulsation Dampeners • positive sealing between the gas and fluid chambers, such as the bladder, serves as a permanent wall between them • zero leakage between both chambers due to bladder barrier • quick pressure response times for pressure control pump pulsation and shock-dampening applications are facilitated by the light weight of the bladder • simplicity of design reduces overall maintenance and operating cost of pumping system

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Figure 9. Flow-through pulsation dampener (Courtesy of Ormet Aluminum Mills Products Corporation, a Hannibal Rolling Mill Division)

Sizing Considerations Suction Pulsation Dampeners and Suction Stabilizers The following sizing procedure will serve two purposes: • evaluation of the pump suction system and subsequent sizing of the pulsation dampener and suction stabilizer with respect to the reduction of pressure fluctuation • improvement of pump suction conditions represented by NPSHa Sample Data Piping • The maximum flow of 200 gpm originates from the storage tank. • Tank is open to atmosphere • Length of the 8” cast iron suction pipe is 106 ft. • Difference of elevation between the storage tank and centerline of pump suction inlet is 23 ft. • Fittings: (1) 8”-90° elbows, (1) 8”45° elbows, (4) 8” Tees

Analysis In order to determinate the Net Positive Suction Head Available (NPSHa) the acceleration head and the friction losses along the suction piping must be calculated. Acceleration head can be calculated as follows: LVnC ha Kg Where: ha = acceleration head, ft. of fluid L = length of suction line, ft V = velocity in the suction line, fps n = pump speed, rpm C = pump type constant, for triplex single acting pump C = 0.066 K = 1.4 for hot water g = 32.2 ft./sec2 Pipe friction loss factor K1 = f

Valve and fitting friction loss factor K2 = (4 valves) x 1.8 + (1-90° elbow) x 1.5 + (1-45° elbow) x 0.8 = 9.50 Total valve and fitting frictional loss factor Kt = 15.81

Fluid • Anaerobically digested municipal sludge conditioned with 10% lime, and 5% ferric chloride. • Fluid temperature = 100°F

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This analysis shows the necessity of installing the suction stabilizer to improve suction conditions. Such stabilizers can reduce the acceleration head, thereby increasing the NPSHa. The unit will also eliminate the acceleration head by minimizing the mass of fluid in the suction pipeline that must be accelerated. The suction stabilizer installed at the pump inlet will reduce the fluid column that must be accelerated to the fluid between the bladder and the pump suction manifold. This fluid mass is normally very small and requires a low-pressure fluctuation. The losses correspond to the acceleration effects and are about 67.5% of the operating pressure. Acceleration losses ha = (0.075) x (23 ft. of w.c.) = 1.73 ft. of w.c. NPSHa for the system with the suction stabilizer is NPSHa = (23.0 + 0) - (1.73 + 6.44) = 14.84 ft. of w.c.

Total pipe frictional losses ht =

Kt V2 (15.81) x (5.12)2 = =6.44 ft. of w.c. 2g 2 x 32.2

Acceleration head ha=

Pump (4 each) • Single action triplex • Max speed = 45 rpm • Flow rate = 200 gpm (800 gpm for all 4 pumps) • Min NPSH = 23 ft. of column of water • Bore = 9.0 in. • Stroke = 6.0 in.

0.020 x 106 L =6.31 = D 0.336

than NPSHr. The actual suction head is not adequate to operate the pump at the 23 ft. of w.c. NPSH requirement.

(106) x (5.12)2 x 45 x 0.066

=34.96 ft. of w.c.

1.4 x 32.2

The maximum static elevation head hs = 23.0 ft. of water column. Net Positive Suction Head Required NPSHr = 9.87 psig @ 45 rpm = 22.8 ft. of w.c., say 23 ft. of w.c. Ptank = 0 psig = 0 ft. of w.c. Net Positive Suction Head Available NPSHa = hs + Ptank - ha + hf = (23 + 0) - (34.96 + 6.44) = - 18.40 ft. of w. c. Since NPSHr = 23.0 ft. of w.c., the NPSHa = 18.40 ft. of w. c. and is less The Pump Handbook Series

As can be seen from the earlier analysis, the installation of a suction stabilizer will effectively reduce the acceleration head and provide a much greater NPSHa to the pump. This will prevent local cavitation and increase pump efficiency. To provide the required Net Positive Suction Head in our example, the sludge conditioning tank should be raised only 8.0 ft. (versus to 18.4 feet) to compensate for the friction losses between the tank and pump suction manifold. The actual volume of pulsation dampener/suction stabilizer can be calculated using the following sizing procedure: Volume = 10 x displaced flow/ revolution Displaced flow = (0.7854 x B2 x stroke x no. pistons x action)/231, gallons

(b)

(c)

(a) Figure 10. (a) no dampener installed (b) appendage type hydro-pneumatic dampener (c) flow-through type dampener Top chart: downstream of dampener Bottom charts: upstream of dampener

Where: B = pump cylinder bore, in. stroke = pump piston stroke, in. no. piston = 1 for simplex, 2 for duplex, 3 for triplex. suction = 1 for single acting pump, 2 for double acting pump. By installing the pump data in the formula, we get the required actual volume of pulsation dampener/suction stabilizer. Volume = 10 x (0.7854 x 92 x 6 x 3 x 1) / 1 x 231 = 49.6 gallons The 50-gallon suction stabilizer installed at the pump inlet will reduce the pressure fluctuation on the inlet pipeline to 1.73 ft. of w.c., thereby providing a NPSHa of 33.24 ft., greater than without a suction stabilizer.

Sizing of Discharge Pulsation Dampeners In order to achieve optimum results from the dampener, its size (capacity) must be properly calculated. The following simple formula can be applied for sizing a pulsation dampener for ether a simplex, duplex or triplex single acting pump. Size of dampener = plunger area (sq. in.) x plunger stroke (in.) x 50 x pump type factor Notes: 1. 50 = pressure factor for about 65% pulsation control 2. pump type factor = 0.25 for duplex single acting pump

Example Determine the size of a pulsation dampener to control pulsations caused by a duplex single acting plunger chemical metering pump handling 50% sodium chloride. Data • plunger diameter (D) = 13⁄4 in. • plunger stroke (L) = 3 in. • max. pressure = 150 psig • pump type factor = 0.25 Size of pulsation dampener = plunger area (sq. in.) x plunger stroke (in.) x 50 x pump type factor =D2/4 x 3 x 50 x 0.25 = 90.2 cu. in. Therefore, a 0.5-gallon size dampener should be selected.

Material and Design Specification Guidelines for Pulsation Dampeners and Suction Stabilizers The next step is proper specification of the dampener’s material of construction with respect to working conditions. The diaphragm or bladder and all other seals in the dampener shall be EPR (Ethylene Propylene) compound compatible with 50% sodium hydroxide (NaOH). The metal parts of the dampener shall be of 316 stainless steel. The selected dampener in this example shall dampen pulsations within 5 to 10% of the mean line pressure.

tion process of pulsation dampeners. 1. Maximum design pressure rating of the dampener must be within the operating pressure limit of the system. 2. Liquid port size of the dampener should be equal to the discharge port of the pump or equal to the discharge line size. 3. Dampener should be charged with a neutral gas (usually nitrogen) to approximately 60% of the mean system pressure. 4. Appendage type dampeners: pipe or nipple connecting the dampener to the line should be as short as possible. The diameter of connecting pipe should not be smaller than the size of the dampener liquid port. 5. The dampener should be installed as close as possible to the pump discharge port. 6. A periodic check of nitrogen gas pressure in the dampener should be made to make sure the dampener is in operable condition. 7. The dampener’s rubber and metal parts must be compatible with the pumped liquid. " V. Larry Beynart is an Application Engineer with Fluid Energy Controls, Inc., a Los Angeles manufacturer of hydro-pneumatic accumulators and pulsation dampeners. He holds a Bachelor of Science Degree in Mechanical Engineering and Hydraulic Certification from Fluid Power Society.

Conclusion The following factors must be thoroughly considered during the selecThe Pump Handbook Series

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POSITIVE DISPLACEMENT PUMPS HANDBOOK

Precision Solutions for Low-Flow Handling When you need less than 100 gallons per hour in increments as small as milliliters, metering pumps are the answer. By Ken Gibson, ProMinent Fluid Controls, Inc. he world of low-flow applications is very different from that of high flow counterparts. It is a very precise field, and, over the years, has experienced many changes and developments. Low-flow handling can be a very critical step in such applications as water treatment and chemical injection. To ensure that it is accomplished well, users must be aware of the process involved, and they must choose the correct tool(s) for the job. A low-flow application normally involves the injection of a small amount of fluid into a process. Since the term “lowflow” can mean different things to people, this article will concentrate on applications that are less than 100 gallons per hour. The injection of process chemicals into applications takes an accurate and precise metering pump. Besides the pump itself, there are certain accessories that should be considered to enhance its performance. This article will explore the area of low-flow applications using metering pumps and how they can be done simply and correctly.

T

Low-Flow Pump Designs There are many low-flow pump types to choose from; they come in all shapes and sizes. Some popular versions are the reciprocating

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1. Plunger 2. Packing 3. Flushing 4. Piston shaft 5. Valves 6. Flushing connector

Figure 1. Packed plunger metering pump

diaphragm, the packed-plunger/piston (Figure 1) and the peristaltic. The reciprocating diaphragm pump can be actuated by one of the following mechanisms: • solenoid • mechanical gear • hydraulic fluid • air Solenoid-driven pumps are some of the most widely used in low-flow situations (Figure 2). Some models can accurately meter flows well below one gallon per hour (gph), while other models can exceed 20 gph. The mechanical gear pump is a higher flow pump and can easily hit flows higher The Pump Handbook Series

than 600 gph. The hydraulicallyactuated pump is capable of high pressures and low flows. Air-driven pumps can be useful, since many applications have pressurized air available. Metering pumps have come a long way in the past ten years, especially the solenoid versions. The better models are microprocessor-based (kind of like having a mini-computer in your pump) and are extremely accurate and very dependable. They can operate manually or by an external source such as a pulse contact or a 4-20 mA analog signal. Fault relays also are common. They notify the

Cutaway view of ProMinent Gamma/L solenoid-driven metering pump 4

5

6

8

9

10

12 2 11 3

7

1

1. 2. 3. 4. 5. 6. 7. 8. 9. 10.

Housing Liquid end Diaphragm Backplate Solenoid Solenoid coil Solenoid axle Armature Cover Stroke adjustment screw 11. Stroke adjustment axle 12. Stroke adjustment knob

Figure 2. Solenoid-driven metering pump

operator when the pump has experienced a problem such as loss of power or low chemical level. Many models have digital displays that show stroking frequency and stroke length. One manufacturer just released a model that displays flow in gallons/hour or liters/hour and gives a totalizing count of the amount of chemical pumped over a period of time. All of these features enable the operator to handle low-flow chemical feed with fewer hassles and more confidence. Applications and Pump Accessories Examples of low flow applications requiring a metering pump include industrial/municipal water and wastewater treatment, cooling towers, boilers, reverse osmosis, swimming pools, car washes, chemical process and laboratories. Industrial customers include food and beverage, pulp and paper, semi-conductor, chemical manufacturers, pharmaceutical and metal finishing/plating. Municipal customers include your local waterworks authority, where drinking water and subsequent wastewater are treated. If you visited these facilities and spent some time walking around, you might be surprised at the number of metering pumps clicking away.

Depending on the application and space allowance, pumps can be mounted on chemical drums, prepackaged and mounted onto plastic or steel skids, or tucked away in areas of the plant where they are forgotten until a flow rate changes or a failure occurs (Figure 3). Regardless of location, the installation should consider flow rates, backpressure and the position of the pump in relation to the material being fed. This is one of the reasons accessories are essential to ensure proper feed and prohibit costly downtime (Figure 4). Foot Valves If you are pulling liquid up from a drum or tank, a foot valve is recommended. A foot valve is a check valve that enables the pump to pull liquid up into the suction line. At the same time, it prevents fluid flow back into the tank once the pump is stopped. This helps keep the suction line primed and makes it easier on the pump when it is turned back on. Injection Valve An injection valve is used on the discharge side of the pump to provide a little backpressure and to inject liquid into a water stream or tank. The injection valve can function like a quill and be secured into a pipe. This enables the liquid to be injected The Pump Handbook Series

directly into a water stream and not dribble down the sides of the pipe, which could cause corrosion. Backpressure Valves In any diaphragm-design positive displacement metering pump, a constant backpressure is recommended. This enables the pump to constantly work against the same pressure and not be subjected to the changing pressures that typically occur in processes. A backpressure or loading valve will do the trick, enhancing metering pump accuracy and repeatability. These valves are normally spring-loaded diaphragm valves that are adjustable up to about 150 psi. The backpressure valve normally is located close to the injection point of the liquid.

Figure 3. Many different arrangements for metering pumps are available. This particular pump is mounted on the tank it is drawing from.

Pressure Relief Valves The pump must be capable of overcoming both the valve and the system pressure. Pressure relief valves also are used in many applications. Similar to the backpressure valve in design, the pressure relief valve has an extra port that enables the pressurized liquid to be diverted in the event that the pump experiences overpressure. This valve protects the user’s investment against the metering pump

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Metering monitor

Cable

Foot valve

Multifunction valve

Injection valve

Figure 4. Some important metering pump accessories

operating outside of its designed pressure rating, such as closing a valve on the discharge side of the pump, also known as a “dead head.” Pulsation Dampeners Another very important accessory for low-flow pumps is a pulsation dampener. This device actually dampens the pulsating flow of the pump, resulting in a more laminar-like flow. It acts much the same way as a surge tank does in a water system, preventing waterhammer, which will cause vibrations that result in damage to equipment, with untimely and avoidable expense.Pulsating flow of a reciprocating metering pump can be equated to waterhammer. A pulsation dampener also is recommended when the discharge line of the pump is extremely long—25 feet or more. Since the pump has to push against all of the fluid contained in the long pipe length, the dampener will hold a certain amount of fluid and then push the fluid through the pipe. This takes a lot of the load off the metering pump by reducing the frictional losses associated with pushing the column of fluid through the discharge piping. For installation purposes, the pulsation dampener should be close to the discharge of the pump, followed by the pressure relief valve.

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Multi-Function Valves Another helpful accessory is the multi-function valve. It is a combination valve that has the properties of backpressure, antisiphon and pressure relief. The valves are normally installed directly on the discharge valve of the pump.

Metering Monitors A metering monitor is an accessory that enables the operator to see if there is flow going through the pump. The monitor runs off the pump’s electrical power and detects each pump stroke. The monitor can be programmed to shut the pump off if there is no fluid being transferred. Pressure Gauges A simple device, often overlooked but very important in the information that it provides, is a pressure gauge installed on the discharge side of a metering pump. Pressure gauges are standard operating tools in hydraulic processes under pressure. A metering pump application is a hydraulic system under pressure. The gauge will assist the operator in adjusting to, and determining the loading set points of both the pressure relief and backpressure valves. It will also give an indication as to whether the metering pump is operating within its designed maximum pressure specification.

Photo 1. These ProMinent Fluid Controls metering pumps are injecting phosphate at a water treatment facility. The Pump Handbook Series

3-way valve (for self-fill calibration column) Isolation valve (for direction flow) Backpressure valve SCR drive Pressure gauge Pulsation dampener Calibration column Pressure relief valve

Motorized diaphragm metering pump Y-strainer

Figure 5. Pump manufacturers now have pre-packaged pump systems with a backup pump.

3-way valve (for self-fill calibration column) Backpressure valve

Pulsation dampener Pressure gauge Calibration column Pressure relief valve Solenoid diaphragm metering pump

Figure 6. A complete metering pump system package

Calibration Column One additional tool, which will assist the operator in providing precise outputs from metering equipment in low-flow applications, is the calibration column. This graduated column, scaled in milliliters or gallons, allows

for determination of metering pump output under operating conditions. It also enables precise adjustments to pump output in a safe and reliable manner. Some or all of these accessories may be needed to ensure precision and The Pump Handbook Series

accuracy in low-flow situations. It is best to check with the pump manufacturer to select the proper accessories for your metering application. There are also software programs available that can help you select the proper accessories.

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Maintenance and Operation In order to keep your low-flow metering application running smoothly, you should set up a maintenance schedule. Most diaphragm metering pumps are low maintenance to begin with. They are normally durable and consistent if they are treated properly and maintained. For solenoid pumps, there should be a routine check on the suction and discharge valves, as well as the diaphragm. If the valves are worked hard, they may need to be replaced due to the constant pounding of the valve balls on the ball seats. Also, O-rings become worn and some are subject to chemical attack of the liquid being metered. A good durable diaphragm will have a steel core center encased in Teflon® and an elastomer backing. This type of diaphragm can last for years if the pump itself is properly maintained. If the pump needs to be repaired, the cost is usually minimal. In a situation where a pump goes down, it is a good practice to have a spare one ready to go. If a spare is not available, some manufacturers have a “loaner program” where they will loan the operator a pump while the broken pump is being fixed at the factory. Still, other end-users prefer to throw the broken pump away and purchase a new one. This practice is a waste of time and effort that could be prevented if a better-quality pump were being used. Most metering pumps have a standard one to two year warranty that covers most parts and labor—as long as the pump is not abused. Normally the pump is sent back to the manufacturer for evaluation. Before shipping a used pump, it is extremely important to flush the liquid end thoroughly for safety reasons. After a thorough check, it then can be determined if the problem was due to manufacturer’s defect, chemical attack or abuse. The best scenario is to have a backup pump or system. Some pump manufacturers now offer pre-packaged pump systems with a backup pump (Figure 5). If the

198

main pump goes down for any reason, the backup pump will automatically start up and continue where the other one left off. This is a wise choice when metering a chemical into a critical process such as sodium hypochlorite for disinfection, or hydrofluosilicic acid for fluoridation in drinking water. Fluoridation is an excellent example of the need for a precision metering pump in a low-flow application. In high concentrations, fluoride ingestion is toxic to one’s health. In low concentrations in drinking water (11.5 mg/liter), fluoride is effective in minimizing tooth decay. The packaged systems can be simple—such as a pump and calibration column, or with most of the accessories mentioned previously. A complete metering pump system package will offer the greatest opportunity for providing safety, accuracy, and reliability (Figure 6).

Metering Goes High-Tech Technology has certainly helped the metering pump industry and subsequently, low-flow applications. Solenoid pumps now come equipped with microprocessors that control all electrical functions in the pump. Control options such as analog (4-20 mA typically), fault relays, timers, flow monitoring and batch counters enable the operator to set up an entire feed application with just a keystroke. New technology has redefined the features and benefits of the metering pump, saving time and providing vital information and precision pumping. Motor-driven pumps have also jumped on the technology train. Some models come with built-in microprocessors, enabling some of the same functions found in the solenoid models such as analog control and fault relays. Pumps are now smarter and better than they were in past years. While some end-users still prefer the old technology, there are new ways of handling low-flow applications—specifically with metering pumps and innovative technology. The Pump Handbook Series

With low-flow situations increasing (better control—lower costs), we now have a better way to handle them and keep all parties happy. ■ Kenneth J. Gibson is the Industrial Market Manager for ProMinent Fluid Controls, Inc. He has been with the company five years and also conducts product training for ProMinent authorized representatives. Prior to this position, Ken was an Equipment Product Manager for Calgon Corporation, where he worked for 20 years. Ken holds a B.S. degree in Chemistry from the University of Pittsburgh.

POSITIVE DISPLACEMENT PUMPS HANDBOOK

Metering Pump System Design for Dependable Performance Accessories such as backpressure and check valves, standpipes and flow controllers can make your metering system run smoothly, but only if you know how to use them correctly. By Ed Warwick, Flow Control Division, Milton Roy Co.

What Is a Metering Pump? Most metering pumps are used to inject a relatively small flow rate of a liquid chemical additive into a larger process, such as adding: • acid or caustic to control process pH, • corrosion inhibitor downhole to protect expensive well casing, • sodium hypochlorite into water to control bacteria, • alum or polymer into wastewater to aid removal of solids, • oxygen scavenger into boiler feedwater to control corrosion, • fertilizers and pesticides into irrigation water for better crop yield. The list continues considerably. The value of metering pumps is that they pump a predictable, adjustable flow rate that remains stable even when system pressure conditions change. The flow from a centrifugal pump changes significantly with varying discharge pressures; flow rate from a metering pump remains quite constant. Metering pumps come in many shapes and sizes. A gear pump is a metering pump. Vane pumps, progressing cavity pumps, peristaltic (“hose”) pump and reciprocating plunger/diaphragm pumps are as well. The smallest pumps produce flow rates as small as a few milliliters per hour, while larger ones produce

as much as 3,000 gallons per hour and more. Operating pressures from 0 to 15,000 psi are common. Although most metering pumps are driven by electric motors, air motors, solenoids or a reciprocating air cylinder drive others. Drivers range in power from sub-fractional to 25 hp. This article is primarily concerned with reciprocating plunger/diaphragm Figure 1. Metering pump construction metering pumps. connected to tubing or pipe to deliver Metering Pump Construction the liquid chemical to the process (Figures 1 and 2) injection point. Imagine a motor and gearbox Metering Pump Operation (usually referred to as the pump “drive Assume that the piping and liquid end”) driving a crank arm to cause a ends are filled with the chemical. On plunger (like an automobile piston) a suction stroke, the plunger is drawn to move back and forth (reciprocate). by the drive to the right. As the The plunger reciprocates inside of, plunger moves, liquid flows in to fill and is part of, the second major the void left by the receding plunger. component of a metering pump, the Liquid tries to flow in from all direc“liquid end.” Also part of the liquid tions, but the discharge check valve end are the inlet and outlet check blocks it. It does, however, lift the valves, which are usually ball-style suction check valve ball and flows in valves (rather than flapper or poppet from the suction piping to follow the style). These check valves allow flow plunger back and fill the pump cavity. in only one direction and block (check) After reaching its back dead center flow in the other. The inlet to the position, the plunger is driven to the pump (the suction connection) is left on a discharge stroke. It now connected with pipe or tubing to a pushes into the liquid in the pump supply tank of liquid chemical. The cavity. (This cavity is called the outlet (discharge connection) is The Pump Handbook Series

199

electronic (4–20 milliamp) or pneumatic (3–15 psi) control signal. The second primary method of controlling flow rate is to use an adjustable speed motor to drive the pump. This changes the strokes per minute and thus the pump flow rate. Variable speed motor systems also can be automated to respond to an electronic control signal. If the pump is to be operated within the upper 50% of its rated flow Figure 2. Elementary liquid end of a metering pump range, then either of the above two methods of flow displacement chamber.) The displaced rate control is usually suitable, reliable liquid tries to flow out in all directions, and effective. If, however, the pump but it is checked from flowing back will be operated at less than 50% out from where it came by the suction capacity or will be pumping “gassy” check valve. The discharge check liquids such as sodium hypochlorite, valve’s ball is lifted from its seat as then variable speed control is generdisplaced chemical flows out the ally preferred. This is due to the liquid discharge connection into the disend’s greater susceptibility to getting charge piping. “air bound” when operated in the Note that during the suction stroke shorter stroke length region. there is no flow in the discharge pipe, Liquid End Designs and that during the discharge stroke The liquid end is the portion of the there is no flow in the suction pipe. pump that carries the reciprocating This alternating flow is described as plunger and is in direct contact with pulsing flow, and it is responsible for the liquid chemical. Liquid ends are some of the piping system components available in three basic designs: necessary for reciprocating plunger packed plunger, mechanically metering pumps, which are not actuated diaphragm and hydraulically needed in the smooth, continuous actuated diaphragms. To generalize: flow generated by centrifugal pumps. • Packed plungers find most use in Controlling Flow Rate pressure systems above 3,000 psi. There are two primary means to They are simple to understand and control the flow rate of these pumps. maintain, but usually require Nearly all designs have a means of weekly inspection and packing adjusting the length of the reciproadjustment. The plungers also allow cating stroke, usually a handknob for some chemical leakage to the manual adjustment. This changes the outside environment. volumetric displacement of each • Mechanically actuated diaphragms stroke and therefore the pump flow are uncomplicated, economical and rate. Typically, this method can be allow no chemical leakage. They are used to adjust flow rate from usually used in applications below maximum down to 1/10 of maximum. 150 psi, and the diaphragm should It can be automated by adding an be replaced annually. electric or pneumatic actuator that will • Hydraulically actuated diaphragms adjust stroke length in response to an are the most expensive and complex

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The Pump Handbook Series

choice, but they allow zero chemical leakage. The diaphragm has a long life—often well over five years, some over 15 years of operation— and they can be used with pressures greater than 10,000 psi.

Characteristics of Pulse Flow Pulse flow is generated in both the suction and the discharge piping. There are two characteristics of pulse flow that need to be dealt with. Peak Momentary Flow The peak momentary flow rate generated every pump stroke is more than three times the average flow through the pipeline. The reason for this is that the plunger does not just “slap” back and forth on suction and discharge strokes. It acts more like a train pulling out of one station for a short trip to another station. Starting “at the station” at zero speed (therefore producing zero flow rate), it begins to move and pick up speed, reaching maximum speed midway. Then it slows down as it approaches the destination station to come to a full stop. It reverses travel to complete the suction/discharge stroke cycle. As noted in “Metering Pump Operation,” on page 45, while discharging into the discharge pipe, there is no flow in the suction pipe and vice versa. So now, this “3x peak flow” begins to make sense. If the pump is producing flow only half the time, it stands to reason that when it is producing flow, it has to produce at least twice the average. In practice, at the beginning and end of each stroke, it isn’t even up to average for a short time. Therefore, when all is taken into account, the peak flow has to be more than twice as high as average flow; it is actually about 3.14(π) times higher (Figure 3). Starting and Stopping The second characteristic is that, with all this starting and stopping of flow in the pipe with each stroke of the pump, the liquid is subject to accelerations at the beginning and end of

Figure 3. Metering pump system accessories

each suction and discharge stroke. Accelerations are always associated with forces, and the pumped chemical exerts its forces on the pump and piping during each stroke. An example of this is “waterhammer.” With a metering pump, this waterhammer happens every stroke of the pump.

Minimizing the Effects Discharge System Design (Figure 4) The design and selection of discharge piping and components affect accuracy, stability of flow rate and the degree of pressure and flow pulsations. They also can affect protection of personnel and the system from harmful high pressure. Backpressure Valve See Figure 4 for the location of the backpressure valve. These metering pumps require that the pressure at the discharge of the pump be greater than pressure at the inlet of the pump. There are generally two reasons for this. If sufficient discharge pressure is not present, liquid can flow freely through the pump in an uncontrolled manner, and flow rate accuracy is lost. Why this happens can be seen in

Figure 2. Notice that both the inlet and discharge check valves enable flow in the same direction. If the pump is discharging into an open stream (low discharge pressure) and the chemical supply tank level is at a higher elevation, then the chemical will “siphon,” or “drain” through the pump. There could be a flow rate even if the pump is not running. This is uncontrolled flow. If the pump design is a “hydraulic lost motion” diaphragm style, it has to internally bleed or bypass some of the hydraulic oil displaced by the plunger every stroke to operate at any capacity setting less than 100%. The motive effort to cause the required bypass comes from discharge pressure. So, if there is not enough discharge pressure, no bypass will happen and the pump will continue to pump 100% flow even though it is at a lesser setting. Under this condition it is not unusual for the pump to pump 100% flow rate even when set to 0%. Most backpressure valves offered by metering pump manufacturers contain a diaphragm that isolates some of the internal valve parts from the corrosive chemical. These designs usually should be operated with no The Pump Handbook Series

pressure on their outlet. Their principle of operation is a balancing of hydraulic pressure force acting on a small seat area against a spring force. Pressure on the outlet acts on the diaphragm, which presents a much larger area for the hydraulic pressure to act on, therefore downstream pressure prematurely lifts the spring and enables chemical to flow through at a much lower pressure. If backpressure is needed and there is downstream pressure, use an inline backpressure valve. This usually has its spring in contact with the chemical. Exactly how much greater the discharge pressure should be over the suction pressure is not easy to determine with much precision and it is usually not too important. In general, if discharge pressure is more than 20 psi higher than suction pressure, there will be good control of the flow rate. Unless there is some overriding consideration, though, 60 to 100 psi is a more comfortable value. Pressure Relief Valves First, a disclaimer. In many applications, a pressure relief valve provides an important personnel safety function. Comply with applicable codes or known practices concerning your use of safety valves. See Figure 4 for the location of a system pressure relief valve and follow these important guidelines: 1. Realize that these metering pumps are capable of generating thousands of psi pressure against a deadhead system within just a few seconds. 2. Set the relief pressure for 1.15 times the maximum expected operating pressure, or 25 psi above the maximum operating pressure, whichever is greater. Make sure all other system components are OK to operate at pressure greater than this set pressure. 3. Locate the valve early in the discharge line. It should be located between the pump and any pressure or flow restricting elements

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such as backpressure valves, strainers or shutoff valves. Also, locate it downstream of a pulsation dampener if one is used. 4. Most hydraulically actuated diaphragm metering pumps contain an integral hydraulic oil pressure relief valve. Many users utilize this as a system relief valve. Using an external relief valve located in the discharge piping in addition to the integral relief valve provides more system protection. Realize that the internal pump relief valve will limit the maximum pressure created by the pump itself, but will not relieve system pressure caused by other elements such as a reactor vessel, extruder or boiler. These systems must have an external relief valve for safe operation. 5. If a relief valve design is used that contains a diaphragm to isolate the pressure-setting spring from the pumped chemical, the valve discharge must be free of any pressure. Even pressure caused by a few feet of liquid head may cause valve chattering when activated, or cause the actual actuating pressure to be significantly less than the pressure originally set. Avoid piping the relief valve discharge back to the pump suction. The pump suction line almost always has pressure pulsations. Pulsation Dampeners A pulsation dampener smoothes out pulsating flow from the pump. In doing so, the pressure and force pulsations exerted on the piping are minimized. Waterhammer and discharge pipes jumping around are eliminated. Also, if a flowmeter is used in the discharge line, its output results will be more stable and accurate. (As a point of interest, new flowmeter designs are now available with software that integrates each discharge flow pulse and gives flow readout accuracy in the 1% range.) When installing a pulsation dampener, follow the rules below:

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1. Locate it near the pump discharge connection. It only dampens flow downstream of itself. It does not dampen flow between itself and the pump. 2. Install a pressure gauge downstream of the dampener. 3. Size the connect pipe to be equal to or larger than the connection size to the dampener, and make it straight. Don’t use elbows. Small pipes and elbows reduce the good effect of the dampener. 4. Dampeners need proper gas precharge to operate. a. Dry nitrogen is the preferred charging gas because it doesn’t cause rust of the metal components or react to deteriorate the rubber internals. Most nitrogen sources are bottled at 2,000 psi, and require knowledge and equipment in good repair to use safely. To avoid this, many users of low pressure pumping systems use air from shop compressors or even a bicycle pump. Air contains moisture that will reduce the overall life of the dampener, but some users accept that tradeoff. b. Precharging is a very misunderstood procedure and is either not done at all, or is done incorrectly in probably 75% of installations. First, determine the lowest expected operating pressure of the system. Then— with no (zero) chemical pressure in the pipe—pressurize the dampener to about 80-85% of that lowest pressure. If that doesn’t work, then—with the system operating and generating flow and pressure pulses— pressurize the dampener to a pressure more than what is showing in the system. After that, slowly bleed gas out of the dampener until pressure pulses are minimized. 5. The size of the dampener affects how well it works. Usually, “bigger is better.” A more rational way to properly size the unit, however, is The Pump Handbook Series

to have the total internal volume of the dampener equal to the volume of 15 or more pump strokes. This yields residual pulsations of 5% or less. Of course, the dampener must be rated to operate at system pressure and temperature, and it must be resistant to corrosion from the process chemical. Pipe Size Pipe size on the discharge side of the pump may be smaller than that on the suction side. If a pulsation dampener is used, a pipe 1⁄2 to 3⁄4 the size of the pump outlet connection is usually quite adequate. This has potential to take advantage of lower cost of smaller pipe, even tubing. If, however, a pulsation dampener is not used, pipe size should be at least the same size as the pump connection or even larger.

Designing the Suction System The design and selection of suction piping and its components affect the accuracy and stability of flow rate and the degree of pressure and flow pulsations experienced. Suction Piping Of all of the metering pump system components, selection of suction piping has the greatest effect on successful pump operation. To select pipe size using only a “rule of thumb” approach, use pipe that is a minimum of one pipe size larger than the pump inlet connection. (For example, if the pump inlet is 11⁄2”, use 2” or larger pipe.) This approach is not recommended, but we know that probably 90% of installations are designed and built without further analysis. Probably 25% of these have unsuccessful startup and need reworking. A more rational and economical approach is to do an NPSH analysis. This analysis must recognize the two characteristics of metering pumps discussed above. First, peak flow rate each stroke—and therefore peak pressure drop—is more than three times the

Characteristic

Effect

Solution

Peak flow in the suction line every stroke is 3x the expected average flow.

Higher friction pressure drop than expected “starves” the pump inlet and results in low flow.

Increase suction pipe diameter. Make it one or more sizes larger than the pump inlet connection size.

Acceleration in suction line every stroke.

Unexpected acceleration pressure drop results in starved inlet and low flow.

Same as for peak flow, above.

Acceleration in discharge pipe every stroke.

Excessive water hammer makes pipes jump around.

Same as for peak flow, above.

Pump flow rate is low because relief valve in the pump or in the discharge piping is relieving for no apparent reason.

Add a pulsation dampener to discharge line. Volume should be equal to or greater than 15 volumes of one pump stroke.

Flow rate is significantly greater than expected and does not change when % stroke of pump is adjusted.

Discharge point is lower than chemical liquid supply level. Liquid is “siphoning” through. Add a backpressure valve in the discharge line.

Pump check valves enable flow in the same direction.

Shorten inlet piping. Move pump and tank closer together.

Add a standpipe (with a diameter greater than pump plunger diameter) to the suction line.

Pulse acceleration is causing unseen significant pressure pulses. Add a pulsation dampener as above.

Inertia effect of the pulse flow acceleration is carrying the flow through the pump at end of stroke. This is called “flow-on.” Add a backpressure valve to the discharge line, or add more backpressure.

Table 1. Flow characteristics of metering pumps: their effects and remedies

Figure 4. Metering pump discharge system The Pump Handbook Series

203

average flow rate. Second, pulsing flow of metering pumps causes another type of pressure drop in addition to the frictional pressure drop, acceleration pressure drop. Friction is affected by pipe inside diameter raised to the fourth power (increasing pipe diameter by a factor of two will decrease pressure drop by a factor of eight). Acceleration is affected by pipe diameter squared. Realizing this should make anyone a believer in using the largest pipe size possible. Standpipe The standpipe is an ideal component. It requires no calibration and no maintenance, and it is supremely effective. A standpipe installed in a suction system eliminates flow pulsations and therefore eliminates acceleration pressure drop, and reduces frictional pressure drop by a factor of three. It has the effect of moving the main supply tank close to the pump. It eliminates pulse flow only between itself and the supply tank, so locate it as close to the pump as possible. To prevent overflow and spillage from the top, it should be capped and vented back to the headspace of the supply tank. The top of the standpipe must be full diameter all the way up past

204

the highest liquid level expected in the tank. It will not work if liquid is in the smaller diameter vent tube. It can be used in most systems, but it will not work where the supply tank level is lower than the pump centerline. If it is to be applied in a system, using chemical with more than 50 centipoise viscosity, a pressure drop analysis should be made to ensure that it will not “suck air.” Calibration Column A calibration column is used to measure pump flow rate. It is essentially a small vessel with volume calibration marks. It provides an inexpensive, accurate way to periodically determine pump flow rate. To use it with the pump running, valves are opened to let supply tank head pressure fill the column with chemical. When the valve to the main supply tank is closed, the only source for the pump to draw from is the column. The liquid level will fall with each suction stroke of the pump. Begin timing at a convenient starting volume mark, and stop timing at another mark. Now, a known volume has been pumped over a known period of time. A math calculation will reveal the pump flow rate. The column should be sized to hold at least a one-minute supply of liquid for

The Pump Handbook Series

the pump set at maximum flow rate. Flow rate can be measured at startup as well and recorded at various pump flow settings (such as 100%, 75%, 50% and 25% stroke settings). Plotting the results will let the user know which settings are required to get a given flow rate. It also can be an important tool for troubleshooting system problems. Columns are available in designs ranging from armored glass to simple clear plastic.

Summary A metering pump is always used with additional system components. Component selection always should be made with an understanding of the unique operating characteristics of metering pumps. Starting with piping, and perhaps continuing with pulsation dampener, pressure relief valve, backpressure valve, standpipe and calibration column, good selection will ensure reliable operation. ■ Ed Warwick is a Technical Support Engineer for the Flow Control Division of Milton Roy Company in Ivyland, PA. He is a graduate mechanical engineer from the University of Florida and has worked in various design application engineering positions for the company since 1965.

POSITIVE DISPLACEMENT PUMPS HANDBOOK

Twin-Screw Pumps vs. Centrifugal and Reciprocating Pumps By Stephen Smith, Screw Pump Specialist, Flowserve Corporation

In this article, the twin-screw pump will be explored and compared to centrifugal and reciprocating pumps. Primary areas of explanation will include development of energy, limitation of energy generation, effect of viscosity and flow characteristics.

The Centrifugal Pump The centrifugal pump is the most widely used pump in industry, is well understood by users and is accepted for most applications. The pump is typically low in price and available in a wide selection of sizes and configurations. Centrifugal pumps generate energy by accelerating the pumped fluid to high speed through the

Figure 1. Centrifugal pump head

spinning impeller and converting the velocity energy to pressure energy. As a result, the pumps operate at high speeds, which causes high fluid shear rates. Due to the generated head being a function of the energy imparted to the fluid by the spinning impeller, centrifugal pumps produce dynamic head, which is related to pressure as a function of the fluid’s specific gravity. This dynamic head can be expressed as the force exerted on the bottom of a column of fluid in terms of meters of fluid. The head produced is a combination of impeller tip speed and pump design, and it is generally not a function of the pumped fluid, especially with thin fluids. For

instance, if a pump produces a head of 100 meters, it will produce 100 meters of head regardless of the fluid pumped (Figure 1). This is a general statement and it must be noted that fluid viscosity and the presence of solids will affect the total dynamic head. As the flow rate increases within the centrifugal pump impeller and case, there are additional energy losses such as eddy currents and friction that decrease the produced head. This results in the familiar shape of the typical centrifugal pump head/capacity curve (Figure 2). Due to the high fluid shear rates, centrifugal pumps, as compared to

Figure 2. Centrifugal pump curve The Pump Handbook Series

205

twin-screw pumps, can become Both high axial and radial loads can extremely inefficient when pumping be present and must be considered fluids of higher viscosities. Figure 3 when making final bearing type and illustrates a typical effect of viscossize selections. Typical anti-friction ity on efficiency of a centrifugal pump bearing design lives are 20,000 to vs. a twin-screw pump. In an appli100,000 L10 hours with lubrication by either grease or oil splash. cation where a centrifugal pump Advantages of centrifugal pumps requires 1000 kW, a twin-screw pump include wide acceptance for low may only require 670 kW. Based on viscosity applications, the ability to the power difference energy curves in handle various inlet conditions, operaFigure 4, the centrifugal pump would tion at induction motor speeds, high require an additional annual energy temperature capability and being well cost of $100,000 for 8,000 hours/year understood by the users. The pump at $0.04/kW/hr. can handle a wide range of non-lubriSingle-stage overhung centrifugal cating, lubricating, abrasive and corropumps in low pressure applications sive fluids. At low viscosities, the are typically lower in price than twinpump exhibits high mechanical screw pumps. At pressures above efficiency. Installation and mainteapproximately 10-14 Bar, the pumps nance costs are typically low. Disadmust be manufactured in multiple vantages include low efficiency at high stages, which begins to rapidly approach the price of positive displacement pumps. For that reason, single-stage centrifugal pumps are typically used at pressures below 14 Bar in low viscosity fluid applications. The reduction in efficiency at even modest viscosities of 80-100 cSt requires alternative pumping technology be considered for any application, regardless of the pressure. Centrifugal pumps are not selfpriming, and if gas is present in the fluid the pump can fill with air/gas and quit pumping (vapor lock). Centrifugal pumps have Figure 3. Typical efficiency curves relatively high NPSH requirements as the flow increases and the pump approaches the best efficiency point (BEP). NPSH requirements range from the capability to handle vacuums to requiring positive pressures and are a function of overall pump design and size. Common centrifugal pump materials of construction include steel or cast iron bearing frames, steel, stainless steel, iron or NiResist wetted parts, depending on the application. In general, most centrifugal pumps use anti-friction bearings. Figure 4. Power difference energy cost

206

The Pump Handbook Series

viscosities, high fluid shear rates, limited pressure capability and vapor locking at moderate gas content or when suction pipe leaks are present.

The Reciprocating Pump The reciprocating pump is a constantspeed, constant-torque pump with a constant output and is available in piston and plunger versions. Piston pumps are typically limited to flows of 100-120 m3/hr and pressures of 70 Bar or less with plunger pumps covering the 70 to 2,070 Bar range. This article is restricted to the piston pump, since it more suitably covers the same hydraulic range as the twin-screw pump. A series of pistons are driven through an eccentric crankshaft with output determined by overall piston size, piston speed and number of pistons. Pump speeds are typically low, in the severalhundred rpm range. The pump generates energy by pushing a slug of fluid into the pump discharge piping system at whatever pressure is required to perform that action. As the piston retracts in the bore, the pump chamber volume increases, which creates a low-pressure region and opens the inlet valve and enables fluid to enter the chamber. At the end of the suction stroke, the low pressure region ceases to exist and the inlet valve closes. The piston then begins the discharge stroke, which reduces the discharge chamber volume and forces the fluid to exit through the outlet valve and into the piping system. The ability to generate pressure is only a function of the structural capacity of the pump and driver power. If system pressures become too high, catastrophic failure of the pump, piping or drive components can occur. For that reason, all piston pumps

Figure 5. Piston pump cross section

should have a relief valve installed in the discharge piping to prevent overpressurization. Volumetric efficiency is a function of valve slippage, packing leakage and (in some instances) fluid compressibility. Slip flow occurs within the pump as a result of backflow through the valves as they close. Viscosity, pump speed and pressure can also have an effect. Typical slipped flow ranges from 2 to 10% of total flow. Overall mechanical efficiencies are relatively high, in the 80–90% range, depending on design and pressure differential. The resulting speed Figure 6. Typical piston pump curve capacity curve is essentially a fixed flow at a fixed speed (Figure 6). Due to the start/stop operation of the piston, there are extensive variations in the flow rate, which results in pressure pulsations. This variation in flow can be reduced by increasing the number of pistons. For that reason, piston pumps are typically manufactured in single chamber and multiple parallel chambers of two through seven chambers or more. A common arrangement found in oil fields is the three-chamber or triplex pump. Flow variations for a three-chamber pump Figure 7. Triplex pump flow pulsations The Pump Handbook Series

are shown in Figure 7. Table 1 shows the effects of piston number vs. flow variations. These flow variations can result in pressure pulsations so severe that pulsation dampeners (accumulators) are required to smooth the pulses to a level at which other system components, such as piping, valves and structural components, are not adversely effected. Fatigue failure of these critical components is a major concern when subjected to the cyclic loading of the pressure pulsations. NPSH requirements are generally similar to other pumps, but an acceleration head factor must be included to ensure that the fluid is capable of filling the bore as the piston retracts or pump knocking will result. In general, positive inlet pressures are required. Common materials of construction include steel frames, steel, iron or bronze pistons and NiResist cylinder liners. Pistons and liners can also employ various coatings such as chrome or ceramics to improve abrasion resistance. Piston pumps are designed in both sleeve and anti-friction bearing configurations. Sleeve bearings should have infinite life when properly installed and lubricated. However, sleeve bearings are designed to operate within a limited speed

207

Effect Of Piston Numbers To Flow Pulsations Type Duplex Triplex Quaduplex Quintuplex

Number of Pistons 2 3 4 5

% Above Mean 24 6 11 2

% Below Mean 22 17 22 5

Total % 46 23 33 7

Table 1. Effect of piston numbers to flow pulsations

range. Operation outside of this range (either higher or lower in rpm) can result in loss of the lubrication film and bearing failure. To start a pump with sleeve bearings, the pump must be first brought up to speed through a bypass line under no pressure load. Anti-friction bearing pumps can be started without the need for a bypass line, and typical bearing design lives are 30,000 to 50,000 L10 hours. Antifriction bearing lubrication is typically by splash if sufficient pump rpm is present. At low speed, a force feed lubrication system might be required to ensure proper bearing lubrication. Advantages of the piston pump include being relatively well understood by the users, high mechanical efficiency and large acceptance by the oil industry. Disadvantages include pulsation levels, resulting vibration problems and low operating speed, which requires gear box or V-Belt drives. Disadvantages also can include low-flow capability, depending on

requirements. The installed cost can be high due to the need for multiple pumps, speed reducers and accumulators. Repair cost can be high and the pump can wear excessively when exposed to non-lubricating fluids (water) and abrasive slurries such as sandy crude oil. Leakage of pumpage past the packing rings on the piston rod also poses environmental concerns.

The Twin-Screw Pump The twin-screw pump, also called the two-screw pump, is a doubleended positive displacement rotary pump. Both product-lubricated and oil-lubricated gear and bearings versions are available (Figure 8). Some manufacturers lubricate the non-gear end bearings with grease. The pump incorporates either a one-piece integral body with the screw bores machined as part of the body or an external case with the screw bores machined in a replaceable liner.

Figure 8. Product and oil lubricated twin-screw pump configurations

208

The Pump Handbook Series

Normal pumping action splits the incoming suction flow at the inlet flange into two equal portions that are directed to the inlet end of the screws. Transfer chambers form as the meshed screw set rotates, which conveys the pumped fluid axially to the discharge chamber in the center of the pump. Special high viscosity pump designs exist with the fluid entering the center of the screw set and discharging at the screw ends. In doubled ended screw pumps, the axial hydraulic forces cancel out. If used, helical timing gears generate axial shaft forces that must be carried by the axial locator (ball or spherical roller) bearings. Hydraulic forces within the meshed screw set generate radial forces that bend the screws toward the bore. These forces are a function of pressure, screw pitch, shaft strength, bearing span and to a minimum degree, the screw profile. Shaft bending direction is predictable, and the direction depends on the

orientation of the bores (Figure 9). The amount of bending is a direct function of pressure, screw pitch and shaft material modulus of elasticity, shaft diameter to the fourth power and bearing span to the third power. The most important of these is the bearing span, and for this reason, manufac-

Figure 9. Shaft bending direction

turers offer short and long bearing bracket pump versions (Figure 10). The long-bearing pumps have longer bearing brackets, which enable operation at higher pumpage temperature and installation of exotic sealing arrangements such as double mechanical seals or packing. In short-bearing pumps, the bearing brackets are shorter, which results in the bearing being held close to the pump body to minimize shaft bending. Short-bearing pumps are typically used for lower temperature applications that do not require exotic mechanical seals. To maximize

pressure capability prior to the screws contacting the bore, manufacturers increase the distance between the screws and bore by machining extra metal (scalloping) out of one side of the bores or offsetting the screws to one side of the bores to give extra room for the screws to bend. Scalloped and offset bore and screw arrangements are shown in Figure 11. To increase shaft overall strength and minimize bending, manufacturers have begun to change from shell or pinned screws to integral (one-piece) screws. The integral screw is stronger, as multiple components are not involved with tolerances and fits, and the overall shaft diameters at the seal areas can be maximized (Figure 12). The advantages of the shell design, including the ability to mix and match materials, are minimal due to increases

Figure 10. Long and short bearing designs

Figure 11. Scalloped bores and offset screws

Figure 12. Pinned vs. integral shafts The Pump Handbook Series

209

Figure 13. Typical twin-screw pump curve

in material technologies. The shell screws have also proven next to impossible to repair in the field due to difficulties in controlling runout and thread start locations. Therefore, virtually all repaired pinned screws are treated as integral screws during the repair process or are returned to the OEM to have new pinned screws made and installed on the shaft. The twin-screw pump typically operates at induction motor speeds. The pump generates energy by trapping a pocket of fluid and transferring it from the suction chamber to the discharge chamber. As the fluid pocket traverses down the bore, the pressure increases linearly until discharge pressure is obtained in the discharge chamber. Due to internal clearances within the pump, some flow slips back from the higher pressure discharge chamber to the suction chamber. This slipped flow is a function of internal pump clearances, differential pressure, number of screw turns, viscosity and pump speed, which results in a straight line performance curve (Figure 13). The ability to generate pressure is a function of volumetric efficiency (which should not drop below 50%) and the structural capacity of the pump. As with the reciprocating pumps, a relief valve should be installed in the discharge piping to prevent over-pressurization. Typical

210

Figure 14. Typical short-bore/short-bearing design

maximum flows for twin-screw pumps are as high 3,400 m3/hr with pressures to 100 Bar. NPSH requirements are a function of pump design, pump size, speed, screw pitch and viscosity. NPSH requirements are typically significantly less than centrifugal pumps and may be one of the primary reasons for selection of this pumping technology. Pump efficiencies are typically high but decrease as viscosities increase due to increased frictional horsepower. However, efficiencies do not decrease as rapidly as centrifugal pumps (Figure 3). By design, there is no metal-tometal contact inside the pump. However, the screws may be designed to run in boundary-layer lubrication conditions against the bores, as described elsewhere in this article. Torque is transferred from the drive shaft to the driven shaft through a timing gear set located either in the product or an isolated cooling and lubricating chamber with the bearings. These gears synchronize the rotation of the pumping screws and hold the meshing clearance so no internal metallic contact occurs between the screws. The externally housed gears and bearings typically require no additional systems for cooling and lubricating. In some cases, high horsepower pumps may require auxiliary bearing and gear oil cooling systems. The Pump Handbook Series

There are two different fundamental design variations of the twinscrew pump that can be categorized based on the bore length. For purposes of labeling, the author categorizes the pumps into either the short-bore design or long-bore design. Details of the short-bore and long-bore designs are outlined next. Short-Bore Design While each manufacturer has modifications to the basic design, Table 2 and Figure 14 detail the basic short-bore design configuration. Due to the short bores, these pumps are typically designed as either noncontacting between the screws and bores or light-contacting. As pressures increase, there is the possibility that the screws may run in boundary layer lubrication conditions against the bore. Screw and bore coatings or treatments are typically used only to improve corrosion or abrasion resistance. Advantages of the short-bore design include reduced costs (because of the shorter bores themselves) and the ability to mix and match screw and liner materials to eliminate the necessity of screw and/or bore coatings. Long-Bore Design Table 3 and Figure 15 detail the basic long-bore design configuration. Due to the long bores and bearing span, the shafts may bend more in this

Short-Bore Design Configuration Locator bearings 2 ball anti-friction Radial bearings 2 roller anti-friction Short bore length (3/4 to 1 times bore dia.) Light to non-contacting screw/bore Corrosion- & abrasive-resistant screw coatings

Integral body or replaceable liner Spur, helical or herringbone gears Circular or scalloped bores Screws centered in bore dia.

Table 2. Short-bore design configuration

Figure 15. Typical long-bore/long-bearing design

design compared to the short-bore design for equivalent size pumps under similar conditions. This may require the screws to run against the body bores in boundary layer lubrication conditions and necessitate extensive usage of screw outer diameter coatings and body coatings/treatments. Limitations to the screw/body bore loading are based on manufacturer’s priority empirical data for materials of construction and past experience. Advantages of the long-bore design include longer bores, which allow longer screw pitches and, therefore, greater flow capacity in a smaller diameter screw. At shorter leads, more screw turns are present, which results in lower pressure differentials per turn. This reduces slipped flow and wear rates. Some applications, such as high pressure thin fluids (i.e. naphtha), require the long-bore design to enable sufficient screw turns to minimize slipped flow and create acceptable volumetric efficiency. Bearings are also

Long-Bore Design Configuration Locator bearings - 1 or 2 spherical roller or ball Anti-friction Radial bearings - 4 or 5 heavy duty roller Anti-friction Long bore length (1-1/2 times bore diameter) Non-contacting to heavy contacting screw/bore Extensive use of screw coatings and Treatments

Integral one-piece body Spur or herringbone gears Circular bores Screws offset to one side of body bore Extensive use of body bore coatings and Treatments

Table 3. Long-bore design configuration

Figure 16. System head requirement curve

Figure 17. Pump system performance The Pump Handbook Series

211

Typical Performance Overview at 330 cSt Pump Type

Quantity Pumps

Flow m3/hr

System 1 Power kW

Energy Cost

Flow m3/hr

System 2 Power kW

Energy Cost

Centrifugal

1

600

485

$155,200

485

492

$157,440

Piston

6

600

255

$86,100

600

346

$110,720

Twin-Screw

1

600

309

$98,800

575

379

$121,280

Table 4. Typical performance overview at 330 cSt

Twin-Screw Pump Advantages Dry running (for limited time periods) Self priming, capable of stripping services and slugging flows Lubricating, non-lubricating, corrosive and abrasive fluids Abrasive and non-abrasive slurries Operate at induction motor speeds Low NPSH requirements High temperature capability High flow and pressure capability High efficiency (lower electrical cost) Low fluid shear rates Multi-phase capability

Low noise and fluid pulsations Can handle high gas content fluids Screw pitch changes allow different flow rates No metal-to-metal contact Flow varies with speed Multiple sealing arrangements Slow pressure buildup Internal and external bearing arrangements Horizontal and vertical pump arrangements Reverse operation possible Low starting torque

Table 5. Twin-screw pump advantages

typically heavier duty when compared to the short-bore design pump. Disadvantages of the twin-screw pumps include cost, a relatively complicated design, use of four mechanical seals for external bearing designs and not being understood by most users.

System Requirements Hydraulic systems, piping, valves and filters require energy to move fluid through the system. This is a combination of velocity energy to accelerate the fluid, friction and elevation change. As the flow increases, the energy requirements increase (Figure 16). Note that the curve starts at 50 meters of liquid, which in this example is the elevation change in fluid level. Overlaying the different pump curves onto the system curves shows how each type of pumping technology reacts as the system changes. This is illustrated in Figure 17, in which the centrifugal pump’s flow varies 20% depending on the system requirements. The twinscrew pump varies less than 5%, and the

212

piston pump’s flow remains approximately the same. Based on a viscosity of 330 cSt, the centrifugal pump is the most inefficient, while the piston pump is the most efficient. Assuming an energy cost of $0.04 per kW-hr and an 8,000-hour year, the centrifugal pump would have an additional energy cost of approximately $56,400 over the twin-screw pump. The piston pump is the most efficient, but approximately six pumps in parallel operation would be required to produce the 600 m3/hr flow rate. Table 4 outlines the performance of the selected pumps based on the projected performance data assuming an 8,000-hour year with an energy cost of $0.04/kW-hr. (Note: Energy cost demand charge is not included in the above cost projections.) ■

References: 1. Karassik, Igor J. Pump Handbook. McGraw-Hill Book Company, New York (1976) 2. Westaway, C.R. and Loomis, A.W., Cameron, Hydraulic Data, The Pump Handbook Series

16th Edition, Ingersoll-Rand, Woodcliff Lake, NJ (1979) Section 1 Hydraulics Stephen Smith has more than 28 years of pump experience as both a manufacturer and user. He is a graduate of Columbus State University and Wingate University and a registered Professional Engineer in the state of North Carolina. Mr. Smith is currently the Screw Pump Specialist for Flowserve Corporation. Previous positive displacement positions included Chief Engineer-Twin Screw Products for Imo Pump and Warren Pumps and Engineering Manager at Brunswick Defense for a Reverse Osmosis Water Purification Unit, which used a variety of pumping technologies including reciprocating and centrifugal pumps. Prior to that, Mr. Smith was V.P. Engineering for Pekor Pumps, a manufacturer of centrifugal slurry and dredge pumps.

Using PC Pumps and Other Rotary PD Pumps for Metering Advances in technology now make these pumps an affordable and cost-justifiable option for just about any metering application. Michael L. Dillon, President, seepex, Inc.

hey don't pulsate. They can pump thick liquids. They have no valves to foul. They're rotary positive displacement pumps, and these days they're being used more and more in metering applications. That's because the combining of rotary pumps, like progressive cavity and gear pumps, with newly available drive, instrumentation and flow measuring technologies has resulted in PD pumps capable of delivering similar, or even superior performance to that of reciprocating pumps.

in a defined cavity (Figure 1). These cavities are carried through the pump into the discharge piping, where the cavity is exposed to the operating discharge pressures. If the cavity is completely filled on the inlet side of the pump and there is no internal leakage in the pump, the pump will consistently deliver fluid from suction to discharge. The level of consistency of the discharge or the lack of variance in the discharge is generally regarded as the measure of the pump's ability to meter the volume delivered. Traditionally, reciprocating pumps have been used to deliver variable Here's How They Work controlled flow volume. This was generAll positive displacement pumps, both ally achieved by changing the stroke rotary and reciprocating, capture liquid length of the piston movement to alter

T

the size of the pump cavity (Figure 2). The stroke adjustment mechanism was typically less expensive than a mechanism that changed the number of strokes delivered per unit of time. Most rotary pumps don't have the ability to change the volume of the cavity, but they can change the number of cavities delivered per unit in time. Because changing speed was generally more expensive than changing the stroke length, reciprocating pumps were more economical than rotary pumps for metering. Over the last few decades, though, electronics technology has been reducing the cost of varying the speed of electric motors, and increasing the use of automated or centralized control

Figure 1. All positive displacement pumps capture liquid in a defined cavity. The Pump Handbook Series

213

Figure 2. Stroke adjustment mechanism in a controlled volume reciprocating pump

systems. The need for users to automate and control has tended to equalize the economics between the reciprocating and rotary pump designs, since both now use the same technologies to vary the delivery rate.

Additionally, many rotary pumps can reverse the direction of flow by reversing the direction of motor rotation. Due to internal valves, this can't be done with reciprocating pumps. One of the problems with reciprocating pumps is pulsating flow. Because the piston or diaphragm in a reciprocating pump is actuated by a cam or eccentric section of a rotating shaft, the actuating component during each stroke accelerates to a maximum velocity in mid-stroke and then decelerates to the end of the output stroke. The same action occurs on the inlet stroke, but there's

no fluid output. The velocity of the fluid is similarly affected. The fluid has zero velocity and flow, followed by increased velocity and flow, followed by decreased velocity and flow, followed by a long period of no velocity or flow (Figure 3). Normally, users request a certain Design Differences repeatability of flow or degree of variance from a metering pump. While reciproReciprocating pumps generally: cating pumps may be very good at • have the ability to reach higher producing repeatable flows over long differential pressures; periods of time--like a minute or an hour • are not dependent upon fluid viscosor a day-they're not very good at producity to form a seal between the stationing repeatable flows in short periods of ary and moving parts in the fluid end time, like a second or a fraction of a of the pump; and second. This problem could be • are especially good at preventminimized (but not eliminated ing back-flow or "slip" in the entirely) using multiple pistons or pump, making output more diaphragms. Such methods, though, linear with speed. would increase the cost of the pump significantly and remove the possiRotary pumps generally: bility of using conventional stroke • have reduced pulsation of the length adjustments to vary flow. flow; There are, however, some radial • have better net positive suction piston designs that use a variable head capabilities that enable swash plate to vary stroke length-but, them to handle higher viscosity at an increased price. fluids; A common alternative is the use • may be able to handle solids; of pulsation dampers, which act and like mufflers. The more the user • have no valves to clog, foul or dampens the output, the higher the leak. demand he/she places on the efficiency and longevity of the Rotary pumps also have the pumping system. capability for wider turndown or Gear, lobe, sliding vane and flexiflow range capabilities when comble impeller pumps all discharge pared to reciprocating pumps that multiple cavities during each Figure 3. Changes in velocity in the reciprocating member of only use variance of the stroke a pump translate into varying flow velocities in the pump revolution, so they have performlength to achieve different flows. discharge. ance equivalent to reciprocating

214

The Pump Handbook Series

rigid materials like metals or thermoplastics. Amount of slip in all cases is a function of viscosity, internal tolerances and differential pressure. In the case of pumps with elastomeric components, hardness of the elastomer also affects the amount of slip in a given application.

air or mechanically operated double diaphragm designs, do not use springs at all to effect proper seating of the valves. Seating is accomplished with simple ball check valves or rubber flapper valves that rely on gravity and the differential pressure to properly seat. In these units, there is always some slippage in the pump that is affected by fluid viscosity and the differential pressure. Valves on the intake side of the pump also impede the flow of material into Controlling "Slip" the pumping chamber, increasing the The presence of slip net positive suction head required, in rotary PD pumps is NPSHr (Figure 4). This is accentuated often cited as a reason to with higher viscosity fluids. avoid them on metering applications. However, Viscosity Issues All positive displacement pumps reciprocating pumps are also prone to back flow have to run at a lower speed as the Figure 4. Valves act as a restricting orifice into pumping chamber. or slip. This is basically viscosity of the fluid increases. The a function of how well inlet into the pumping chamber is an pumps with multiple pistons. Screw and how quickly the check valves seat orifice. The more restrictive the design pumps and progressive cavity pumps and seal. If the discharge valves do not of the orifice, the longer it takes for have discharge flow profiles that essen- immediately seat during the intake material to fill the pumping chamber. tially have no pulsation. There are really only two ways to stroke, a portion of the pumped liquid Although this article focuses on will slip back into the pumping solve this problem: either slow the progressive cavity pumps, with some chamber. pump to a lower speed to allow enough limitations, most of the following Most of the more expensive recip- time to fill the pumping chamber, or techniques are applicable to other rocating metering pumps will use change the design of the entry orifice rotary positive displacement pumps. strong springs to effect positive sealing so there is less restriction. The usual in the valves. Some will even employ way of altering this design is to use a What's Different About PCs? tandem valves to minimize back-flow larger pump that has a larger entrance Progressive cavity pumps are on both the suction and discharge sides port and operates at a slower speed. somewhat unique because they gener- of the pumping cavity. While this Some rotary pumps (lobe pumps and ally employ a compression fit between ensures a more predictable flow, it also progressive cavity pumps) are availthe rotating pump member and the increases the amount of work done by able with optional square or rectanstationary pump member that is most the pump, as well as manufacturing gular inlets that significantly open up often constructed of a synthetic elas- costs and absorbed power require- the inlet into the pumping chamber so tomer. Other rotary pumps use similar ments. Moreover, adding components very viscous materials can be pumped construction, including rotary lobe to any mechanical system increases the and metered (Figure 5). If the pumping pumps and gear pumps which coat the likelihood of mechanical failure. cavity is not filled and the pump rotating member of the pumping Certain reciprocating pumps, such as cavitates, flow can't be controlled. elements with an elastomer. The compression fit reduces slippage or backflow in the pump, so the performance curve is more like a reciprocating pump with valves on thin viscosity liquids. As fluid viscosity increases, a tolerance fit can be used, and both the rotating and stationary components in the pumping elements can be made of Figure 5. Open hopper progressive cavity for viscous materials The Pump Handbook Series

215

Cavitation Issues Cavitation can be very difficult to recognize in reciprocating pumps. In PCs, gear pumps and screw pumps, cavitation is normally recognized by pulsing flow. Since reciprocating pumps pulsate anyway, cavitation is normally recognized only after the pump has been damaged, or often by the presence of "knocking noises" in the pump. Knocking noises occur as material vaporizing from the pressure drop coming through the entry orifice collapses back into a liquid as it is pressurized in the pumping chamber. Because many metering applications are for low flows and use small pumps knocking noises often go undetected. While valves have the definite ability to eliminate slip, making the prediction of the flow rate easier, they also contribute to cavitation, making the prediction of the flow rate more difficult. Valves also add another mechanical component to the pumping system that can fail. Any problem with the valves negates the metering capability of a reciprocating pump.

Case History: Hypochlorite Metering A large East Coast electric utility had been using mechanically actuated diaphragm pumps to feed sodium and calcium hypochlorite into its intake water lines to kill algae and mussels. But these pumps would routinely vapor lock when the liquid in the feed tanks went below a certain level. A secondary problem with this application was the free chlorine gas that was frequently present when these pumps were disassembled. Maintenance personnel complained of headaches, nausea and nasal or lung irritations after repairing the pumps. After checking references from a hypochlorite producer that had replaced centrifugal pumps with progressive cavity pumps due to high shear, the utility elected to try the PC pumps. Since installing the PC pumps, the utility has experienced no "gas" locking problem at this site. The utility has since installed more PC pumps on hypochlorite service in other power plants.

The Real Problem

(like fuel transfer from tankers), or for very aggressive chemicals that attack elastomers.

Controlling Slip Electronically Electronics technology now offers a range of solutionsvarying in complexity, cost and degree of accuracy-for controlling slip in rotary PD pumps. As these technologies become more common, the expense and complexity of these solutions will decline.

Variable Speed Drive Options The least expensive method of controlling slip is to use a variable speed drive, so that flow is predictable even though not directly correlated with speed. This can be done with a calibration tube on the inlet side of the pump to measure actual flow rates. Using a programmable calculator and matrix algebra, the performance of the pumping system on a particular liquid can be defined. Then a curve can be generated that shows the speed required to meet a specific flow requirement. The author described this technique almost two decades ago in a previous article (Ref. 1). By calculating the flow curves for two different speeds, with a minimum of three pressure data collection points for each curve, flow rate can be calculated through interpolation if the speed and differential pressure are known. This is probably the least expensive way to meter differing flow rates with a rotary PD pump. A different calibration should be made for each different pumped fluid to allow for the effect of differing viscosity on slip and performance. Some sort of tachometer is required to accurately measure the rpm. It can be part of the drive, an inductive pick-up on the motor, gear or pump shaft or a hand-held portable unit; but it must provide a discrete parametric output. The key to any metering system for a rotary pump is the feedback system that assures that the pump is operating at the required speed to produce the desired flow rate. Variable speed AC and DC drives have become very sophisticated, reliable and economical, with the capability to add a number of feedback signals. In smaller power requirement applications, DC drives are the best choice. They are typically less expensive in these ranges and offer very high turndown ratios with excellent starting torque capabilities with a wide variety of commonly available options and accessories. Typically, a SCR (Silicon Controlled Rectifier) or a PWM (Pulse Width Modulated Rectifier) changes AC current into DC current in the motor controller. The SCR unit is usually the least expensive option. Today it’s common to find SCRs capable of providing a 50:1 turndown ratio. PWM units can typically offer wider turndowns (100:1 is common) and use less energy.

It's apparent that the real problem with rotary PD pumps is slip. Pump design, differential pressure and the material being pumped have dramatic effects on the amount of slip. For materials of moderate viscosity (200-2,000 cPs), gear pumps can be used reliably for metering applications. Due to internal clearances, though, they're not reliable for metering thin fluids like water, alcohol or ketones. On the other hand, rotary PD pumps with a compression fit between the stationary and rotary parts (ie. certain lobe pumps, some gear pumps and PCs) can be used for these applications. There are some large screw pumps and gear pumps used for metering of thinner fluids that do not use a compression fit in the pumping elements. By using very precise machining methods, the internal tolerances are held to a very small limit. These pumps tend to be quite expensive and Drive Accuracy Is Important When selecting a DC drive, it is important to recognize are used for either very high flow rates or differential pressures

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that the accuracy of the drive is defined by its ability to hold a given speed. One of the problems with DC motors is that they tend to slow down as loads increase. In a motor, this slowing is also referred to as "slip". Load will change in the motor if viscosity or pressures change in the pumping system. While the drive may be promoted as having no more than a 1% or 2% variation in speed, the fine print will define this as a percentage of the base motor speed. If speed of the motor is 1,750 rpm, then the drive is only guaranteed to have a speed fluctuation of no more than 15.5 rpm. With a SCR that has a 50:1 turndown ratio, at the low speed of 35 rpm, a 15-rpm fluctuation could mean a speed or flow inaccuracy of 43%. Two types of control feedback accessories are commonly used to counteract slip. The less expensive system is a "current feedback loop". This device compares the current used in both the rotating (armature) and stationary (field) windings in the motor to match current with load demand to provide stable motor speed. This will generally enable the drive to be accurate within 2% of the selected speed over the drive's entire turndown range. A second type of feedback is a tachometer feedback loop. Here, a tachometer is furnished with the motor and the actual motor speed is provided to the drive electronics. The drive will then modify the current to supply the correct speed. While both analog and digital tachometers and drives are available, they have the same inherent function. Just make sure that a digital drive is used with a digital tachometer and vice versa. Programmable meters enable users to not only see readout performance in rpm, but other engineering units as well. Almost all variable frequency drives (VFDs) offer internal feedback loops to minimize motor slip, which also occurs in AC motors. Users don't have to worry about purchasing a number of options to make the drive suitable for a metering pump application. The main limitation of AC drives is their turndown ratio. Without external constant speed blowers

on the motor, most VFDs used on fancooled motors are limited to a lowest operating frequency of 10 Hz. The motor runs at such a slow speed that the fan will not blow sufficient air across the motor to cool it. These units can only provide a 6:1 turndown ratio. Positive displacement pumps are constant torque devices and VFDs used to drive them need to provide constant output torque. (A constant torque drive can be used up to 100 Hz as a constant horsepower drive. If used as such, with a four-pole induction motor, operating at 3000 rpm, this will allow a 10:1 turndown ratio. This is actually better for the motor as it will run in a cooler environment.) A more expensive, high performance version of the VFD, called a Vector Drive also has become available. Their integral feedback systems make them very accurate. They also use TENV (totally enclosed non-ventilated) or TEBC (totally enclosed blower cooled) motors. These drives offer maximum performance with turndown ratios as high as 1000:1. The new AC and DC drives with higher turndowns allow speed reduction to be accomplished electronically with enough performance to also provide flow variation, thus eliminating the need for mechanical speed reduction. The customer can gain a higher performance variable speed system with fewer components at a price comparable to a fixed speed system. All of these systems require users to occasionally, but regularly, re-calibrate their pumps. This is true for any pump. Internal components, including valves and valve seats, can wear, or fluid rheology may change, affecting the pump performance. The final word on using rotary pumps for metering applications is the use of a flowmeter feedback loop to the drive. Most electronic flowmeters offer output signals that are compatible with both AC and DC electronic drives. Although some drives have integrated proportional integral derivative (PID) controllers, stand-alone PIDs can be used on any drive that will accept an The Pump Handbook Series

external follower signal. This is, most commonly, a 4-20 ma (analog) signal or a 5-volt DC (digital) square wave signal. Again, it's important to make sure that all of the sending and receiving devices are compatible. With the use of these electronic technologies, a pump can be as accurate as a flowmeter, some of which advertise accuracies with as little as <0.1% variation.

Case History: Soft Drink Production One of the country's largest soft drink bottling plants was becoming "locked-in" by residential development of surrounding properties. In the past they had always produced a syrup from concentrate and stored it in large vats. From the vat it was sent to the bottling line where it was diluted with carbonated water. As the number of products to be bottled increased (regular, diet, caffeine free, etc.) and as the company bottled other types of beverages, it no longer had adequate physical space to store syrup in vats. At this point the company considered the use of continuous blending systems to make the syrup. Because consistency in taste and quality are so important in the soft drink industry, the pumping systems used must be extremely accurate and repeatable. For this project, an outside consulting/manufacturing firm provided four skids, each with seven PC pumps. The pumps were all equipped with either vector drive AC or PWM controlled DC motors. The output of each pump was monitored with a coriolis type flow meter and flow rate was determined by computer. Coriolis meters recalibrate a system several times per second, but they can't tolerate any pulsation. Reciprocating and several rotary pumps could not be used in this application because the system constantly would be correcting itself from reading varying flows. With the new PCs, each different soft drink "recipe" can programmed into the computer and different ingredient "totes" can be rapidly connected to the system. As a result, this plant no longer requires vats to store syrup

217

and has an unlimited capacity for new ically controlled rotary PD pumping types of products. system will self-correct for wear, slip, changes in viscosity and power fluctuSummary ations. Electronic advances have gotten New electronic drives now enable these systems down to such an affordrotary positive displacement pumps to able and cost justifiable level that they offer relatively pulsation-free perform- should be considered for just about any ance with turndown ratios as high as metering application. 1000:1. These new systems allow rotary pumps to compete favorably in price References 1. Chemical Engineering, Vol. 92, with reciprocating metering pumps. In many ways, they're more reliable than Issue 15, July 22, 1985 Dillon M., St. reciprocating pumps. Moreover, in a Clair K., Kline P., "Predicting Flow Rates number of applications, rotary pumps from Positive Displacement Rotary offer more exact metering performance Pumps", Chemical Engineering, Vol. 92, than reciprocating pumps. Coupled with No. 15, pp. 57-60, 1985. a flowmeter feedback loop, an electron-

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Michael L. (Mike) Dillon is President of seepex, Inc., in Enon, OH. A graduate of Ohio University and Wright State University, Dillon is the author of numerous articles on positive displacement and progressive cavity pumps. This article was adapted from his presentation at PumpUsers Expo 2000. For more information on the pumps and applications referenced in this article, e-mail him at [email protected].

Multiphase – The Final Pumping Frontier

For many applications, the ability Petroleum Engineering Department at Texas A&M University has conducted

Pressure builds for combined gas/liquid capabilities.

to pump a gas-liquid mixture is research on the application of multiphase twin-screw pump technology. Multiphase phenomena in twin-screw pumps will be examined as well as potential future Over the past decade the pump indus- applications of multiphase pumping try has made a concerted effort to develop technology. this capability. Today, the oil & gas industry is finding many advantages to this Applications in the technology and, in some applications, Oil & Gas Industry multiphase pumping has been adopted To the oil & gas industry, multiphase as a “best practice.” The following article pumping is simply the ability to boost discusses how and why multiphase pressure without the need to separate the pumping is being selected for use in the liquid and gas phases. This provides oil & gas industry, and also how the advantages in many types of development, pump industry has risen to the meet the in some cases allowing access to energy special challenges posed by multiphase resources that were completely out of pumping. For the past few years, the reach a few short years ago. The modern

highly desirable.

Dr. Stuart L. Scott, Associate Professor of Petroleum Engineering, Texas A&M University Ana M. Martin, Graduate Research Assistant, Texas A&M University Photo 1. A California steam-flood project The Pump Handbook Series

219

ogy since this has been the emphasis of work at Texas A&M. The technology is well suited to multiphase oil and gas production due to its ability to handle nearly 100% gas-volume fractions (GVF’s). Most oilfield applications operate in the 80-98% GVF range. Photo 2 shows a multiphase pump installed in a Canadian application (Apache’s House Mountain Field, located northwest of Edmonton, Alberta).

Multiphase Phenomena in Twin-Screw Pumps In most configurations, the multiphase mixture enters one end of the pump, is split into two flow streams, and then passes through chambers created by the interlocking of the twin screws. An examination of the fluid behavior indicates complex multiphase flow. At low pump speeds, the gas phase separates from the liquid and Photo 2. A multiphase pump at House Mountain Field travels along the top of the screws. Photo 3 (top) shows this effect in a low-pressure multiphase production system transports separator, single-phase pumps and large model of a twin-screw pump. At higher the “full well-stream” from the wellhead fin-fan cooling units. A multiphase pump speeds, the centrifugal force imparted by to a centralized facility, which is often replaces all this equipment, significantly the rotating screws forces the liquids to the many miles from the well. The full well- reducing the surface footprint, capital costs outside, while the gas moves as a separated stream includes all produced materials - and permitting requirements, while phase along the shaft (Photo 3, bottom). oil, natural gas, water, sand, etc. This improving performance. For multiphase flow, efficiencies and distribution (flow pattern) of liquid and gas development mode provides for a within the volume cell change with gasminimum impact on the environment and Multiphase Pumping Technology a reduction in costs. However, this long Three development paths being followed volume fraction and liquid viscosity. Pump multiphase flow line also creates backpres- by the pump industry are the screw pump, efficiency is determined by the liquid seal sure on the reservoir, which can reduce the helicoaxial pump and the piston pump. formed in the small gaps between the two flow rates and recoveries. In many appli- This article highlights twin-screw technol- screws, and between the screws and the containment chamber wall. cations, pressure boosting Unlike in single-phase flow, through multiphase pumping large amounts of liquid slip has been found to be an effecfrom one chamber to another tive means of eliminating this as gas volumes are reduced. difficulty, allowing the full Twin-screw pumps are potential for multiphase production to be realized. positive displacement Another force, driving pumps, consisting of two change in many on-shore fields, intermeshing screws placed is the need to reduce environinside the pump’s casing. The mental impact of operations. meshing of the screws creates This includes reducing the sealed cavities, which move “footprint” of operations, axially from the suction port improving practices and elimito the discharge as the screws nating flaring. Photo 1 shows rotate. For multiphase applithe traditional, large footprint cations, these pumps are required to produce oil (from a commonly configured with Texaco steam-flood project in dual inlets, external timing California). Included are a large Photo 3. Low speed (Top) and high speed (Bottom) multiphase flow gears and external bearings. 220

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The dual-suction design ensures axial- through these clearances, so the throughforce balancing, reduction in the put is no longer the displacement times required NPSH and, very importantly, the RPM. The main driving force of the allows the mechanical seals to function slip flow is the pressure difference across at suction pressure. The timing gear the pump. The larger the pressure drop, transmits the torque from the driving the larger the slip-flow rate. The green screw (connected to the motor) to its curve in Figure 1 corresponds to a typical companion screw, synchronizes the two pump performance curve. It can be seen screws and prevents direct contact that, as the pressure difference increases, between the screws. For this type of the delivered flow rate diminishes (since positive displacement machine, each the slip increases). The black line correturn of the screws will result in a finite sponds to the theoretical, or geometrical, volume displacement. Twin-screw flow rate described previously (at a pumps don’t actually develop pressure constant RPM, the flow rate is constant as dynamic machines do, so the result- regardless of pressure difference across ing discharge pressure is a function of the pump). the upstream pipeline system (i.e., the Under multiphase flow conditions, the energy required to move the fluid pump will exhibit significantly differthrough the pipeline, overcoming the ent behavior. The pump delivery is not pipe friction, valves, etc., at the delivered flow rate). The volume “moved” per revolution is termed the pump “displacement.” The displacement is a function of the geometrical parameters of the pump, such as root and bore screw diameters and screw pitch. This means that, theoretically, the pump’s flow rate will be only a function of its size and the operating RPM; the only ways to control the flow rate are by either varying the pump’s Figure 1. Typical twin-screw pump performance curve speed or using a by-pass at the discharge. A controlling concept in twin-screw controlled so strongly by the pressure pumps is that of slip. Screws are milled difference, and the performance is closer to provide a clearance, a separation from to theoretical (as depicted by the red the case or lining, to avoid contact. Also, curve on Figure 1). In this case, the a clearance must exist between the screws volumetric efficiency of the pump (calcuthemselves, to avoid direct contact lated at suction conditions) increases with between them. The dimension of these the presence of gas in the fluid. It has been clearances is highly proprietary, and takes determined experimentally that, for into consideration many factors such as ordinary screw pumps and for GVF’s up screw deflection, temperature and metal to 85%, only liquid slips through the expansion, service (single or multiphase clearances. The main multiphase-pump flow), fluid viscosity and pressure differ- manufacturers now offer a design with ential. The presence of these clearances an internal recirculation system, which eliminates the theoretical seal line, since allows operation with up to 100% GVF now the screws are neither in contact with for a limited time. This recirculation each other nor with the liner. As a conse- system guarantees that enough liquid will quence, part of the flow will slip back be present at all GVF’s to ensure a liquid The Pump Handbook Series

seal in the gaps. Pumps with an effective recirculation system will have only liquid slipping through the clearances at any GVF (true only for a limited time at very high GVF). Since the liquid is the only fluid leaking back through the clearances, the mass of gas captured in each cavity, as it is transported from the suction port to the discharge, will remain constant. The liquid leakage will flow from the discharge toward the suction, due to the pressure difference. As it passes to the first upstream cavity, part of this leakage will flow back to the next upstream cavity and so on, making its way to the lower pressure at the suction. Some, however, will remain in the first cavity. Since the gas is compressible, it will reduce its volume to accommodate the new incoming liquid. This scenario will repeat serially in the rest of the cavities within the pump. Thus, it is the slipping liquid that compresses the gas in the pump. Since not all the liquid that started slipping will make it to the suction side, more fluid is being pumped (compared to the only-liquid case). This is why the volumetric efficiency, referred to suction conditions, increases with the presence of gas. To help understand this process, one could make an abstraction, thinking of each of the cavities formed within the screws as small tanks (as shown in Figure 2). These tanks would be connected by a pipe, which represents the clearance between the screws. The flow through this pipe from tank to tank is equivalent to the slip flow or leakage through the clearances and gaps. It is easy to see that, if only liquid is flowing through the gaps, the gas will be compressed by part of the slip flow. The remaining leakage will flow to the next upstream tank, where part will remain and will compress the gas. In practice, multiphase production is characterized by wide fluctuations in the gas and liquid mass-flow rates. 221

During periods of substantial gas flow, temperature becomes the critical variable determining the performance of the pump. Without the liquid phase to remove the gas-compression heat, temperatures begin to rise, and may cause a high temperature shut-off or even damage the pump. Predicting the length of time a pump can effectively operate under 100% GVF conditions is a multiphase-flow problem that is strongly dependent on the re-circulation system, and the amount of liquid that is retained in that system. To date, pump temperature performance with 100% GVF has only been investigated experimentally, and only a very small empirical database has been established.

Texas A&M Multiphase Pump Research In 1999, Texas A&M began a project to investigate multiphase pumping from an application point of view. That is, how do these pumps function under multiphase conditions and how can they be incorporated into existing reservoir, wellbore and pipeline models of the total oil & gas production system? Twin-screw pumps were the focus of this work. By 2000, a Joint Industry Project (JIP) had been formed to assist industry with the effective implementation of multiphase pumping technology. Current members of this JIP include BP, Chevron and Marathon. The goal of this work is to develop a mechanistic model of pump performance Figure 2. Gas compression by the leak flow that can be used to generate the tables necessary for use in steady-state and transient multiphase-pipeline simulation codes. To better understand multiphase flow behavior in pumps, a large-scale outdoor laboratory has been created. Termed the A&M Multiphase Field Lab, this facility features a full-size, twin-screw, multiphase pump. The facility’s horizontal flow loop consists of 2000 ft of 2-inch coiled tubing. Using this facility, a wide variety of flow conditions can be studied. These include gas-volume fractions (GVF’s) from 0 to 100%, fluid viscosity from 1 to 200 cp., and a wide range of pumps speeds and differential pressures. This allows a much more complete evaluation of pump performance than can be performed by the pump vendor, which typically generates only a few water-air tests to provide to the user. Also, using this facility, the pump 222

can be configured with the coiled-tubing flow line to investigate pump/flow-line interactions, and the effects of flow patterns (such as slug flow) on pump performance.

The Future The next five years promise to be an exciting time for multiphase pumping. The technology has moved out of the research laboratory and is well established in many types of oil and gas operations. Clearly, there will be continued growth in the number of installations for heavy-oil and steam-flood projects. Many pumps are already scheduled for delivery in Venezuela, Canada and California during the remainder of 2001 and well into 2002. Key future areas for application of this evolving technology include wet gas compression (96-100% GVF) and wellhead compression. Also, interest in the subsea application of multiphase pumps is growing. Deepwater and ultra-deepwater development schemes make extensive use of subsea, completed wells. Access to a pump located on the seafloor is much less expensive than repairing a pump located in a subsea wellbore. Therefore, accessibility is driving the serious consideration of subsea multiphase pumping. Subsea multiphase pumping is an active appliedresearch area, where several projects are underway or planned for field demonstration. These include several successful installations of helicoaxial multiphase pumps by Framo, for Statoil, ExxonMobil and BP. In general, the helicoaxial technology has a substantial lead for subsea applications. Several subsea deployments of the twin-screw technology also have been undertaken. These include the SBMS-500 project by Westinghouse/Leistritz with Petrobras, and the Kvaerner/Bornemann Demo-2000 project for Norsk Hydro. Twin-screw manufacturer Flowserve (formerly Ingersol-Dresser Pumps) has partnered with DSND for subsea work. At Texas A&M, construction of a large-scale test facility, to promote the qualification of multiphase pumps for subsea application, has been proposed. This facility would provide the ability to test subsea pumps at conditions closely representing deepwater (34 deg. F temperature and up to 4,400-psi external pressure). To examine the effect of subsea conditions on a working pump, this facility would provide dynamic testing

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P&S0701P32

capability (testing under flowing conditions with substantial lengths of flow lines upstream and downstream of the pump). This will provide the ability to simulate the stress of slugging on pump integrity and reliability. ■

Dr. Stuart L. Scott is an Associate Professor of Petroleum Engineering at Texas A&M University. Scott holds a B.S. (1982) degree in Petroleum Engineering, an M.S.(1985) degree in Computer Science and a Ph.D.(1987) degree in Petroleum Engineering, all from The University of Tulsa. Before

C M Y K

joining Texas A&M, he was an Assistant Professor at Louisiana State University and also worked nine years for Phillips Petroleum Company. Winner of the 1987 Society of Petroleum Engineers (SPE) graduate student paper contest, he is an active member, having served as Chair of the Production Operations Technical Committee (2000 & 1992) and overall Chair for the 1st SPE Forum on Multiphase Flow, Pumping and Separation Technology. In 2000, Scott was selected as the Chairman of the Production Committee for the ASME Petroleum Division. Dr. Scott leads the Texas A&M efforts in multiphase production

The Pump Handbook Series

system research and Chairs the annual Texas A&M Multiphase Pump User Roundtable (MPUR). Ms. Ana M. Martin is a Ph.D. Candidate in Petroleum Engineering at Texas A&M University. She holds B.S. (1995) and M.S. (1999) degrees in Mechanical Engineering from Universidad Simon Bolivar (Caracas, Venezuela), where she worked as a Research Assistant in the Mechanical Energy Conversion Laboratory. She is currently engaged in the modeling of Twin-Screw Multiphase Pumps as part of her research at Texas A&M.

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MECHANICAL SEALS & SEALING SYSTEMS HANDBOOK Table of Contents Seal Environmental Controls.......................................................................................1 Mechanical Seals Keep Pace with User Demands .....................................................9 Extending Mechanical Seal Life ................................................................................11 Effect of Bearing Performance on Seal Life .............................................................14 Magnetic Liquid Seals Contain VOCs .......................................................................16 Just What Is a “Totally Engineered Sealing System”? ..............................................19 Off-Design Operation and Seal Performance ...........................................................22 Testing Seals in Adverse Situations ...........................................................................25 A Quick Reference Guide to the API 682 Standard.................................................29 Gas-Barrier Seals Establish Beachhead ....................................................................35 Bellows Seal Repair: The Plant Perspective..............................................................41 Cartridge Seals: Pros, Cons and the Human Factor.................................................43 Seal Face Application: Little Room for Error ...........................................................46 Dual Sealing Systems; a.k.a Double and Tandem....................................................50 Seal Reliability at Chevron ........................................................................................55 Seal Chamber Design Affects Reliability, Emissions................................................58 Seal Protection: Guarding Against Air and Abrasives ..............................................62 Mechanical Seal Installation .....................................................................................66 The Importance of Seal Failure Analysis..................................................................70 Zero-Leak Seals Cut Emissions..................................................................................72 High Temperature Sealing Systems.................................................................76 Applying Dry Gas Sealing Technology to Pumps............................................79 Shaft Sealing for Pulp & Paper.......................................................................83 Reliability Through Fluid Sealing Management .............................................86 Choices in Emission Control for Rotating Equipment....................................90 Maximize Seal Flush Performance for Longer Seal Life .................................98 Estimating Heat Generation, Face Temperature and Flush Rate for Mechanical Seals .................................................................................108 Mechanical Seal Failure Analysis..................................................................114

MECHANICAL SEALS & SEALING SYSTEMS HANDBOOK Table of Contents Materials for Seal Faces ...........................................................................................120 Packing in the 21st Century.............................................................................................124 Built to Last: Working with a New Mechanical Seal Standard.............................130

All materials © 2002 Pumps & Systems, LLC. No part of this publication may be reproduced without the written consent of the publisher. The publisher does not warrant, either expressly or by implication, the factual accuracy of the articles or descriptions herein, nor does the publisher warrant the accuracy of any views or opinions offered by the authors of said articles or descriptions.

MECHANICAL SEALS & SEALING SYSTEMS HANDBOOK

Seal Environmental Controls BY: COKER GOLDEN he performance of a mechanical seal is directly governed by the conditions under which it operates. Often overlooked or misunderstood are timeproven methods of improving the operating environment of mechanical seals. The “Piping Plans” that have been documented in industry standards like API-610 and ANSI B73.1 provide a basic approach for designing seal environmental controls. The upcoming API-682 standard for mechanical seals will be even more detailed. Mechanical seals face increasingly stringent demands as the business and regulatory climates of the ‘90s assign more and more value to safety, reliability and protection of the environment. Although many challenges face the user, seal environmental controls can be effective tools for optimizing seal performance. To understand the influence of a controlled environment on a mechanical seal, first examine the nature of the seal. The most important aspect about a mechanical seal is that it’s really a bearing! The faces of a seal operate as a thrust bearing; their value comes from their ability to operate with an acceptably low lubrication (leakage) rate. As with any bearing, a mechanical seal will only perform if adequate lubrication is present and if loads are stable and controlled. These conditions are rarely met in a seal installation if environmental controls are not utilized. Stuffing box pressure, which loads the seal, will vary widely in many pumps. Pumped fluids are seldom ideal as lubricants since they are so often corrosive, abrasive or operating near their boiling point. To simplify, the basic functions of seal environmental controls are to protect the seal from attack by corrosion and abrasion, control hydraulic loading, and optimize the lubricity of the fluid film between the seal faces. To be thorough, however, consider additional demands on the seal and

PHOTO COURTESY OF SPECTRUM PROCESS, INC.

T

API-Plan 54 its systems: • Start-up, shut-down, and standby conditions may present difficulties. • Upset conditions should be considered (dry runs do happen). •

Leakage products may compromise the operation of the seal or may be dangerous.

•

Normal seal emissions may be under scrutiny by regulatory agencies.

•

Seal failure may create a safety hazard.

At the same time, try to anticipate seal and system component failures and to add back-up or redundant elements.

IMPROVEMENTS IN DESIGN AND MATERIALS In many cases, corrosion and abrasion are not a problem due to recent developments by seal material producers as well as pump and seal manufacturers. Computer-based The Pump Handbook Series

techniques such as finite element analysis and computational fluid dynamics have resulted in product advances that seem to hold great promise. New and improved face materials (that is, better quality and consistency) mean seals will better resist the effects of corrosion and abrasion. Advances in seal design allow the application of previously difficult to use alloys for critical components. High performance alloys produce seals that are much more stable and corrosion resistant than in the past. And credit the pump manufacturers for doing exciting work in seal chamber design. New chamber configurations promise terrific improvements in the ability to operate seals in severe environments. Even so, seal environmental controls offer ways to improve or completely disguise the required seal environment. Pressure and shaft speed combine to form the basic “load” of the mechanical seal face pair. The industry standard “PV” factor, the product

1

of pressure and sliding speed, is the benchmark defining the load capacity of a given seal face combination. Normally, you cannot influence shaft speed, although seal environmental controls can provide a variety of tools for control of the seal load’s pressure component. Improving the pressure situation may involve reducing the pressure to a less demanding level or, in some cases, increasing the pressure to yield improvements in seal face lubrication. Also, environmental controls can stabilize the pressure being sealed, effectively isolating the seal from pump and system transient effects. New designs have also yielded meaningful improvements in the ability of seals to withstand higher/varying pressures while maintaining consistent leak and wear rates.

FIGURE 1

LUBRICATION—A DIFFICULT PARAMETER Given these improvements, the lubricant film between the seal faces usually emerges as the most difficult parameter of seal operation to control. Consider the complicating factors: •

Often, we try to seal a pumped fluid that exhibits poor lubricity.

API-Plan 32, external flush

FIGURE 2

•

The lubricant film will almost certainly be under high thermal and mechanical stresses between the seal faces.

•

The process may either add or remove heat from the film.

•

The lubricant may be prone to flashing, or de-gassing, or disassociation into other compounds.

API 11

API 41

API-Plan 11, basic discharge recirculation; API-Plan12, adds strainer; API-Plan 21, adds heat exchanger; API-Plan 22, adds strainer and heat exchanger; API-Plan 31, adds cyclone; and API-Plan 41, adds cyclone and heat exchanger

2

The Pump Handbook Series

Any loss of lubrication will increase face temperatures, and heat is the sure precursor to reduced seal life or increased leakage. So analysis of all the factors affecting the quality of seal face lubrication can become challenging in a hurry. In many cases, mechanical seal manufacturers are the best resource for guidance here. Also, seal environmental controls, carefully applied in these cases, can yield the greatest improvements. Standard control methods offer a variety of techniques for solving the lubrication problem. Start and stop conditions should be examined thoroughly because environmental controls can be very helpful in a number of typical situations; i.e., hydraulic transients during pump start-up and valve actuation, dry start-

ups, heat soak during standby, vacuum conditions at the seal chamber during standby, etc. If the pump application has enough history for the user to predict upset conditions, seal environmental controls can help. Dry running or temperature and pressure excursions are often unavoidable. Some systems require a lengthy period to recover from upsets, resulting in surging flows and varying conditions at the mechanical seal. Here again, careful analysis can yield dividends.

FIGURE 3

LEAKAGE IS MANAGEABLE Mechanical seal leakage can never be zero, so sealed fluids always leak to the external environment. In many applications, this can be a primary problem. The leakage products may thicken or turn to crystal forms, causing seal failures. Simple environmental controls solve most of the problems of single seals; other cases require the use of tandem or double seals for management of the leakage products. These seals require auxiliary support systems. Containment is the watchword. Seal environmental controls have a successful history of providing effective containment over long operating periods. Emissions from operating seals are currently the hottest topic. This is another area where manufacturers are making rapid advances. Some of the new single seal designs are proving successful in services where emissions must be greatly reduced without compromising seal reliability and

FIGURE 4 OPTIONAL ROTAMETER FOR GASES

API-Plan 13, recirculation/vent to suction

life. New combinations of multiple seals are also being introduced. Some advanced seals depend less on environmental controls than their predecessors; with others, environmental controls are a critical element of their enhanced performance. Over time, the seal, pump, or some seal environmental control component is going to fail. Some users may be able to afford control systems that monitor seal conditions and provide notification of failures while protecting systems and personnel. These cases require a cost-benefit analysis of the more complex control systems.

TOOLS FOR SEAL OPTIMIZATION

DRAIN / VENT

API-Plan 62, external fluid/gas quench with API bushing (reverse connections for heavy gas)

The tools for optimizing mechanical seal environments are many and varied. First, there are the basic API-610 Piping Plans. (Each API plan has a corresponding ANSI plan; the ANSI numbers add “73” before the API number.) These plans are described by the familiar schematic representations of their basic components. The old plans are still the basic building blocks of our best efforts with seal environments. Next there are new ideas, such as enhancements or modifications to the old standards. Finally there are system layout recommendations.

The Pump Handbook Series

3

PLAN 32 Plan 32 (Figure 1) is for the injection of a suitable fluid (defined as a flush fluid) from an external supply to a point near the process seal. This plan assumes that “dilution is the solution.” Injected fluid becomes the main component of the lubricant film, and it carries away seal-generated heat as it enters the process stream. The flush will also act to dilute and carry away abrasive particles. The addition of a restriction bushing at the throat of the pump’s stuffing box can reduce the amount of required liquid and retain pressure in cases where this would benefit the seal. The plan’s drawbacks are that process dilution is taking place, regulation of injection rate may depend on small valves or orifices that could be prone to clogging or wear, and an automatic shut off valve may be necessary if pump operation is intermittent.

PLAN 11, 12, 21, 22, 31, 41 Plan 11 (Figure 2) and its associated plans (that is,12, 21, 22, 31 and 41) are schemes that use the high pressure process fluid available at the pump discharge to improve the seal environment. In these plans the process fluid is reinjected similar to Plan 32. Again, this injected flow has the benefit of providing enhanced temperature and pressure control at the seal faces. The recirculation rate must be controlled because excessive flow rates produce harmful velocities in the confines of the seal chamber. But carefully applied, these plans have been beneficial over the years in a myriad of seal applications. Often, the factory supplied ports on the seal cavity have been poorly located; newer designs are much better in this regard. The injection port should be large enough to minimize velocity and should not inject the fluid stream at the working parts of the seal, where the impingement could be harmful. And as with Plan 32, a throat bushing can be part of the scheme for pressure and flow regulation. Consider eliminating the API orifice by sizing the recirculation line small enough to provide the required throttling effect. In low and medium pressure situations, the absence of an orifice will enhance long-term flow

4

stability. The addition of the extra components (for example, strainer, heat exchanger, and cyclone separator), which defines the other plans, would provide obvious benefits where analysis suggests application. Use filters and strainers only when anticipated loading is near nil; otherwise, filter loading will diminish system performance and require constant attention. As mentioned earlier, manufacturers are introducing new seal chamber designs. Some of these are successful in mimicking the performance of discharge recycle plans by inducing process flow in and out of the seal area and reducing abrasive wear. If these new designs become well-accepted, they will lessen dependence on Plan 11 and its derivatives.

PLAN 13 Plan 13 (Figure 3) is an excellent scheme that is often overlooked. Most pumps exhibit a higher pressure in the seal cavity than at the suction port. This plan uses the pressure differential to pump process fluid from the seal cavity back to the suction side. It also does a terrific job of purging gases from the seal chamber. The flow also acts to remove seal generated heat from the stuffing box. Before applying Plan 13, ask the pump manufacturer for a definition of the seal chamber pressure at your operating points to ensure that the needed pressure above suction is always available. Since Plan 13 may actually lower the seal chamber pressure slightly, make sure the pressures do not drop so low that the lubricant film between the faces is reduced. Plan 13 is ideal for many vertical pumps, where it acts as a high point vent. Also, this plan is often ideal for large, slurry pumps where entrained gases can concentrate at the seal faces because of cyclonic separation in the chamber. One warning though: Slurry pumps often feature rear pump-out vanes on the impeller to reduce stuffing box pressure. (This is a holdover from designs intended for packed stuffing boxes.) Most slurry pump vendors are willing to give advice on ways to achieve higher seal chamber pressure to improve seal lubrication and provide the necessary pressure. Plan 13 is typically simple The Pump Handbook Series

and trouble-free. As with Plan 11, using the line size instead of an orifice for flow control is a good idea, especially on slurries.

PLAN 62 Plan 62 (Figure 4) is the external Quench Plan. In seal terminology, a quench fluid is one that has passed once through the mechanical seal assembly and then drained or exhausted to the outside. The routing of this fluid ensures that it never enters the process zone. The use of a quench plan typically incorporates the familiar quench and drain ports on the mechanical seal gland, as well as the API disaster bushing which closely encircles the rotating shaft. Quench fluids provide a variety of benefits for the mechanical seal. The fluid will always have superior heat transfer capacity as compared to ambient air. Thus, quench steam is commonly used to heat and flush out leakage products that would otherwise solidify before exiting the seal and cause a hang up of flexibly mounted seal parts. Steam is also used as a quench medium in very hot processes where it provides additional cooling at the seal faces and an inert blanket to prevent coking of leakage residues. Inert gases are also used for anti-coking benefits. They don’t provide the cooling of steam, but they avoid the accompanying condensate stream. Liquid quenches can also be used to dissolve and flush out crystalline leakage products, such as those that are caustic. They may also provide face lubrication to single seals under vacuum. The drawback here is that a process waste stream, albeit small, is produced. The drain port and disaster bushing are also beneficial when the seal fails or wears out. The bushing and porting contain leakage products and direct them to a desired location; this is better than the case where leakage is flung outward as it exits along the shaft. Many Plan 62 systems are in service. Often they are the difference between success and failure of single mechanical seals.

BARRIER SYSTEMS Barrier systems are external closed loops required for the support of multiple seals. The barrier fluid is usually different from the process fluid, and it may or may not enter the process depending on the relative system pressures. The first step in barrier system selection is to analyze the required heat transfer. Barrier systems will involve multiple heat sources and heat sinks. The seal faces will always be a source of heat, and the pump and process will be either a heat source or a heat sink. The external loop is most commonly designed to remove heat, but in some cases may serve as a source. Every situation deserves a detailed review of heat flow and optimum temperatures with an eye toward controlling seal face lubrication quality. The seal manufacturer will be the best source for analysis and recommendations at this point in the design. All barrier systems are designed to take advantage of the thermosyphon effect - that is, where the difference in fluid temperature between the hot and cold barrier loop legs will initiate a continuous flow around the loop. Loop flow can be enhanced by the addition of an integral pumping element in the multiple seal. Auxiliary heating or cooling of the fluid is commonly desired and usually accomplished by a tubing coil internal to the pressure reservoir. Thermosyphon systems without auxiliary cooling are ineffective in rejecting heat unless their operating temperature is more than 150° F above ambient. Systems with auxiliary cooling, with or without the pumping element, can successfully control the temperature in nearly all low and medium performance pumps. Certain installations may call for a circulating pump in the barrier loop (Plan 54). The resulting higher flow rates will lead to much greater heat exchange efficiency and consistency.

PLAN 52 The choice of barrier system pressure determines which API plan to follow. Plan 52 (Figure 5) governs cases where the barrier system pres-

FIGURE 5 VENT OR FLARE ORFICE UNION HIGH PRESSURE SWITCH

HIGH LEVEL SWITCH

LOW LEVEL SWITCH

COOLING COIL

DRAIN VALVE

Plan 52 - Double seal reservoir system (Plan 54 adds circulating pump) sure is less than the stuffing box pressure. With this pressure bias, the mechanical seal is defined as a “tandem seal,” and the process liquid lubricates the inboard seal. The inboard leakage is captured in the barrier loop, where it is vented or contained (as desired). The outboard seal runs in a lightly loaded condition, cooled and lubricated by the barrier fluid at low pressure, until the inboard seal fails. At this point Plan 52 is isolated to allow process pressure to build and be sealed by the outboard seal, which will carry the full system load until shutdown. So the plan normally functions as an

The Pump Handbook Series

installed back-up seal, which can take over primary duty at any time. The transfer to the backup seal is automatic. All systems should have instrumentation to signal the change-over. Where vaporizing products are being pumped, normal leakage vents through an orifice, usually to a vent or flare system. The pressure switch signals an alarm or closes the normally open vent valve or does both when system pressure rises in response to a large inboard leak. Plan 52 systems on nonvolatile fluids may be run with the vent valve closed. An alarm will be generated by either the pressure switch or high-level switch when inner seal

5

failure occurs. The operator can respond by draining off excess fluid or manually venting to reduce pressure. Normal inboard leakage will require periodic drain-off. Inboard failure will, of course, quickly generate a repeat alarm, signifying that the outboard seal must now carry the primary load. The low level switch is required to signal outboard seal failure. Plan 52 is an effective answer to emissions problems. A suitably sized pressure reservoir will be slowly diluted over time with liquid leakage products. The outboard seal will remain at a low pressure; thus the dilute solution will have a low leak rate to the outside. The dilution results in very low emissions at the outboard seal. Plan 52 will slowly become more rich in the process product, so it will periodically need to be drained and refilled with fresh barrier fluid to reduce emissions. The plan can be drained and refilled while the system is in operation. In essence, this plan is an installed spare seal concept with the ability to monitor system and seal status, to contain process liquid, and to handle failures in a controlled way.

FIGURE 6

PRESSURE REGULATED GAS SUPPLY LOW PRESSURE SWITCH

LOW LEVEL SWITCH

COOLING COIL

PLAN 53 If you elect to run the barrier system pressure higher than the process pressure, design according to API Plan 53 (Figure 6). This pressure relationship defines the “double” seal. A higher pressure in the external loop almost creates a complete new environment for the inboard process seal. Both seals are now cooled and lubricated by the chosen barrier material, which should be selected with low face wear in mind. A moderate amount of barrier material will migrate into the process through the inboard seal, so barrier fluid compatibility with the process is necessary. System pressure is typically achieved by pressurizing the overhead space of the reservoir with inert gas. Failure of either the inboard or outboard seal will result in a loss of liquid in the reservoir, so a single low-level switch will suffice. Since loss of gas pressure is undesirable, a pressure switch in the overhead space is strongly recommended. The positive pressure bias contains

6

Plan 53 - Tandem seal reservoir system

the process liquid completely. Thus, a Plan 53 system will have zero emissions of the process liquid. The usefulness of Plan 53 is even greater if the installation includes a method of refilling the pressure reservoir while in service. With an adequate refill method, normal liquid loss over the life of the seal can be replenished without interrupting operation. Refills can come from: • a direct plant supply of the barrier liquid (source pressure must be higher than gas pressure) •

a hand-operated pump, which can be permanently mounted and piped to the reservoir (popular in explosion hazard areas)

The Pump Handbook Series

•

an auxiliary fill tank, which can be mounted so as to drain into the main reservoir

The fill tank is typically vented for filling with refill liquid, then pressurized from the regulated gas source of the main reservoir. A block valve between the fill tank and the main tank allows gravity transfer. All refill systems should be designed so that new fluid is introduced at a slow rate in order to avoid thermal shock to the mechanical seal. Plan 53 can come very close to creating a completely idealized environment for the process seal, ensuring that it is cooled and lubricated by an optimized fluid. In many cases this plan enables the double seal to sur-

vive dry running indefinitely. And as with Plan 52, Plan 53 allows constant monitoring of the performance of the seals and provides containment and control when seals fail.

ported valves are fine, but check carefully. On Plan 52, select valves with stem emissions in mind. 4.

The seal supply and return lines must slope continually upward to the pressure tank. (Any bubbles that do not automatically vent will ruin syphon flow.)

5.

In most cases, insulating the hot leg will increase leg flow, thus reducing ∆T of the barrier fluid passing through the seal. Overall barrier temperature may rise, so check to avoid a too high temperature. (High winds can actually chill a thermosyphon, reducing or stopping flow! Insulation on the hot leg will prevent this problem as well.)

PLAN 54 Plan 54 adds a circulating pump to Plan 53. This plan is most often needed to ensure heat exchange in large heat load applications, and in low speed and standby applications where thermosyphon flow is not consistent. Plan 54 is typically very stable and predictable, even with upset conditions. This plan can serve multiple seals; however, in most shared systems it is difficult to identify a leaking seal when fluid loss occurs. The circulating pump may be centrifugal or positive displacement; hermetic pumps are often well-suited here, since the barrier fluid has been selected as a good lubricant. The forced circulation frees Plan 54 from the location constraints of thermosyphon systems and this plan may well be located at some distance from the seal.

SYSTEM LAYOUT Barrier systems must be laid out carefully. Faulty design will totally defeat system performance and create maintenance headaches. Some important points to follow: 1.

Location and layout of the reservoir and supply and return lines are critical. The reservoir should be as close to the seal as possible with the bottom head of the reservoir a minimum of 18 inches above shaft centerline height.

2.

The liquid lines to and from the seal should be tubing, not pipe. Lay out the lines for the absolute minimum number of bends. Tubing should be 1/2 inch outside diameter (OD) as a minimum. Seal manufacturers have reported incidences of starvation of integral pumping elements when tubing smaller than 5/8 inches OD is used on the seal supply.

3.

Isolation and drain valves near the seal may make pump and seal maintenance much easier. Select valves for the minimum obstruction to loop flow. Most standard

6.

7.

Use only precision gas pressure regulators on Plan 53. They must be of the relieving type to allow for liquid expansion. Check for low measure rise as the regulator begins to relieve.

tions and increases the stability and operating range of these seals. •

Plan 52 on volatile services is often improved by the addition of a low cracking pressure check valve ahead of the orifice. A constant pressure builds in the reservoir, aiding outboard seal lubrication and reducing the foaming tendencies of some barrier fluids. The check valve also isolates the system from pressure transients in the vent or flare system.

•

Manufacturers have also introduced tandem seals featuring dry running outboard seals that eliminate the need for Plan 52 systems. Trials have not been perfect, but these new products suggest the possibility of installed standby seals that do not depend on a liquid barrier system to survive.

•

Air-cooled Plan 52 and 53 systems are beginning to appear. Air coolers must be effective with the limited flow and head available in the typical thermosyphon loop. Successful air coolers could do away with the complexity and expense of water cooling, which in some services must be a oncethrough water service. An added benefit of air is that there is no possibility of water and process cross-contamination. W a t e r ingress into hot processes can be very dangerous because the water will flash to saturation pressure.

•

Plans 53 and 54 have typically used gas as the pressure source. In high performance applications, gas pick-up in the barrier fluid can hurt system performance. Conditions can lead to barrier fluid foaming due to the temperature swings in the loop, and dissolved gases can also compromise face lubrication in high temperature and pressure seals. Two types of gas-free systems have appeared. One uses a back pressure regulator valve on the discharge of a positive displacement pump to pressurize the barrier fluid. The other uses an air driven piston pump that is dead-ended in “stall” mode with a precision regulator providing static

Carefully inspect the barrier liquid porting at the seal. Unfortunately, many seals have severely restricted porting and external fittings. The loop flow on these seals will be a fraction of the desired flow. Also, seal glands should be ported tangentially whenever possible. Test results show a 4x increase in flow over radially ported seals. API 682 will address this by requiring minimum port diameters.

MODIFICATIONS TO THE PLANS There are new and interesting modifications to the classic plans. These have been developed by OEMs and creative end-users. They deserve a quick review. • Single seals in emission sensitive services have been modified by adding a flush distribution network internal to the seal. These new high performance seals route the Plan 11 fluid to the seal gland instead of to the existing pump port. The fluid is then introduced evenly around the circumference of the seal faces. It has been proved that even distribution of fluid reduces thermal face distorThe Pump Handbook Series

7

•

8

response to pressure bias changes. pressure control. Both concepts The use of a dome-loaded regulaeliminate gas entrainment in the tor, in either conventional or gasbarrier fluid. The all-liquid sysless systems, enables tems have the addithe barrier pressure to tional benefit of track box pressure replenishing lost fluid and remain at a fixed automatically, and the Every bias above it. The level switch is on the resulting stability at atmospheric reservoir, situation the inboard faces can where it should be greatly improve high easier to maintain. deserves a pressure seal perfor“Dome Loaded” regumance. detailed lators offer special advantages on high CONCLUSION review of pressure double seal Seal “environmental installations. Where heat flow controls” range from simthere are varying ple to complex, and each stuffing box pressures, and concept has its strengths the barrier pressure and limitations. Careful optimum must be fixed at a analysis and application of value above the hightemperature. sound practices will yield est pressure expected. measurable benefits for As box pressures the seal user. change, the pressure Review your seal probbias across the critical lems carefully, and let seal inboard seal will vary, usually manufacturers be a key resource for increasing leakage and face wear experience and recommendations. Wellas the faces deform slightly in

The Pump Handbook Series

designed environmental controls may be your most economical and reliable way of maximizing pump and seal life while achieving process containment and improving plant safety. ■ Coker Golden is President of Spectrum Process Inc., a manufacturer of modular process systems in Greensboro, NC. He received a mechanical engineering degree from North Carolina State University in 1972 and holds several patents for mechanical seals.

MECHANICAL SEALS & SEALING SYSTEMS HANDBOOK

Mechanical Seals Keep Pace With User Demands missions. Worker safety. Operating costs. Every plant wants to improve all three. And all three depend on mechanical seals. In response, seal manufacturers are employing new materials and new designs to improve the performance of single and double mechanical seals.

E

SEARCH FOR THE IDEAL FACE MATERIAL Ideally, seal faces would be constructed from materials that are selflubricating, could run at any speed and any pressure with no wear and no change in flatness,” says Bill Boro, marketing manager for Calnevar Seals (Oxnard, CA). “No one has found that perfect material yet, but manufacturers are getting closer.” One set of materials commonly used is siliconized graphite running against silicon carbide (SiC). SiC is not new, but bonding SiC to parts is now cost effective because of a new process — reaction bonding. The pairing of graphitized compounds and SiC or tungsten carbide yields hard faces resistant to chemicals and providing good lubricity. Most major manufacturers offer seals constructed of these compounds. Of course, these materials are not ideal for every application. Manufacturers offer a wide range of material, one of which will match the needs of most any application. Ordinary materials such as stainless steel or various elastomers still have their place, as do exotic materials such as titanium nitrate. The wide availability of high performance materials is pushing the limits of seal designs.

SINGLE SEAL DESIGN The basic design of a single mechanical seal hasn’t changed dramatically. Single mechanical seals consist of a rotating and a stationary seal with face pressure maintained by a spring or metal bellows. Several years ago manufacturers encapsulated the components into a cartridge to simplify installation and developed radially split designs to facilitate installation in large, splitcase pumps. But the basic design remained the same.

New designs of single mechanical seals represent fine tuning. For example, Flex-a-Seal (Essex Junction, VT) recently patented a design that replaces O-rings with self-aligning metal bellows on both sides of the face. “Now users don’t have to worry about a $1,500 seal failing because of a 15 cent O-ring,” says Flex-a-Seal’s Greg Campbell. A.W. Chesterton (Stoneham, MA) recently introduced a patented alignment mechanism to prevent off-center mounting. “Cartridge seals can be hard to align, but the Self-Centering Lock Ring eliminates the problem,” says Chesterton’s Doug Bridge. Manufacturers have also turned to computer methods to optimize the design of seal faces. For example, Finite Element Analysis lets manufacturers model the pressures and temperatures on a seal face. Then designers can test their ideas on the computer and find the design that minimizes thermal distortion, pressure distortion and heat generation. The result: seal faces that stay flat and parallel and therefore minimize wear and leakage. These and other design features have improved single mechanical seals. Manufacturers report MTBF (mean time between failure) in excess of three years and are working toward five year unattended life. And single mechanical seals can achieve emissions below 500 ppm routinely, with some manufacturers claiming 100 ppm emissions for their single mechanical seals.

DUAL SEAL ARRANGEMENTS For applications that demand a safety backup or near zero emissions, single mechanical seals aren’t enough. The solution in both cases is to add a secondary seal. If the secondary seal is rated for the temperature, pressure and other operating conditions of the process, it can be thought of as an installed spare in case of failure of the primary seal. As one user in a Texas Gulf coast ethylene plant says, “Even if it weren’t for the new emissions regulations, we’d still be using dual seals for worker safety reasons.” The Pump Handbook Series

For many users emission control is the major reason for installing dual seals. To handle emissions from the primary seal, the space between the seals can contain a barrier fluid or be vented to a flare or other vapor handling system. In systems with a barrier fluid, the fluid can be below the pressure of the process fluid or at a greater pressure than the process fluid. In nonpressurized systems, the primary and secondary seal face the same direction (i.e., front-to-back in a tandem arrangement). Leakage from the primary system follows the pressure gradient, but then the barrier fluid must be handled as a hazardous material. In pressurized systems, the seals face each other in a double seal arrangement. The pressure gradient prevents leakage of process fluid but forces a small amount of barrier fluid into the process. A pressurized system lets no hazardous material into the barrier fluid, but the pressurization system is more complex and the barrier fluid must be compatible with the process. A recent innovation from John Crane Inc. (Morton Grove, IL) achieves the advantage of forcing the barrier fluid into the process but doesn’t require a pressurized barrier fluid. This design, called upstream pumping, moves a small quantity of barrier fluid from the low pressure side of the seal to the high pressure process side of the seal. Spiral grooves in one of the seal faces provide the pumping action for the barrier fluid. The pumping action also maintains a liquid film between the two seal faces, thereby reducing wear and heat buildup. (Maintenance of liquid between the seal faces is an ongoing challenge in seal design, especially in applications for which the process liquid vaporizes at seal face temperature, contains abrasives, or otherwise is a poor lubricant.) To eliminate the need for an elaborate support system for the barrier fluid (and the attendant maintenance required), other manufacturers offer systems in which emissions from the

9

primary seal are vented to a vapor handling system. One such design features a dryrunning secondary seal. A dry-running design from EG&G Sealol (Cranston, RI) illustrates the principles involved. The secondary seal is a full-contact face seal that uses welded metal bellows with a low spring rate to maintain face loading. The face loading is controlled to maintain contact without causing undue wear. Carbon graphite resin against silicon carbide provides low friction and low wear rates. Many manufacturers offer dry running secondary seals. And users are taking note. As one user says, “Since I don’t know whether our barrier fluid might appear on an EPA hit list, I’m closely watching the progress of these designs.” Another recent innovation in dual seal arrangements is to eliminate all wear at the secondary seal by using a magnetic fluid as the secondary seal. BW/IP International, Inc. (Temecula, CA) and Ferrofluidics Corporation (Nashua, NH) have developed a magnetic secondary seal for use on API pumps. The principle of ferrofluidic sealing is fairly simple: a fluid containing magnetic particles is held in place by a magnet (Figure 1). The fluid provides a hermetic seal without wear. But the design has some limitations.

Pressure greater than 7 psi can dislodge the fluid. (But fluid can be recharged even while the pump is operating). Also temperatures greater than 250° F weaken the magnetic field, therefore high temperature applications require a liquid cooling system or more frequent recharge of the magnetic fluid. Process liquids may not be compatible with the magnetic fluids, so seal chambers must have a liquid drain. BW/IP recommends this design for applications in which a primary seal can provide adequate performance except for occasional upsets (which is the case in many applications, according to data cited by BW/IP). This point illustrates the fact that all these advances in dual arrangement seals should build on the advanced designs used in the primary seals. As Chesterton’s Doug Bridge says, “the best solution isn’t always just to catch the emissions from a primary seal whose design is 20 years old.” Seal manufacturers are making seals that last longer and leak less. Seals are no longer the weak link in the pump reliability chain. But users must remember that seals are precision parts. They require careful operations and maintenance. O&M staff must pay close attention to the seal chamber monitors and to the pump itself.

FIGURE 1

HOUSING

MAGNET POLE PIECE FLUX PATH

N

S Ferrofluid

SHAFT

Design of Magnetic Fluid Seal “As with any piece of equipment, you must be careful,” says one user at a major chemical plant in the eastern U.S. “For example, upstream pumping is a good idea, but we had one pump fail because the operations staff didn’t make sure it was primed. You can’t forget upfront engineering!” A user at another plant agrees: “New seals help us improve mean time between planned maintenance but only when we are careful about selecting the right pump for the job and impress on our operations staff the importance of closely monitoring the equipment.” ■

Advances in Seal Chamber Design Help Previous API and ANSI standards required mechanical seals to fit into stuffing boxes designed for compression packing. The need to fit into such a tight space seriously compromised seal performance due to heat buildup, inadequate cross-sectional dimensions of faces, installation and assembly difficulties (especially for dual seal arrangements). Revised standards permit enlarged seal chambers. Therefore manufacturers are researching optimal designs for seal chambers. One such project is underway at Durametallic Corporation (Kalamazoo, MI) where investigators use high speed video and laser doppler velocimetry to monitor conditions in an acrylic pump that duplicates a standard 2 x 1-10 ANSI B73.1 pump. The study, recently presented at Texas A&M University’s 10th International Pump User’s Symposium, shows which seal chamber designs are best for which applications. The study shows, for example, that a four- or fivedegree taper is best for applications that involve entrained gasses. Another conclusion is that an open,

10

enlarged seal chamber helps remove heat generated by seals. However, a larger bore can also result in a concentration of large, high-velocity particles at the corner of the gland face and chamber bore. To predict the erosion caused by these particles, Durametallic researchers developed an erosion index based on shaft diameter and speed, seal chamber dimensions, percent of solids in the fluid, size of the particles and specific gravity. This erosion index value can then be used to select the optimal shaft diameter and speed and the optimal chamber dimensions for a given fluid. The study also shows a strong link between seal life and pump design. Proper matching of seal chamber and impeller design reduces cavitation in the seal chamber and lengthens seal life. For more information, see “Enhanced mechanical seal performance through proper selection and application of enlarged-bore seal chambers,” William Adams, Richard Robinson, James Budrow, Proceedings of the 10th International Pump User Symposium, March 1993.

The Pump Handbook Series

MECHANICAL SEALS & SEALING SYSTEMS HANDBOOK

Extending Mechanical Seal Life BY JIM DOWNING ecently, there has been increasing concern about leakage and fugitive emissions from mechanical seals installed on pumps, mixers, and other rotating equipment. This is especially true in plants handling chemicals listed in the Federal Clean Air Act Amendments of 1990. Most of these plants have rotating equipment expected to operate twenty-four hours a day. Although some of this equipment operates for several years without failure, most maintenance personnel will admit to a large number of “seal failures” after only five to seven months (or less). Clearly, improvements are in order. One solution is to require equipment manufacturers to redesign their products to handle any imaginable installation or operational error (pipe strain, misalignment, cavitation, dryrunning, etc.) without damaging the mechanical seal, bearings, bushings, etc. Another is to require all installations and operations to be flawless. However, neither of these approaches is realistic. A more reasonable approach is to take elements of both and combine them using the common sense of maintenance and operations personnel. The balance of this discussion will be limited to centrifugal pumps; however, many of the recommendations are also applicable to other rotating equipment. To extend the life of mechanical seals, we must regard each seal application as unique and consider everything that can affect the seal in that application (pipe strain, alignment, etc.). The mechanical seal is only one part of a system that includes the pump, bearings, baseplate, foundation, and piping. Each part of the system is affected by everything that happens to another. For example, misalignment between the motor and the pump shaft is detrimental to the shaft coupling, motor bearings, pump bearings, and mechanical seal. Therefore, if a seal fails in an application, the entire pump system should be considered in

R

determining why the seal did not last longer. Seals fail for a variety of reasons, but most failures can be attributed to excessive heat, excessive movement, or improper application. Extended seal life can be obtained by proper application, installation, and operation. • Proper application includes using an appropriate seal configuration (single, double, tandem), materials (metal parts, faces, elastomers, or other secondary seal), and environmental controls (flush, recirculation, barrier fluid system, heating/ cooling) installed in a pump suitable for the service conditions (pump type, size, materials, shaft strength, bearing protection, and other appropriate options). • Proper installation includes inspecting the pump to find problems like a bent shaft, failed or loose bearings, an out-of-balance impeller, or a bearing housing that could allow contamination. Also included are alignment between the pump shaft and driver, using a flexible shaft coupling, and minimizing pipe strain by using supports and expansion joints. Obviously, the mechanical seal must be installed according to instructions, and the pump base and foundation must adequately support the equipment. • Proper operation means avoiding running the pump dry or with the suction or discharge valve closed, or while the pump is cavitating. If the correct size pump has been applied and installed, it will be running close to its best efficiency point (BEP). Centrifugal pumps are designed to minimize radial loading on the impeller at the BEP; however, radial forces increase as the pump is operated farther away from its BEP (more or less flow, Figure 1). These forces can run into the thousands of pounds as the pump approaches zero flow (shut-off or deadheading). High radial The Pump Handbook Series

loads can cause increased shaft deflection. ANSI B73.1M-1984 paragraph 4.5.4 specified the maximum dynamic shaft deflection at the stuffing box face to be 0.002 in. (0.05 mm) at “maximum load” for the A70 and smaller pumps and at “design load” for A80 and larger pumps (with a 1.0 specific gravity liquid). This has been changed slightly in ANSI B73.1M1991, which specifies maximum allowable deflection at the impeller centerline of 0.005 in. (0.13 mm). Yet it is not uncommon to see removable and non-removable shaft sleeves scored from contact with the “throat” of the packing stuffing box (where the typical clearance is on the order of 0.040 in. [1 mm] or more). Either there are many high specific gravity liquids being pumped, or in many pumps these specifications are only met with solid shafts. Most manufacturers of ANSI pumps now recommend solid shafts to achieve longer seal life. Many brochures, manuals, and trade articles outline the use of the ratio L3/D4 to evaluate and compare

Photo 1. Tapered seal chamber with optional bushing.

11

insert, or stationary face) fracture them. In either event, liquid is now the rotating face. or vapors leak. This minimal clearThis concept has been ance also results in little liquid volused for years in other ume available to keep seal faces cool, industries for higher rpm clean, and lubricated. or other critical or tough Because of these problems, manapplications. ufacturers have made oversize stuffBest Efficiency Point When the bellows is ing boxes or seal chamber designs (B.E.P.) in the rotating part of the available (Photo 1). A study by seal, the rotating face must Barnes, Fitney, and Nau of BHR oscillate back and forth Group Ltd. reports on tests of conwith each shaft revolution. ventional stuffing boxes and various The stationary design configurations of seal chambers. RADIAL offers nonoscillating alignThe study discusses problems FORCE ment of seal faces and found with conventional stuffing CURVE automatically compenboxes and includes the finding that a sates for shaft deflection “flared housing enhances heat transand misalignment fer and removal of gas, vapor, and between the shaft and seal solids, significantly improving perforchamber face or gland. mance relative to all other housing 0 CAPACITY These stationary seals can designs tested.” The authors also cauhandle a higher L 3 /D 4 tion that “vortex modifiers” are needed in some applications involving Radial force and total dynamic head vs. capacity. ratio, perhaps as high as 75, and still last a long solid particles.1 time. Seal faces are usually made of pumps for specific service. This ratio Because many manufacturers carbon-graphite, aluminum oxide has been called the shaft flexibility fachave only recently made this design ceramic, tungsten carbide, and silitor because the higher the number, available, and because of increascon carbide. Relatively brittle, these the more flexible the shaft and the ed demand for cartridge-mountmaterials can’t withstand much tengreater the deflection. It is used pried seals, most of these designs are sile force, but they can accept commarily for pumps with overhung also cartridge mounted. Cartridgeparatively much higher compressive impellers (such as ANSI pumps and mounted seals are preferred by most forces. It would therefore be logical most other process pumps). L is the customers due to the simdistance from the center of the plicity of installation and FIGURE 2 impeller to the center of the first bearthe protection afforded the ing, and D is the diameter of the shaft seal faces during handling DIAMETER in the seal area (Figure 2). If the shaft D = SHAFT and installation. Simply (NOT SLEEVE DIAMETER) has been cut down and a sleeve used, switching from “compoD is the diameter of the shaft under nent” seals to cartridgethe sleeve (not the sleeve). Obviously, mounted seals has the ratio will be smaller for a solid increased the mean time shaft, and the smaller the ratio the between failure (MTBF) less the shaft deflection will be for a of pumps in many plants, specific set of conditions. Although as has switching from there is not an ANSI spec for the rotating seals to stationary maximum allowable ratio, a number seals. of industry engineers have suggested Although shaft dethat a maximum ratio of 50 is preflection may be minimal L ferred to achieve longer life with at the stuffing box face, it most seals. is much higher inside the Most major seal manufacturers stuffing box, where most Shaft length (L) and shaft diameter (D) defined for have recognized the advantages of a seals are located. The the ratio L3/D4 for pump service evaluation. seal design to handle shaft deflection clearance between the (and gland or stuffing box face misO.D. of the seal and the alignment). This seal design is comI.D. of the stuffing box is typically to place these faces in compressionmonly called the “stationary” design 0.062 inches (1.6 mm) or less. loaded situations. because the springs or bellows do Contact between the seal and the Brittle materials such as these are not rotate with the shaft (Figure 3). stuffing box or build-up in the stuffalso sensitive to stress risers like It is also known as the “rotating seat” ing box is not uncommon. Contact notches, slots, and holes. Most maindesign because the larger seal face can shock the faces out of flat or even tenance personnel involved with (in previous designs the “seat,” mechanical seals know that if a face RADIAL FORCE IN LBS. TOTAL DYNAMIC HEAD

FIGURE 1

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The Pump Handbook Series

FIGURE 3 1. Prior to start-up springs adjust for stuffing box face misalignment.

2. At start-up springs adjust for shaft deflection.

3. During Operation - No spring movement except to compensate for shaft end play or additional shaft deflection. Resulting In: • NO VIBRATION • NO WOBBLE • REDUCED EMISSIONS • LESS FACE WEAR • INCREASED LIFE

Spring loaded stationary face. has broken in service, it broke at a stress riser. Stress risers are especially vulnerable to impacts at start-up, in intermittent service, and in services involving liquids that set up or are sticky or viscous. Common sense would say to avoid seal designs that put faces in tension or have stress risers. Any effort to extend the life of mechanical seals should bring up the question “Which failed first, the seal or the bearings?” Failure of one can lead to failure of the other. Fluid from a leaking mechanical seal can contaminate the lubricant and damage the bearings. Tests have shown that lubricant contaminated with 0.002% (200 ppm) water causes a reduction in bearing life by 48% and contamination to only 6% causes some 83% reduction. 2 Contaminants more corrosive or abrasive than water reduce bearing life even faster. Bearing lubricant can also be contaminated with wash-down water as workers clean the area or from leaks above the pump. These liquids can enter through the breather in many pumps, but the primary entry path is between the shaft O.D. and the bearing housing.

Some pumps utilizing grease lubricated bearings have no sealing protection at this opening, and others have just a shaft mounted flinger device. However, most pumps use an elastomeric lip seal to seal this opening. Manufacturers of these seals have published expected life as two to three months or less. For example, page 27A of Catalog 4570101, Handbook of Seals, published by CR Services (Elgin, IL), states, “Under optimum conditions, an average conventional seal can run 1,350 hours to failure while Waveseal® can run 3,000 hours to failure under the same operating conditions.” Under conditions found in the real world, the lip seals can fret or score the shafts in as little as two weeks (336 hours) of operation. As this occurs, bearings are unprotected and contaminants move freely between the worn lip seal and the damaged shaft. Many engineers and maintenance personnel long ago recognized the need for improved bearing protection. For example, in 1984 a major pump manufacturer contracted with a bearing protector manufacturer to stock protectors to make them an option to standard lip seals. The early designs were two-piece; however, manufacturers have recently developed a “cartridge” design to ease installation and handle axial shaft movement without separating. (To prevent electric motors from failing due to contaminated bearings, also consider bearing protectors on motors.) A common sense approach involves looking at all components of the pump system, not just the mechanical seal. This approach can include installation of cartridge bearing protectors in pump bearing housing and in electric motor end bells. A solid shaft should be specified to ensure the L3/D4 ratio is no higher than 50 (75 if a stationary seal is used), and shaft runout ratio and end play should be minimized. Rotating elements should be balanced to minimize vibration. Cartridge-mounted stationary-design mechanical seals can be installed with appropriate environmental controls in a flared/conical/ tapered seal chamber (with vortex The Pump Handbook Series

modifier if solids are present). A flexible shaft coupling between pump and driver can be used and the shaft can be properly aligned. External loading on the equipment can be reduced by minimizing pipe strain and ensuring the equipment base is adequate. Attention to these items is a reasonable approach to extending the life of each component of the pump system, including the mechanical seal.

REFERENCES 1. N.D. Barnes, R.K. Flitney, and B.S. Nau. Mechanical Seal Chamber Design for Improved Performance. Proceedings of the Ninth International Pump Users Symposium, Texas A&M University, Houston, Texas, March 1992, pp 61–68. 2. E.L. Armstrong, W.R. Murphy, and P.S. Wooding. Evaluation of Water-Accelerated Bearing Fatigue in Oil-Lubricated Ball Bearings. Lubrication Engineering, Vol. 34, No. 1, 1977, pp. 15–21.

ACKNOWLEDGEMENT The author wishes to acknowledge the consultation of Robert J. Russel, Ph.D., metallurgical engineering, Purdue University, and C.O.O. of Five Star Seal, on the subject of brittle seal faces discussed in this paper. ■ Jim Downing is Vice President— O.E.M. Sales and Technical Support for Five Star Seal Corporation. He received his B.S. in Mechanical Engineering from Larmar University in 1970 and his MBA from the University of Dayton in 1988. Jim has 23 years of experience with pumps and seals as a maintenance engineer, sales engineer, product manager, and sales, training, and technical support.

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MECHANICAL SEALS & SEALING SYSTEMS HANDBOOK

Effect of Bearing Performance on Seal Life BY BOB GOODENBERGER echanical seals play a role in containing emissions as well as providing a safe environment for plant personnel. Proper selection, operation, and maintenance of the seal system are critical to good seal performance. Pump dynamics and bearing performance will also have an effect on seal performance.

M

SEAL DESIGN Seal designs account for temperature, pressure, corrosion, wear, and normal pump dynamics. All of these parameters must be within normal limits for the seal to operate satisfactorily. Pump dynamics include shaft movements—both radial vibration and axial movement (Figure 1)—and temperature changes. Temperature changes should be limited to normal process temperature swings. Temperature transients—swings from ambient to high temperature—should be avoided whenever possible, because this can cause thermal distortions and lead to wear and leakage.

housing and its support. Because the seal is closer to the applied load, it sees more movement than the bearing. By the time the bearing begins to fail, the seal may have been subjected to movements that exceed its normal design. There are many different causes for the loads on the bearings, and many are hard to detect. Seal leakage, however, is easily detected. This results in the identification of the seal as being bad, and no further investigation is performed to see if the leakage has some underlying cause.

lubrication all result in the loss of lubrication. Any of these conditions can cause the bearing to fail, which may result in seal leakage and the need for pump maintenance. Bearing overload is any load imposed an the bearing that is greater than the design load of the bearing. This load can be due to constant loading or continuously changing loads. Overload of the thrust bearing can be caused by changes in the flow, which changes the hydraulic pressure profile on the pump, or by changes to the pump hydraulics themselves.

LUBRICATION

PUMP WEAR RINGS

The major cause of bearing failure is loss of lubrication. This sounds simplistic, and it is. Ultimately bearing overload, underload, and improper

One of the least understood causes of thrust overload is the maintenance of pump wear rings. Often when pumps are maintained, the

FIGURE 1

PUMP DYNAMICS There are many times, however, that pump dynamics are not as stable as the pump and seal designers would like. The result is that seal performance is affected, even though there may be other areas of the pump that are beginning to fail but are not recognized because the seal is more obvious. The more common dynamics that cause seals to leak are vibration, pressure pulsations, and shaft movements. Any of these can directly cause the seal to leak. They may also cause the bearings to weaken and fail, which in turn causes the seal to leak or be damaged.

MECHANICAL SEAL RADIAL BEARING

CASE SIDE WEAR RINGS

HEAD SIDE WEAR RINGS

BEARINGS AND LOADS The bearings of a pump are the major support element for the rotor, and as such are also the major support element for the seal. When the shaft experiences a process load, either radial or axial, the force must travel past the seal before getting to the bearings, where the load is transferred to the

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A typical overhung pump, without bearing bracket.

The Pump Handbook Series

THRUST BEARING

wear rings need to be replaced. It often seems to be cost effective to machine one wear ring to clean up any wear, and to manufacture a mating ring and machine it to match. However, this changes the hydraulic balance of the pump and may overload the bearings, depending on the original loads on them and the new loads caused by the changed rings. An example of this wear-ringinduced load would be a pump with 300 psi differential pressure and an eye wear ring diameter of 6 inches. If over a period of years the case ring were cleaned up several times and the impeller ring were replaced to maintain clearance, the resulting diameter could be as large as 6.25 inches. If no other changes were made, this enlargement would result in a thrust change of 721 pounds. This could very well overload the thrust bearing, which could result in vibration or axial displacement of the shaft. Either of these could cause the seal to be less effective, and a “seal failure” would be the diagnosis; but the root cause would not be recognized.

PROCESS CONTROL One of the major causes of radial bearing overload is process control. If a pump is being controlled by a level either on the suction or the discharge side of the pump, there are often times when the flow through the pump is reduced below the minimum stable flow. This results in large radial loads due to high pressures on one side of the impeller. If this condition happens frequently or for extended periods of time, the radial bearing can be overloaded and fail. In this case, however, the radial displacement at the seal is greater than the displacement at the bearing, and the seal often drips or is damaged before the bearing failure is obvious.

FIGURE 2 RADIAL FORCES

AXIAL FORCES

Axial and radial forces

CAVITATION Cavitation is another cause of bearing wear that affects the seal, both directly and indirectly. Cavitation causes high-frequency pressure pulsations that destabilize the seal. This creates loads that are transmitted along the shaft to the bearings. Continuous cavitation in a high energy pump can cause the bearings to fail due to constantly shifting or high amplitude loads. These loads usually affect the radial bearing, but on some impeller designs the thrust bearing can also be affected.

SUMMARY There are many influences in a pumping system that cause seal leakage, and many times these have little to do with the seal itself. In an effort to increase the mean time between maintenance (MTMB), the investigation into the cause of seal failure often needs to go beyond the seal itself. The

The Pump Handbook Series

hidden causes need to be understood and corrected. Maintenance programs that are designed to find and eliminate the root cause of failures can reduce overall plant emissions, while at the same time increasing MTBM and decreasing cost and downtime. Bob Goodenberger is a regional engineer for John Crane Inc. in Houston, Texas. He is a graduate of Northern Arizona University and received his Bachelor of Science degree in mechanical engineering in 1980. Upon graduation Bob went to work for a major petrochemical manufacturer in the Houston area as a general engineer and then as a rotating equipment engineer. As a rotating equipment engineer Bob has been involved in all aspects of pump installation and maintenance. Bob is a member of the American Society of Mechanical Engineers (ASME) and the Society of Tribologists and Lubrication Engineers (STLE).

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MECHANICAL SEALS & SEALING SYSTEMS HANDBOOK

Magnetic Liquid Seals Contain VOCs BY WILL MRAZ ero leakage rotary shaft seals based on magnetic liquid technology are used successfully in sealing high vacuum in semiconductor processing equipment to exclude contaminants from billions of computer disk drives. The proven features of magnetic liquid sealing (zero leakage, long life, and low cost) seemed to be an obvious answer for eliminating emissions of VOCs, but only if the technology could be adapted to meet the unique challenges found in refinery and chemical processing industries. In June of 1990 Ferrofluidics Corporation, BW/IP Inc., and Chevron Research and Technology Co. initiated a multiple-phase program to pursue the application of Ferrofluidic® (magnetic liquid) sealing technology for use in a real-world refinery installation. This program ultimately developed a sealing system that exceeds current and future regulations of the 1990 Clean Air Act pertaining to the emission of VOCs and volatile hazardous air pollutants from process pumps. This new sealing system utilizes a ferrofluid (magnetic fluid) to form a “liquid O-ring” that maintains a perfect seal while in contact with the rotating shaft and stationary magnetic structure at normal operating speeds and accommodates typical pump shaft run out (Figure 1). This seal is part of a dual seal system that uses a conventional mechanical seal as the primary liquid seal. Vapors leaking past the mechanical seal are contained by the magnetic liquid seal and are vented through a port in the mechanical seal gland for appropriate disposal. The magnetic liquid seal virtually eliminates the escape of all fugitive emissions. The design challenge was to incorporate the well-documented zero leakage features of magnetic liquid seals into a package that would satisfy the following additional requirements: • retrofit onto existing pumps to protect the significant investment in

Z

16

the installed base of process pumps • provide a minimum 2–3 year service life to reduce costs of maintenance • have field replaceable parts and fluids to minimize downtime for repairs • lower the cost of compliance to provide an alternative to costly magnetic drive pumps and complex tandem sealing arrangements Before committing to the project of developing this new seal, BW/IP engineers conducted a proof-of-principle test to verify the performance of magnetic seals at speeds, pressures, and temperatures typical in refining pumps.

A commercially available Ferrofluidic hollow-shaft sealed spindle was installed in a dual chamber test configuration as shown in Figure 2. The sealing conditions of a typical process pump were duplicated by pressurizing the process chamber with propane gas at 240 psig. Vapors escaping past this primary mechanical seal were contained in the Inter Seal Chamber and applied to the magnetic liquid seal. The Ferrofluidic seal was designed for vacuum service and had a rated sealing capacity of 40 psig, so to prevent an over-pressure condition, the pressure in the Inter Seal Chamber was kept below 26 psig. The shaft was rotated at 3600 rpm and VOC sniffer readings were taken at the magnetic liquid seal/shaft

FIGURE 1

Housing Pole Piece N S

Magnet

Ferrofluid Flux Path Shaft

Magnetic liquid O-rings around the shaft produce a nonwearing, maintenance-free seal The Pump Handbook Series

FIGURE 2 Inter Seal Chamber

Propane 240 PSIG 110°F

Sniffer Monitoring Point

Magnetic Liquid Seal Assembly

Primary Mechanical Seal

Shaft Speed 3600 RPM

Mounting a magnetic liquid secondary seal behind the mechanical seal eliminated VOC emissions

junction at 30-minute intervals. During a 99 hour test, no detectable VOC emissions were recorded. This proved the feasibility of using magnetic fluid seals on light hydrocarbon vapors, even though the seal assembly used in the test was too large to fit on the typical process pump. Also, due to the speed limitations of the close-toleranced Ferrofluidic seal and ball bearings, the seal assembly needed to be water cooled, and the unit could not be field recharged with magnetic fluid. Despite these limitations, early results convinced the companies involved that a magnetic liquid seal system could trap fugitive emissions. The final design incorporates the basic magnetic liquid seal in a compact configuration that has a width of about .75 in. and a diameter equal to that of the seal gland plate of the primary mechanical seal to which it attaches (Figure 3). This low-profile seal features a patent pending selfcentering provision that ensures seal/ shaft concentricity and allows the seal to be easily installed on most process pumps conforming to API 610. Refilling or recharging the seal with additional magnetic liquid is easily accomplished, even while the shaft is rotating at full speed. The design has been tested at speeds to 55 ft/sec with no loss in sealing capability. Projected maximum speed is estimated to be greater than 100 ft/sec.

The temperature limits of the seal are compatible with typical light hydrocarbon processes. The magnetic seal can be run continuously at 200°F for extended periods with no loss of sealing capability. Higher temperatures can be accommodated with water

cooling schemes or special ferrofluids. In December 1991 this seal was installed on a United centrifugal process pump at the Chevron Products Company refinery in El Segundo, CA. The pump, which has been in continuous service since that date, pumps sour butane and has been inspected via EPA test method 21 quarterly (3 times) for hydrocarbon emissions according to the South Coast Air Quality Manage-ment District (SCAQMD) Rule 1173. Readings for these formal tests were 0, 20, and 10 ppm hydrocarbon leakage against a background concentration of residual hydrocarbons that often exceeds 20 ppm. Extensive monitoring by the refinery engineers verified that these readings have remained below the 50 ppm level that the EPA considers to be “zero leakage”. The field tests provided real world conditions that allowed performance monitoring and validated the following design goals: • easy installation and self centering • easy fluid fill and refill

FIGURE 3 Mechanical Seal Gland

Vent

Primary Mechanical Seal

Bushing

Secondary Ferrofluidic Seal

Drain

The magnetic liquid seal installs on the seal gland of the primary mechanical seal

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(EMSTM), through their extensive worldwide sales force. Initial market acceptance has been positive, and many refineries and chemical processing plants have successfully evaluated their first units and are committing to additional installations to coincide with normal maintenance schedules. Chevron Products has initially specified these magnetic liquid sealing systems for use in 25 applications in some of their eight refineries and seven chemical processing facilities. Ferrofluidics Corporation is engineering and modifying these specialty sealing systems for nonpumping applications. While the initial design efforts focused on pump sealing, Zero-leakage magnetic liquid seal it was recognized that other developed to retrofit process pumps rotating equipment could benefit from zero leakage sealing. Compressors, blowers, and • reliability that exceeds one year other gas-handling units typically uti• virtually zero VOC emissions lize packing configurations or dry detected running mechanical seals and are Results of both the lab and field obvious candidates for variations of tests convinced the companies the nonwearing magnetic liquid seal. involved that magnetic liquid seals Such equipment often experiences could provide significant benefits to higher differential pressures than the the petroleum and chemical processEMS seal was designed for and may ing industries. BW/IP International require multiple magnetic seals in (BW Seals) decided to introduce the series for added pressure capacity. new sealing concept, designated the For high torque/ high pressure sealing Emission Management System Seal applications such as in mixers and

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The Pump Handbook Series

agitators, the magnetic seal can be incorporated in a hollow-shaft sealed spindle with heavy-duty bearings. Because of the inherent flexibility of Ferrofluidic sealing technology, zero leakage seals can be installed in demanding applications where other sealing methods do not deliver the required performance. The transfer of magnetic liquid sealing technology to pumping applications has been accomplished smoothly and rapidly. The opportunities for applying this new sealing technology to other equipment that must comply with VOC emissions regulations are extensive throughout the refinery and chemical processing industries. As equipment designers and users gain more understanding of the capabilities of magnetic liquid sealing, this technology could become the most cost effective method of preventing VOC emissions. ■ Will Mraz is Engineering Manager of the Components Division (rotary seals) of Ferrofluidics Corporation. His primary area of responsibility is in new market applications for magnetic liquid sealing technology. He joined the company in 1989 and was previously employed as a Design Engineer at Sikorsky Aircraft. Mr. Mraz received his B.S.M.E. from Clarkson University and is the author of many technical papers and articles on Ferrofluidic sealing technology.

MECHANICAL SEALS & SEALING SYSTEMS HANDBOOK

Just What Is a “Totally Engineered Sealing System”? An end-user’s answer. BY RICHARD L. ENLOW THE QUESTION The proposed API-682 document, Shaft Sealing Systems for Centrifugal and Rotary Pumps, uses the terminology “totally engineered sealing system” (TESS). I’ve asked numerous representatives from major seal and API pump manufacturers to define a TESS. Their answers have left me feeling uncertain. I’ve also attended two API-682 seminars, but these haven’t given me any new insight into a TESS. Because I’ve been frustrated and believe others are also, I offer my answer to the question, “What is a totally engineered sealing system?”

ENGINEERED SEAL Today’s technology has made great strides into computer engineered seal designs (see Pumps & Systems, January 1993.) However, the use of a computer model to design a seal is one thing, validating the model is another. API-682 will require seal manufacturers to qualify seals by using controlled test environments, which will ultimately validate the computer models. Seals which fall out of the API-682 scope will require special engineering. The purchaser can then pay for a qualification test if desired. Is a properly engineered seal alone a TESS? No!

ENGINEERED PUMPS Manufacturers provide pumps that meet or exceed their generic pump curves. Yet pump salesmen typically don’t care what seal goes into the pump. The choice is whatever the specifier wants. As a specifier, I always have something to say about seal designs. However, I don’t always know about the pump’s true performance characteristics, the final piping arrangement, the final throttle bushing design, or the orifice design. In the past, a pump manufacturer would specify the pump’s minimum flow as the calculated “temperature rise –

minimum flow” (see “Ask the Experts,” Pumps and Systems, March 1994). The industry is now wiser and has introduced concepts such as minimum continuous stable flow (MCSF), suction and discharge recirculation, and suction specific speed. The bottom line is that pump manufacturers can no longer claim pump performance ranges as wide as in the past. Can this affect seal life? Should this be considered in the TESS? Yes.

THE SEAL ENVIRONMENT The best engineered seal is often laid to waste because someone failed to recognize that a good chamber environment is necessary for long seal life. API-682 will define sealing systems to have a high probability for three years life. Here are some typical scenarios that shorten seal life: Example 1. The API-610 data sheet says “solids.” The solution: use a cyclone. The probable end result: the cyclone is removed due to erosion or clogging. Someone failed to realize that cyclones must be selected carefully and are successful only within certain parameters. Gone with the cyclone are the small orifice ports, which may have regulated the seal chamber pressure and flow. Subsequently, the seal becomes a repeat offender. Example 2. The user specifies a normal temperature range of 135°F through 150°F. The specifier fails to consider the start-up and upset conditions. Subsequently, the seal fails. Example 3. The seal manufacturer’s drawing calls for 3 gpm flush. The system’s designer uses 3 gpm to size the cooler but does nothing to regulate the flow. Subsequently, the seal fails. Example 4. The seal manufacturer writes in small print on the seal drawing, “Flush must be 25 psig above the vapor pressure.” The project designer fails to recognize and ensure this condition. Subsequently, the seal fails. The Pump Handbook Series

Are these items part of the TESS? Yes! Enough examples.

INSTRUMENTATION When projects are installed, equipment from the original equipment manufacturer (OEM) may lack documentation. Instruments seldom arrive with datasheets detailing the selection (i.e., ISA datasheets). Troubleshooting is difficult without information. Consequently, a large portion of the end-user’s time is spent trying to reconstruct what instrument was supplied and why. Just ask a question about a pressure switch setpoint, an orifice size selection, a pressure relief setting or corrosion protection. Often, the answers you’ll get neglect basic engineering design principles and can be summed up by, “Whatever the user wants!” Is instrumentation part of the TESS? Absolutely!

A TYPICAL PROCUREMENT Dateline: Years ago...A responsible project designer needs a pump. He fills out the necessary datasheet. The seal must be from manufacturer A or B to minimize stocking and maximize vendor response. (These are some good points!) The pump OEM is then selected through competitive bids. The OEM provides the seal’s model number, price and delivery. The quote includes the standard seal accessories. The designer believes the pump salesman will take care of the pump and seal. Pump, yes. Seal, no. The OEMs provide only what the contract specifies. Even then, the purchaser has to hound the pump salesman to meet the job specifications. What about the seal environment? The pump salesman places the seal order. To fill the order, the seal is pulled off the shelf or manufactured. During this period, the pump salesman has virtually no contact with the seal representative. Discussions are limited to physical sizing and porting.

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Finally, the seal is delivered to the pump OEM. Included is a seal drawing and maybe a parts list. It is placed on the shelf until use. Ditto for seal accessories. Nobody notifies the local sales rep that the seal has been sold. The pump undergoes performance testing. If the seal does not physically fit, the pump is changed on the spot. After a successful run, the pump is shipped to the end-user. Minimal documentation is sent along with the equipment. Months later, the seal survives the start up (miraculously) and performs for three months. Now the fun begins. While troubleshooting, the end-user finds the meager system documentation, throws his hands up in the air, and calls the local seal representative. The rep arrives quickly and begins identifying the problems. The rep outlines several problems, “The specifier failed to mention this and that. Oh, the box pressure is not controlled. Oh, the coolers are too small. Oh, the orifice is too large. Oh, cyclones won’t work here. What you need is an external flush.” And here it comes ... “Try this seal! Company X uses it in a similar service.” Scenarios like these occur every day. Just what is a totally engineered sealing system? And who is responsible to ensure it?

RESPONSIBILITY Several responsibilities are outlined in the upcoming API-682 and API-610, 8th Edition. When followed and documented, seals will have a better chance of trouble-free operation. Specifiers need to take datasheets seriously and provide the proper process details. (These are the ones that are bringing up the next generation of specifiers.) The industry needs to address pump factors that affect the seal’s environment at MCSF and 110% BEP. Additionally, pump salesmen need to have hydraulic/seal type experts readily available to address seal installations. Likewise, seal salesmen need to have designers that really understand the pump’s effect on seal life. These items point to the industry need for better teamwork,

20

communication and clear accountability.

THE PROCESS SPECIFICATION The first step in achieving the TESS is a properly filled out datasheet. For flashing liquids, I suggest that a vapor pressure curve be included. The curve should cover pressures over the full temperature range. Show the horizontal lines representing the maximum and minimum suction pressures. In this way, you’ve created a visual aid that illustrates the environment’s sensitivity to temperature and pressure. Solids and corrosives are difficult to design for. The specifier should not be satisfied with merely the declaration that solids are present. Efforts should be made to identify the solid’s characteristics. Crystallization (such as, from caustic materials) needs to addressed at the design stages, not post-installation. To achieve a TESS, the specifier must find the best available flush stream process information. Computer technology allows process parameters to be quickly estimated. Curves may be generated for viscosity, specific gravity, and pressures as they vary with temperature. Unfortunately, most specifiers are satisfied with “normal operating” parameters and don’t seek more information.

SEAL DESIGN SPECIFICATION As an end-user, my expertise is not in seal design. Yet like some users, I want to stretch my knowledge and search for information about a seal’s design limitation. Some end-users believe that someone somewhere is taking care of the seal. API682 will increase this confidence. Even so, the seal is selected by a person, whether it is a specifier, enduser, or pump manufacturer. Surely, success is limited by the selector’s seal knowledge. Procuring a correct engineered seal is the second step in a TESS. Recently, I issued several pump specifications that requested everything under the sun in seal design. Here are some requested items (I know it is extensive and technical, but just what is an engineered seal?):

The Pump Handbook Series

•

seal contact profiles

•

film thickness profiles

•

expected face deformation

•

leakage rate

•

contact pressures

•

face pressure profiles

•

balance ratios and shifts

•

seal chamber bulk temperature

•

seal chamber bulk pressure

•

face temperature plots

•

fluid thermodynamic plots

•

barrier fluid cavity temperature rise

•

barrier fluid tank temperature rise

Did I receive all that was requested? No, but I did not expect to. Did I understand everything that I received? No. However, what amazed me was that some seal vendors proudly explained how they could theoretically examine a seal and even described their modeling accuracy. Some have offered qualification tests to prove their modeling. On the other hand, some vendors did not read the specification, did not take it seriously, and did not allow adequate time to address the requests. Subsequently, I’m not working with these vendors. Even the successful vendor needs prompting, if you truly want information about the seal. Vendors may neglect previous promises once a manufacturing schedule is in place. I feel that vendors think I am not serious or not paying attention. Wrong. I am not recommending that seal specifications request massive amounts of technical plots and calculations, but I encourage end-users to establish minimum design documentation requirements for seals. Technology is growing.

THE PUMP SPECIFICATION The third step for a successful TESS is to understand the hydraulic and mechanical limitations for the selected pump. The specifier must find pump information that can affect the seal life. As a starting point, use the API-610 datasheet format.

NO MAN’S LAND—ACCESSORIES AND ENVIRONMENTAL CONTROLS The fourth step for a TESS is to bridge the pump and seal. Who is responsible for this coordination? This is not about who supplies the physical components, rather it is about who is responsible for selecting the components to provide the environment as specified by the seal OEM. Is it the pump vendor, seal manufacturer, specifier or the end-user? I believe it is the pump OEM. In September 1993, I brought together seal and pump reps for one of my projects. I fully described my expectations for seal environment documentation. The pump used an API Plan 11 for a low vapor pressure product, which was fairly straightforward. After several months, I have yet to be impressed with their efforts and capabilities. Because end-users have not asked for this type of documentation, pump manufacturers don’t have it. Modeling hydraulic losses can be completed even with standard spreadsheets, but have yet to be developed. These won’t be developed unless specifiers request that supporting documentation becomes part of the totally engineered sealing system. One of my favorite questions to ask seal and pump manufacturers is, “On seal pots and dry running seal vent systems, what will the vent orifice diameter be and why?” The typical answer is, “Whatever the client wants.” I interpret this as “I don’t have the foggiest idea!” If that’s the case, why don’t they just say so. Another question to ask is, “To what extent must seal leakage progress before the pressure alarm goes off on dual secondary sealing

systems?” I usually don’t receive an answer. Perhaps, the answer doesn’t matter. Completed documentation that supports accessory selection and sizing should be the primary responsibility of the pump OEM (only if they are procuring all components). The OEM, in turn, may request documentation from the seal manufacturer. For example, the seal manufacturer may provide the pump manufacturer with information about the theoretical floating throat bushing, orifice, and seal entrance losses. The exchanger manufacturer may provide the exchanger cooling calculations. Taken together, this information may then be used to complete the sealing system modeling. Modeling is also a good way to show that someone looked at the sealing system. But, it will take several years to achieve accurate models. Model verification will be required.

VERIFYING THE ENVIRONMENT MODELING DURING THE PERFORMANCE TEST Many pump orders require a performance test using job seal components. Why not use the same test to validate a particular hydraulic model? Suppose an API Plan 11 model was used to size the seal components for a hydrocarbon stream. By drilling one additional port on the seal gland and installing a pressure gauge, the test stand seal chamber pressure can easily be recorded. Then by substituting the actual test water conditions, the theoretical water value can be calculated. Correlation between the theoretical water value and the actual testing raises the assurance that the

The Pump Handbook Series

seal’s hydrocarbon environment will be achieved.

TESS—DEFINED A totally engineered sealing system is one that includes all the physical components required to maintain the seal chamber environment to within the seal manufacturer’s specifications. The pump’s performance curve is considered through the complete temperature range. Calculations or selection reasoning is fully documented to support the seal design, hydraulic losses, heat transfer equipment and any other engineering design that maintains the proper environment. Instrument documentation includes detailed datasheets and calculations supporting the alarm setpoints.

CONCLUSION The term “totally engineered sealing system” will be explored in the years to come by manufacturers and specifiers alike. Manufacturers will provide seals tested to API-682. OEMs will develop computer models to raise the certainty that seal accessories will maintain the correct seal environment. Instruments will be marketed to extend the TESS concept. But, no single component can be called a totally engineered sealing system. A TESS can only occur by a team effort between the specifier and the manufacturers. ■ Richard L. Enlow, P.E., has been in the oil industry for 11 years. He is currently a Rotating Equipment Specialist with Mobil Oil Corporation, Joliet, IL.

21

MECHANICAL SEALS & SEALING SYSTEMS HANDBOOK

Off-Design Operation and Seal Performance BY WILLIAM V. ADAMS, LEE A. WALING, RANDY R. DINGMAN, AND JOSEPH C. PARKER istorically, seal users and manufacturers had to adapt mechanical seals to chambers that were originally designed for use by compression packing. The radial dimensions of the chamber cavities were established based on the needs of compression packing. Pressure in the conventional stuffing box was minimized to reduce loads on the packing and minimize leakage past the packing to atmosphere. These restrictive dimensions remained the industry standard until the late 1980s. Although such chamber conditions were best for compression packing, they did not necessarily extend the life of mechanical seals or the mean time between planned maintenance (MTBPM). Current application guides recommend use of enlarged bore seal chambers to improve MTBPM. However, more than 90% of pumps remain equipped with conventional stuffing boxes that limit the cavity space for seal application and provide poor pressure and flow circulation in the seal cavity. Earlier studies determined the dissipation of seal-generated heat in various seal chamber designs (Ref. 1) and the effects of these designs and their applications on mechanical seal performance (Ref. 2). Although seal chamber studies have advanced the longevity of mechanical seal designs and provided an environment for more creative designs, evaluation of pump operation as it relates to mechanical seal MTBPM has not been adequately addressed. This article discusses further the effects of offdesign pump operations on various types and arrangements of mechanical seals and seal chamber designs. Past evaluations (Ref. 1 and 2) were conducted under recommended pumping conditions of 100% of best efficiency point (BEP) flow. However, it is common for pumps to operate under off-design settings that can induce conditions such as dry running, gas ingestion, vibration, cavitation, and internal recirculation,

H

22

TABLE 1

Mechanical Seal Selection Guide Single Seal

Double Seal

Conventional Box API Plan 02

Enlarged Bore Chamber API Plan 02

API Plan 11

Conventional Box API Plan 53

Enlarged Bore Chamber API Plan 53

100% BEP Flow

3

2

1

1

1

10% BEP Flow

4

3

2

1*

1*

Low NPSH, 127% BEP

5

3

2

1*

1*

Low NPSH, 100% BEP

5

3

2

1*

1*

3% Gas, 100% BEP

3

2

2

1

1

Dry Run, No Flow

4

5

4

1

1

Low Boil PT Margin

4

3

2

1

1

Operating Condition

1 = Optimum Choice, 5 = Less Optimum Choice

* Vibration May Cause Higher Barrier Leakages

Mechanical seal selection guide which can be detrimental to mechanical seal life. To understand in greater depth how industrial pump operations affect MTBPM of seals, studies were conducted on how various mechanical seals and seal chamber configurations withstand the off-design or upset conditions most frequently experienced in industry. Evaluations were performed by cycling the pump through the following user suggested conditions (Figure 1): • 100% of BEP flow • 10% of BEP flow • 127% of BEP flow at a low NPSH condition • 100% of BEP flow at a low NPSH condition • 3% entrained gas by volume at BEP • dry running operation Various mechanical seal designs and arrangements were tested to determine which presented specific positive or negative effects. Testing was conducted on a standard 2 x 1-10 ANSI B73.1 pump incorporating 316 stainless steel for all wetted parts. An The Pump Handbook Series

acrylic pump was also used to obtain visual records of the fluid flow, gas entrapment, and cavitation effects for selected test conditions. Space limitations prohibit a complete description of the test equipment and methods here. However, the entire testing procedure is detailed in our paper, “The Role of Off-Design Pump Operation on Mechanical Seal Performance,” presented at the 1994 International Pump Users Symposium.

SEAL SELECTION GUIDELINES Observations and analysis of the testing regimen confirm that the selection of the proper seal chamber and mechanical seal design for various pump operating conditions can significantly extend MTBPM of mechanical seals. Table 1 offers further assistance in this selection process. This seal selection guide provides recommendations for each off-design operating condition. Recommendations are specified for applications using various combinations of seal and seal chamber designs. They are not meant to illus-

single seal is required for this application, then the best configuration would be a single seal with a bypass flush (per API Plan 11). If a bypass flush is not applicable, then a single seal with an enlarged tapered-bore seal chamber would be the next best selection. A bypass flush, where the flush is directed over the seal faces, will always result in improved single seal performance as long as the flush is not excessive enough to cause seal face erosion. If an operator is not sure of the conditions within a seal chamber, a bypass flush is recommended for all single seal applications.

trate a comparison between impeller designs/seal types (i.e., pusher versus welded metal bellows). They are the best recommendations based on the combination of equipment installed, and even though some of these recommendations appear to be equivalent, performance levels will vary accordingly. For example, if a pump is known to run dry after its feed tank has been drained, then the best seal design for the application is a double seal with a barrier fluid circulating feature in either an enlarged tapered-bore seal chamber or a conventional stuffing box. If a single seal is required for this application, then the best configuration would be a single seal in a conventional stuffing box. If a pump operates at low NPSH because of high flow or a restriction in the suction line where there is a potential for pulling a vacuum in the seal chamber, the best configuration for the application is a double seal with a barrier fluid circulating feature. (Note that the asterisk on the selection guide indicates that the barrier system needs to be monitored more often at this condition since higher vibrations may cause higher barrier fluid leakage rates). If a

RECOMMENDATIONS BASED ON TEST RESULTS

emissions can be further improved. Our studies support the following conclusions: 1. It is imperative that careful consideration be given to the pump fluid characteristics and the heat rates of mechanical seals. Ensure that the fluid in the seal chamber is always in a liquid state and capable of providing good lubrication to the mechanical seal. Past guidelines have always stressed operating 25°F (14°C) away from the flash temperature of the fluid in the seal chamber. This recommendation appears to be frequently overlooked or ignored, but it is still applicable. 2.

Open, enlarged-bore seal chambers are more forgiving to the survival of single mechanical seals than conventional stuffing boxes during both normal and off-design pump operation. Seal face-toproduct boiling point margins can be increased by as much as 40°F (22°C) without the aid of a bypass flush. Use double mechanical seals, with barrier fluid circulating features, to achieve adequate protection to survive off-design pump conditions including low NPSH cavitation and dry-running operation independent of the seal chamber design.

3.

Low NPSH cavitation conditions can cause significant damage to single mechanical seals when there is a vacuum in the seal chamber. Apply a bypass flush to reduce the seal face temperatures considerably in this situation. However, the best method is to use a close-clearance throat bushing in combination with a bypass flush to ensure a positive pressure in the seal chamber.

4.

Dry-running conditions can also cause significant damage to single mechanical seals. Operating with a conventional stuffing box is less severe during dry running than with an enlarged taperedbore seal chamber. A good indicator of dry-running operation is a squealing sound emanating from the seal gland ring. There is no known “fix” for keeping a single mechanical seal cool during

A good marriage between the pump, impeller, and seal chamber is needed to provide an adequate environment in the seal chamber for mechanical seals and achieve maximum MTBPM for the total pumping system. With pump and seal manufacturers working more closely, opportunities for substantial improvements appear feasible. The uptime of centrifugal pumps, the longevity of mechanical seals, and the minimization of fugitive

FIGURE 1

Pump Performance Curve 600 10% BEP

3% Entrained Air

500 400 Total Developed Head (ft)

100% BEP 300 127% BEP 200 Low NPSH (Cavitation)

Dry Running 100 0 0

25

50

75

100

125

150 175 200 225

250

Flow @ 3600 RPM (US GPM)

Pump performance curve The Pump Handbook Series

23

dry-running operation; the only alternative is to use an external flush with a close-clearance throat bushing or use a double mechanical seal. 5.

6.

7.

24

Based on our results, off-design operation of typical ANSI pumps at 10% of BEP flow has no direct harmful effects on mechanical seal performance. However, the accompanying vibration levels may damage the seal and reduce overall long-term reliability of the equipment. Gas entrainment, at volumes up to 3% in the pumping stream, has no apparent damaging effects on the performance of single mechanical seals, even though flow visualization studies show the presence of gas bubbles in the liquid around the mechanical seal. Flow visualization studies also show that strakes added to the seal chamber bore may have some beneficial operating effects with various impeller designs. Pump impeller design can have significant effects on the seal chamber environment and the operating performance of mechanical seals. For pumps equipped with enlarged taperedbore seal chambers, impellers containing balance holes tend to bleed off gases from the seal chamber during cavitation or gas entrainment operation. Improper setting of the impellers containing back pump-out vanes can cause high vacuums in the seal chamber even when operating at 100% of BEP flow, which would result in significant damage to mechanical seals.

8.

9.

Metal bellows seals equipped with close-clearance vibration dampeners can tolerate vibrations from off-design pump operation many times longer than those not so equipped. Dual mechanical seals can tolerate low NPSH and dry-running off-design conditions with less damage than single mechanical seals. In addition, these seals equipped with circulating features will achieve adequate protection to survive operation during low NPSH and dry-running, off-design conditions.

10. Even though mechanical seals can tolerate these off-design conditions for short periods, the post-test carbon surface profiles indicate that the mechanical seal could have appreciably higher leakage when, and if, the pump is returned to normal operating conditions. In the development of future pump and seal chamber standards, every effort should be made to ensure a net positive pressure in the seal chamber during low NPSH and off-design pump operation. 11. Maintaining a positive pressure in the seal chamber per API Plan 11 has beneficial effects on the performance of single mechanical seals during off-design pump operations. These studies, evaluating the effects of off-design pump operations on the performance of mechanical seal designs, have resulted in a seal selection guide you can use to optimize mechanical seal life. ■

The Pump Handbook Series

REFERENCES 1.

Davison, M. P., “The Effects of Seal Chamber Design on Seal Performance,” Proceedings of the Sixth International Pump Users Symposium, Turbomachinery Laboratory, Department of Mechanical Engineering, Texas A&M University, College Station, Texas, p. 3-8 (1989).

2.

Adams, W. V., Robinson, R. H., and Budrow, J. S., “Enhanced Mechanical Seal Performance through Proper Selection and Application of Enlarged-Bore Seal Chambers,” Proceedings of the Tenth International Pump Users Symposium, Turbomachinery Laboratory, Department of Mechanical Engineering, Texas A&M University, College Station, Texas, p. 15-23 (1993).

Note: For a complete copy of the testing procedures that formed the basis for this paper, contact Texas A&M University, Turbomachinery Laboratory, College Station, TX. William V. Adams is Vice President of Technology for the Durametallic Corporation and a member of the Editorial Advisory Board of Pumps and Systems magazine. Lee Waling is a Senior Engineering Specialist in the Research and Development Department at Durametallic Corporation. Randy Dingman and Joe Parker are both Design Engineers in the Research and Development Department at Durametallic Corporation.

MECHANICAL SEALS & SEALING SYSTEMS HANDBOOK

Testing Seals in Adverse Situations BY DAVID A. PFAFF AND DEAN R. GERBER

H

istorically, many pumps have been equipped with carbongraphite bearings and seals. Over time, pump users needed materials that were more resistant to chemicals and abrasions, could carry higher loads and withstand higher temperatures. Manufacturers and users both began specifying silicon carbide vs. silicon carbide (SiC vs. SiC). Typically the specified silicon carbide materials are either self sintered or reaction bonded. Many magnetically coupled and canned motor pump manufacturers now specify silicon carbide for radial bearings, journal sleeves and thrust bearings. Mechanical seal manufacturers and users are applying more SiC mating pairs. Silicon carbide materials exhibit outstanding characteristics that provide excellent performance and long life in “regular service.” However, field results indicate that “regular service” does not always determine seal or bearing life. Often unusual or upset conditions cause premature failure. SiC is sensitive to operating conditions, particularly to lack of lubricating product flow to the seals and bearings. If product lubrication is lost or the pump is started without fluid, this “run dry” condition can result in pump failure. This can happen when the pump or a valve is incorrectly positioned or there is a loss of suction pressure. In these instances, the “dry run survivability” of either seal or bearing material is critical — what will happen and how long can the components survive under the intense frictional heating of dry running? We tested SiC pairs in start up dry running, process-upset dry running and marginally lubricated running. We looked at several types of silicon carbides for the purpose of providing pump designers and users with a variety of material selection possibilities and to guide our material development efforts toward improved performance. A comprehensive evaluation of your application and system

FIGURE 1. DRY RUNNING TEST 12.70

Driver

X Wear Plate Test Bearing Thrust Washer

31.75 13.89 Dia. Dia.

Inconel Spring Test Bearing Holder Insulator

25.4 Dia. 17.53 Dia. T.C. Well

Bearing Housing Torque Reading Point Bearing Bearing Housing Base Plate Mounted On Drill Press Table

29.08 Across Flats

5.56 1.70 X

Dimensions in MM. “X” Surfaces Lapped Flat and Smooth

Details of thrust bearing test assembly.

FIGURE 2 Water Out

Stationary Primary Seal Ring

Water Pressure Sensor

Water In

Rotating Mating Ring

Fluid Temperature Sensor

Test head of marginally lubricated mechanical seal test rig The Pump Handbook Series

25

should be completed before deciding which material combinations are best suited for a particular application.

TEST MATERIALS Four different types of silicon carbide were compared against carbongraphite materials: • self sintered silicon carbide (SiC) • reaction bonded SiC (SiC + Si) • a variation of reaction bonded SiC with graphite added (SiC + Si + G) • siliconized graphite (SiG) Carbon-graphite, on the other hand, exhibits self-lubricity, good chemical resistance, moderate thermal conductivity and a lower modulus of elasticity. Table 1 provides a description of the test materials, and Table 2 lists their pertinent physical characteristics.

THE RESULTS In the dry running tests, the bearing and seal materials were placed on a modified drill press (Figure 1), a testing method used for nearly 40 years to evaluate friction and wear performance of carbon-graphites and other materials. Material wear, friction coefficient and temperature rise were measured during each 30 minute test. The test can be run dry or submerged. The carbon-graphite provided the lowest friction coefficient, a relatively low temperature rise and low wear regardless of the mating material. This result was not surprising, based on carbon-graphite’s inherent selflubricating properties. On the other hand, both reaction bonded and self sintered silicon carbide when run against each other exhibited high friction, temperature rise and wear. We found that combining graphite with other materials made for longer lasting, cooler running bearings and seals. Reaction bonded SiC, when combined with graphite, showed moderate friction, temperature rise and wear. Graphite, in this case, helped with lubricity. Siliconized graphite which has free graphite in its matrix had moderately low friction, temperatures and low wear, approaching the result of carbon-graphite alone (Table 3).

26

To simulate the loss of fluid in a pump, tests of SiC combinations were conducted on the same test rig as the dry running test. Each test was started with the bearing and wearplate fully submersed in demineralized water. After 15 minutes of operation, the system was drained and the mating pairs were allowed to run dry for 45 minutes. Again, we measured temperature rise, friction coefficient and material wear. We saw the same general trends in material rankings that we saw in the dry running test. Self sintered and

reaction bonded SiC pairs showed the highest friction wear and temperature rise during both the wet and dry running parts of the test. All other combinations gave relatively low friction coefficients, temperatures and wear (Table 4). Pump users today are placing significantly more hardface pairs in mechanical seal assemblies. Traditionally, a carbon-graphite primary ring and hardface mating ring form the mating pair in a mechanically sealed pump. Using carbongraphite as one member of the pair

TABLE 1. DESCRIPTION OF TEST MATERIALS Material

Material Description

PS10138

Self Sintered Silicon Carbide (SiC)

PR9242

Reaction Bonded Silicon Carbide Bonded through the Infiltration of Molten Silicon SiC + Si)

PG9723

Reaction Bonded Composite of Silicon Carbide, Graphite and Silicon (SiC + Si + G)

PE7130

Chemical Vapor Reaction Siliconized Graphite (SiG)

PE8148

Siliconized Graphite Impregnated with Resin (PE7130 & Resin) (SiG + R)

PE10295

Siliconized Graphite Impregnated with Antimony Metal (PE7130 & Antimony) (SiG + Sb)

PE10339

Siliconized Graphite Impregnated with PTFE (PE7130 & PTFE) (SiG + PTFE)

P658RCH

Corrosion Resistant Carbon-Graphite Composite (C-G)

P658RC

Corrosion Resistant Carbon-Graphite Impregnated with Resin (P658RCH & Resin) (C-G + R)

TABLE 2. TYPICAL PHYSICAL PROPERTIES OF TEST MATERIALS Material

PS10138 PR9242 PG9723 PE7130 PE8148 PE10295 PE10339 P658RCH P658RC

Compressive Hardness C.T.E. Strength (ksi) (Vickers) (In/In/°F x 106) 575 400 80 10 12 30 10 31 34

3000 2400 1500 2000 2000 2000 2000 95( l.) 95 (l.)

The Pump Handbook Series

1.9 2.5 2.3 2.4 3.2 3.1 2.4 2.6 2.7

Thermal Temp. Limit Cond. in Air (°F) (BTU/hr ft 2 °F/ft) 75 3000 85 2500 90 1000 30 850 30 450 32 750 30 700 5 600 5 500

adds a strong measure of survivability during dry running, either at start up or during process upset. On the other hand, when hardface mating pairs, such as SiC vs. SiC are installed under these conditions, catastrophic failure can occur. Tests were conducted to simulate a marginally lubricated condition bordering on dry running, because of the high contact pressure and speed. The hardface mating pair tests were conducted on a mechanical seal test rig (Figure 2). We measured friction and wear of various pairs running in demineralized water at 1820 sfpm for 100 hours. Again, we found that self sintered and reaction bonded combinations are limited to lower contact pressures (about 50 psi). At higher pressures, catastrophic failure and/or unacceptable wear occurred. As with the dry running test, we found that composite SiC materials survive and have low wear, similar to carbon-graphite vs. reaction bonded (Table 5).

TABLE 3. DRY START UP RESULTS AT 20 PSI AND 442 SFPM Bearing Material PS10038

Mating Material P658RCH PE10295 PE7130 PE8148 PG9723 PS10138 PR9242

PR9242

• Many siliconized graphite SiC materials have better dry running results against any SiC material we reviewed. Under certain conditions, the results were comparable to a carbon-graphite material running against the same SiC materials. • A reaction-bonded composite of SiC and graphite performed moderately well in dry running tests. • Under marginally lubricated conditions, the siliconized graphites and a reaction bonded SiC composite with graphite showed the highest capability and low wear approaching the capability of carbongraphite running against any SiC

Temp. Rise Ave. Wear (in x 10-4) Avg. Max. Bearing Mating Material Material 78 90 0.4 0.0 94 122 0.1 0.0 104 120 1.0 0.3 89 120 1.0 0.0 162 198 0.4 0.0 212 243 17.2 19.5 251 290 31.1 27.0

0.05 0.13 0.15 0.16 0.40 0.54 0.66

0.09 0.21 0.18 0.37 0.48 0.59 0.70

P658RCH PE10295 PE8148 PE7130 PG9723 PS10138 PR9242

0.07 0.14 0.14 0.26 0.45 0.58 0.63

0.09 0.22 0.22 0.31 0.62 0.65 0.70

75 88 97 141 189 255 269

80 97 107 230 230 285 319

0.8 0.9 1.0 0.8 0.4 14.8 23.1

0.0 0.0 0.5 0.0 0.5 26.5 20.0

PR9723

P658RCH PE10295 PE8148 PR9242 PE7130 PG9723 PS10138

0.10 0.28 0.32 0.33 0.34 0.34 0.40

0.11 0.53 0.51 0.42 0.44 0.57 0.51

109 127 142 163 151 155 179

133 175 195 212 190 183 230

0.5 0.9 0.5 0.3 2.4 0.7 0.3

0.0 0.0 0.5 0.5 1.2 1.0 0.0

PE7130

PE7130 PS10138 PE8148 P658RCH PE10295 PR9242 PG9723

0.13 0.14 0.19 0.21 0.22 0.29 0.35

0.13 0.18 0.40 0.26 0.35 0.53 0.53

95 87 110 135 129 134 161

105 95 160 165 180 170 220

0.2 1.2 2.2 1.0 2.2 3.2 2.2

0.2 0.0 0.0 0.0 0.5 0.0 0.0

PE8148

P658RCH PE8148 PE10295 PS10138 PG9723 PR9242 PE7130

0.08 0.14 0.19 0.22 0.23 0.25 0.29

0.09 0.18 0.35 0.51 0.64 0.43 0.35

105 105 117 103 116 115 131

110 118 128 153 125 138 145

1.0 0.5 0.4 0.4 0.7 1.5 2.8

0.0 0.0 0.0 0.0 0.0 0.5 1.2

PE10295

P658RCH PS10138 PE8148 PR9242 PE10295 PE7130 PG9723

0.11 0.14 0.17 0.18 0.23 0.37 0.40

0.13 0.38 0.22 0.44 0.32 0.44 0.66

97 96 97 96 112 128 166

120 123 107 137 144 160 213

0.0 0.5 1.2 0.2 0.5 1.4 2.1

0.0 1.0 0.5 0.5 0.0 0.5 0.0

PE10339

PE10295 PE8148 PS10138 PE10339 PG9723

0.13 0.18 0.18 0.24 0.33

0.22 0.31 0.22 0.35 0.70

98 97 98 123 153

115 110 115 140 245

1.6 2.2 1.2 2.8 1.6

0.0 0.5 0.5 0.5 0.5

P658RCH

PE8148 PR9242 P658RCH PG9723 PS10138 PE10295 PE7130

0.08 0.10 0.11 0.11 0.11 0.12 0.18

0.13 0.13 0.13 0.13 0.13 0.13 0.22

84 91 83 92 110 82 76

100 105 100 105 130 90 90

0.8 0.0 0.0 0.2 1.0 0.8 2.0

0.0 0.0 0.0 0.0 0.0 0.0 0.0

WHAT IT MEANS When specifying or designing pump seals and bearings, and you’re considering using silicon carbon for either of the components, keep the following in mind: • Self sintered and reaction bonded SiC have the poorest dry running and marginally lubricated capability when run against each other or against themselves.

Friction Coeff. Avg. Max.

The Pump Handbook Series

27

TABLE 4. PROCESS UPSET TEST RESULTS AT 20 PSI AND 442 SFPM Bearing Mating Material Material

Avg. Friction Coeff. Dry Temp. Rise (°C) Average Wear (in x 10-4) Dry Dry Wet Avg. Max. Bearing Mating Avg. Max. Avg. Material Material

PS10138 PE8148 PE10295 PG9723 PR9242 PS10138

0.24 0.26 0.27 0.40 0.47

0.37 0.46 0.43 0.51 0.54

0.11 0.09 0.12 0.28 0.15

103 106 128 146 191

160 170 207 208 255

2.2 0.7 0.9 5.1 11.6

0.0 0.0 0.0 2.5 20.0

PR9242

PG9723 PE10295 PE8148 PS10138 PR9242

0.22 0.23 0.24 0.48 0.62

0.32 0.41 0.32 0.56 0.73

0.10 0.12 0.19 0.26 0.21

83 108 105 164 203

135 168 168 235 298

0.1 0.4 0.3 8.1 11.2

0.1 0.0 0.0 18.0 4.0

PR9723

PR9242 PS10138 PG9723 PE10295 PE8148

0.25 0.28 0.32 0.34 0.37

0.46 0.46 0.45 0.53 0.55

0.09 0.07 0.06 0.07 0.12

106 105 101 148 146

164 183 155 224 236

0.8 0.9 0.7 0.7 0.4

1.0 1.0 0.0 0.0 0.5

PE8148

PS10138 PE8148 PG9723 PR9242 PE10295

0.15 0.21 0.21 0.29 0.31

0.22 0.41 0.40 0.44 0.50

0.13 0.07 0.07 0.11 0.07

87 109 91 112 128

128 183 150 172 228

0.7 0.07 0.07 0.2 0.6

0.5 0.0 1.0 0.0 0.0

PE10295 PS10138 PR9242 PE8148 PG9723 PE10295

0.17 0.18 0.26 0.29 0.31

0.31 0.26 0.51 0.62 0.53

0.10 0.10 0.08 0.06 0.09

71 80 108 89 116

90 107 157 167 182

0.2 1.8 1.2 1.0 0.4

0.0 1.5 0.0 0.0 0.0

material evaluated. Of course, you need to take the normal precautionary testing and design care when you are using laboratory trends in actual pump bearing and mechanical seal applications. ■ David A. Pfaff is a Product Engineer, Pump and Meter Components at Pure Carbon Company, St. Marys, PA Dean R. Gerber is a Product Engineer at Pure Carbon Company.

TABLE 5. MARGINALLY LUBRICATED MECHANICAL SEAL TEST RESULTS FOR POPULAR MATINGS Test Conditions:

Primary Ring

Self Sintered SiC (PS10138) Reaction Bonded SiC (PR9242) SiC-Graphite (Composite) (PG9723)

Siliconized Graphite (Resin Impregnated) (PE8148) Siliconized Graphite (Antimony Impregnated) (PE10295) Hard Carbon-Graphite (Resin Impregnated) (P658RC)

28

Test Duriron — 100 hours; Rubbing Speed — 1820 sfpm (1750 rpm); Sealed Medium — Demineralized Water Mating Ring Contact Pressure (PSI) Self Sintered SiC, PS10138 SiC-G CompositePG9723 Reaction Bonded SiC (PR9242) Nickel Bonded WC Self Sintered SiC, PS10138 Siliconized Graphite, PE8148 Nickel Bonded WC SiC Graphite Composite, PG9723 Reaction Bonded SiC, PR9242 Self Sintered SiC, PS10138 Reaction Bonded SiC, PR9242 Siliconized Graphite, PE8148 Self Sintered SiC, PS10138 Reaction Bonded SiC, PR9242 Siliconized Graphite, PE10295 Alumina (99.7%) Siliconized Graphite, PE8148 Self Sintered SiC, PS10138 Reaction Bonded SiC, PR9242 SiC-G Composite, PG9723 Reaction Bonded SiC, PR9242

50 220 50 110 167 220 275 275 275 110 167 220 220 220 275 75 220 220 220 275 275

The Pump Handbook Series

Water Primary Mating Ring Temp. Ring Wear Wear at Inlet (°F) (IN. x 10-4) (IN. x 10-4) 155 115 115 115 115 115 115 140 140 115 115 115 140 140 140 115 115 115 115 140 140

0.0 1.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.5 0.0 0.6 0.0 0.0 0.5 4.0 4.4 0.5 0.0 1.1 5.0

2.0 0.0 0.0 0.0 0.2 0.0 0.0 0.8 0.0 0.2 0.7 0.0 0.1 0.5 0.3 0.0 0.0 2.0 0.0 0.0 0.2

Average Coeff. Friction 0.05 0.05 0.09 0.05 0.04 0.05 0.04 0.04 0.05 0.06 0.07 0.05 0.06 0.07 0.05 0.07 0.06 0.09 0.08 0.05 0.6

MECHANICAL SEALS & SEALING SYSTEMS HANDBOOK

A Quick Reference Guide to the API 682 Standard INTRODUCTION Widespread acceptance of recently developed sealing technology, improved seal life and increasing concerns about safety and environmental issues have made the mechanical seal a major component part in the pump industry. Therefore, it is imperative to have an established means for classifying sealing systems so that clear guidelines about their use can emerge. Previous standards focused primarily on pump design. The API 682 standard, entitled Shaft Sealing Systems for Centrifugal and Rotary Pumps, is the first API standard published strictly to address issues relating to mechanical seals. This article, the first in a two part series, is meant to serve as an educational tool for the reader who wants to understand the content of the 682 standard without studying the complete document. The paragraph numbers given in parentheses along the left hand column correlate with API sections and paragraph numbers. API standards are laid out in sections numbered 1, 2, 3 etc. with each paragraph and subsection numbered 1.1, and 1.1.1, etc. Text taken directly from API 682 is indicated by italics here.

(1.1.3) The standards set within the API 682 document are considered ”default” standards. These standards specify commonly supplied equipment that has a high probability of meeting a 3 year uninterupted service objective while complying with emission regulations. When a standard option exists, it can be regarded as equivalent and acceptable if it provides a sealing system that facilitates the reliability, maintainability and standardization of the equipment. The standard option seal design is required to be tested and qualified under service conditions equivalent to the conditions in which it will be employed. The standard (default) exists to establish consistent guidelines while the standard options encourage evolving technology. (1.2) Alternative Designs (1.2.1) The standard also specifies acceptable alternatives to the guidelines provided. Although it is not clearly defined as such, API 682 is considered a metric document. All units of measure are given first in metric units followed by imperial units in parentheses. The alternative design paragraph allows substitution of equivalent imperial dimension, fasteners and materials with purchaser approval. This paragraph also encourages new concepts in state-ofthe-art sealing technology.

SECTION 1 - GENERAL

(1.4) Definition of Terms

(1.1) Scope

Terms common to the sealing industry are also defined in section 1. The standard has five and a half pages of definitions, but particular attention should be given to four new terms:

Section 1 provides general information about API 682. It outlines the coverage of the standard in terms of seal sizes, pressure range, and temperature. These specifications are as follows: (1.1.1) • Seal sizes: 30-120mm (1.54.5”) •

Pressure: 0-34.5 bar (0-515 psia)

•

Temperature: -40° to 260°C (-40° to 500°F)

(1.4.3) Barrier Fluid - A fluid which is introduced between dual mechanical seals to completely isolate the pump process liquid from the environment. Pressure of the barrier fluid is always higher than the process pressure being sealed. (1.4.5) Buffer Fluid - A fluid used as a lubricant or buffer between dual mechanical seals. The fluid is always at The Pump Handbook Series

a pressure lower than the pump process pressure being sealed. (1.4.8) Dual Mechanical Seal - A seal arrangement using more than one seal in the same seal chamber in any orientation which can utilize either a pressurized barrier fluid or a non-pressurized buffer fluid (previously referred to as a double or tandem seal). (1.4.10) Flashing Hydrocarbon Service - Any service that requires vapor suppression by cooling or pressurization to prevent flashing. This category includes all hydrocarbon services where the fluid has a vapor pressure greater than 1 bar (14.5 psia) at pumping temperature.

SECTION 2 - SEAL DESIGN (2.1) Standard Seal Types and Arrangements Section 2 outlines the seal design specifications that meet API 682 requirements, emphasizing that: All standard mechanical seals regardless of type or arrangement shall be of the cartridge design. The standard defines a cartridge design as: A mechanical seal unit, including sleeve, gland, primary seals, secondary seals, etc. that can be tested and installed as a unit. One exception to this rule is noted. Hook sleeve cartridge units are not considered cartridge seals under these guidelines. API 682 recognizes three different seal types and three different seal arrangements. They are characterized as follows: Seal Types • Type ”A” pusher 1.

(Appendix B) rotating flexible element, multiple spring, O-ring secondaries

2.

reaction bonded silicon carbide versus premium grade blister- resistant carbon

3.

fluoroelastomer O-rings

29

4.

hastelloy™ C springs

7.

shaft speeds <5000fpm

5.

type 316 SS sleeve, gland, retainer and other metal parts

•

type ”C” bellows

6.

premium carbon floating bushing in gland

1.

(Appendix B) stationary bellows, flexible graphite secondaries

7.

shaft speeds <5000fpm

2.

•

Type ”B” bellows

reaction bonded silicon carbide versus premium grade blisterresistant carbon

1.

(Appendix B) rotating bellows, O-ring secondaries

3.

flexible graphite secondaries

4.

inconel™ 718 bellows

5.

type 316 SS sleeve, gland and other metal parts

6.

premium carbon floating bushing in gland

2.

reaction bonded silicon carbide versus premium grade blister-resistant carbon

3.

fluoroelastomer O-rings

4.

hastelloy C bellows

7.

bronze anti-coking baffle

5.

type 316 SS sleeve, gland and other metal parts

8.

shaft speeds >5000fpm

6.

premium carbon floating bushing in gland

FIGURE 1. ARRANGEMENT 1, SINGLE SEAL

Seal Arrangements (2.1.3) • arrangement 1, Single Seal (Figure 1) includes: 1. single inside mounted cartridge type ”A” pusher seals 2. single inside mounted cartridge type ”B” low temperature bellows seals 3. single inside mounted cartridge type ”C” high temperature bellows seals.

The flexible element of the type ”A” or type ”B” seal rotates with stationary flexible elements which are chosen from a data sheet selection. The FIGURE 2. ARRANGEMENT 2, DUAL SEAL UNPRESSURIZED flexible element of the type ”C” seal is stationary with a rotating flexible element also selected from a data sheet. In API 682, type ”C” bellows seals are specified specifically for high temperature applications because of their advantage in dealing with ‘out of perpendicularity’ between the gland and shaft. A stationary bellows can deflect to match the rotating face when out of perpendicularity occurs. Rotating bellows, on the other

30

The Pump Handbook Series

hand, are not specified for high temperatures because when out of perpendicularity conditions exist, these designs would have to flex and change positions once per shaft revolution to accommodate the runout of the stationary face. (2.1.5) • arrangement 2, Dual Seal Unpressurized (Figure 2) includes: 1.

unpressurized dual inside mounted cartridge seals with two type ”A” pusher seals

2.

unpressurized dual inside mounted cartridge seals with two type ”B” bellows seals

3.

unpressurized dual inside mounted cartridge seals with two type ”C” bellows seals.

The two rotating flexible elements and two mating rings are in series (sometimes referred to as a tandem seal) with stationary elements chosen from a data sheet. The inner seal must be designed with a positive means of retaining the sealing components when the buffer fluid is pressurized to 40 psig. •

(2.1.6) Arrangement 3, Dual Seal Pressurized (Figure 3) includes:

1.

pressurized dual inside mounted cartridge seals with two type ”A” pusher seals

2.

pressurized dual inside mounted cartridge seals with two type ”B” bellows seals

3.

pressurized dual inside mounted cartridge seals with two type ”C” bellows seals.

The two rotating flexible elements and two mating rings are in series with stationary elements selected from a data sheet. The inner seal is internal (reverse) balanced to withstand reverse pressure differentials without opening. (2.2) Seal Design - General (2.2.1) This section provides a simple set of guidelines for using component parts. The standard

FIGURE 3. ARRANGEMENT 3, DUAL SEAL PRESSURIZED

does not cover the design of component parts of mechanical seals; however, the design and material of the component parts shall be suitable for the specified service conditions. (2.2.2) It is the responsibility of the seal manufacturer to provide component parts that meet API specifications. Minimizing seal face heat generation leads to extended seal life. For this reason the seal manufacturer must supply the rate of face generated heat and an estimated heat soak for each design. (Appendix E of the standard provides a complete breakdown of heat generation and heat soak calculations.) (2.2.3/4) Seal vendors must also provide thrust loads created by the seal and the maximum axial movement each seal can withstand without damaging the seal. (2.2.7) Any seal used in vacuum service must have all seal components positively retained to prevent them from becoming dislodged in service. (2.3)Seal Chambers And Glands API 682 provides a detailed breakdown of seal chamber and gland guidelines. An overview of this information is divided into six areas as follows: 1. Seal chamber types (2.3.2) The standard recognizes three types of seal chambers: traditional, externally mounted and internally mounted. API 682 defines the standard chamber as the traditional type, a chamber which is integral to the pump head and is supplied by the pump manufacturer. 2. Minimum dimensions (2.3.3) Seal chambers must conform to a set of minimum dimensions supplied by tables in the API 682 document. The seal dimensions provided in the tables stipulate 1/8” minimum between the O.D. of the seal and the I.D. of the seal chamber.

3. Design stress (2.3.3) All seal chambers, glands and bolting must be in accordance with the ASME Boiler and Pressure Vessel Code Section VIII, Division I.

•

To isolate the seal chamber fluid.

•

To control the flow into or out of the seal chamber.

4. Pump interface design (2.3.5) Correct alignment between the pump and seal cartridge unit supplied by the seal vendor is accomplished by either an inside or outside registered fit to the pump. The registered fit surface on the pump must be concentric to the shaft within 125 micrometers (0.005”). Also, to maintain squareness between the cartridge seal unit and the shaft, a maximum runout of 15 micrometers (0.0005”) per 3 cm (1”) at the interface should be maintained on the pump. Mating joints between the pump and bolt on seal chambers or glands must have a controlled compression, confined gasket to prevent blowout. Also, to prevent distortion of the gland, there must be a metal-to-metal contact inside and outside the gland bolts.

Floating throttle bushings shall be installed in the seal gland or chamber and positively retained against pressure blowout to minimize leakage if the seal fails. Throttle bushings are supplied when specified in dual seals due to axial space limitations.

5. Bushings (2.3.12) Throat bushings can be employed in a number of different situations. API 682 references the following uses: • To function as a replaceable wearing part. •

To establish differential hardness between rotating and stationary parts.

•

To increase or decrease seal chamber pressure. The Pump Handbook Series

6. Connections (2.3.15) Seal chamber and gland connections must be marked by the appropriate symbol. Symbols should be marked on the component in some permanent form (such as stamping, casting or chemical etching). Table 1 gives the symbols along with corresponding connections, location and type. (2.3.19) The specifications default to 3/4 national pipe thread (NPT) connections on the process side with 1/2 NPT as a standard option, 3/8 NPT is specified for atmospheric connections. This is to provide adequate strength at the connection joint and to eliminate confusion between process versus atmospheric connections. (2.3.19/11) ”Drill throughs” from gland connections to the seal chamber must be at least 5 mm (3/16”) in diameter except for multiport distributed flush ports, which may be a minimum of 3 mm (1/8”).

31

TABLE 1 Symbol

Connections

Location

Type

BI BO C D F H Q I O

Barrier/Buffer Fluid In Barrier/Buffer Fluid Out Cooling Drain Flush Heating Quench In Out

180° 0° -180° 0° -90° ---

Process Process Process Atmospheric Process Process Atmospheric ---

(2.3.18/16) A means for completely venting vapors from the seal chamber before pump startup is required. Unused connections must be plugged with a solid round (bull) or solid hex plug. (2.4) Shaft Sleeves API 682 provides a complete breakdown of shaft sleeve requirements. The elements of this section include the following: (2.4.1/2) 1. A one piece sleeve of wear, corrosion, and erosion resistant material shall be furnished by the seal manufacturer. The sleeve shall be relieved along its bore with locating fits at each end. (2.4.4) 2. Shaft to sleeve sealing device shall be O-rings located at the impeller end of the sleeve or flexible graphite at the outboard end of the sleeve. When flexible graphite gaskets are supplied they must be captured between the sleeve and the shaft. (2.4.6) 3. Sleeves must have a minimum radial thickness of 2.5 millimeters (0.100 inches). The seal sleeve thickness in proximity of set screw locations must use the minimum thickness listed on the table below (Table 2 here; Table 5 in API 682). This is to prevent distortion due to tightening of the screws. 4. To assure minimal sleeve runout (50 micrometers (0.002”)) on a shaft mounted sleeve, the sleeve to shaft clearance shall be 25 to 75 micrometers (0.001 to 0.003 inches). Further it must be finished throughout its length with a concentricity of 25 micrometers (.001”) TIR [total indicator reading]. (2.4.3) 5. Shaft sleeves shall have a shoulder, when a sleeve attachment is used it will be axially fixed by a split ring, and when key drives are supplied they will be posi-

32

tively secured to the shaft.

(2.5) Mating Rings (2.5.1/2) Flatness of the seal faces is paramount for achieving low emission and good seal performance. With this in mind, clamped mating rings and overlaid or coated mating rings are disallowed by this standard. Only mating rings constructed from one homogeneous material, with anti-rotation devices designed to minimize face distortion, are allowed.

SECTION 3 - MATERIALS Because material selection is so critical to reliable operation of a mechanical seal, this section and Appendix F are both dedicated to materials. This section is broken down into subsections containing information on the following: 1. general materials information 2. seal faces 3. seal sleeves 4. springs 5. secondary sealing components 6. metal bellows 7. gland plates 8. bolt-on seal chambers 9. miscellaneous parts (3.2.2/4) Reaction bonded silicon carbide is specified as the standard mating ring material, and self sintered silicon carbide is identified as an acceptable alternative. A premium grade blister resistant carbon graphite with suitable binders is specified for the matching face. (3.2.1) Temperature limits for these materials when used in hydro-

carbon service are provided in Table F-3 (not included here). (3.5.1/3) Fluoroelastomers (Viton™) O-rings are specified for low temperature secondary seals and flexible graphite for high temperature applications. Table F-4 (not included here) in Appendix F gives the temperature limitations for secondary seals. Graphite filled 304 SS spiral wound gland gaskets are specified for services over 150°C (300°F). (3.6) Two materials, Hastelloy™ C and Inconel™ 718 are specified for Type B low temperature bellows and Type C high temperature bellows respectively. (3.7/9/8) Sleeves, glands, bolt on seal chambers and miscellaneous parts should be constructed from 316 SS. Outside drive screws may be constructed from 316 SS or hardened carbon steel.

SECTION 4 - ACCESSORIES Because accessories play such an important role in a sealing system, this section highlights auxiliary systems, defined as piping systems. The standard divides these systems into 3 groups: (4.1.1) Group I – Sealing fluid (including barrier/buffer fluid) Group II – Steam/water injection or quench Group III – Cooling water (4.1.2) Included in the above piping systems are: tubing, piping, isolating valves, control valves, relief valves, temperature gauges and thermowells, pressure gauges, sight flow indicators, orifices, barrier/buffer fluid reservoirs and all related vents and drains. Materials for these components are listed in Table 6 (not included here) of API 682. (4.5/6) Cyclone separators, seal flush coolers, internal circulating devices, external circulating

TABLE 2 Shaft Diameter mm/in < 56mm or 50mm (< 2.250”) 50 - 80 mm (2.250 - 3.250”) > 80 mm (>3.250”) The Pump Handbook Series

Minimum Sleeve Radial Thickness mm/in 2.5 mm (0.100”) 3.8 mm (0.150”) 5.1 mm (0.200”)

TABLE 3. SEAL QUALIFICATION TEST PARAMETERS (TABLE 9 IN API 682) Qualification Test Conditions Base Point Cyclic Ranges (Dynamic and Static) Pressure Pressure Temperature Temperature (BAR) (BAR)

Test Fluids

Barrier/Buffer Test Fluids for Dual Seals

Water

Glycol/Water

4

80°C (180°F)

1-4

Propane 20% NaOH

Diesel Glycol/Water

18 8

30°C (90°F) 20°C (70°F)

11-17 1-8

Mineral Oil (-6 to 150°C Applications) Mineral Oil (150 to 260°C Applications)

Diesel

8

20°C (70°F)

1-17

Mineral Oil

8

260°C (500°F)

1-17

pumps and external seal flush systems are also addressed in subparagraphs under accessories and Auxiliary Systems Components.

SECTION 5 - INSTRUMENTATION API 682 offers this section as a default when the purchaser’s inquiry or order does not reference a detailed instrumentation or instrument installation specification. Instruments covered in this section are: temperature indicating and pressure gauges; thermowells; level and flow indicators; alarm, trip, control, pressure and level switches; and relief valves.

SECTION 6 - INSPECTION, TESTING AND PREPARATION FOR SHIPMENT Seal manufacturer qualification and air testing and hydrostatic testing for pressure containing parts are covered in this section. A standard and alternate pump manufacturer seal test are also provided. (6.3.1) The seal manufacturer qualification testing is intended to instill a degree of confidence in the end user that each seal/system will perform in accordance with the scope of API 682 and has been suitably tested prior to market availability. The intent is not to test every individual seal size in all fluids; but to qualify the overall design. The qualification test does not constitute an acceptance test. Two seal sizes 50mm (2”) and 100mm (4”) of each seal type (A,

20-80°C (70-180°C) 30°C (90°F) 20-80°C (70-180°F) 20-90°C (70-200°F)

psig for non-flashing fluids. Reestablish base point. •

Drop fluid temperature in the seal chamber to the minimum cyclic test temperature as shown in the table. Re-establish base point conditions.

•

Raise the fluid temperature to the maximum shown in the table. Re-establish base point conditions. For the mineral oil tests, after the base condition is reached, raise the fluid pressure to the maximum shown in the table (Table 3).

•

Turn off the seal flush for one minute.

•

The test shall be shut-down for a minimum of ten minutes.

•

Establish base point conditions.

•

Repeat steps b through g, three additional times.

•

Repeat steps b through e.

•

Re-establish flush and allow test seal to reach equilibrium conditions (including emissions for hydrocarbons) at the base point.

•

The test shall be shut-down. Maintain base point conditions for a minimum of ten minutes.

150-260°C (300-500°F)

B, and C) should be tested in an appropriate test rig. Each test seal should be of the same configuration, type, design and material as specified in appendix B. To fully qualify a seal, it must be tested on fluids from each application group, non-hydrocarbon and flashing and non-flashing hydrocarbon, for which its use is specified. The qualification test for each fluid consists of three phases, dynamic, static and cyclic. These three phases must be run consecutively, without disassembly of the seal, in accordance with the following test parameters: The dynamic phase consists of a 100 hour minimum 3600 rpm test at constant temperature, pressure and speed with a base point as specified in Table 3. The static phase is a 4 hour test at 0 rpm applying the same basepoint temperature and pressure as the dynamic phase. The cyclic phase is tested at varying temperatures and pressures including start-ups and shut-downs. For flashing hydrocarbons, the cyclic test phase includes excursion into the vapor phase and back to liquid. The cyclic phase is performed after the dynamic and static phase as outlined in the steps below: •

Operate at the base point condition until equilibrium is established.

•

Drop pressure to cause all fluid in the seal chamber to vaporize or 0

The Pump Handbook Series

The qualification test does not have a pass/fail requirement. Rather, the purchaser must evaluate the results of the qualification tests to determine if the tested seal type has met all requirements of API 682. The seal manufacturer will supply the results of the qualification test with each proposal. A certification of conformance, stating that testing was completed to API 682 specifications, should also be supplied. (6.3.2) A hydrostatic test should be performed on all pressure containing parts such as seal chambers, seal glands, dual seal reservoirs and heat exchangers including auxiliary components. Each test shall be conducted using a liquid at 1 1/2 times the maximum allowable working pressure but not less than 1.4 bar (20 psi) for a minimum of 30 minutes.

33

TABLE 4. APPENDICES

(6.3.4) Each cartridge seal— including the seal chamber, if it is supplied by the seal manufacturer—and gland must pass an air test prior to shipment. The test consists of pressurizing the seal to 25 psi for a minimum of 5 minutes. Acceptance is less than 2 psi pressure drop during the 5 minute test. During the pump manufacturer’s performance test, the seal, as configured for the application, should be used, providing the seal manufacturer and purchaser agree.

SECTION 7 - MANUFACTURER DATA This section outlines the data that should be supplied by the seal manufacturer in the proposal and for subsequent contracts.

APPENDICES Table 4 list the appendices with their corresponding titles for easy reference. ■

Appendix A Data Sheets - The data sheets are divided into two sections. Section 1 should be completed for all seals purchased in accordance with API 682. It provides basic seal specifications, options and testing requirements. Section 2 should be completed for seals purchased in accordance with API 682 and when completed API 610 data sheet is not included with the purchase documents (such as in a seal retrofit case). The data in Section 2 provides information on the pump service and design.

Appendix B RECOMMENDED SEAL SELECTION PROCEDURE This appendix is a recommended seal selection procedure for service conditions within the scope of API 682. When followed it provides a recommended seal type, arrangement, and piping system. However, it is the responsibility of the purchaser or seal vendor to ensure the selected seal and auxiliaries are suitable for the intended service. When the service conditions fall outside the scope, a more detailed engineering review is specifically recommended.

Appendix C STANDARD FLUSH PLANS AND AUXILIARY HARDWARE This appendix contains drawings of standard flush plans and auxiliary hardware which have historically been used in industry. While not all of these plans are recommended or referenced in API 682 they may have application in special cases with purchaser approval. Short tutorials of recommended seal flush plans 11, 13, 23, 31, 32, 52, 53, 54, and 62 are given at the end of Appendix B.

Appendix D STANDARD SEAL CODE DESIGNATIONS

ABOUT THE AUTHORS: Patrick M. Flach is the western hemisphere technical sales manager and group leader for new product development in the Industrial Division of EG&G Sealol. He has more than 28 years of experience in pump design and applications. Mr. Flach served on the manufacturers task force during the writing of the 6th and 8th editions of the API 610 standard. David P. Casucci is engineering manager for the Industrial Division of EG&G Sealol in Cranston, RI. Mr. Casucci was a contributing member of the API-682 First Edition Task Force. He has co-authored several publications on mechanical metal bellows seals.

A standard seal coding system, new in API 682, is given with examples.

Appendix E HEAT GENERATION AND HEAT SOAK CALCULATIONS

Appendix F TABLE OF MATERIALS AND MATERIAL DESCRIPTIONS Used when referenced in the standard.

Appendix G SEAL TESTING SEQUENCE CHART Used during testing by the seal vendor.

Appendix H INSPECTOR CHECKLIST Three levels of inspection are given: LEVEL 1 - Typically used for pumps in general refinery services. LEVEL 2 - Comprises performance and material requirements. More stringent than Level 1. LEVEL 3 - Should be considered for critical services.

Appendix I MANUFACTURER DRAWING AND DATA Consists of a simple distribution record (schedule), followed by a representative description.

Appendix J PURCHASER CHECKLIST Used to select the purchaser’s specific requirements.

34

The Pump Handbook Series

MECHANICAL SEALS & SEALING SYSTEMS HANDBOOK

Gas-Barrier Seals Establish Beachhead Driven by strict emissions requirements, gas barrier seals are winning new victories and converts in mechanical sealing technology.

as-barrier seals are a revolutionary milestone in the mechanical sealing of rotating equipment for the process industries. Adoption of gas-barrier sealing technology has been spurred by the fugitive emissions restrictions contained in the 1990 Clean Air Act Amendments (CAM) Title lll; specifically, the U.S. EPA’s proposed National Emissions Standard for Hazardous Air Pollutants (NESJ IAP) or Hazardous Organic NESHAP (HON), subpart H of which addresses equipment leaks.1 Gas-barrier seals can be either contacting (based on boundary lubrication) or noncontacting (based on full fluid-film lubrication). Although gas-lubricated seals have been produced in single and tandem designs, it is the gas-barrier double seal that, to date, has seen the greatest application potential (Photo 1). Users are increasingly specifying gas-barrier seals as they gain familiarity with the technology.

G

effective for many routine services, liquid-lubricated seals in critical services are limited in terms of performance, maintenance and economics. A brief look at traditional liquid-lubricated seals will help highlight the performance benefits of gas-barrier

sealing technology. Single seals. The most common type of seal is the basic single, liquidlubricated seal design. Mounted inside a seal chamber, a single seal can economically restrict VHAP (volatile hazardous air pollutant)

TRADITIONAL ALTERNATIVES REVIEWED Before examining appropriate uses of gas-barrier seals, let’s look at the traditional alternatives. Until recently, mechanical seals used with process pumps invariably have been lubricated with a liquid, either the pumped product or a barrier or buffer fluid in its liquid state. While

Photo 1: An example of a gas-barrier double seal (GB-200 Dura Seal) featuring a contacting design.

By William V. Adams and Randy R. Dingman The Pump Handbook Series

35

emissions to less than 1000 ppmv, provided the specific gravity of the pumped product is 0.4 or greater. Single seals can accommodate environmental controls and can be balanced to withstand high seal chamber pressures. Dual seals. Dual seals are used when the leakage rate or safety of single seals is not adequate for a given application, such as toxic liquids whose leakage into the environment would be hazardous. Dual seals are available in either pressurized or non-pressurized designs. Non-pressurized dual seals. Nonpressurized, tandem seal arrangements are often specified where there are health, safety and environmental concerns. Under normal operating conditions the inner, or primary, seal is designed to withstand full product pressure. The outer seal operates in a suitable buffer fluid at atmospheric or very low pressure. Pressurized dual seals. With pressurized, dual (double) seals, a compatible barrier fluid is injected into the seal chamber at a point near the inner seal and exits at a point near the outer seal. The fluid is maintained between the two seals at a pressure higher than that of the product behind the impeller. Pressurized dual seals are required when the specific gravity of the pumped product is 0.4 or lower. Liquid-lubricated dual seals, however, require a supply tank or reservoir of fluid that is circulated between the inner and outer seals. In addition, American Petroleum Institute specification API 682 requires that all liquid-lubricated dual seals incorporate a pumping device to circulate the liquid in a closed loop from the seal cavity, between the inboard and outboard seals, to the supply tank reservoir and back again. Depending upon operating conditions, cooling coils may have to be added to the reservoir to remove excess seal facegenerated heat. Moreover, improper operation of liquid buffer/barrier systems has been identified as a major cause of failure for liquid-lubricated dual seals. With their pumping rings and supply tanks, liquid-lubricated dual seals can also leak barrier fluid into both the process fluid side of a pump and the atmos-

36

phere. Typical leak rates range from 2 to 5 cc/day in both directions. As a result, periodic maintenance on the supply tank reservoir is necessary. Maintenance cycles usually involve a weekly check of the liquid level in the tank reservoir and quarterly refills.

GAS-BARRIER SEALING OVERVIEW Compared with liquid-lubricated seals, gas-barrier sealing technology offers reliability, simplicity and lower operating costs. Gas-barrier seals embody more than 20 years’ experience with gas-lubricated, “dry-running” sealing technology. Until recently, this technology could not be applied realistically to seals for process pump services, although dryrunning seals had been developed and used for many years on equipment such as compressors, mixers and agitators. But, with the evolution of seal technology, advanced seal face materials, and exhaustive laboratory and field testing, seal designers have succeeded in applying gas lubrication technology to mechanical seals on typical ANSI and API process pumps. Historically, the phrase “dry running” refers to adverse seal operating conditions such as pump cavitation or loss of barrier fluid. Under such circumstances, dry running could quickly lead to seal damage or destruction. In the context of gas-barrier sealing technology, however, dry running refers to the fact that the fluid lubricating the seal faces is in a gaseous state rather than a liquid state. Thus, the following principles typically

apply: Seal faces are made of self-lubricating, non-galling materials. One face is typically made of carbongraphite, while the mating face is a relatively hard material with good thermal conductivity, such as silicon carbide or tungsten carbide. Seal face flatness is critical to reducing gas leakage. Seal face deflection, due to pressure or thermal gradients, is precisely controlled to reduce leakage and prevent failure.

GAS-BARRIER SEALS EXAMINED Gas-barrier double seals meet the most stringent (i.e., zero) emissions control requirements. With a gas-barrier seal, an inert gas such as nitrogen or instrument air serves as the seal face lubricant and coolant in place of a conventional liquid barrier. The gas-barrier also prevents leakage of pumped product to the atmosphere. Moreover, gas-barrier double seals eliminate pumping rings, supply tanks, and the maintenance associated with conventional dual-seal systems. In the case of a double seal, the gas is injected at a pressure differential of 20 psig to 50 psig (138 kPa to pressure. A purge through the seal is not required since the barrier is deadended. Inner Seal. The inner seal of a typical contacting gas-barrier double seal uses conventional face sealing technology to minimize barrier gas leakage into the pump. The low differential pressures on the inner seal allow the use of conventional sealing

Application Report A chemical manufacturer in the southeastern United States had a critical process pump that cavitated continuously. This manufacturer decided to evaluate a gas-barrier double seal for this critical application, but first wanted to see if it would perform on a less critical condensate pump with a similar cavitation condition. The condensate pump was a Goulds 3196 MT with a 1.750 in. (45 mm) diameter shaft, rotating at 1750 rpm. The condensate was at a temperature of 250 F (121 C). After attempting a pusher seal that gave them only one month of service life, the chemical manufacturer switched back to compression packing which performed better, but still lasted for only three months. They then installed a contacting gas-barrier double seal with a 40 psig (276 kPa) nitrogen barrier pressure. This seal has since been running successfully for more than two years.

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Photo 2: Inert gas flowing through the grooves of the outer seal rotor provides lubrication and cooling. faces. Inner seal leakage into the pump is less than 0.05 standard cubic feet per hour (scfh) at a 50-psig differential pressure. The inner rotor is made of specially formulated carbon with the same chemical resistance as traditional carbon, but with improved thermal and physical properties. The inner stator an be constructed of either tungsten carbide or silicon carbide. This type of inner seal also can tolerate pressure reversals. If the barrier gas is lost, the seal continues to operate as a true single seal; a shift of the O-ring on the rotor causes the balance diameter to shift, allowing the seal faces to remain closed due to the process fluid pressure. Outer Seal. With the contacting gas-barrier seal it is the outer seal that incorporates advanced sealing technology that is the basis of gas-barrier operation. The outer seal uses a bidirectional slotted carbon rotor that allows the seal to operate “dry” up to 200 psig (1380 kPa). As shown in Photo 2 angled slots or feed grooves on the rotor feed the pressurized gas to an annular groove to provide hydrostatic balancing. The total pressure drop occurs across the dam where all the sealing takes place. The feed grooves constantly force the gas through the faces to keep them cool and clear of wear debris that might be present. The pad on the outer seal rotor has two functions: 1) to act as a bearing support to reduce the overall unit loading and 2) to develop a slight lift due to thermally induced seal-face waviness in order to reduce the unit loading across the face. This design reduces seal face loading to one-quar-

ter of that of a conventional seal design.

EPA and OSHA regulations • improved product quality.

GAS-BARRIER SEALS TESTED IN THE REAL-WORLD

REDUCED INSTALLATION AND OPERATING COSTS

Experience gained with dry-running mixer seals has been applied in producing gas-barrier seals that can consistently withstand the rigors of day-to-day operating conditions. Key areas addressed in developing gasbarrier double seals include: Pressure reversals. The ability of a gas-barrier seal to handle pressure reversals has been verified by laboratory and field testing. Wear particle contamination. Wear particles from seal faces are the inevitable consequence of start-up and shut-down procedures. The groove depth on the outer rotor effectively accommodates such particles; the volume of gas flow through the grooves keeps the seal faces cool and clear of debris. Gas barrier quality. Gas-barrier seals can be forgiving to the presence of moisture in the gas supply; auxiliary devices such as coalescing filters represent added cost and maintenance items. Thermal transients. The faces of gas-barrier seals have been designed to withstand the effects of thermal transients and the higher-than-normal face deflection that may occur. Speed considerations. Ideally, the sealing capability of a gas-barrier seal should be independent of shaft speed (i.e., capable of running at any speed from zero to its upper limit without excessive barrier leakage). Bidirectional rotation. In addition to reducing the need for spares inventorying, the bidirectional rotation capability of a gas-barrier seal makes it impossible to install the wrong seal in the wrong pump. Gas barrier leakage. Compared with conventional liquid-barrier seals, leakage of the barrier gas from a gasbarrier seal during normal operation is negligible. The influence of gas-barrier sealing technology on process operations is seen in a number of key areas: • reduced operating costs • increased operating reliability • extended mean time between planned maintenance (MTBPM) • more effective compliance with

Compared to liquid-lubricated dual seals and sealless pumps, gas-barrier seals provide significant opportunities to reduce installation and operating costs. The gas-barrier seal can be connected to a plant source of nitrogen or instrument air. All that is required is a simple control system (i.e., a pressure gage, flowmeter, needle valve and regulator), Figure 1. Preengineered and assembled gas control panels are available from gas-barrier seal manufacturers. Because the gas supply from in-plant headers is continuous, there is no need to monitor and replenish supply tanks; nor is there risk of seal failure resulting from a depleted supply tank. Typical leakage rates, depending upon the type of gasbarrier sealing technology used, can vary from 0.1 scfh to 3 scfh. Gas-barrier seals eliminate the need for barrier fluid cooling, handling and disposal. In addition, the cost of the gas is inconsequential compared to liquid barriers — even water. Estimated cost of nitrogen for operating a gas-barrier seal is $3 per year; estimated cost of cooling water for the barrier fluid of a liquid-lubricated double seal is $1,700 per year. Moreover, operating power consumption of gas-barrier seals, whether contacting or noncontacting, is a small fraction of that required for liquid-barrier seals and sealless pumps. A liquid-barrier double seal operating at 3600 rpm and 200 psig, for example, would require approximately five times as much power as a typical gas-barrier seal (approximately 1.400 hp vs. only 0.280 hp for the gasbarrier seal). A sealless pump would require about double the power to achieve the same flow and head as a conventional pump sealed with a gas-barrier seal.

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INCREASED OPERATING RELIABILITY The cartridge design of gas-barrier seals simplifies installation and virtually assures proper operation. Gas-barrier double seals also have been designed to withstand off-design pump operation, cavitation and dryrunning conditions — all of which

37

can cause rapid failure of liquid-lubricated seals and sealless pumps. With the cost to repair a pump averaging $3,500 2, the forgiving behavior of gas-barrier seals can translate into significant savings and improved productivity. For the same reason, gas-barrier seals can provide a viable alternative to typical low emissions single-seal applications in which the chance of catastrophic failure exists. Transfer pumps, which are frequently subject to dry running, are a case in point.

FIGURE 1: GAS BARRIER SUPPLY SYSTEM DIAGRAM

EXTENDED MTBPM Unlike conventional liquid-lubricated seals, gas-barrier seals eliminate seal face-generated heat as a failure mechanism. With a contacting seal design, for example, seal face-generated heat is only 1/10 to 1/100 that of a liquid-barrier pusher seal. Heat generated at the seal faces is dissipated by gas passing between them, thus avoiding excessive bulk or seal face temperatures. This, plus the ability to accommodate upset, cavitation and dry-running conditions significantly reduces potential failure modes. With a design service life of up to five-plus years, gas-barrier seals are suited for numerous applications in the chemical process, petrochemical/refinery, pulp and paper, power and pharmaceutical industries.

REGULATORY COMPLIANCE Gas-barrier seals effectively eliminate the need for costly, time-consuming emissions monitoring of pumps handling highly toxic, hazardous substances. Either new or existing pumps can be ouffitted with a gas-barrier seal and enlarged-bore seal chambers. The same zero-emissions/process fluid-containment capabilities of gas-barrier seals further facilitate compliance with OSHA Process Safety Management requirements.3 Sealless pumps and liquid-lubricated double seals can also qualify for exemption from EPA-mandated emissions monitoring. However, conversion to sealless pumps represents a major expenditure for equipment changeout including costly instrumentation, plus high horsepower consumption. Liquid-lubricated double seals require elaborate barrier fluid systems and, for exempt status, no

38

TABLE 1. TYPICAL OPERATING SPECIFICATIONS FOR GAS-BARRIER SEALS Contacting

Non-contacting

Temperature: -40 °F to 250 °F (-40 °C to 121 °C)

-40 °F to 600 °F (-40 °C to 315 °C)

Pressure:

Full vacuum to 200 psig (1,380 kPa)

Full vacuum to 500 psig (3,450 kPa)

Speed:

0 to 3,000 fpm

50 to 10,000 fpm

Shaft sizes:

1.000 in. to 2.750 in.

1.000 in. to 6.000 in.

VHAPs can be used as the barrier fluid. In contrast to these more costly options, gas-barrier seals offer a reliable, economical solution.

IMPROVED PRODUCT QUALITY In most instances, application of a gas-barrier, such as nitrogen, instrument air, steam or any other inert gas between double mechanical seals is not viewed as a source of contamination of the pumped process fluid. This is seen by many users as having a substantial advantage over liquid-lubricated seals which, in the case of an inboard seal failure, can dump large amounts of barrier fluid into the process system and contaminate the product. Typical leakage past the inner seal of a gas-barrier seal is 0.05 scfh at 50 psig; across the outer seal, typical leakage is 1.1 scfh at 200 psig.

THINGS TO KNOW ABOUT GAS-BARRIER SEALS The two major situations in which gas-barrier seals should be considered are: • the handling of hazardous and/or toxic process fluids • any services critical to overall plant The Pump Handbook Series

operations. For installation, gas-barrier seals require the enlarged-bore seal chambers as defined by the ANSI/ASME B734 and API5 standards. Gas-barrier seals are not recommended for use with process fluids containing more than 10% solids. The selection of contacting vs. noncontacting gasbarrier seal designs also depends on the application. Contacting gas-barrier seals are capable of bidirectional rotation and of accommodating upset conditions. They can also maintain a positive seal over the entire range of operating speeds, from 0 rpm to their maximum speed limit. Table I illustrates typical operating specifications for both contacting and non-contacting gas-barrier seal designs. Field reports indicate solid and growing acceptance of gas-barrier sealing technology. Economical operation, forgiving designs and zero leakage combine to make the gas-barrier seal a significant extension of mechanical sealing technology. ■

REFERENCES: 1 “National Emissions Standard

for Hazardous Air Pollutants for Source Categories: Organic Hazardous Air Pollutants from Synthetic Organic Chemical Manufacturing Industry and Other Processes Subject to the Negotiated Regulation for Equipment Leaks,” EPA 40 CFR Part 63, Subpart H, 1994. 2 Ehlert, D., “Bearing Lubrication Trends and Tips,” Pumps and Systems, December 1993. 3 “Process Safety Management,” OSHA 29 CFR 1910.119. 4 American National Standards Institute, American Society of Mechanical Engineers, New York, NY 5 American Petroleum Institute, Washington, DC. William V. Adams is Vice President of Technology for the Durametallic Corporation. He is also currently chairman of the WG-3 Emissions Task Force of STLE’s Seals Technical Committee and a member of the API 682 Seals Standards Committee. Randy R. Dingman is a Design Engineer in the Research and Development Department at Durametallic Corporation.

More on Non-Contacting Seal Design By Pumps and Systems Staff The non-contacting type of mechanical seal design is based on the concept of hydrodynamic lubrication and incorporates geometry changes to the seal faces such as spiral grooves. The only heat that is developed is that of shearing gas at the seal faces. This type of sealing system can run on the process fluid or neutral barrier fluids such as nitrogen, purified air or steam. The non-contacting gas lubricated sealing system was initially developed for compressors, but it is now being applied to difficult pumping applications to control emissions or maintain process fluid purity.

DRY SEAL DESIGN Successful operation of a seal in a liquid or gas environment is dependent on the development of

a fluid film and the removal of unwanted heat. Understanding the processes that occur at the seal faces provides the basic fundamentals necessary to design a seal to operate totally dry. An ordinary contact seal can only operate in a liquid-lubricated condition. If operated in a gas environment, the frictional heat would destroy the seal in a matter of minutes. Consequently, frictional heat must be eliminated to operate successfully in a dry environment. This is no easy task since the seal must also restrict flow to the environment. The non-contacting design accomplishes this by providing one seal face with a set of spiral grooves. When the shaft begins to rotate, pressure is built up within the spiral grooves, sepa-

FIGURE 1: CUT-AWAY OF A TYPICAL NON-CONTACTING GAS LUBRICATED SEAL.

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the seal faces. Usually, this is just a two or three degree temperature rise at the seal faces. A small amount of barrier fluid is used to create the non-contacting feature of this technology. Figure 2 illustrates the consumption at various pressures and speeds. This is a notable development in seal technology and a major shift in thinking that all seals for pumping applications must be liquid-lubricated. According to James Netzel, chief engineer at John Crane, Inc., another advantage of the non-contacting seal design is that horsepower losses at the seal FIGURE 2. GAS BARRIER CONSUMPTION faces are extremely small since frictional 1981 0.07 contact is eliminated. 1698 0.06 This results in significant energy cost sav3000 FPM 1415 0.05 PM F 00 ings. 25 1132

0.04

PM

M

0.03

P 0F

0

15

0.02 500

F 00

10

566

FPM

283

0.01 0.0 0

849

0.0 25

50

75 100 125 150 175 200 225 PRESSURE (PSIG)

NOTE: Data based on single seal arrangement

BARRIER CONSUMPTION (ML/L)

BARRIER CONSUMPTION (SCFM)

rating the seal faces. A sealing dam at the seal face limits the flow of gas. When the shaft stops turning, the seal faces close, maintaining the seal in a static condition. When this concept is used in a double seal arrangement, a barrier gas at higher pressure prevents static and dynamic leakage of the process liquid to atmosphere. Also, since there is no contact, there is no frictional heat. As a result, a non-contacting seal requires no circulation of fluid for cooling at normal pump shaft speeds. The only heat developed is that of shearing the gas film at

APPLICATION TO CONVENTIONAL PUMPS Compared to typical liquid-lubricated double seal systems, gas lubricated seals are less complex, provide greater reliability and have proven to be

TABLE 1: REPRESENTATIVE LISTING OF FLUIDS SEALED WITH GAS LUBRICATED SEALS Representative listing of fluids sealed with gas lubricated seals Acetic Anhydride Acrylonitrile Alkyl Resins Anhydrous Cyclohexane Benzene Benzene Toluene Bisphenol Acetate Butadiene Dichlorobutadiene Dichloroethane Ethyl Chloride Ethyl Hexanol Bottoms Formaldehyde Gasoline Gasoline Glacial Acrylic Acid Hydrocarbon Kerosene Latex Cold

40

122°C 38°C 240°C 39°C 153°C 150°C 174°C 40°C -24°C 80°C 24°C 202°C 122°C 231°C 38°C 21°C 116°C 228°C 16°C

Latex Hot LPG Methanol Molten Chemical Tar Nitric Acid (57%) Organic Acid Phenol Plastic Pigment Propane/Butane Propane with Caustic Recycle Gas Sanotherm 59 Slurry Water Sodium Carbonate - saturated Spent Methanol & Chemical Slurry Styrene Sugar Solutions Therminol Therminol #59 Toluene DI ISO Cyanate Wastewater with Hydrocarbon Wet Gas

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80°C 50°C 21°C 172°C 31°C 94°C 151°C 21°C 47°C 122°C 113°C 174°C 66°C 56°C 60°C 52°C 217°C 307°C 300°C 192°C 74°C 46°C

more cost effective. The barrier system for a gas lubricated seal consists of a compatible barrier gas source capable of providing gas at a pressure level 25-50 psig greater than the pump box pressure and a locally mounted gas control panel. Plant header nitrogen is normally used as the source and piped directly to the gas control panel, through which it is fed into the seal gland. The gas control panel serves several basic functions: • regulates source gas pressure • filters particulate and removes moisture • measures/monitors gas flow and consumption • measures gas pressure • provides additional backup to process loss in event of seal and barrier gas failure. Non-contacting gas lubricated seals handle off performance operation of a pump as well, Netzel emphasizes. “This design is very versatile since the seal is operating independent of the fluid in the pump,” he says. “The pump can cavitate or run dry with no effect to the seals. This is possible since the seals are non-contacting and operating on the gas barrier and not the fluid in the pump.” Documented seal life increases of more than 1000% have been achieved in difficult to seal services as a result of replacing conventional contacting seals with gas lubricated double seals. Non-contacting gas lubricated double seals are currently being used on a wide range of process fluids in the chemical, petrochemical, refinery, pharmaceutical and pulp and paper industries. A representative listing of actual services being sealed with the non-contacting design is provided in Table 1. Gas lubricated seal technology is a major development for sealing pumps. The benefits in installation, operation and maintenance provide solid economic return.

MECHANICAL SEALS & SEALING SYSTEMS HANDBOOK

Bellows Seal Repair: The Plant Perspective ith the cost of bellows seals increasing and federal emissions standards ever tightening, understanding the repair of these seals is as important as ever. Whether sending seals to the original equipment manufacturer (OEM), an aftermarket repair shop, or repairing the seals in-house, the concepts of repair are the same. Plant engineers now must take responsibility when sending out bellows seals for repair. Items such as checking free length and spring testing bellows can reduce repair costs and ensure longer run life in operation. The details of these items are not the important factor, but understanding them (and how they affect repair) is very important. Two types of repair facilities exist: OEM, or authorized repair centers, and aftermarket repair centers. The repair facility chosen must have a specific understanding of each manufacturers’ specifications for free length, spring tension at working length, and seal face holding method.

W

FREE LENGTH All manufacturers have a tolerance for the free length of a bellows seal when it is new. During some operating conditions, the bellows may reach a set. In other words, the bellows could have a finite life and at some point will lose its ability to flex. If a seal in this condition is stretched in order to reset it at working length the integrity of the seal has been compromised. Stretching of the bellows may allow the seal to work for a short period of time, but the bellows may return to the set position. This may not be an immediate problem in some services, but it can be in others. In light hydrocarbon services this may create VOC violations almost immediately upon startup. In the interest of

Mean Time Between Failure (MTBF) and EPA regulations, try to avoid this practice. After a bellows seal has been removed from service, measure the bellows’ free length. Seal manufacturers measure the free length both from the tip of the seal face as well as to the top of the seal face holder. A good practice is to measure the free length without the seal face, to determine if the bellows is repairable, because the seal face may be worn. Charts for this free length can be obtained from your seal supplier. Don’t give up — some vendors may have difficulty obtaining this information.

SPRING TENSION Spring tension is not only important in the repair of bellows-type seals. It can be useful in explaining to pump mechanics the importance of setting bellows at the given operating length. This spring tension has a margin that varies from manufacturer to manufacturer. Each manufacturer has its own tension and tolerance for each size and material of bellows. These can also be obtained from your seal supplier. It is imperative that the vendor repairing the bellows have access to the spring tension and tolerance

charts for each manufacturer of bellows. Do not assume the repair facility understands this component of seal repair. If they cannot provide written procedures detailing this portion of the repair, then work out a compromise by mandating new seals as replacements for the used seals at a given percentage of the new price. If the question of spring tension is not asked by the repair facility, chances are this test is not being done. Depending on the volume of seals at a given facility, an investment in a spring tester may prove cost effective in weeding out those seals unable to pass this test. I will mention again that worn inserts can give a false reading in spring tension. Only a few thousandths of an inch may be required to bring the seal back to the proper tension specifications. A new seal face may be all that is required for the repair.

SEAL FACE HOLDING METHOD Each bellows manufacturer has their own method for holding inserts. Some inserts are cold pressed. Some are held with a shrink fit by the seal face holder.

FIGURE 1. BELLOWS ASSEMBLY Free Length

Seal Size Seal Face

Bellows

The Pump Handbook Series

Seal Face Holder

41

Others have a combination of these with an adhesive. (This adhesive may be incompatible with some products.) If the seals are sent to a shop other than the OEM, the facility must have an understanding of these holding methods. The deflection of the seal faces is affected by the force exerted on them by the seal face holder.

SEAL SHOP EVALUATION When reviewing any seal repair shop, a few pointed questions can determine their understanding of bellows seals. Here are some sample questions to ask: • Where is your spring tester for testing bellows tension? • What methods do seal manu-

42

■ facturers use for holding inserts? • Do you mind if I get a copy of your repair procedures for bellows seal repair? • Can you tell me the allowable tolerance in spring tension for a specific manufacturer, size, and material bellows? • Do you check free length and spring tension in developing an estimate of repair for me?

SUMMARY Effective bellows seal repair requires a basic understanding of free length, spring tension, and seal face holding methods. Although the repair is not limited to these items, they are often misunderstood or left out as a part of the repair procedure.

The Pump Handbook Series

ABOUT THE AUTHOR: Eddie Mechelay is Mechanical Supervisor at Colorado Refining Company, Denver, CO and a member of the Pumps and Systems’ User Advisory Team.

MECHANICAL SEALS & SEALING SYSTEMS HANDBOOK

Cartridge Seals: Pros, Cons and the Human Factor nstalling cartridge seals requires a working knowledge of their advantages and limitations as well as the human element. Factors such as API 682, cost, inventory and ease of installation must be considered in deciding whether to install cartridge seals on pumps. Once this is determined, other criteria can help in the process of selecting and specifying the proper seal for the application. The scope of this discussion is to examine factors that must be considered if a cartridge seal is to perform to its potential – not to determine if conversion to cartridge seals is appropriate in any given situation.

I

ADVANTAGES/DISADVANTAGES The most common reason for using a cartridge seal relates to its quickness and ease of installation. This is true for most any application. It eliminates the errors involved in determining the seal set with scribe marks on the shaft. Another advantage is its reliability. Most seal manufacturers bench test the cartridge prior to shipment, thus ensuring proper operation at startup. Essentially the only sealing element that can be damaged during installation is the o-ring or grafoil ring on the cartridge sleeve. The only drawback to this is the unknown. If a seal is installed on a critical or hazardous piece of equipment, the pump mechanic doesn’t know if the seal faces were damaged in shipment. Fortunately, with today’s packing methods, this rarely occurs. The high initial cost of converting to cartridge seals is probably the biggest deterrent to converting a seal application. Repair costs, however, typically are not much higher than component seal repair. (Note: with component seals the shaft sleeve is considered a pump part and thus not a seal repair expense. The cartridge sleeve is always considered in the seal repair cost. This can lead to mis-

leading repair cost comparisons.) Another disadvantage of cartridge seals is the reduced clearance in the stuffing box. Since the cartridge seal requires a sleeve, the outside seal diameter is typically at least 1 ⁄ 16 ″ larger. Although this may not appear excessive, in a typical ANSI pump this could reduce the clearance to the stuffing box from 3⁄32″ to 1⁄32″. In certain applications this reduced clearance can cause excessive heat, which can result in flashing across the seal faces and reduce seal life. Other measures can be taken to eliminate this increased heat, but they may increase the initial cost. The instructions for many semiopen and open vane impeller pumps require that the impeller clearance be set after the pump has been installed. A cartridge seal allows this flexibility. After the pump is assembled, the mechanic does not set the seal to the shaft with the set screws until after the impeller adjustment has been made. Component seals must be set prior to this adjustment, and any impeller adjustment after this set will either increase or decrease the tension on the seal. Plant inventory can be either an advantage or a disadvantage. If the plant has an extensive stock of component seals, conversion to cartridge seals can represent a considerable capital cost. On the other hand, with creative thinking a single cartridge design can be used in a variety of pumping applications. This would necessitate stocking only one part instead of the many parts required for a component seal. For this to be successful, an extensive site survey with complete pump data is required, along with cooperation from the seal vendor.

RETROFIT CONCERNS Many factors must be considered in retrofitting an older pump for a cartridge seal. Numerous roadblocks can be eliminated if the end-user is The Pump Handbook Series

flexible and communicates with the seal vendor. Because the cartridge seal typically requires more room in the stuffing box and distance to first obstruction for the gland, the enduser may need to make pump modifications. Following are some ideas designed to create more room: Bore the stuffing box out 1⁄16″ per side. Many packed pumps that are being converted to cartridge seal already have corrosion pits that are at least this deep. Contact the pump vendor if there are special concerns about the amount of metal that may be removed. Remove “extra” parts. Some stuffing boxes have removable cooling jackets. If those jackets are no longer in service, consider removing them. This may require the stuffing box to have an insert installed at the bore of the stuffing box. Remove the old shaft sleeve. This is highly recommended. This not only increases the stuffing box OD with respect to the seal, but also eliminates a potential leak path. Many old shaft sleeves were sealed with packing or paper gaskets. These sealing methods can prove to be troublesome in light hydrocarbon service when trying to comply with EPA emissions requirements. The following items should be considered when removing the shaft sleeve: 1. If the shaft sleeve is a hook sleeve, it will have to be cut off just past the impeller end of the cartridge seal. This sleeve is a spacer for the impeller and also an integral part of the clearance for the throat bushing. 2. On older split case pumps the pump seal packing may have been used in the deflection calculations for the shaft. In other words, the packing acted as a bushing. Removing the shaft sleeve in such applications may cause excessive shaft deflection, resulting in premature seal failure. One cure for this is to manufacture a new shaft with an increased OD. The

43

pump vendor can assist in sizing this new shaft to ensure minimal deflection. Some cartridge seals use a pilot fit to center the seal. This fit may need to be corrected at the time of installation. This is highly probable on pumps that previously had packing. Increase the distance to first obstruction. 1. Many pumps use flingers or slingers at the bearing housing to prevent contaminants from entering the oil reservoir. These flingers can protrude from the bearing housing as much as 1⁄2″. They can be removed and replaced with a number of products on the market such as magnetic

or labrynth seals. Lip seals also can be used but are not recommended unless the bearing housing is in oil mist service. 2. Machine the face of the stuffing box. Although this is not frequently required, sometimes an extra 1⁄8″ or 1⁄4″ can be obtained by machining back the stuffing box face. This also creates a good sealing surface. Challenge the seal vendor to think creatively. Many times seal designers become entrenched in thinking only one way. Demand creativity on the designer’s part. Some questions to ask might include: 1. Does the carbon disaster bushing need to be as long as it is

1. Cutoff off hook sleeve. Still acts as a spacer for the impeller and left for proper clearance to throat bushing. 2. Cartridge sleeve mounted directly to pump shaft. 3. Stuffing box repair by boring and sleeving existing stuffing box.

THE HUMAN FACTOR Although the seal may have been designed properly, it now must be installed correctly. Cartridge seal installation is relatively simple, and therein may lie the problem. Reactions from mechanics often range from “Don’t you trust us in setting the seal?” to “Boy, these cartridges are a breeze to install.” In either case, action must be taken to educate the mechanics. In the first case, the mechanic has a point unless the person responsible for converting the cartridge seal takes the time to educate him. Circumstances such as a reduction in inventory, parts availability or vendor warranties may be all that needs to be explained. Or the mechanic may require more detailed information. It is important to take the time to explain the reasoning behind the decision. One technique that may prove helpful is to review the seal drawing with the mechanic and explain the different parts and stackup of the seal. Once he has an understanding of the seal, he may no longer think of it as just another part. He may still think he could do it “just as well,” but at least he understands the integrals of the seal and does not consider himself a “parts changer.” In the second case, the mechanic may tend to become complacent when installing the seal. The cartridge seal can be labeled a “cureall.” Items such as shaft runout and shaft-to-sleeve fit may be forgotten unless the importance of these and similar measurements is stressed.

OTHER CONSIDERATIONS

4. Cartridges seal pilot fit. 5. Grafoil rings for sealing sleeve to shaft. 6. Drive screws for setting seal to shaft. 7. Retaining bolt clips. Note: These can be either hex or allen bolts. Accessibility can be difficult on retrofit pumps. FIGURE 1. Single cartridge seal with stationary bellows

44

drawn? 2. Have you considered modifications that other plants have made to similar pumps? 3. Have you considered all seal types: bellows or pusher? 4. Can I install the flush line in the stuffing box instead of the gland?

The Pump Handbook Series

Another consideration is that the location of the grafoil ring on the shaft sleeve can become a problem. In hot applications, where a perflouroelastomer is not sufficient, grafoil rings are used. These must be crushed after the gland bolts and drive screws are tightened. Depending

on the room available between the gland and the bearing housing, accessibility to these bolts can be difficult. Cut off Allen wrenches or otherwise modified tools may be required to obtain enough torque to crush these rings properly. In addition, with the sleeve sealed on the gland end, product or coke may migrate underneath the cartridge sleeve, making removal of this sleeve difficult. With some vertical pumps the seal can be replaced without removing the pump and in some cases both the pump and the driver. Care should be taken when making this decision. If the seal fails, factors in the pump or driver may have contributed to the failure. If this sign is ignored, a more

costly failure could occur. Although the newer API 610 pumps have large stuffing boxes, many ANSI pumps do not come standard with large bore stuffing boxes. They do, however, offer large bore boxes as an option. When purchasing a new ANSI pump with a cartridge seal, specify this large bore box. Some ANSI manufacturers offer retrofit large bore boxes for older model ANSI pumps. Cartridge seals are a part of every plant. Understanding the benefits and limitations can lead to a cost-effective conversion. Allowing the seal vendor flexibility in design and offering pump modifications will increase the probability of suc-

The Pump Handbook Series

cessful conversion. And as always, successful installation depends upon educating plant employees.

ACKNOWLEDGMENT: The author would like to thank Patrick Guy, Tim Askey and Art Quirk for their help in developing this article. ■ Eddie Mechelay is mechanical supervisor at Total Petroleum—Denver Refinery, Denver, CO and a member of the Pumps and Systems User Advisory Board.

45

MECHANICAL SEALS & SEALING SYSTEMS HANDBOOK

Seal Face Application: Little Room for Error Critical safety and operational issues depend on precision mechanical face seals. By Laurence J. Thorwart echanical face seals are on constant duty protecting the plant and environment from product leakage. Whether it is a refinery or chemical process pump, paper mill or boiler feed pump, primary coolant pump in a pressurized water reactor, steam turbine, compressor, mixer, agitator or centrifuge, the mechanical seal is critical to safety and efficiency. Mechanical seals are engineered and designed for various applications. Therefore, consultation with the seal company is crucial to ensure proper application.

THE MANY FACES OF MECHANICAL SEALS

PHOTO COURTESY PURE CARBON COMPANY

M

ery. The seal faces are exposed to high pressures in pipeline service pumping natural gas liquids, and up to 2000 psi in boiler feed pumps. The seal faces are lapped flat, often to within one helium lightband (.000011″), which doesn’t allow much room for error.

Likewise, the heart and lungs of a mechanical seal, the seal faces, also must be selected and applied properly for the various environments in which they perform. The seal faces are almost always exposed to the product, which may be at very high or low temperatures, corrosive, abrasive, susceptible to flashing or extremely viscous. Even in buffered or barriered arrangements, the seal faces must be capable of “withstanding” the product as well as the barrier or buffer fluid. If the fluid film is lost, the faces of a mechanical seal may rub against each other at surface speeds up to several hundred ft/s in turbomachin-

Various carbon graphite and silicon carbide seal faces

46

The Pump Handbook Series

A wide selection of engineered materials is available for medium and heavy duty applications. Carbon graphite materials are often found at one side of the seal pair, while silicon carbides, tungsten carbides and aluminas are usually the mating face. Today, the increasing demands of applications often put these hardface materials in combination with each other, eliminating the carbon face, but more on that later. Other face material options include metals such as Ni-Resist cast iron, stainless steels and bronze, while coatings such as chrome, chrome oxide, chrome carbide and tungsten carbide are available for specific niche applications. Powder metal, composites of Teflon and other composites are used in a smaller number of applications. Table 1 lists some of the basic physical and chemical resistance characteristics of the many available faces. Table 2 outlines the basic advantages and disadvantages of the most common face materials.

CARBON GRAPHITE Graphite’s self-lubricating characteristic combined with the wear resistance of amorphous carbon make carbon graphite a suitable material for most sealing applica-

COMPATIBILITY WITH MATERIAL

CERAMICS Alumina - 85% Alumina - 99.5% Carbon - hard Carbon - soft

HARDNESS MOHS

YOUNG’S TENSILE MODULUS STRENGTH E, PSI (103) PSI

THERMAL EXPANSION IN/IN/°F (10-6)

THERMAL CONDUCTIVITY BTU/HR. FT.2

DENSITY G/CC

TEMPERATURE LIMIT, °F (a)

ACIDS

CAUSTIC OILS

HALOGENATED SOLVENTS

°F/FT.

8 9 4 1

30,000 50,000 3,500 2,900

20,000 35,000 7,000 3,000

3 4 3 2

7 14 6 8

3.4 3.9 1.8 1.6

2,500 3,100 500 500

Y Y Y ?

? Y Y Y

Y Y Y Y

Y Y Y Y

2 9 8

1,500 3,000 22,000

5,000 5,000 15,000

2 3 2

40 30 85

1.8 1.9 2.8

1,000 500 1,000

Y Y Y

Y Y ?

Y Y Y

Y Y Y

9.7 9.7

60,000 55,000

35,000 30,000

2 2

80 85

3.1 3.1

3,000 2,500

Y Y

Y ?

Y Y

Y Y

METAL & CERMETS Gray Cast Iron Hastelloy “B” M-2 Tool Steel

5 6 7

15,000 30,000 30,000

30,000 100,000 200,000

6 10 6

25 25 14

7.2 8.9 8.2

400 1,400 900

N Y N

Y Y Y

Y Y Y

? Y ?

“Niresist” Stainless Steel (316) Stainless Steel (440C)

4 4 5

15,000 29,000 29,000

60,000 100,000 100,000

10 9 6

8 9 14

7.4 8.0 7.8

900 1,100 1,100

? ? N

Y Y Y

Y Y Y

Y ? ?

“Stellite” Tungsten Carbide/Co Tungsten Carbide/Ni

7 8 8

32,000 90,000 85,000

150,000 200,000 100,000

8 3 3

8 60 55

8.4 15 15

1,800 2,000 2,000

N ? ?

Y Y Y

Y Y Y

? Y Y

COMPOSITES Phenolic/carbon Teflon/gas

2 3

100 50

2,900 1,500

25 50

1 1

1.8 2.1

200 300

? Y

? ?

Y Y

Y Y

Graphite - fine grain Siliconized Graphite Silicon Carbide Graphite Comp. Sintered Silicon Carbide Reaction Bonded Silicon Carbide

(a) Severe oxidation in air, or significant loss of hardness, or charged microstructure

Y = Probably Satisfactory

N = Not Recommended

? = Depends on Conditions

TABLE 1. Physical properties/compatibility – primary and mating ring materials

tions. A wide range of grades is available, many of which are unique because of the impregnant used to enhance the material while rendering it impervious. Table 3 lists common impregnants used in the industry. This is not to underestimate the tribological and chemical characteristics of the wide variety of carbon graphite base materials which the manufacturer has at his disposal. For example, alkylation units used in the production of unleaded gasoline require pumping hydrofluoric acid. When a carbon graphite containing residual minerals is exposed to HF acid, the minerals swell and put the carbon seal in tension. This can eventually create a series of mud cracks. The carbon manufacturer can recommend a special grade without these residual minerals for HF service. However, the special grade used in HF acid may wear rapidly in a seal used in a heat transfer fluid pump. Without the minerals present to polish the mating surface and maintain equilibrium, the graphite

Primary Advantage

Primary Disadvantage

Carbon Graphite

• Lubricity • Corrosion Resistance

• Abrasion

Silicon Carbide

• Abrasion Resistance • Stiffness • Corrosion Resistance • High Performance (P-V)

• Brittleness

Tungsten Carbide

• Toughness • Abrasion Resistance • Stiffness

• Corrosion

Alumina

• Abrasion Resistance • Stiffness

• Thermal Shock

Metal

• Fabrication

• Low Performance

Coatings

• Application

• Delamination • Wear-through

TABLE 2. Primary and mating ring materials

can overfilm, causing rapid seal wear. Thermoset resins are the most common impregnant used to seal the pores of the carbon graphite face. They are chemically resistant and impart increased physical characterThe Pump Handbook Series

istics but, contrary to popular belief, do not bind the carbon together. The carbon graphite is, in fact, a carbon bonded composite. Antimony, an interesting metal impregnant that has become very popular recently, improves the

47

IMPREGNATION

SERVICE

Thermoset Resin

General duty to 500°F in water, coolants, fuels, oils, light chemical solutions, food & drug

Antimony

Hot water, steam, light hydrocarbons

Copper or Silver

High pressure service to 3000 PSI

Carbon

Highly corrosive environments

Film Former (Fluorides, etc.)

Dry running—vacuum & cryogenics

Oxidation Inhibitor (Phosphates, etc.)

High temperature & speed—turbine engine applications— to 800 ft/sec—and 1000°F.

TABLE 3. Carbon graphite material application selection

strength and thermal conductivity of the material, but more importantly it reduces friction and wear. Manufacturers of carbon materials believe that the antimony oxide film transferred to the mating face improves the tribology at the interface due to the solid lubricant-like properties of the oxide film. Improving the conductivity of the seal material to remove frictional heat is almost always beneficial. Antimony-impregnated carbon is most commonly used in hot water, steam, hot oils, higher temperature fluids and light hydrocarbon service. The low Young’s modulus or “stiffness” of carbon graphite is a double-edged sword. On one hand, it allows compliance to the seal face and offers a small measure of “forgiveness.” However, to prevent deformation of the carbon in high pressure applications, the designer may have to retain the carbon in a metal housing to improve stiffness, or substitute a higher modulus seal face material.

SILICON CARBIDE Silicon carbide is an excellent rubbing surface for carbon graphite because it is extremely hard and abrasion resistant, and at the same time highly conductive and very stiff. This advanced ceramic is available in four basic types including: siliconized graphite, reaction bonded, sintered and composites with graphite or spherical porosity. Friction and wear testing, as well as field application, have proven silicon carbide’s versatility and improved pressure-velocity capability. Figure 1

48

shows a relative comparison of the most popular mating face materials when run against a premium resin impregnated carbon graphite. Although not dry-run capable in the strictest sense, composites of SiC with graphite, controlled porosity or lubricating additives have been greatly improved to allow a level of dry-run survivability. While these composites allow running in marginal service such as poorly lubricated applications or in dry-running startup or upset conditions, catastrophic loss can occur with solid silicon carbide.

TUNGSTEN CARBIDE This cermet is a workhorse because of its extreme toughness, stiffness and abrasion resistance. Although the use of silicon carbide has grown rapidly, tungsten carbide

MATING RING

ALUMINA

METALS

ALUMINA When aluminum oxide came onto the seal face scene a generation ago, it heralded a new age of hard, abrasion and corrosion resistant material that rapidly advanced seal performance and life. Several purities are available. The lower purity of 85% or more is suitable for low duty, high volume applications such as hot water circulation or coolant pumps. The high purity, 99.5% variety is better suited to higher P-V and chemically aggressive duty. Alumina is still popular in CPI as well as water pump and appliance applications.

HARDFACE MATING PAIRS When carbon is not suitable as one of the seal faces due to its physical limitations (abrasion, pressure or temperature), the user now has an alternative. Hard materials can be run in combination with each other if designed and applied properly. The most popular hardface com-

CARBON GRAPHITE PRIMARY RING

SILICON CARBIDE TUNGSTEN CARBIDE

faces still can be found in a large range of applications from oil well drilling and mining to papermills. Two basic varieties are available: cobalt and nickel bound. The presence of these metals aids in the sintering and binding of the composite and varies by grade in its content. Typically for seal faces, 6% metal binder content is used. Cobalt generally is tougher, while nickel offers improved corrosion resistance.

CARBON-GRAPHITE-RESIN IMPREGNATED

CARBON-GRAPHITE-RESIN IMPREGNATED

ANTIMONY IMPREGNATED

ANTIMONY IMPREGNATED

C-G-R

C-G-R

PRESSURE VELOCITY LIMIT (P-V) FIGURE 1. Relative comparison of mating pair capability The Pump Handbook Series

binations are silicon carbide running against a sister variety, or in combination with tungsten carbide. The advanced materials available today designed into engineered seals from the mechanical seal OEMs now allow safer use of hardface mating pairs. However, the user still must be cautious because these hardfaces are not yet truly dry-run capable. Nearly every seal manufacturer offers a hardface combination for abrasive duties such as oil drilling mud, chemical slurries, paper stock, food products such as candy, or pre-processed crude oil. Generally, two dissimilar materials run better in combination with each other—such as silicon carbide versus tungsten carbide or one type of silicon carbide versus another.

The SiC composites of graphite, or with additions of spherical porosity, are extremely effective in improving the rubbing characteristics of the combination by helping to reduce friction through retention of fluid in the “pockets” at the interface. The resultant hydrodynamic fluid film is critical in keeping the seal’s faces from intimate contact. Note that suitability of hardface combinations in certain applications depends greatly on seal design and service.

SELECTION AND IMPLEMENTATION

of the user, but mechanical seal manufacturers should always be consulted. Select manufacturers of face materials have invested in friction and wear testing capability and can assist the seal supplier in recommending the optimum “mating pair” for your particular service. ■ Laurence J. Thorwart is Vice President–Marketing & Product Engineering for Pure Carbon Company, St. Marys, PA. He holds a B.S. in Chemical Engineering from The Pennsylvania State University and is a registered Professional Engineer.

Today’s advanced face materials used in engineered mechanical seals provide excellent service and reliability. Proper application of face materials ultimately is in the hands

The Pump Handbook Series

49

MECHANICAL SEALS & SEALING SYSTEMS HANDBOOK

Dual Sealing Systems; a.k.a.Double and Tandem New terminology or old, dual sealing systems offer proven solutions for today’s emissions requirements. By Patrick M. Flach f the term “dual sealing systems” doesn’t mean much to you, don’t panic. The wording is new to almost everyone. If you are familiar with double and tandem seals, however, you will soon feel comfortable with the new terminology. Although the terms “double seals” and “tandem seals” describe arrangements of mechanical seals, they have come to refer to pressurized or non-pressurized seals as well.

I

OLD NOMENCLATURE AND DEFINITIONS The following familiar names and definitions come from the Fluid Sealing Association (FSA) Mechanical Seal Handbook: Double Seal. Two mechanical seals mounted back-to-back, face-toface or in tandem, designed to permit a liquid or gas buffer fluid between them. (See Figures 1a, 1b, & 1c.) Tandem Seal. A multiple seal arrangement consisting of two seals mounted one after the other, with the faces of the seal heads oriented in the same direction. (See Figure 1c) Buffer Fluid. A fluid that is introduced between two seal elements, quite often at a pressure that is higher than the pressure of the fluid on either side of the seal assembly (also called a barrier fluid).

A

B

C

NEW NOMENCLATURE AND DEFINITIONS The following new terms are found in the Society of Tribologists and Lubrication Engineers (STLE) Special Publication SP-1 A Glossary of Seal Terms and the American

50

FIGURE 1. Double Seal: A. Back-to-back; B. Face-to-face; C. Tandem The Pump Handbook Series

Petroleum Institute API Standard 682. Dual Mechanical Seal. A seal arrangement using more than one seal in the same seal chamber in any orientation that can utilize either a pressurized barrier fluid or a nonpressurized buffer fluid. (Previously referred to as a double or tandem seal.) Barrier Fluid. A fluid that is introduced between dual mechanical seals to isolate the pump process liquid from the environment. The pressure of the barrier fluid is always higher than the process pressure being sealed. Buffer Fluid. A fluid used as a lubricant or buffer between dual mechanical seals. The fluid is always at a pressure lower than the pump process pressure being sealed. Bellows Seal (metal). A type of mechanical seal that uses a flexible metal bellows to provide secondary sealing and spring-type loading. Pusher Type Seal. A mechanical seal in which the secondary seal is mechanically pushed along the shaft or sleeve to compensate for face wear. From these new definitions a new terminology has evolved: Dual Seals Pressurized or Dual Seals Unpressurized. A Dual Seal Pressurized uses a barrier fluid and was previously referred to as a double seal. A Dual Seal Unpressurized uses a buffer fluid and was previously referred to as a tandem seal.

EARLY DUAL SEALS During the late 1940s and early 1950s soft packing was the preferred method for controlling leakage from centrifugal and rotary pumps. As mechanical seals made their way into industry, questions were raised about their safety. Union Pump Company (formerly Union Steam Pump), a manufacturer of centrifugal pumps for the refining industry, answered these questions by developing and obtaining a patent for its Uni-Lok System (Figure 2). This system consisted of two spring loaded rings, soft packing and a buffer tank in series with a primary single mechanical seal. Theoretically, if the mechanical seal with the new technology failed, the two rings of packing would act as a backup and control the leakage. Uni-Lok can be considered one of the first dual seal concepts com-

HARD SURFACING UNDER UNI-LOK PACKING FIGURE 2. Uni-Lok System

mercially available. We have come a long way since its inception, but many modern sealing systems still use a buffer/barrier tank. The difference is that contemporary systems employ more than one seal in the stuffing box or seal chamber.

FEDERAL EMISSION REGULATIONS Mounting concern about emissions of hazardous compounds into the atmosphere has led to stricter regulations on leakage from pumps and other equipment. The 1990 amendments to the Clean Air Act (CAA) are examples of such regulations. Under Title III of the CAA the U. S. Environmental Protection Agency (EPA) has signed the National Emission Standard for Hazardous Air Pollutants (NESHAP). This rule controls the amount of emissions of certain volatile hazardous air pollutants (VHAP) from manufacturing processes producing synthetic organic chemicals. The rule is referred to as the Hazardous Organic NESHAP or HON. Subpart H of HON covers provisions for equipment leaks that affect pumps and other equipment in VHAP services. Leakage standards take the form of work practices and are not equipment mandates. The standards do define a leak and require that less than 10% of the pumps in a process be classified as leaking. Pumps in VHAP services must be visually checked each The Pump Handbook Series

week for drippage, which is defined as a leak. Seal leakage also may be detected during monthly tests, if results exceed the HON rules. Pumps in light liquid services equipped with dual mechanical seals are exempt from monthly monitoring if the barrier fluid is not a VHAP and the barrier pressure exceeds the process pressure. Dual non-pressurized seals are also exempt from monthly monitoring provided that the buffer fluid system is monitored and the buffer fluid reservoir is vented to a closed-vent system.” Rules pertaining to refineries are forthcoming.

DUAL SEALS The following dual seal systems are presented in STLE’s Special Publication SP-30 Guidelines for Meeting Emission Regulations for Rotating Machinery with Mechanical Seals. Non-Pressurized Seals Contacting Dry-Running Pressurized Seals Contacting Non-contacting Low-leakage, contacting mechanical seals are described as having faces that have found an equilibrium position such that they are barely touching. A very thin film of the product or sealing fluid lubricates the seal faces. This is referred to as “mixed

51

one must remember that the inner seal of a dual non-pressurized arrangement is lubricated by the product. As the product migrates across the faces to lubricate them, it mixes with the buffer fluid. As more and more product enters and mixes with the buffer fluid, some product fluid can pass by the outer seal to the atmosphere. This is why we say near zero emissions. During system upsets the buffer can lubricate the inner seal somewhat, and the outer seal acts as a backup seal containing emissions. Dual seal arrangements with non-pressurized buffer systems are not subject to EPA monthly monitoring. FIGURE 3. Containment seal with metal bellows provides long life for dry-running conditions.

2

PI

3

PAH

1 TO VAPOR RECOVERY SYSTEM

VENT “OUT” GLAND

PLANT NITROGEN (OPTIONAL) VENT “IN” FIGURE 4. Piping plan for dry-running outboard seal.

lubrication.” With the faces barely touching and mixed lubrication, friction is low, and little heat is generated. The result is acceptable wear over a long seal life. A non-contacting seal arrangement is a special design in which the faces are configured to be lubricated by a full-fluid film of the product or sealing fluid. The seal faces do not touch each other during operation. A dry-running seal is described as a dual non-pressurized seal, arranged in series, with a low emission inner seal and a dry-running full-contact outer seal. Non-Pressurized, Contacting. Dual (tandem) sealing systems have been used for years to control emissions to the atmosphere and as safety

52

back-ups. The inner seal of such an arrangement is cooled by piping similar to that described in API Plan 11, 13 or 32. A buffer fluid provides cooling and lubrication to the outer seal and buffer emissions. This is usually accomplished by means of an API Plan 52 or some similar piping arrangement. Circulation of the buffer fluid is accomplished by means of a pumping ring or other internal circulating device. Advantages • near zero emissions to atmosphere • contains emissions during system upsets • not subject to monthly monitoring To explain “near zero emission,” The Pump Handbook Series

Disadvantages • requires fluid buffer system higher initial cost higher maintenance cost buffer fluid maintenance • hazardous process fluids can migrate to buffer fluids hazardous waste The main reason for using a nonpressurized fluid buffer system with dual seals is to lubricate the outer seal. These systems consist of a seal pot or tank, connecting piping and associated valves and pressure gauges. They are usually made from stainless steel and have a high price tag. Although the cost of the buffer fluid can be an issue, the cost of maintaining these systems is a bigger concern. The more upsets in an operating system, the more migration of product to the buffer fluid, and the more it needs to be changed. After the buffer fluid has been changed, it becomes a hazardous waste and must be dealt with as such. Pressurized, Contacting. These dual (double) seals have also been used for years to control emissions to the atmosphere and as safety mechanisms. They are used in applications where the product being sealed is dirty, abrasive or polymerizing. The barrier fluid is used to cool and lubricate both sets of seal faces and therefore is at a pressure higher than the product being sealed (usually 20-25 psid). The barrier fluid enters the seal chamber, or stuffing box, near one set of seal faces and is removed near the other set of faces. To ensure circulation, an internal circulating device should be supplied. Seal setting is so

important to the performance of this type of seal that cartridge arrangements are strongly recommended. API Plan 53 and 54 are the most common pressurized piping arrangements used for these seals. Advantages • near zero emissions to atmosphere • contains emissions during system upsets • not subject to monthly monitoring As stated, this is a pressurized barrier system. Therefore, the barrier fluid tends to migrate into the product. The 20–25 psi pressure differential keeps most of the emissions in the pump, however, and does not allow them to migrate into the barrier fluid. Emissions to atmosphere are usually much lower than in a non-pressurized buffer system arrangement. Pressurized barrier seal systems operate better during plant upset conditions than non-pressurized buffer systems do because the inner seal faces are lubricated by the barrier fluid as opposed to the product. Dual pressurized seals also do not require monthly monitoring. Disadvantages • requires a fluid barrier system high initial cost high maintenance cost barrier fluid maintenance • hazardous process fluids can migrate to barrier fluid hazardous waste • barrier fluid can enter fluid • complex barrier fluid pressuring system Dual pressurized seal arrangements require barrier systems. They are like the non-pressurized systems except that they are more complex. Complex systems like these have a higher initial cost and are more expensive to maintain. During some system upsets, like losing barrier fluid pressure, hazardous process fluids can enter the barrier system and contaminate the barrier fluid which then becomes a hazardous waste that must be addressed. During normal operation some barrier fluid can enter the product because the barrier fluid is at a higher

pressure. Also, if the inner seal fails, the barrier fluid will enter the product; thus a compatible barrier fluid is a must and can be hard to find. Non-Pressurized, Dry-running. A non-pressurized, dry-running, secondary containment seal is a product of modern sealing technology. This type of seal is supplied as a cartridge arrangement that helps maintain proper face loads and allows for easy installation. Containment seals like the one in Figure 3 use a metal bellows that also helps maintain low face loads, the secret to long lasting dry-running seals. If a pusher seal were used as the containment seal, the force necessary to slide the Oring could be greater than the minimum face load required for low face wear and long life. These seals can be used with a bellows or pusher inside seal that is cooled and lubricated by an API Plan 11, 13, 23 or 32. The piping plan for the dry-running outboard seal is shown in Figure 4 and is discussed under advantages. Advantages • no buffer fluid system required low initial cost low maintenance cost • low emissions to atmosphere • no hazardous waste • contains emissions during systems upsets • not subject to monthly monitoring • full safety/containment feature Unlike non-pressurized or pressurized contacting and pressurized non-contacting dual seal arrangements, which require a fluid buffer/ barrier system to lubricate their faces, dry-running seals do not require fluid lubrication. This lowers initial as well as maintenance costs. Emissions passing by the inner seal faces are vented to a flare or vapor recovery system. In this way the pressures across the dry-running seal are minimized and very little process fluid emissions pass to the atmosphere. Also, the problem of hazardous waste, which is generated from contaminated barrier/buffer systems used on pressurized and non-pressurized contacting seals, is eliminated. The Pump Handbook Series

A dry-running seal contains emissions during system upsets and acts as a full safety backup seal. In addition, when connected to a Closed-Vent System (vapor recovery system or flare), a dry-running seal meets the requirements of the National Emission Standard for Equipment Leaks and is not subject to monthly monitoring. Disadvantages • requires a vapor recovery system, flare or other closed-venting system The National Emission Standard for Equipment Leaks requires this type of seal to be connected to a flare or vapor recovery system. Pressurized, Non-Contacting. Technology in the form of sophisticated computer programs and new manufacturing techniques were used during the development of gas compressor seals. Today this technology is being applied to pump seals in the form of grooved faces. These offer both hydrostatic and hydrodynamic lift-off capabilities. A hydrostatic seal maintains a film thickness between the faces by means of pressure, even without relative motion. The pressure is provided either by external source or by the pressure differential across the seal. A hydrodynamic seal has special geometric features on one or both of the mating faces. These features are designed to produce separation between the faces. This “lift” arises solely from the relative motion between the stationary and rotating portions of the seal faces. By pressurizing these dual seals with nitrogen or other inert gases, separation of the seal faces occurs at a very low rpm. Because the faces are open, very small amounts of nitrogen leak into the product being sealed and to the atmosphere. Also, there is no contact at the faces; therefore, very little heat is generated. Advantages • zero emissions to atmosphere • no liquid barrier/buffer system required • low power consumption • low heat generation • low wear–long life • not subject to monthly monitoring Release of fugitive emissions to the atmosphere is eliminated by

53

maintaining a gas barrier system pressure of 20-30 psi over the product pressure being sealed. The pressure of the gas barrier fluid also provides hydrostatic lift, which helps separate the faces. During normal operation minimal power is required to drive the seal, so very little heat is generated, resulting in low face wear and long life. Monthly monitoring of emissions is not required for these pressurized dual seals. Disadvantages • requires nitrogen or other inert gas supply • requires a gas supply regulating system For proper operation of this type of seal, it is essential to maintain and regulate a supply of inert gas at 20-30 psi over the product pressure.

54

CONCLUSION The technology of controlling pump leakage has come a long way since the days of soft packing. The first step was the development of single mechanical seals. Then came systems like the Uni-Lok, dual non-pressurized contacting (tandem) seals, and dual pressurized contacting (double) seal. Dual non-pressurized and dual pressurized seals are being used in many applications today and will continue to be used because of their performance history and service life. Emission regulations have inspired seal designers and manufacturers and users of pumps to work together in advancing technology even further. Today, as a result, pump users can choose either a pressurized, non-contacting or a nonpressurized, contacting seal as the solution to particular sealing applica-

THE PUMP HANDBOOK SERIES

tions. ■ Patrick M. Flach is the western hemisphere technical services manager and group leader for new product development in the Industrial Division of EG&G Sealol. He has more than 28 years of experience in pump design and applications. Mr. Flach served on the manufacturers task force during the writing of the 6th and 8th editions of the API 610 standard.

MECHANICAL SEALS & SEALING SYSTEMS HANDBOOK

Seal Reliability At Chevron An innovative, multi-pronged approach throughout all plants has drastically increased mechanical seal reliability, saving millions in maintenance costs. By V. Ray Dodd and Robert C. Watkins hevron U.S.A. Products Company consists of six U.S. refineries with a total rated crude capacity exceeding 1,000,000 barrels per day. In total, Chevron operates 26,000 centrifugal pumps and 10,000 shaft seals. During the last decade considerable progress has been achieved in mechanical equipment reliability. By developing and using an equipment reliability program, Chevron was able to reduce maintenance costs on mechanical equipment by approximately 50%. However, during the same period (1979-1987) we made very little improvement on mechanical seal reliability (Figure 1). Average mechanical seal service life was relatively short at only 1.7 years, and the trend showed few signs of improvement. Admittedly, rotating equipment failures due to seal mal-

C

functions at Chevron were less than the industry average, but they still were consuming maintenance dollars at an unacceptably high rate. Industry data indicated that 75% of all centrifugal pump maintenance is the result of mechanical seal failures. Our data agreed.

THE PROBLEMS Recognizing the opportunities, Chevron formed an internal task team in 1988 to attack the problems. No sacred cows were permitted, so the team thoroughly investigated every aspect of: • lack of industry or company standards • seal purchasing practices • seal misapplication • lack of in-house seal expertise • spare parts inventory • maintenance procedures

110% BEFORE

AFTER

Percent of Basepoint

100% 90% 80% 70% 60%

Mechanical Seal Maintenance Cost

50%

Machinery Maintenance Cost

FIGURE 1. Historical machinery and mechanical seal maintenance costs The Pump Handbook Series

1994

1993

1992

1991

1990

1989

1988

1987

1986

1985

1984

1983

1982

1981

1980

1979

40%

• manufacturer relationships • lack of standardization among refineries These and several other factors were all placed on the table. The team’s mission was to quantify current seal reliability status, develop a detailed plan to optimize seal reliability, meet or beat emissions regulations, reduce spare parts inventory, develop in-house expertise through training, and at all times maintain state-of-the-art technology in all seal applications. The program became effective in 1988. Chevron’s experience with other reliability efforts indicated that initial investments in manpower and training would be required. However, we also knew from previous experience that the end results dramatically offset the up front investment.

THE PLAN The task team recognized the need for a change in business philosophy. Chevron had to take responsibility for the reliability of its mechanical seals, with each location required to have a dedicated mechanical seal analyst. The task team needed to develop standard seal designs for refinery services and address all sealing systems, including piping, support equipment and seal hardware. For any mechanical seal reliability program to be successful, you first have to make sure you have a reliable pump. No one can solve a pump problem by simply changing the seal design. In some cases, it may be necessary to modify or even replace older pumps with modern designs.

55

What the task team wanted to find was a “super seal”—one seal to solve all the problems. However, they determined that refineries handle only two basic liquids, oil and water. While it proved unrealistic to try to identify the one seal design that could be applied to all refinery services, it has been possible to find the best seal for most services and, in some cases, a single best seal design for more than one service. The team broke these two categories down as: Refinery Services hot/cold water caustic acid Hydrocarbon (oil) flashing non-flashing By examining seals that were operating reliably in various Chevron facilities, the task team determined that these product categories could be satisfactorily sealed by using only two seal types in four different arrangements (Figures 2 and 3).

Type A Standard

Water • Cold

< 180°F

Acid

< 180°F

Hydrocarbons • Non-Flashing

< 300°F

Hydrocarbons • Flashing

< 140°F

Caustic • Amines

< 180°F

Water • Hot

> 180°F

Type A Dual Arrangement 2 Type A Plan 23

Sour Water Hydrocarbons • Non-Flashing

> 250 psi Type C Standard

Hydrocarbons • Flashing

> 140°F

Hydrocarbons • Non-Flashing

> 300°F < 250 psi

Non-hydrocarbon (water)

ELEMENTS OF SUCCESS Because an investment in manpower and training resources is necessary, a program of this magnitude cannot be accomplished without the commitment of management. Goals for departments and individuals must be changed to reflect the new philosophy. The old method of buying seals has been completely revised. No longer are seals selected and sold by seal sales personnel. Instead, all requirements are reviewed first by Chevron’s professional reliability people and by the selected manufacturer’s technical representative appointed to serve Chevron. The specific seal design that is best for the service (determined and exhaustively factory tested under conditions simulating field conditions) was selected. The initial cost of a seal should not be the major factor; the total cost of ownership should be. So competitive bidding may not result in the best seal for the job. The positive working relationship we developed with our mechanical seal manufacturers also played a major role. In-house reliability training is ongoing with the direct assistance of designated manufacturer technical

56

> 300°F

FIGURE 2. Refinery services seal type and arrangements

Item

Pre-Standardization

Post-Standardization

Reduction

Seal Manufacturers

4

2

50%

Seal Types

19

3

84%

Seal Sizes

21

10

48%

Bellows Designs

7

1

86%

Insurance Stock

100%

40%

60%

FIGURE 3. Seal standardization program typical crude unit (58 pumps with 67 seals)

representatives and led by the most knowledgeable Chevron reliability experts. Three to ten full-time reliability professionals assigned to each refinery, who have become expert in identifying the root causes of seal problems, are equipped to evaluate them and determine the best possible solution. This has enabled us to assume responsibility for our mechanical seal reliability and not be dependent on seal sales people. Coupled with maintenance training, quality maintenance procedures and check lists, the program has produced outstanding results. Our centralized computer database provided several benefits. First, it allowed us to identify that we had a problem. Otherwise, we might have been satisfied with a two year service life. It also provides the ability to spot “bad actors” in all refineries. The methodology here is to “flag” these seals on their second failure, allowing us to re-engineer the seal design for the service. In this way, seals exhibiting excessive failure rates are being eliminated. Another benefit of this centralThe Pump Handbook Series

ized database is automatic, companywide notification of problems solved. After all, if you solve a vacuum column bottoms pump seal problem in one refinery, the same solution should be good for all vacuum column seals. Consequently, each refinery no longer has to reinvent the wheel for itself, saving untold dollars and grief. Used universally and consistently, the Chevron database is a great tool to measure and track progress of our seal program. The data driven management decision process allowed us to make major changes in the way we purchased and maintained our mechanical seals. In addition to tracking the data, we also published the results company-wide. Not only did Chevron see the need to handle seal problems in the same manner at all six refineries, but it recognized that seal manufacturers needed to provide the same service level and the same sealing solutions to all six refineries. Drawing on past experiences with other equipment, Chevron requested that manufactur-

cost reduction of nearly 40%. Through seal standardization, spare parts inventory has been reduced drastically. Not only has the number of different seal manufacturers, styles, sizes and bellows designs been reduced by 48 to 86%, but also the value of spare parts inventories has been cut by 60%.

6

Service Life (Years)

5

4

CONCLUSION

3

1994

1993

1992

1991

1990

1989

1988

1987

1986

1985

1984

1983

1982

1981

1980

1979

2

FIGURE 4. Mechanical seal service life (MTBR)

ers provide one person to handle the technical aspects of their business on a national level. This one person at the seal manufacturer would be responsible for all communication regarding seal designs, troubleshooting, and other technical problems for all six Chevron U.S. sites. Relatively informal standardization agreements with preferred seal manufacturers allow for audits to assure that Chevron obtains the best possible pricing provided to the seal vendor’s best customers. In addition, these preferred manufacturers must agree to commit service center and factory resources to provide complete replacement cartridge seals from their inventory and completed seal drawings within 72 hours for pre-engineered seals. This permits Chevron to retrofit pumps removed from service with minimum downtime.

THE 682 CONTRIBUTION The publication of API 682 has resulted in adopting the standard as the Chevron standard. Full compliance is required with all provisions of the new “user-driven” API 682,

the first API standard strictly to address issues related to mechanical seals. It requires cartridge seals and includes “fluid-handled” seal selection steps that lead to default or preselection of the best seal for the job. These five services cover about 95% of Chevron refinery mechanical seal services. (The remaining 5%, of course, require special engineering.) Figure 2 lists the mechanical seal services used to designate standard seals by design and type that are now universally used in Chevron refineries. In some services, alternate seal designs and types are permitted, especially where spare parts inventories of the alternate design are high.

SOLUTIONS AND RESULTS The results of the task team’s efforts have been substantial and are ongoing. The mean time between repair (MTBR) for the 10,000 mechanical seals throughout Chevron refineries has increased from less than two years in 1985 to almost six years (Figure 4). This produced an overall seal-per-year maintenance

The Pump Handbook Series

As a result of this concerted effort, management support and the active assistance of the seal vendors’ technical representatives, Chevron has been able to increase mechanical reliability by a factor of three, drastically reducing excessive mechanical maintenance costs, and at the same time, meet or exceed current emissions regulations. This gives us the best of all worlds, a reliable and affordable sealing system, lower operating costs and a cleaner environment. ■ V. Ray Dodd is the senior equipment reliability specialist for Chevron U.S.A. Product Company. He has been with Chevron for more than 35 years, working in operations, maintenance, inspection and engineering. Mr. Dodd established Chevron’s Equipment Reliability Program. Robert C. Watkins is the mechanical equipment reliability specialist for Chevron U.S.A. Products Company’s Refining Division. His career at Chevron encompasses 27 years, including 10 years in maintenance and 17 years in the machinery reliability field. He currently coordinates the Refining division’s Machinery Reliability Groups and is the Machinery Best Practices Master.

57

MECHANICAL SEALS & SEALING SYSTEMS HANDBOOK

Seal Chamber Design Affects Reliability, Emissions Varied seal chamber designs help tailor pump to process. By Nick Ganzon ndustry is making a concerted effort to improve the mean time between planned maintenance of centrifugal pumps. A major part of this effort focuses on reducing the failure rate of mechanical seals by enhancing seal environments. As a result, a proliferation of new mechanical seal chamber designs has become available to users. These advancements enable users to tailor the pump to the process. This article focuses on different seal chamber designs and how they are used to optimize performance. Three chamber styles are covered: the conventional stuffing box or standard bore, the enlarged bore and the tapered bore. Variations on these three designs, including jacketed seal chambers and removable seal chambers, are discussed also.

I

Seal chambers can influence the seal environment in four basic ways: 1. Pressure 2. Solids handling 3. Vapor removal 4. Temperature To understand how the seal chamber affects the sealing environment, it is helpful to understand each basic design: Standard Bore. More commonly known as a stuffing box, this seal chamber was originally designed for soft packing and therefore has radial and axial dimensions based upon standard packing sizes. A restriction at the bottom of the chamber limits the interchange of fluid between the seal chamber and the pumpage.

58

Enlarged Bore. This is similar in design to the standard bore seal chamber, but it is specifically designed for mechanical seals. It has increased radial clearances that create a larger volume of fluid encasing the seal than does the standard bore. The bolt circle for the gland is also enlarged. Tapered Bore. This seal chamber features the increased radial clearance of the enlarged bore; however, the bottom of the chamber is exposed to the impeller, and the walls of the seal chamber are tapered to promote self venting upon shutdown and self draining during disassembly. Most current designs include vanes or ribs to modify the flow within the chamber for improved handling of solids and vapors. Refer to Figures 1 and 2 for nomenclature and general cross sectional views. Seal chamber selection goes hand in hand with mechanical seal selection, and how critical the intended service is drives both of these choices. Having explained the three basic seal chamber designs, we can now discuss their interaction with the sealing environment.

PRESSURE Pressure is perhaps the most important seal and seal chamber selection criterion. To prevent flashing of the product, chamber pressure should be about 25 psi higher than the vapor pressure of the process fluid. If this pressure differential cannot be assured through pump design, a seal chamber flush should be used to provide it. This is especially important in low NPSHa applications. The procedure for calculating The Pump Handbook Series

the seal chamber pressure is similar for all designs, and it requires knowledge of both the operating conditions and basic pump design. It is necessary to identify the location of the seal chamber relative to the impeller. As simple as this may seem, it is extremely important in determining the pressure adjacent to the seal chamber, which is typically considered to be the seal chamber pressure itself. The seal chambers on between bearing pumps are generally located next to the pump suction. This is especially evident with double suction pumps, in which both seal chambers are adjacent to the suction bays. Seal chamber pressures in these cases are simply suction pressure. Most multi-stage between bearing pumps include bypass lines from the high pressure seal chamber back to suction, effectively creating suction conditions in this chamber. In end suction pumps the seal chambers are behind the impeller on the discharge side. The seal chamber pressure here is the suction pressure and a percentage of the total developed head. The design of the impeller is a major factor in this percentage. In general, the seal chamber pressure of pumps with different impeller configurations may be expressed by the equation: Psc = Psuction + K x TDH x 2.31 Pumps with pump-out vanes

SG Where

Psc = seal chamber pressure, psi

K = impeller design factor enclosed impellers (with no balance holes) = .7 impellers with pump-out vanes = .15 impellers with balance holes = .10 between bearings pumps = 0 TDH = total developed head, feet SG = specific gravity The impeller design factors (K) listed here are general values usually sufficient for seal selection. They change with speed, impeller diameter and clearances. If a more accurate seal chamber pressure is required, contact your manufacturer. Pump-out vanes and balance holes are most often used for two purposes: decreasing seal chamber pressure and reducing axial thrust. Open impellers usually have pumpout vanes while enclosed impellers use balance holes. There are open impeller designs that use both pumpout vanes and balance holes; in these cases use the balance hole K factor. On standard bore and enlarged bore seal chambers with restricted

throats, an increase in seal chamber pressure may be observed when a seal flush is applied. This pressure increase depends on the inlet pressure, flush rate, and the clearance at the throat to shaft area. The throat clearance is normally large enough to prevent pressure build-up. However, close clearance bushings may be installed to reduce leakage and build pressure as required for the application.

SOLIDS HANDLING Solids within the pumpage are a major obstacle to achieving maximum mean time between planned seal maintenance. Even in applications where no solids are intentionally introduced into the system, particles from corrosion or improper cleaning of pipes may be prevalent enough to damage the seal or seal chamber. The ideal seal chamber prevents solids from entering the seal environment and thus eliminates erosion problems from the equation. To accomplish this requires a seal chamber with a restricted throat, such as the standard bore or enlarged bore with a flush. The flush provides a flow from the seal chamber into the pumpage and prevents solids from entering. In

Face of seal chamber

First obstruction Flush

applications where the concentration of solids is above 10% by weight, this is the preferred method of seal environment control. The enlarged bore seal chamber is preferred over the standard bore because increased radial clearances allow any solids that may infiltrate the seal chamber to be centrifuged to the periphery of the chamber and kept away from the seal. The standard bore seal chamber cannot keep the solids away as effectively as the enlarged bore. One drawback is that once solids do enter the seal chamber, they have no natural exit and erode the chamber. Additional flush only moves the rotating ring of solids axially. It does not expel them from the chamber. Solids removal requires drawing the solids off through a tap or shutting down the pump, dismantling it and removing material by hand. On applications with solids concentrations below 1-2%, it may not be necessary to prevent solids from entering the seal environment – only to prevent their accumulation. In these situations the tapered bore seal chamber with vanes or ribs is preferred. The advantage is that flushing often can be eliminated, avoiding the cost of auxiliary equipment and flushing costs. For solids concentrations up to

Bottom of seal chamber

Psuction = suction pressure, psig

Restriction

Radial Clearance Shaft OD

Clear length or distance to first obstruction

Depth

Figure 1. Seal chamber nomenclature The Pump Handbook Series

59

2A

2B

10%, there are more specialized tapered bore seal chambers that refine the flow within the seal chamber even further. An example of this is a design that uses angled vanes as depicted in Figure 2c. This design can also be applied without a flush and may accommodate solids concentrations higher than 10% with a flush. The tapered bore seal chamber received much attention and criticism in recent years because it had a high incidence of erosion failures in services with solids (see Erosion sidebar). Current designs of tapered bore seal chambers have eliminated these problems, and this approach now provides satisfactory environments for the seal in a wide variety of conditions without a flush.

VAPOR REMOVAL 2C

2D

Figure 2. Seal chamber styles A. Standard bore (packed box) is characterized by long, narrow cross section. Originally designed for soft packing, mechanical seals were forced into cavity envelope. Requires an API/CPI flush plant for optimal performance. B. Enlarged bore features increased radial clearances over the standard bore. This chamber design enables optimal seal design. Restriction at bottom of the seal chamber limits fluid interchange. Requires an API/CPI flush plan for best performance. C-D. Tapered bore features increased radial clearances similar to the enlarged bore, except there is no restriction at the bottom of the cavity and is open to the impeller backside. Current designs include vanes or ribs to provide solids and entrained gas handling. Flushing is often not required as design promotes cooler running seals by providing increased circulation over faces.

60

Entrained air or gases create problems for the mechanical seal as well. While solids are spun to the periphery of the seal chamber and cause erosion, gases accumulate around the shaft and seal preventing proper seal face lubrication. If these gases are not removed, or a flush is not provided directly on the seal faces, dry running and seal failure will occur. Venting the seal chamber prior to start-up is essential for enhanced seal life. Tapered bore seal chambers inherently provide this by design while the standard bore and enlarged bore seal chambers require manual venting through a high point. During operation it is necessary to prevent gas from accumulating around the shaft. Tapered bore seal chambers with vanes or ribs accom-

plish this by breaking up the flow around the shaft and preventing pressure gradients from forming. The gases stay entrained in the pumpage and circulate through the seal chamber without accumulating. This feature makes the tapered bore seal chamber the ideal design for handling entrained gases. The standard bore and enlarged bore seal chambers can also be applied in these services. However, great care must be taken to ensure proper venting prior to start-up, and that a flush is always present. If any gases enter the restricted bore seal chambers, removal can be accomplished only upon shutdown or by insertion of a bypass to suction that originates near the shaft. If the pump experiences dry running, no seal chamber has a clear advantage. Seal life depends on maintaining a flush on the seal faces. Tapered bore seal chambers will evacuate the chamber quickly once pumpage is reintroduced while restricted bore seal chambers require manual venting.

TEMPERATURE Research shows that enlarged bore seal chambers run substantially cooler than the standard bore seal chamber when no flush is applied. From what has been observed so far, it is clear that the restricted bore seal chambers require a seal face flush for all the conditions discussed, whether or not seal chamber temperature will allow it. Tapered bore seal chambers circulate pumpage through the chamber, maintaining seal chamber temperatures close to pumpage temperatures. Testing shows this tem-

Service

Standard bore

Enlarged bore

Tapered bore with vanes/ribs

Ambient water with flush Solids 0-2% w/o flush Solids 0-10% w/o flush Solids 0-10+% w/ flush Paper stock 0-5% w/o flush Paper stock 0-5+% w/ flush Self venting and draining High boiling point liquids w/o flush Seal face heat removal Molten polymerizing liquid w/ flush

B C C C C C C C C C

A C C A C A C C A A3

A A1 A1,2 A A1 C A A1 A B

Table 1. Seal chamber selection guide Notes: 1. Dependent upon seal chamber to vapor pressure margin 2. Only on tapered bore seal chambers with vanes/ribs designed for high solids concentrations 3. With jacketed cover feature The Pump Handbook Series

perature rise to be only about 1°F, while seal face temperatures were only about 5°F higher based upon a single inside mechanical seal. Flushing may be required on tapered bore seal chambers when pumpage temperature approaches vapor pressure and the heat generated by the seal may exceed this value. In situations where the seal and seal chamber temperatures must be controlled closely, the enlarged bore seal chamber should be selected. The enlarged bore chamber has a greater fluid volume to dampen temperature fluctuations. The seal faces operate in a non-confined volume and tend not to experience drastic temperature excursions due to poor circulation as they would in a standard bore seal chamber.

APPLICATIONS Table 1 is a general application guide for seal chambers. Generally

the enlarged bore seal chamber should be applied in services where close monitoring and control of the seal environment is desired and an API or CPI flush plan is used. The tapered bore seal chamber with vanes or ribs should be used in services in which entrained gases and minimum solids can be permitted. Specialized vane or rib configurations should be used on greater solids concentrations. The enlarged bore seal chamber should be selected over the standard bore design for optimum life. The increased radial clearances not only allow for better environment control, they also free seal designs from the tight envelope constraints.

sealing environment. Use of the standard bore or stuffing box is no longer necessary, and when it comes to providing optimal sealing conditions, is not the preferred choice. Improved mean time between failure can be achieved by carefully considering the unique modifying effects each seal chamber design has on the sealing environment. ■ Nick Ganzon is a product engineer for the Engineered Products Division of Goulds Pumps. He was recently awarded a patent in seal chamber design and is a graduate of Worcester Polytechnic Institute, MA.

CONCLUSION Seal chamber designs have improved considerably. It is now possible to select a chamber for a specific application to provide the optimum

The Pump Handbook Series

61

MECHANICAL SEALS & SEALING SYSTEMS HANDBOOK

Seal Protection: Guarding Against Air and Abrasives By Joe Dunford anufacturers are steadily eliminating seal failures that are specific to design and installation. In spite of this, most seal failures occur before the faces are worn out; a need to improve seal life through enhanced operating conditions has thus been identified. EnviroSeal has been investigating the nature of the seal environment, examining factors such as seal cavity flow patterns, process conditions and fluid characteristics. Research and product development have been targeted to extend seal life while simultaneously reducing or eliminating the need for external flush and control systems. This article reviews some of the results obtained from laboratory testing and testing in the field, where we have worked with customers to solve problems in handling specific chemical groups. Due to the extent of the data, only basic concepts of the work on regular bore, horizontal seal cavities are presented here. However, these concepts can be applied to problems with taper bore and other seal cavity designs. The initial focus of our investigation was the elimination of abrasives from traditional seal cavities, as most pumps in service today utilize the basic packing style arrangement, with the throat or “packing” stopper at the bottom. The objective was to develop a new set of low cost products that would allow mechanical seals to operate more effectively in this large body of existing pumps and equipment. The symptoms of abrasives in

M

62

the seal cavity are well known, and it is universally accepted that the presence of abrasives results in shortened seal life. It became apparent during testing that air in the seal cavity, upon flooding or operation of the pump, can also be a large problem, complicating the protection of the seal from abrasives. Many pump users know that air is undesirable in the seal cavity, but they may fail to appreciate how severely air affects seal life. A basic understanding of how air and abrasives interact is essential to proper selection of seals and environmental controls, and this behavior is determined by the fluid flow in the seal cavity.

rotational and axial components combine to determine the behavior of abrasives and air in the seal cavity. The speed of the rotational component is very near to that of the pump and results in a cyclonic environment in the cavity. Cyclones are used to separate high and low specific gravity contaminants, and this separation is exactly what occurs in the seal cavity (Figure 2). Abrasives are centrifuged to the cavity bore, and air is forced to the shaft because it is lighter than the fluid. Once separation has occurred, the behavior of the abrasives and air is dependent upon axial flow.

SEAL CAVITY FLOW PATTERNS Fluid flow in any seal cavity is friction driven by the motion of the shaft, seal or other rotating components (Figure 1). Although mostly rotational, the flows also have very important axial components. The

Axial

Abrasives (out)

Air (in)

Figure 2. Effects of rotational flow (Centrifugal)

Rotational (Angular)

Figure 1. Seal cavity flow components The Pump Handbook Series

In any seal cavity design the axial flows are driven by the frictional effects of any rotating, radial surfaces exposed to the fluid. Most axial flows generated within the cavity form closed loops, as shown in Figure 3. To determine the path of the loop, find the major rotating radial surface exposed to the fluid and, on a cross sectional sketch, draw an arrow radially outward toward the bore. This arrow represents the driven flow. Next, draw an arrow

Discharge recirculation

or

Clean flush

Radial rotating surface

Figure 3. Axial flow patterns generated in a seal cavity using a standard throat restriction

along the inner axial surface, such as the shaft, that would indicate the source (suction) of the fluid. The general flow loop can now be determined by joining the two arrows with a third arrow, which curves around the inner surfaces of the cavity. If the driven flow strikes the cavity at right angles, it will split to create a lower velocity loop in the opposite direction. This secondary loop is driven along the bore towards the seal face and returns along the surface of the seal. These are typical flows found in the traditional deadended seal cavity.

ABRASIVES IN THE SEAL CAVITY Abrasives enter the seal cavity during flooding of the pump, during stop and start sequences, through fluid exchange during pump operation, and through dirty flush injection. Once in the seal cavity, the heavier specific gravity abrasives become trapped by the throat due to centrifugal forces within the seal cavity. Centrifuge throws the abrasives to the bore, where the axial flows at the bore push them towards the ends of the seal cavity (Figure 4). The fluid is capable of turning radially inward and returning along the inner surfaces to complete their flow loops, but centrifuge prevents the abrasives from following the fluid inward. Thus, abrasives rotate in torroidial patterns at the ends of the cavity. Because they become trapped at these locations, it doesn’t take long for seal cavity erosion or seal damage to occur. It is not necessarily the quantity of abrasives that cause erosion problems, but the hardness of the contaminants. If 10 pieces of very hard abrasives are trapped in a cavity and are spinning at 3000 rpm, then there

Figure 4. Accumulation of abrasives in a standard seal cavity

Figure 5. Abrasives accumulate even if a flush or discharge recirculation are used

are 30,000 revolutions of abrasives every minute. This small amount of contaminant can quickly do a great deal of damage. Experience has shown that if small amounts of abrasives can be removed rapidly, then erosion and seal life are dramatically improved. As shown in Figure 5, even if a flush is injected at the gland, the flow travels in a rotational flow pattern towards the exit at the gap between the shaft and the bushing. Abrasives in the flush centrifuge to the bore and tend to be pushed toward the bottom of the cavity. If a dirty recirculation from discharge is used, the abrasives can accumulate at the bore to such an extent that they actually close in on the seal and prevent it from responding to axial shaft movements. If the throat of the seal cavity is removed, the cavity is fully exposed to the back of the rotating impeller. The impeller then becomes the dominant force driving the axial flow component along the bore towards the seal face, (Figure 6.) Now when abrasives are centrifuged to the bore, the axial flow drives them to the seal face end, where they are trapped because centrifuge will not allow them to return with the fluid along the inner surfaces. This usually leads to erosion, or to the accumulation or packing of fine particulate or sludge in the seal components. Knowing that abrasives accumulate at the bore, and that they will follow the direction of any axial flow component, EnviroSeal developed SealMate and SpiralTrac. Both products remove abrasives through flow control at the bore. The flow patterns created by SealMate are shown in Figure 7; those generated by SpiralTrac appear in Figure 8. The flow

patterns of both devices draw fluid from the bore of the cavity, away from the seal. SealMate establishes a high exchange of fluid into and out of the cavity, which serves to replace flush and recirculation lines. SpiralTrac uses a recirculation of the fluid within the seal cavity to separate and expel the abrasives. One customer, a manufacturer of sulfuric acid in the Houston area, was experiencing seal failure in a scrubber pump every 8 hours, with an actual replacement every 2-3 days. He reported having to replace the seal cavity every 6 months due to erosion. After the installation of SealMate, the pump has now been running for more than 3 years with no problems reported. The customer was experiencing severe erosion—not because he was dealing with a slurry, but because small amounts of very abrasive contaminants were becoming trapped and continuously recirculating. The seal protector in this case simply established flow patterns to remove the abrasives from the bore rapidly. This worked so effectively that no further evidence of erosion has been found.

The Pump Handbook Series

AIR IN THE SEAL CAVITY Abrasives are certainly a major problem in the seal environment, but in certain products and process conditions, air or gas can introduce other complications. Air becomes trapped in the cavity in three principal ways: during the initial flooding of the pump, through air (gas) carried in the process fluid, and when air is introduced with the flush fluid. In most continuous process loops, air is well purged from the rest of the system as a standard routine, but specific venting of the seal cavity is

63

rarely carried out. In these systems some of the most severe problems with seal failures are associated with the lack of proper venting of the cavity on initial flooding of the pump.

AIR TRAPPED AT FLOODING OF THE PUMP When the pump is flooded, the process fluid rises rapidly in the volute area but very slowly in the seal cavity because it has to pass through the gap between the throat and the shaft. As a consequence, surface tension of the external fluid bridges across the small gap at the top, (under the bushing), preventing full venting of the cavity, even to shaft level. Repeated testing with water always results in an air bubble in the seal cavity, with at least 1/8” of the shaft exposed. This means that at startup in a dead-ended seal cavity, approximately 1/3 of the cavity volume is air; since air centrifuges inwardly, it can surround the seal. During operation, air accumulates at the lowest radial points first, usually between the back of the impeller and the back of the seal, then spilling over to the next “highest” radial points. The air trapped in the cavity at flooding is enough to reach the critical point, where it is spilling over and enclosing the seal components, resulting in wildly fluctuating and greatly increased seal face temperatures. There is no capacity in the cavity to withstand additional air from elsewhere without severe damage occurring to the seal. If the cavity is properly vented at the initial flooding, however,some additional air from system piping and other sources can be tolerated from the seal face temperature point of view. The existence of this air has important implications in the selection and application of mechanical seals for many pumps in critical services. In the typical seal arrangements used to handle acids, caustics and other chemicals, users not wanting to introduce flush for process or cost considerations often run the seals in a dead-ended cavity. With acids, if the seal faces run at high temperatures due to the presence of air (200 - 250°F), then the aggressiveness of the product in that area can be changed to such an extent that the material selections for the application become inappropriate and seal failure results. If abrasives are pre-

64

sent and hard seal faces are required, the presence of air is increasingly more critical, since air will force temperatures to increase rapidly to even higher levels. Proper venting of the seal cavity is critical to seal survival in most acid applications. A case history with a chemical company in England is a good example. A nitro benzene recovery unit was experiencing a short seal life, of about 6 weeks, in nitro benzene, sulfuric acid, caustic and sulfurous solutions, mostly at elevated temperatures. Because of the dangers of possible leaks from small fittings, no discharge recirculations were used, and due to the nature of the chemicals, clean flush was not an option. The products had solids in suspension, and hard faces were necessary to obtain even the aforementioned service life. This was an excellent application for retrofit of SealMate because the cavity was automatically vented, the continuous internal recirculation kept the seals still cooler, and with the abrasives removed, a better seal face combination was used (carbon on silicon carbide). A program to retrofit 12 pumps was initiated in May 1994. No seal failures have been experienced to date. In caustics, brines and other crystallizing products, the level of air in the seal cavity is critical even with the pump standing still. If the level of air is such that the interface of caustic and air goes axially through each side of the seal, then ideal conditions for product crystallization on the seal components exist. If the pump is permitted to stand still, even for a short period, crystals can form on the faces, between sliding components and in springs. When the pump is started, these crystals can cause the seal momentarily to fail to respond to an axial movement, allowing product to pass through to the atmospheric side of the faces. Further crystallization will occur, and the repeated cycle will lead to severely shortened seal life. Many users are forced to use hard faces and to inject clean flush to obtain satisfactory seal life. One customer in a chloralkali plant with an application of 50% caustic in 240°F supersaturated brine was injecting 6 gpm of hot condensate to achieve 18 months service life. The cost of the condensate and the subsequent evaporation was more than $50,000 per year. If the The Pump Handbook Series

Figure 6. Flow patterns and abrasive accumulation with no throat restriction

flush were cut off, the seal life would be reduced to between 2 days and two weeks, “depending upon luck." Two pumps were equipped with SealMates, and are still running 2 years later, with no flush. SealMate was designed to vent the cavity while the pump is standing still, preventing the crystals from forming, and to establish a flow pattern to remove crystals during pump operation.

AIR FROM THE PROCESS A much more complex condition exists where the process fluid carries a lot of air or gas, or where tanks are emptied and the pumps run dry. In these instances, the pump is frequently the only centrifugal component in the system, and some of the air (gas) will accumulate at the center of the centrifuge, which in this case will be the seal cavity. The air in the pumpage is broken into very small bubbles, and water, for example, will turn gray instead of its normal clear color. As the small bubbles pass through the pump, those in the fluid near the impeller and behind it become caught in that localized centrifugal environment and are forced inward, behind the impeller, to form into a single bubble tight to the shaft. This accumulation of air displaces the fluid, starting from the shaft and expanding radially outward. The air can then envelope the seal, causing the faces to overheat. This problem will occur even more quickly if the cavity is not properly vented at startup. Depending on the amount of free air present in the system, different approaches can be taken. With clean pumpage and small amounts of air, a taper bore cavity could be used. This would take advantage of the fluid flow driven off the back of the impeller, along the bore and towards the gland

rotating flow, to draw abrasives into a spiral cut into the front face, and to convey them to a single exit groove between the bushing and the shaft. Cartridge double seals can now be used with soft faces and a low pressure barrier, even if air is a problem, because the abrasives are looked after by SpiralTrac. Tangential injection should be used for the flush or discharge recirculation, to avoid driving contaminants directly into the seal components. Figure 7. Flow patterns and particulate removal using SealMate™ Seal Protector If required, use a flush for seal face cooling only.

Figure 8.Flow patterns and abrasives removal using SpiralTrac™ Throat Bushing

and seal, to push air back from the faces. Also, with small amounts of air, there may be sufficient space between the impeller and the back of the seal to act as a reservoir, if the cavity (of any design) was properly vented at the start. With greater amounts of air, double seals could be used in any cavity design to ensure proper face lubrication. If abrasives are present, double seals with a SealMate, or SpiralTrac equipped cavity would provide protection from both air and abrasives, and this arrangement would allow the use of less sensitive and cooler running, soft faces. When a discharge recirculation or flush can be used with a narrow bore traditional seal cavity, the flow can be arranged to push air away from the seal face and toward the throat at the bottom of the seal cavity. Some of the air will be pushed out to the area between the back plate and the impeller. If there are any abrasives in the flush, they will accumulate at the bore. However, these abrasives can now be removed almost as quickly as they enter by the simple pressing of a SpiralTrac bushing through the bore and against the cast in throat. The bushing is designed to make use of the

AIR AND FINE PARTICULATE IN THE CAVITY The handling of fine particulate in applications such as lime slurry, clays and polymer powders is usually characterized by "packing" of the particulate in the seal cavity. The fine, light particulate tends to dewater in the seal cavity, concentrate and envelope the seal, causing failure. All of these products are successfully handled without flush by SealMate. Even in taper bore cavities, SealMate is successful because the flow pattern drawing from the bore prevents accumulation and packing of the particulate. In applications with very fine polymer powders, the radial shaft clearance to the SealMate is tightened from 0.015” to about 0.007”, enhancing the tangential acceleration and separation of the incoming particulate. SealMate does not remove all of the fine particulate, but customers report that seals are clean and free of most of the accumulated contaminant and seal life is dramatically improved. In a chlor-alkali facility in the southern USA, SealMate was installed in two pumps handling 80% lime slurry in supersaturated brine at 240°F. The customer reported having to inject up to 6 gpm of hot condensate to achieve 6 month service life. If the flush was turned off, the seal would fail within 2 hours. After SealMate was installed, the first pump was removed from service after 6 months due to erosion problems in the discharge piping; the second pump was removed from service for seal repair after 3 years. This was accomplished without using flush. The savings in evaporation alone were estimated at $40,000 per pump per year. The extreme difference in service life can be attributed to the combination of venting, circulation and grit removal. Installation of SealMate, has achieved similar success throughout The Pump Handbook Series

the recausticizing area in pulp mills, including caustic, lime muds, unclarified green liquor and black liquors with single seals. Similar results have been obtained in high density polyethylene in hexane, polypropylene in toluene and in other polymers. SealMate separates and rejects incoming particulate by drawing the fluid through the small gap at the shaft. This accelerates the particulate, creating a form of tangential separation. Again, SealMate does not remove all of the fine particulate, but seal life is greatly extended while the need for the flush is completely eliminated by the rate of removal that is achieved.

SUMMARY These concepts represent only a brief glimpse of the technology involved; however, using the principles outlined, the behavior of the air and abrasives in other situations can be anticipated. For air, arrange for venting of the cavity prior to startup, then determine if the process introduces large amounts of air during operation. If it does, double seals may be required as discussed. For eliminating abrasives, consider the potential for entrapment at points where the fluid is forced to turn radially inward to complete its flow loop, then change the direction of the flow in the area. This can be accomplished with flush injections, SealMate, SpiralTrac or even some of the newer seal cavity designs. Keep in mind that it is the rotating radial surfaces that drive the axial flows, not the seal cavity. Using these principles, many seal failures can be predicted and avoided, the need for clean flush can be eliminated or greatly reduced, and tremendous operating and maintenance savings can be achieved. ■ Joe Dunford is President of EnviroSeal Engineering Products, Ltd. (Nova Scotia, Canada). He is a professional engineer with a Bachelor of Applied Science degree from the University of Waterloo, and has spent his last six years conducting research and development on methods for improving the mechanical seal environment of pumps.

65

MECHANICAL SEALS & SEALING SYSTEMS HANDBOOK

Mechanical Seal Installation Here's what to do when manufacturer's instructions aren't available. By Eddie Mechelay nstallation of mechanical seals is a seemingly simple topic that actually involves many issues, each of which must be addressed to insure success. To some readers, some of these considerations may seem a bit overwhelming; to others they may seem petty. All of these points, however, should be considered in any mechanical seal installation and many of them should be a part of a written procedure for plant mechanics. Others will become a part of the repair process without specific written or oral instructions. However, implementation of OSHA CFR 1910 - Mechanical Integrity should include a written procedure pertinent to your facility. The procedures provided in this article, while applicable to most situations, are not meant to be all-inclusive.

I

PUMP CONDITION To limit the scope of this discussion, we will assume that the bearing frame of the pump has been repaired properly and that all runouts are within 0.002" at the seal area. Before installing the seal, the condition of the pump shaft must be noted. If the shaft has a hook sleeve, the surface of the gasket seats should be free of any nicks or burrs. If the seal has a sleeve with an o-ring, the shaft should be free of pits or scratches where the ring seats.

Seal Type

Seal Set

Multi-Spring Single Spring Bellows

1/8" Half the Spring Length Working Length*

*In non-hazardous services a set of 1/8"5/32" for bellows less than 4", 3/16" for seals above 5".

in such a case. Note that the table is not exact and should be used with caution in hazardous services.

COMPONENT SEALS Here are a few tips for installing the stationary hardface into the gland: 1. For o-ring type seats lubricate the o-ring with a product-compatible grease to ease the pressure of installation. (Dow Corning 111 Valve Lubricant and Sealant works well.) This is especially true for silicon carbide seats, which are brittle. (See hint on removing silicon carbide hardfaces from glands.) 2. For seats with grafoil, chamfer the gland to prevent the square edge from damaging the grafoil. This can be done with tool steel. The process does not require a lathe. The following steps should be taken in installing component seals when the seal drawing is available (Figure 1):

BASIC INSTALLATION PROCEDURES The general rule of thumb is, "Install per the manufacturer's instructions" or "Install per the seal drawing." Obviously this is the preferred choice. Any time the manufacturer's drawing is available, the mechanic should use it to set the tension on the seal, and follow the printed installation instructions accompanying the drawing. However, this simple task can become confusing if a drawing is not available. Table 1 gives a guide for setting seals

66

1. Scribe a mark on the shaft or shaft sleeve that lines up with the face of the stuffing box (Line A). Be certain to have all gaskets on the seal

Scribe Mark B

Scribe Mark A

Figure 1 The Pump Handbook Series

chamber/backplate installed before determining this location. 2. Scribe a second mark on the shaft (Line B) for the seal-setting dimension determined from the seal drawing. 3. Place the gland on the shaft close to the bearing frame, taking care not to bump the stationary seat. 4. Place the gland gasket on the shaft. 5. Install the rotary unit using Line B as the reference for the set point and tighten the set screws evenly. 6. Complete the pump assembly and tighten gland bolts evenly. If a seal drawing is not available, use the following steps to determine Line B (Figure 2): 1. Determine the seal working length. (This number is often stamped on the rotary unit.) a. Measure the free length of the rotary unit. b. Using Table 1, determine the working length by subtracting the seal set from the free length. 2. Scribe a mark on the shaft or shaft sleeve that lines up with the face of the stuffing box (Line A on Figure 2a and 2b). Be certain to have all gaskets on the seal chamber/backplate installed before determining this location. 3. Determine the location of the stationary seat. The seat will either protrude into the seal chamber, as in Figure 2a, or be recessed in the gland (Figure 2b). 4. Referring to Figures 2a and 2b, scribe a line (Line C) on the shaft where the stationary face will operate with the gland bolted in place. This is dimension 'x' on the drawing. a. If the hardface protrudes into the seal chamber, measure from Line A into the seal chamber and scribe Line C. b. If the hardface is recessed

in the gland, measure from Line A away from the seal chamber and scribe Line C. 5. Measure into the stuffing box from Line C the working length of the seal and scribe Line B. 6. Install the rotary unit, using Line B as the reference for the set point, and tighten the set screws evenly. 7. Complete pump assembly and tighten gland bolts evenly.

CARTRIDGE SEALS The set of the seal on cartridge seals is inherent in the design and established at the factory as complete assemblies are produced. A given assembly will contain center tabs or spacer washers that must be removed after the seal assembly is bolted in position and the sleeve is locked to the shaft. The following are typical installation instructions: 1. Install the seal assembly on the shaft. 2. Assemble pump backplate and impeller. In the case of open or semi-open impellers, make impeller clearance adjustments at this time. 3. Slide seal assembly into the stuffing box and tighten gland bolts. For seals with o-ring secondary seals: 4. Tighten sleeve collar set screws to the shaft. 5. Remove centering tabs or spacers. (Retain for pump disassembly.) For seals with grafoil secondary seals: 4. Crush grafoil rings with bolts provided. 5. Tighten sleeve collar set screws to the shaft. 6. Remove centering tabs or spacers. (Retain for pump disassembly.)

OTHER CONSIDERATIONS Secondary Seals - Gaskets, orings, or grafoil are all secondary seals for shaft sleeves. The following

are questions to ask when installing these: Is the seal material the right material? Is the sealing surface in good condition? Am I installing it over threads? Installation over threads is a common occurrence on vertical pumps. One simple cautionary approach is to wrap the shaft with masking tape and then slide the seal over the shaft. This minimizes the chance of the threads (especially square groove threads) cutting the oring during installation. Note that masking tape is about 0.005" thick. If the outside diameter of threads is close to the inside diameter of the seal or sleeve, this technique will not work. Pump Types - Pumps with open face impellers usually have the clearance from the volute adjusted in the field. If the seal is a component design, this can be a significant problem. Consider that the mechanic measures the working length of the seal, either spring or bellows, in the shop and installs the seal gland. After he installs the bearing frame in the field, he will need to determine the distance the impeller is away from the volute. Depending on shaft variability, he may need to adjust this distance .032" or more. With a single spring seal application, this variability is more than likely acceptable. With multi-spring seals, or especially bellows seals, this deviation from the recommended working length could lead to premature failure. In light hydrocarbon and flashing services, working length can become even more critical due to the importance of correct face loading, which prevents flashing and dry running of seal faces. One solution to this problem is to use cartridge seals in these services. In this case the impeller adjustment is made prior to tightening the seal gland bolts. Another remedy is to remove the volute with the bearing frame when the pump is overhauled. In this way the distance between the impeller and the volute could then be set without the seal. The pump could be disassembled and the seal installed. Gland Piping - This item may seem obvious but is often overlooked The Pump Handbook Series

Line B

Line C

Line A

Working Length "X"

Figure 2A

Line B

Line C

Line A "X"

Working Length

Figure 2B

- CLEAN the piping! This simple task can prevent orifices from getting plugged and, with dual seals, can prevent contaminants from entering new barrier/buffer fluid in the seal pots. Unfortunately, seal manufacturers can mislabel the gland ports. This is especially true of seals that were built prior to API 682, which standardized porting terminology. Care should be taken to insure that the flush port and the in and out ports on dual seals are labeled correctly. This can be done by connecting low pressure shop air to the seal or by referring to the seal drawing. Split Case Pumps With Dual Seals - These pumps use two seals to prevent leakage at the ends of the shaft. Dual seals incorporating pumping rings or cutwaters to assist in the circulation of the barrier fluid are direction- dependent. In other words, the in and out ports on the coupling end of the pump are opposite the in and out ports on the outboard end of the pump. Since the glands usually are shipped with ports stamped, they may need to be restamped both for the ports and for

67

2-4 PSI Steam

Seal Gland

Pinch Back Block Valve to Allow Condensate to Drain

Figure 3

designating the end of the pump on which each of the seals is to be installed. This can be done by the vendor if the person ordering the seal takes the time and care to specify it at the time of order. Errors in these situations are even more common if pipe fitters and not pump mechanics are used to pipe the seal to the seal pots. An item to note concerning gland piping is that the Eighth Edition of API 610 allows for stainless steel tubing in hydrocarbon services. This is a change. Previously, only pipe was allowed in hydrocarbon service. Seal Pot Piping - The general recommendation of seal vendors is to keep the piping to the seal pots to a maximum of five linear feet. Unfortunately, those of us in industry realize that this is often impossible. In situations where compromise is a must, enlarge the tubing to 3/4" to minimize pressure drop. On dual Gland Ring

Drill and Tap

Insert

Figure 4

68

seals with pumping rings, one seal supplier recommends installing a larger supply line to the seal than the return to the seal pot. Typically, this will mean a 5/8"- 3/4" supply line will be used with a 1/2" return line. Also, many pipe fitters are trained to install what I call "pretty tubing." Although 90º bends are nice to look at, they usually increase the linear footage of the pipe system and will increase the pressure drop. These 90° bends also increase the chance of there being low spots in the piping. Explaining to the pipe fitter the reason for reduced runs and gradual bends can increase buy-in on an installation. Pipe fitters may view as shoddy craftsmanship the arrangements that yield solid pump performance, but the benefits of performance over beauty are clear to most people. Seal Pots - Seal pots purchased from the seal vendors are built to insure proper installation; however, many pots are built by users themselves. The most common problem I have encountered with these is that the return line to the seal pot is above the liquid level in the pot. This scenario prevents proper circulation of the barrier fluid. In applications without pumping rings, thermal convection does not exist, eliminating cooling – a situation that can easily cause premature inner or outer seal failure. Steam Quench - Steam quench pressure recommendations vary among seal manufacturers, but the range is often 2-4 psi. One problem often encountered at these low pressures is that it is not a steam quench at all but more like a condensate quench. In cooler services this may not be a problem because as the condensate enters the gland, it most likely vaporizes to steam. In hot installations, though, the condensate rapidly flashes to steam, which is a safety concern for operations personnel inspecting the equipment. Also, it can be damaging to the seal faces and shaft. Two techniques can be used to minimize this problem. One is to pipe the steam as seen in Figure 3. This setup allows the condensate to The Pump Handbook Series

drain to the sewer while the steam enters the seal. Many seal vendors believe that no steam is better than wet steam. The other method, limited to hot services, is to run the quench line along the flush line (if an API Plan 11 is used). This keeps the steam from condensing as it enters the gland, and it slightly cools the flush. A setup like this has another advantage. In hot oil services it will keep the flush line warm while the pump is idle, thus preventing the oil from setting up in the line.

WHAT GOES WRONG? Although I have addressed some of these issues previously, the following is a list of things I have seen most frequently go wrong during seal installations. Secondary Seals - Many conscientious mechanics have had difficulty with grafoil rings. Let's face it, they are tough to handle without damaging them. A good practice is to ask the seal vendor to supply an extra ring or set of rings with each new or rebuilt seal. Although this may seem excessive, it can help a mechanic at 2 a.m. when he damages one. The sleeve gasket on hook sleeves often is the only source of VOC emissions on a pump. If the environmental department for your company has written-up a seal for repair due to VOC emissions, take the time to ask an inspector from that department to go out to the pump and sniff it with you present. Understanding where the leak is can prevent a lot of wasted time and effort. On old pumps in light hydrocarbon services, the leak will often be entirely under the sleeve. Gland Gaskets - A two dollar part on a $4,000 seal can cause a mechanic an entire day of work if care is not taken. Take the time to insure that the gaskets of the right size and material are being installed. Installation is not difficult if the right parts are available. The lone exception to this is split case pumps, especially older ones. Seldom is the bottom half of a split case pump removed from the field and brought to the shop. Because of this the surface for the gland gasket may often

not be flat. In addition, the case gasket on the pump is a part of the sealing surface for the gland gasket. On split case pumps with a paper style gasket, tabs should be left protruding at the gland sealing surface until the case is bolted tight. After the case is bolted tight and the pump is free to rotate, these tabs can be cut with a razor blade flush with the gland face. Also, in low pressure conditions a product compatible sealant can be used in conjunction with the gland gasket to help fill any voids. If problems persist, other gasket materials, such as flexible graphite, may more readily conform to surface defects. Piping - As stated, piping to the seal is often installed incorrectly. Getting the pump mechanic involved in the piping can help eliminate this problem since he typically has access to seal drawings. Also, he may have access to pump drawings, as in the case of Sundyne pumps where no gland exists and the piping must be installed to numbered ports that are identified only in the manufacturers instruction manual. Bellows Working Length -

Each seal manufacturer has its own working length dimension for bellows, based on the size of the seal. I strongly recommend a chart or wallet size card be made available to mechanics to assist them in installing the bellows with the proper working length. Beware of bellows with grafoil rings for the secondary seal. Each manufacturer references the seal from a different point, a fact that can cause great confusion. Although mechanical seal installation appears to be quite simple, a variety of factors can contribute to failure. Paying attention to details will increase the probability of a successful installation. And, if their benefits are recognized, proven procedures will become routine for the mechanics. ■

Removing Stationary Inserts

seat through the gland. See Figure 4. This allows the mechanic to push out the seat slowly, minimizing the risk of fracturing seats, especially silicon carbide. The author would like to thank Art Quirk of EG&G Sealol for his help in preparing this article. Eddie Mechelay is Mechanical Supervisor at Total Petroleum's Denver Refinery. He is a member of the Pumps and Systems User Advisory Team and a frequent contributor to the magazine.

BIBLIOGRAPHY 1. Seal & Survive, Seal School Handbook, EG&G Sealol, 1991 2. Dura Seal Manual, Ninth Edition, Durametallic Corporation, 1993

One technique to remove stationary inserts is to drill and tap for a small- diameter bolt behind the

The Pump Handbook Series

69

MECHANICAL SEALS & SEALING SYSTEMS HANDBOOK

The Importance of Seal Failure Analysis By William Burke

ost pumps enter the repair shop because of mechanical seal or packing failure, yet few companies re-gularly troubleshoot these components. Ten to fifteen minutes spent at this critical juncture can mean doubling or tripling the life of new seals placed in service later. It is relatively easy to determine the cause of seal failure in 80% of common industrial applications. A company buying mechanical seals from a reliable vendor has the right to expect a sensible analysis on the majority of their seal failures. The seal representative is seldom at the shop when the pump arrives for repair, but the people changing the seals can do a very good job of identifying the reason(s) for failure. Steps taken at this point can substantially improve the reliability of the application and save the company far more money than the cost of the time spent troubleshooting. Failure to address obvious seal problems almost certainly dooms pumps to a short service life upon rebuild. For the purpose of this article, I will not discuss packing failures but will focus on the major causes of seal failure. The following data are compiled from seal records kept at two chemical plants that I have personally inspected. The totals cover a two year plant history and are not application specific. In most cases the seals were reviewed after the pumps were repaired and returned to service. By using the form included in this report, however, the plants have been able to extend seal life and increase Mean Time Between Failure (MTBF) dramatically over the past two years. CLIENT ONE – a large plant with relatively simple processes and more than 600 pumps. This facility's large powerhouse is used for co-generation purposes as well as to supply steam needed for in-house processes.

M

70

A wide variety of seal designs and materials of construction are used, including component seals, split seals, cartridge seals (single and double/tandem). CLIENT TWO – a medium sized plant that standardized cartridge seal designs two years ago. Although it does not have a large powerhouse, the facility runs significantly more involved pump/seal applications (evaporation/condensing/reactionloops), a fact which leads to other concerns.

RESULTS Much can be gleaned from these plant experiences. In each case the largest factor contributing to failure was I.D./O.D. rubbing of the rotary seal member against some object, usually the bore of the stuffing box. Numerous documented studies clearly prove the value of large bore stuffing box/seal chambers for mechanical seals. Both facilities have instituted on-going upgrades for their pump populations. If a seal is removed from service and shows rubbing as the cause of failure, it is immediately returned to the previous pump work order. If that report also states rubbing as cause of failure, the stuffing box is replaced on the spot with a seal chamber, prior to seal replacement. The concept of "Shaft Deflection Ratio" has been discussed at length in previous issues of this magazine. A good rule of thumb is that a low value of 'Shaft Deflection Ratio' will increase seal life in cases where the pump is operating away from its Best Efficiency Point (BEP). Remember that your pump operates where its curve crosses the system curve. Systems age and change over time as a result of such developments as a gate valve being replaced with a globe valve and long horizontal runs of piping building up solids. These can become headaches for maintenance and operations people. The Pump Handbook Series

The second most common cause of seal failures in this study related to o-ring problems. The proper ring for an application must accommodate both the fluid being pumped and the fluid used to clean the system. Viton might be the proper material for normal applications. But what if your system is cleaned at every shut-down with a caustic steam solution? You may lose the seal a week later, and failure analysis will blame o-ring deterioration. Is the connection made between the caustic steam washout and the failure of Viton? You are far better off stipulating more expensive o-rings that can survive both conditions rather than losing seals after your turn-around, when the plant and system are up and have to run. Note the drop at both plants in seal failures due to poor installation. With the decrease in maintenance budgets and personnel, the increasing utilization of "multi-craft" people demands simpler seal designs for installation by off-shift personnel. Component seals can be very difficult to install and are sensitive to dimensional information often unknown and/or unavailable to "back-shift" employees. If your plant has downsized its maintenance department (and whose plant has not?) cartridge designs may justify their additional costs because of their ease of installation. Don't blame an electrician for shortened seal life when complicated seals are being installed! Note also how often bearing problems cause seal failures; the two are inextricably intertwined. Lip seals were designed for automotive water pumps around the time of the Second World War. They have a limited design life that is far short of average bearing life. Bearings seldom fail because of fatigue. More often its because of contamination and water emulsions in the bearing housing. Upgrade your lip seals every time you change bearings on your pumps, and

you will see bearing life triple and quadruple. Upgrade from lip seals to whatever your engineering staff decides is best – labyrinth, full face or magnetic seals. After all, would you buy equipment today for a job in your plant or home that was designed in the 1940s? Finally, TRAIN YOUR OPERATORS! They deserve and need training because they are the people causing the majority of your seal and bearing failures. I have yet to work with a maintenance person who did not care about the quality of his work. Maintenance people are not the ones wrecking the production equipment. Untrained operators regularly run pumps dry, start them with closed suction or discharge valves, and fail to report increased noise or vibrations from the equipment they run. They do not do this to sabotage the plant; they do it from ignorance. Production "owns" the equipment; maintenance just "borrows it" when it needs to be repaired. Stop cursing your maintenance budget and start training your operators. You'll see how quickly maintenance can become a profit center for your plant instead of the ugly step-child it is often regarded as in plant budget and planning sessions.

CONCLUSION Seal failure analysis does not have to be complicated or time consuming. Most of it can be done with reasonable accuracy at the time of failure by your in-house staff in less than ten minutes. Do not lose the data; put it on a form (see sample), and log it into a simple data retrieval system. Marshal the assets you already have in-house to get a better view of your problem pump applications, make a few simple upgrades on repeat offender pumps, and improve your bottom line today. Plants can readily increase their overall profitability with better training of operators and by providing their maintenance people with the freedom to make some simple upgrades at the time of seal/bearing failure. ■ William "Doc" Burke has been a fluid sealing specialist for A.W. Chesterton in the Chicago area for eight years.

FAILED SEAL RETURN FORM 1.

Mechanic's Name:

2.

Work Order #:

Date: 3. Equipment I.D.#: 4. Equip. Model/Size/Name Plate Information:

5. Place an "X" next to the appropriate cause of seal failure if you have examined the failed seal: ❑ ❑ ❑ ❑ ❑ ❑ ❑ ❑

I.D./O.D. rub o-ring failure bearing/pump poor installation lost environmental controls corrosion faces worn/opened pump ran dry

6. Any changes on pump? (impeller adjusted,sleeve/shaft changed,bearings changed, stuffing box bored out, etc.)

Distribution of Failure Causes

7. Environmental controls installed/re-installed? ❑ Yes ❑ No

120 100

8. Additional comments?

Company A

80

Company B

60 40 20

Corrosion

Faces Worn/Opened

Lost Environmental Controls

Pump Ran Dry

Poor Installation

Bearing Failure

O-Ring Failure

I.D./O.D. Rubbing

0

The Pump Handbook Series

71

MECHANICAL SEALS & SEALING SYSTEMS HANDBOOK

Zero-Leak Seals Cut Emissions Zero-leak seals offer alternatives to sealless pumps, but it pays to know the technologies. By Robert C. Waterbury, Senior Editor

ncreasingly restrictive emission control laws are forcing industry to consider environmental protection as a major business objective. In the United States, the Environmental Protection Agency (EPA) spearheaded this effort by passing Clean Air Act regulations requiring U.S. plants to reduce emissions of 189 hazardous air pollutants by 80% in the next several years. The American Petroleum Industry (API) has responded with a standard of its own known as API 682 that seeks to reduce maintenance costs and control volatile organic compound (VOC) emissions on centrifugal and rotary pumps in heavy service. API 682, a pump shaft sealing standard, is designed to help refinery pump operators and similar users comply with environmental emissions regulations. In addition to meeting emissions requirements, the standard also specifies that an API 682-qualified sealing system must operate uninterrupted for at least three years. This is proving attractive to operators because it offers reduced maintenance, increased uptime and lower operating costs. As a result, a number of mechanical seal designs are being offered that promise reduced emissions in compliance with API 682. These are frequently available in single seal, dual seal unpressurized and dual seal pressurized versions. Depending upon their target markets and applications, not all leak-free designs comply with API 682. However, most designs offer some form of dual sealing and pressurization in the form of a gas or liquid barrier to prevent leakage. Following is a look at some commercial products that represent different design concepts aimed at

I

72

achieving zero leakage or leak-free operation in compliance with hazardous emissions regulations.

API 682 TYPE A PUSHER SEALS PA-100,-200,-300. The new PA Series seals from Durametallic are type A Pusher seals that conform to API 682. They include the PA-100 single seal, PA-200 dual in-line seal with a buffer fluid and the pressurized PA-300 dual in-line seal operated with a fluid barrier. These are inside mounted, cartridge pusher designs featuring unitized compression units for simplified installation. The seals are designed to maximize heat transfer away from the seal faces. The rotating seal faces are premium grade blister-resistant carbon; stationary mating rings are reaction-bonded silicon carbide. A specially-designed, high performance flow inducer is featured on the PA-200 and PA-300 dual seal arrangements. The rotating seal faces of the PA Series are driven by axial pins that engage slots on the rotor OD. This configuration minimizes mechanical stresses that can lead to catastrophic failure of the rotating seal face. The unitized compression unit in the PA Series consists of Hastelloy "C" springs. Sleeves and glands are 316 stainless steel and the secondary Orings are Viton. Tungsten carbide stationary faces are available as an option for abrasive services. Operating temperatures for the PA Series range up to 180ºF in water/caustic, a maximum of 300ºF in non-flashing hydrocarbons and up to 500ºF in flashing hydrocarbons. Maximum operating speed for the PA Series is 5000 fpm. Type 48. John Crane is another supplier of single, dual unpressurized and dual pressurized cartridge The Pump Handbook Series

seals that conform to API 682. Its Type 1648, a single cartridge seal, typically controls emissions to less than 200 ppm. The Type 2648, a dual unpressurized cartridge seal, is designed to provide additional safety where emissions must be controlled. And the Type 3648, a dual pressurized cartridge seal, provides essentially leak-free containment of hazardous fluids and light hydrocarbons. Each of these seal designs is based on the proven performance of Crane's earlier Type 48 seal, and features computer-optimized seal face designs for maximum emissions control and long operating life. They offer operating temperatures ranging from -40ºF to +500ºF and speeds to 5000 fpm. In addition to its standard type A multiple spring pusher seal, the Type 48 seal family includes: Type 48 LP. A balanced pusher seal with plain face design for low pressure applications, it can be used as an outboard seal on either dual pressurized or unpressurized arrangements. Type 48 M. A balanced pusher seal with plain face that incorporates a hammerhead to increase primary spring pressure for medium to high pressure applications; it is also used as an outboard seal on dual pressurized or unpressurized designs. Type 48 HP. A balanced pusher seal with hydropadded face design that incorporates a hammerhead to increase primary ring pressure for high pressure; it also may be used as an outboard seal on dual pressurized arrangements. Type 48 RP. A double balanced pusher seal for ID pressure applications, commonly referred to as reverse pressure design, it is used as an inboard seal on dual pressurized

Photo 1. The Durametallic PB-100 Type B bellows seal conforms to API 682 and is available in dual non-pressurized and pressurized versions.

arrangements. Type 48 S. A shorter balanced seal designed for OD pressures only, it is sometimes used as an outboard seal on dual pressurized arrangements.

API 682 TYPE B BELLOWS SEAL The API 682 Type B Series is a companion seal for rotating bellows applications from Durametallic. As mandated in API 682, it also complies with the emissions regulations and provides an expected minimum uninterrupted service life of three years. Likewise, it is offered in a single seal PB-100 design, a non-pressurized PB-200 dual in-line seal with a buffer fluid and the pressurized PB-300 dual in-line seal operated with a barrier fluid. These are all inside cartridge, balanced rotating bellows designs capable of handling pressure reversals. The mating rings of the PB Series are interchangeable with Durametallic's Type A (PA Series) and Type C (PC Series) seals. Moreover, glands and sleeves of the PB Series are interchangeable with PA Series seals. A specially designed, high performance flow inducer on the PB-200 and PB-300 dual seal arrangements improves the circulation of seal flush across the rotating

Photo 2. John Crane's Type 1648 pusher seal available in single and dual configurations is based on tested Type 48 seal technology and conforms to API 682.

carbon faces. The PB-300 pressurized dual seal also consists of an inside balanced seal with two flexible rotors and two mating rings in series. Its inner seal is capable of operating in reverse pressure conditions. The PB Series also incorporates a circumferential flush around the OD of the stationary, silicon carbide seal face. Operating temperatures and speeds are similar to those for the PA Series with the exception of flashing hydrocarbons. High temperature PC Series stationary bellows seals have been developed for high temperature services.

NON-API SEALS The seals discussed above are API type intended mainly for use in the petroleum, refining and heavy industries. But there are other leakfree designs as well for applications in food, pharmaceuticals, petrochemicals, pulp and paper, etc. Type 2800E. The Type 2800E from John Crane is one example. It is a gas-lubricated, non-contacting double cartridge seal applied in food processing, pharmaceuticals and chemicals as well as some petroleum-related industries. It offers advanced, dry-running seal technology for process applications, zero emissions, high performance, extended seal life and reduced pow-

The Pump Handbook Series

Photo 3. Delta 4550 from Inpro features dual-port, direct injection of motive gas into the seal interfaces.

er consumption in a design that eliminates wet sealing systems. The 2800E uses inert gas to positively seal hazardous fluids for zero atmospheric emissions. The 2800E features a dual balance inboard seal design that withstands pressure reversal and prevents process leakage under upset conditions where barrier pressure is lost. The dry-running double seal design allows 100% containment of process fluid in compliance with environmental control regulations. The design key is its patented spiral groove technology. The spiral groove patterns on both sides of the mating ring pump gas toward the ungrooved portion of the sealing face. This compressed gas cushion serves as a sealing dam to block escape of volatile fluid being sealed. When stationary, the sealing dam (ungrooved portion of the face) closes and prevents gas or product leakage. Most non-contacting seals consume gas constantly, even when static. The non-contacting design reduces seal face heat generation and torque/power consumption by as much as 98% compared to a wet seal. Face wear is virtually zero, increasing mean time between planned maintenance up to 1000%. 4550 Delta Gas Purge Seal. The Inpro 4550 Delta P seal is a back-to-back gas-lubricated barrier seal of cartridge design. This design, however, uses magnets instead of

73

coil springs. The repelling force of like-pole, rare earth magnet technology is proven in numerous applications. In this case the repelling magnets are contained in stainless steel, back-to-back, rotating seal faces. The stationary sealing faces are made of Rulon®. Low pressure gas such as air or nitrogen enters through dual-port, direct injection between the stationary sealing faces. This causes the rotary faces to "lift off" the stationary faces. The same gas then migrates to the chamber enclosing the rotary faces. The residual pressure between the rotors positions the rotors to within microns of the stationary faces. A microscopic film of air, gas or liquid separates the sealing faces and thus there is no wear, frictional drag, heat generation or practical upper and lower limit to rotating speed. The 4550 Delta P gas purge seal faces are pressurized and slightly separated even when the seal is in a static condition. The sealing technology can be used for dry running or submerged mixer shafts; and is designed to capture and control hazardous/toxic emissions from single primary mechanical pump seals. Magnetic Fluid Seal. Another alternative zero-leak sealing technique is offered by Ferrofluidics. It forms a seal based on the response of a magnetic fluid to an applied magnetic field. The sealing components of ferrofluidic technology include a ferrofluid, permanent magnet, two pole pieces and a magnetically permeable shaft. The magnetic circuit created by the stationary pole pieces and the rotating shaft concentrates magnetic flux in the radial gap under each pole. When ferrofluid is applied to the radial gap, it assumes the shape of a liquid O-ring and produces a hermetic seal. Ferrofluidic vacuum rotary feedthroughs use multiple rings of ferrofluid in stages formed by grooves machined into either the shaft or pole pieces. A single stage typically sustains a pressure differential of 0.2 atmospheres. Thus, the pressure capacity of the entire feedthrough is roughly equal to the

74

Figure 1. Magnetic fluid sealing forms a ferrofluid or liquid O-ring around the pump shaft with a permanent magnetic field.

sum of the pressure capacities of the individual stages. Ferrofluidic magnetic seals can virtually eliminate leakage of gas and vacuum from around rotating shafts. A wide variety of standard options includes solid and hollow shafts, rotary gas unions, cantilevered and co-axial seals and water-cooled seals and bearings to accommodate many different rotational speeds, pressures and temperatures. When combined with a primary mechanical seal, a magnetic seal provides a barrier to fugitive emissions leaking past the mechanical seal. It thus becomes a simple, low-cost alternative to multiple seal configurations. Ferrofluidic seals are used in semiconductor, fiber optics, robotics, aerospace and chemical/hydrocarbon processing industries. GPA Mechanical Seal. The GPA mechanical seal from Helicoflex is a more conventional mechanical seal available in single and dual configurations. It was designed originally to withstand abrasive suspensions of bauxite and hydrated alumina encountered in aluminum production processes. It consists of a stationary ring of carbon tungsten inserted into a housing, a rotating seal sleeve assembly that rotates with the pump shaft and a spring diaphragm or membrane that connects the rotating ring to the sleeve. The spring diaphragm serves to: • ensure contact between the The Pump Handbook Series

rotating and stationary rings • compensate for wear that may occur as the two ring surfaces contact one another • act as a seal between the rotating housing and the shaft sleeve • position the rotating housing in relation to the pump shaft. The GPAD or double seal configuration consists of two seals fitted into a single shaft sleeve with only one component serving as housing for the two stationary rings. Two diametrically opposed ducts are drilled into the stationary ring housing so that buffer liquid can be injected. The internal seal is adjusted in the same way as the single seal, by positioning and clamping the sleeve on the shaft, and then adjusting the tension of the external spring diaphragm. The GPAD corresponds to a "cartridge" type seal design and is recommended for use on centrifugal pumps handling hot liquids that may cause pitting. It is also used in applications that involve pumping at high suction pressures. 255 and 225 Dual Seal. The 255 and 225 Dual Seal from Chesterton are cartridge designs that feature

Figure 2. The A. W. Chesterton 225 Dual Seal features stationary seal rings in back-to-back configuration for use with maximum pressures inboard and outboard to 600 psi.

a self-centering locking ring that locates the seal barrel precisely and squarely to the shaft. Unrestricted monolithic faces maintain full face contact under fluctuating operating conditions. Both seals use a specially designed fluid barrier system in which positive pressure barrier flow is generated through pumping ring features. Both have hydraulically double balanced inboard seal rings that permit operation in double mode (barrier fluid higher than stuffing box) or tandem mode (barrier fluid lower than stuffing box). Both seals must operate in double barrier fluid mode to achieve leak-free operation. The 255 Dual Seal features stationary seal rings in a back-to-back configuration; and is recommended for use with maximum inboard pressures of 600 psi and maximum outboard pressures of 250 psi. The 225 Dual Seal features in-series seal rings for increased pressure handling and maximum reliability in sealing VOCs and VHAPs. Both inboard and outboard faces can handle pressures up to 600 psi. In double barrier fluid mode, the barrier fluid pressure can reach 250 psi above process pressure. Type 1010/1012. The Type 1010 and 1012 from EG&G Sealol also offer leak-free performance. The Type 1010 is a non-contacting bellows design with static O-rings that reportedly eliminates dynamic secondary seals. It operates on an

and inwardly in the radial direction to conform to thermal expansions, pressure effects and tolerance stackup. This minimizes the radial squeeze on the O-ring and thus reduces and controls the O-ring friction forces. The 1012 is offered in single, tandem, tandem with buffered interstage labyrinth and double non-contacting gas face seal configurations.

LOOKING AHEAD

Figure 3. The Type 1010 noncontacting gas bellows seal from EG&G Sealol operates on an extremely thin (0.000080) film of gas.

extremely thin film (0.000080) of gas that offers high film stiffness for stability and long seal life. Used in conjunction with a buffer gas, the 1010 extends MTBPM and is available in both single and double cartridge configurations. Operating temperatures range from -40ºF to +500ºF; vacuum pressure to 21 bar or 300 psig; and speed to 10,000 sfpm. The Type 1012 is another noncontacting gas seal that accepts either filtered process or buffer gas at a small pressure differential above the suction or balance pressure and generates a stiff cushion of gas approximately 2.5 microns (0.0001") thick at normal operating conditions. It is offered in either unidirectional or bidirectional seal options. A compliant spring plate flexes outwardly

The Pump Handbook Series

Leak-free seals that conform to API 682 or are used in similar industry applications though they do not conform to a particular standard are available to meet most current industry needs. However, refineries and other industries will continue to need seals for extreme services that go beyond what is available commercially today. Thus, while products similar to the ones described here are a good starting point in selecting seals to meet strict emissions requirements, they should be considered a point of departure in developing new, more specialized sealing technologies. The seal manufacturers and various industry segments are building on these technologies to meet everincreasing restrictions on emissions. Where the standards leave off, the seal manufacturers and industry task forces continue to develop new technologies that will offer better performance , reliability and service life.

75

MECHANICAL SEALS & SEALING SYSTEMS HANDBOOK

High Temperature Sealing Systems Applications dictate specific sealing options. By Pat Flach t the turn of the century, automobiles quickly replaced the horse and buggy. This change created great demand for gasoline and lubricating oils, both products of refined crude oil. The process of refining crude oil requires heating it to temperatures of 700-800°F. Thus beginning with the earliest refineries, sealing high temperature oil has been a continuing challenge. Soft packing was used in the early days. However, it was difficult to find materials that would withstand the high temperatures. Cool oils were injected through a lantern ring to create a cooler environment for the packing and prolong its life. Some of the injection oil went into the pump and mixed with the hot crude oil. Some leaked along the shaft to the atmosphere. Although cooling was accomplished, the leakage created environmental and safety problems (Figure 1).

A

Mechanical Seal Development Safety and environmental issues prompted development of the end face mechanical seal. Mechanical seals consist of a rotating and a stationary

Figure 1. Packed pump

76

Closing Mechanism Springs or Bellows

Primary Sealing Point Rotating Face

Stationary Face

Secondary Sealing Point

Figure 2. Basic seal

face, a closing mechanism and secondary packings. The two faces are lapped flat and run against each other, forming a seal between a rotating shaft and the surrounding stationary housing. A closing mechanism and hydraulic pressure keep the faces together. Common closing mechanisms include a single coil spring, multiple springs and metal bellows. One secondary seal is used between the stationary face and the housing. Another is used between the rotating face and the shaft. Stationary seals for low temperatures are typi-cally elastomeric O-rings. The shaft secondary seal arrangement defines the mechanical seal as a pusher seal or non-pusher seal. The difference between a pusher seal and a non-pusher seal is in the closing mechanism. Pusher seals use springs. Non-pusher seals use a bellows made of rubber or thin foil. Pusher seals require the secondary seal to move back and forth along the shaft to compensate for shaft movement or seal face wear. This dynamic seal can damage the shaft, and this is a common cause of seal failure. Secondary seals in non-pusher The Pump Handbook Series

designs are static, and the shaft movement and face wear compensation are accommodated by a bellows (Figure 2). The first end-face mechanical seals used in hot refinery applications were component pusher designs. To survive, they needed a cool environment around the secondary seal, much the same as packing. Typically, this was accomplished by supplying cooling water to a jacket around the stuffing box or cooling the product and then injecting it into the seal area. Cool oil injection into the seal area from an outside source was also used. The disadvantages of this sealing system are reliability and cost. Fouled coolers, cooling water loss and unreliable sources of cool flush oil are all causes of premature failure. Additionally, system operating costs increase with the use of refined product as a cool flush.

Welded Metal Bellows Welded metal bellows seals developed in the mid 1950s for the aerospace industry were introduced to industrial markets in the mid 1960s. Refineries quickly adopted these seals

for high temperature applications because they did not require the process fluid to be cooled. The static secondary packing allowed seal designers to use various metallic and non-metallic gaskets in place of the elastomeric o-rings used in pusher seals. Typically brass, copper, stainless steel or fiber materials such as asbestos were used as the static secondary packing. The stationary faces were clamped between the stuffing box and the gland plate and employed either a fiber material in a flat gasket configuration or a spiral wound stainless steel configuration filled with high temperature fiber (Figure 3). Eventually, distortion of the clamped stationary faces and problems with the static seal along the shaft were identified. This led to the use of flexible graphite as a secondary packing. This packing material allows seal designers to mount the stationary face flexibly and eliminate face distortion. The material is also more forgiving as a secondary shaft seal and provides increased seal life. The flexible nature of graphite, however, demands great care in installation (Figure 4).

Cartridge Seals One design feature that greatly enhanced the performance of mechanical seals was the development of better face materials such as silicon carbide and blister resistant carbons. In addition, switching from component seal designs to cartridge seal designs increased reliability by eliminating assembly problems. Cartridge seals are assembled in a controlled

environment, and special tools and care are used when incorporating flexible graphite secondary packings into the design. Attention to seal set dimensions and the ability to pressure test cartridge seal assemblies before installing them in pumps also increase reliability. Another significant development in designing seals for hot applications was the introduction of a stationary flexible element seal as opposed to the traditional rotating flexible element seal (Figures 5 and 6). A greater degree of out-of-squareness between the shaft and the face of the stuffing box is tolerated by making the flexible element stationary. A stationary face must deflect only once to compensate for the out-of-squareness whereas a rotating flexible element must deflect once per revolution. Thus, higher face surface speeds are achieved because centrifugal forces are not working on the flexible element. All of these design features lead to better face contact and minimal leakage across the seal.

Sealing Arrangements/Applications Different seal arrangements can be used depending on the type of hot product being sealed and applicable federal, state and local regulations. Four applications come to mind when considering hot fluid sealing: crude oil refining, heat transfer fluids, hot water, and polymerizing fluids. Each of these applications has unique requirements that must be considered.

The atmosphere around the mechanical seal on the product side as well as the atmospheric side must be addressed. Various flush plans are used to maintain a stable environment for seal operation on the product side and quench plans on the atmospheric side to minimize build. Today, single metal bellows seals are normally used in hot crude oil applications with a Plan 11 product flush and a Plan 62 atmospheric steam quench (Figure 7a & b). Plan 11 eliminates heat buildup around the seal faces and removes any vapors that may have formed. Plan 62, steam quench, on the other hand, cools any minute leakage across the faces and prevents coking. Coking is the buildup of a hard black abrasive residue. If allowed to build up, coke can prevent the seal faces from flexing and thereby cause premature seal failure. Carbon versus silicon carbide faces are generally used. However, if dirt and grit are present, two hard faces such as tungsten carbide vs. silicon carbide are required. Use of a rotating flexible element versus a stationary flexible element normally is based on personal preferences and to some degree the condition of the equipment being sealed. Depending on the amount of H2S present in crude oil, hydrogen sulfide embrittlement or stress corrosion cracking should be addressed. This is a matter of selecting the correct bellows material, such as Inconel 718. Sealing heat transfer fluids is similar to sealing hot oil except H2S usually is not a concern. The main difference is on the product side. A

Figure 5. Rotating flexible element bellows seals Figure 7a. API Piping Plan 11

Figure 3. Clamped stationary face

Figure 4. Flexibly mounted stationary face

Figure 6. Stationary flexible element bellows seal The Pump Handbook Series

DRAIN

Figure 7b. API Piping Plan 62

77

Plan 02 cooled seal chamber or stuffing box is used rather than a Plan 11. Plan 62 steam quench is used in the same way as hot crude oil applications to prevent coking. Because heat transfer fluids are used over and over, cleanliness of the fluid can be a concern. As with crude oil, both faces should be hard if the heat transfer oil tends to become dirty. In the power generating industries boiler feed and high temperature boiler circulation applications pose a greater challenge. There are two main reasons why these applications are more difficult: higher pressures and lack of lubricity. To seal these applications successfully, the water around the mechanical seal must be reduced

used successfully because only a small amount of water in the seal chamber must be cooled. By starting out with a system temperature of 140°F, sealing can be accomplished successfully up to approximately 180ºF before the coolers must be cleaned. Face materials are generally silicon carbide versus carbon graphite. High temperature flexible graphite secondary seals are the most reliable.

Types of Polymers P1

atmospheric side of the seal because they set up when exposed to air. These applications normally require a pressurized dual sealing system with a Plan 53 or 54 seal flush (Figure 9a & b). One of the major considerations in picking a Plan 53 or 54 is the pump shaft speed. Because these fluids are normally pumped at slow speeds, Plan 54, which uses a recirculating pump, may be required to assure barrier fluid flow. Picking the correct barrier fluid is always a challenge because it must be compatible with the product and handle high temperatures without flashing, and it must be durable. These applications should be reviewed closely by your seal manufacturer because of their uniqueness compared to other applications discussed above.

Looking Back/Ahead T1

Figure 9a. API Piping Plan 53 Figure 8a. API Piping Plan 21

T1

Figure 8b. API Piping Plan 23

to approximately 140°F. This is accomplished in one of two ways: utilizing a Plan 21 or Plan 23 (Figure 8a & b). Plan 21 takes fluid from the discharge through a cooler and injects it into the seal chamber. It is not frequently recommended because of the load on the heat exchanger. On the other hand Plan 23, in which the product is recirculated with a pumping ring in the seal chamber through a cooler and reinjected in the gland, is

78

Figure 9b. API Piping Plan 54

Polymers are available in two basic versions: thermosetting, which harden into a permanent shape, and thermoplastic, which melt to a flowable state with an increase in temperature. Polymerizing fluids can be very tricky to seal because their narrow thermal window is usually 450-750°F, and they have widely varying viscosities. They pose problems on the

The Pump Handbook Series

The sealing industry has come a long way in sealing hot applications. It moved from using soft packing to pusher seals to high temperature metal bellow seals. Where will the future take high temperature sealing? With the combination of non-contacting technology and metal bellows seals, we can look forward to future developments that will provide industry with more reliable seal products and increased mean time between planned maintenance (MTBPM).■ Pat Flach is Technical Services Manager, Western Hemisphere, for EG&G Sealol Industrial Division, and a member of the Pumps and Systems Editorial Advisory Board. He has more than 28 years of experience in pump design and applications, and is a member of ASME, STLE and served on the API 610 manufacturers task force. He holds a degree in mechanical engineering technology from Kellogg Community College and attended the Western Michigan University School of Business.

MECHANICAL SEALS & SEALING SYSTEMS HANDBOOK

Applying Dry Gas Sealing Technology to Pumps Successful field applications demonstrate significant improvement in sealing reliability. By William V. Adams

C

What Are Dry Gas Seals?

(Photo Courtesy of Durametallic Corporation)

Pumps have traditionally been sealed with compression packing or mechanical seals. Both of these methods require liquid lubrication. Single liquid lubricated mechanical seals are quite effective in meeting emission control restrictions. However, some services demand the use of dual mechanical seals. Liquid lubricated double mechanical seals de-pend on a pressurized barrier liquid that circulates in the cavity between the inner

Photo 1. This noncontacting seal uses Advanced Pattern Gas Seal (APGS) face geometry, shown in the cutaway portion of the photograph. There is no seal face contact and, thus, no wear.

and outer seals. The liquid barrier system requires an investment in auxiliary equipment as well as careful monitoring to maintain the seal properly. In addition, the barrier fluid must be carefully selected in order not to contaminate the process fluid. It must also be safe for release into the environment should the barrier system develop a leak. Liquid lubricated seals present additional problems if operating conditions or system upsets cause them to run dry. When the seal chamber is void of liquid, seal face friction increases, and heat dissipation is greatly reduced, resulting in rapidly rising seal face temperature. Rising temperature, in turn, creates excessive seal face wear and thermal damage both to the seal faces and to the O-ring secondary seals. Dry gas seals have become an attractive alternative to liquid barrier systems because of their superior emissions control, because product contamination is minimal, and installation, operating and maintenance costs are lower. Also, the horsepower requirements and annual energy costs are both considerably lower for dry gas seals than they are for liquid lubricated seals. These differences alone can often offset the cost of retrofitting equipment to dry gas seals. Dry gas seals are typically pressurized with inert nitrogen, although dry air or steam are sometimes used when either is non-threatening to the process fluid. The gas barrier is deadended into the cavity, requiring no cirThe Pump Handbook Series

(Photo Courtesy of Durametallic Corporation)

onventional pumps inherently leak. With increased governmental regulations that severely limit the level of fugitive emissions allowed from process equipment, the technology for controlling leakage has advanced radically in the past few years.

Photo 2. This contacting seal face gas barrier seal allows bidirectional rotation, and incorporates a rotor design that contributes to lubrication and cooling of the seal faces.

culating or cooling mechanisms. Should leakage of the gas occur, the loss is automatically replenished by the gas supply which is constant. In addition, the leak of a noncontaminating gas will not jeopardize process stream purity. Nor will it contaminate the environment.

Contacting and Noncontacting Seal Faces There are two basic types of double gas barrier seal designs for centrifugal pumps. One uses contacting seal faces, and the other uses noncontacting seal faces. Both designs require that deformation of the seal faces be kept to an absolute minimum so the faces remain parallel over the entire operational range. Each design performs well in a gas

79

Leakage (SCFH)

0.1

Non-contacting Design 0.400" Face Width

0.01

Contacting Design 0.200" Face Width

0.001 Contacting Design 0.400" Face Width

0.0001 0

1000

2000 RPM

Figure 1. Gas seal leakage The Pump Handbook Series

3000

4000

Total Head, feet of liquid

minimum capacity

as tg cen me per volu by

4

1

gas-free liquid

0 1 2 3

Leakage of dry gas seals is roughly proportional to the cube of the gap between the sealing faces. The gaps between noncontacting seal faces range from 0.00005 inches to 0.00015 inches (0.0013 mm to 0.0038 mm), while contacting seal face designs have gaps approximately three times the combined roughness of the asworn faces. This combined roughness usually measures between 0.000005 inches and 0.000010 inches (0.00013 mm and 0.000254 mm), which would result in a sealing gap of 0.000015 inches to 0.000030 inches (0.00038 mm to 0.00076 mm). A comparison of leakage between contacting and noncontacting seal face designs with equal seal face widths for a 2.000 inch (50.8 mm) diameter shaft operating at 50 psig (345kPa) and from 0 to 3600 rpm is shown in Figure 1. The effect of entrained gas on the performance of centrifugal pumps is illustrated in Figure 2. This factor should be considered when a double gas barrier seal is being specified. The maximum recommended amount of entrained inert gas should be less than 3.0% by volume. Even the highest nitrogen leakage rate for the noncontacting seal face design would produce only 0.075% entrained gas in a pump that is circulating 50gpm of water. This entrained gas, however, must be removed from the pumping system at some time. Various face groove patterns are used in contacting and noncontacting dual gas seals. The two most common groove patterns are used for noncontacting seal face designs. These are the “pumping groove” design and the “pocket groove” design (Figure 3). With the pocket groove pattern, pockets are etched into the seal face. As the pump shaft rotates, these pockets scoop up the gas while the lands

6

80

Leakage Considerations

5

medium where there is a gas at both the inside diameter and outside diameter of the seal faces. For centrifugal pump applications, the inner seal has the process liquid at one side of the seal faces. In many of these applications, the fluid at the sealing interface is constantly changing due to dry running, cavitation and off-design pump operation. Contacting seal face materials can operate in the dry mode as long as the seal face loads are reduced and the heat generated by the seal faces is adequately dissipated. Contacting seal face gas seals use the balancing of hydrostatic forces to reduce the seal face closing loads, but these forces do not create seal face separation. Sometimes a grooved seal face pattern is employed to help provide hydrostatic forces and cool the seal faces. Also, contacting seal faces are wide by design to reduce the overall unit loading on them. Noncontacting seal face designs use various configurations of shallow grooves to provide face separation. When pressure is applied to the seal, the forces exerted on the faces are hydrostatic and constant whether the seal is rotating or stationary. Hydrodynamic forces are generated by viscous shear of the gas film when the smooth counterface is rotating. Wide seal faces are required for noncontacting designs to provide maximum load support at slow peripheral speeds and low pressures. The heat generated by noncontacting seal faces is less than that generated by contacting seal faces. Contacting seal faces are subject to heat produced by gas film viscous shear as well as by the rubbing of seal faces. Noncontacting seal face designs, on the other hand, are subject only to heat generated by the viscous shearing of the fluid between the seal faces. The spring force applied in both noncontacting and contacting dual gas seals is significantly smaller than in liquid lubricated designs. This reduced closing force, along with precise hydrostatic balancing, allows the toleration of the seal face contact that can occur during pump startup or upset conditions. The subsequent reduction in generated heat can be translated directly to less required horsepower and energy cost savings.

Liquid Capacity, gpm

Figure 2. Effects of entrained gas ROTATION

SHALLOW GROOVES

SEALING DAM

POCKET GROOVE DESIGN ROTATION

SHALLOW GROOVES

SEALING DAM

PUMPING GROOVE DESIGN Figure 3. Noncontacting seal face groove patterns

around the pockets keep it from escaping, thereby increasing gas pressure and hydrodynamic lift. In the pumping groove pattern, grooves are slanted in one direction around the seal face. As the pump shaft rotates, the grooves pump gas inward while the dam at the groove’s inner dimension restricts gas flow. This, in turn, generates a pressure in the dam area that provides hydrodynamic lift. Both of these designs, along with most other noncontacting seal face groove configurations, require shallow groove depths – less than 0.0003 inches (0.0076 mm) – to provide proper face separation. Face stiffness is an important parameter that should be considered when comparing the performance of different noncontacting gas seal face designs. Seal face stiffness is defined as the rate at which axial opening and closing forces change with seal face

gap, and is measured in pounds-persquare-inch. To maintain optimum seal performance when the gap between the seal faces suddenly decreases, the seal face opening force must increase sharply. This will prevent contact of the seal faces and allow them to reopen to the original operating gap. A noncontacting seal face should be as stiff as possible to prevent seal face contact. When compared to the characteristics of unidirectional seal face patterns, as illustrated in Figure 3, bi-directional seal face patterns reduce stiffness by as much as 50% in most designs. This is because the bidirectional pattern, in order to accommodate seal rotation in either direction, must allocate fewer grooves to each direction of rotation than can be allocated for a unidirectional seal face pattern. Typical contacting seal face groove designs are shown in Figure 4. Contacting seal face grooves do not need to be as shallow as those for noncontacting seal faces because the grooves for contacting seals do not provide a significant dynamic lift. Contacting seal face groove depth can be as much as 0.062 inches (1.57 mm). The grooves are used mainly to direct the pressure down to the sealBI-DIRECTIONAL ROTATION DEEP FEED GROOVE BEARING PAD

SEALING DAM DEEP ANNULAR GROOVE

ing dam in order to produce hydrostatic force. Some contacting seal face patterns incorporate straight radial feed grooves. Others incorporate angled feed grooves to force the barrier gas through the seal face for a cleansing and cooling effect. The pads between grooves provide a bearing support for the seal to reduce the overall unit load. These bearing pads typically have a waviness that generates a slight hydrodynamic lift. The majority of the contact load reduction, however, is due to the hydrostatic forces.

Dual Gas Seal Barrier Systems Nitrogen and instrument air are the most common barrier gases used for dual gas seals. If compressed air (i.e., plant air) is used, oil and particulate matter must be removed. A simplified barrier system for a dual gas seal is shown in Figure 5. The minimum components required for this system are a regulator and a pressure gauge to ensure that the barrier gas pressure remains 20 psig to 50 psig (138 to 345 kPa) greater than the pressure of the process fluid. A flow meter that indicates changes in gas flow will monitor the seal’s performance. Changes in gas flow can be noted and recorded at a remote site as well as directly at the meter. A needle valve or an orifice can be used to limit the maximum amount of nitrogen leakage in the event of a seal failure. A pressure switch can be used as an alarm or shutdown mechanism for loss of gas pressure, and a flow switch can perform a similar function for excessive gas flow. A 15 micron in-line filter is

ANGLED FEED GROOVE DESIGN

8

BI-DIRECTIONAL ROTATION DEEP FEED GROOVE

7

6 5

9

4

3

BEARING PAD

2 1 GAS BARRIER SOURCE

SEALING DAM DEEP ANNULAR GROOVE

STRAIGHT FEED GROOVE DESIGN Figure 4. Contacting seal face groove patterns

1-FILTER 2-REGULATOR 3-NEEDLE VALVE 4-FLOW METER 5-CHECK VALVE

6-FLOW SWITCH 7-PRESSURE SWITCH 8-PRESSURE GAUGE 9-THREE WAY VALVE

Figure 5. Double gas seal barrier system The Pump Handbook Series

recommended for gas barrier systems that may have foreign particles present. Connecting an alternate gas supply by a three-way valve can maintain barrier cavity pressure to allow the pump and seal to continue operating while maintenance is being performed on the barrier gas system. For process fluids that require the temperature to be maintained above 200ºF (93ºC), a steam barrier gas seal system may be beneficial. A typical steam barrier system diagram is shown in Figure 6. A pressure reducing valve is required in a steam barrier system to maintain an accurate barrier cavity differential pressure above the seal chamber pressure. Steam traps are also required to collect condensate from the piping and the seal and direct it to a condensate return. A steam trap bypass line can be installed to maintain system pressure during steam trap maintenance. STEAM SOURCE

GLAND RING 6 7

4 2

3

5 TO SEAL

9

8

1 CONDENSATE RETURN 1-STEAM TRAP 2-STRAINER 3-GATE VALVE 4-PRESSURE REDUCING VALVE 5-GATE VALVE

CONDENSATE RETURN 6-THERMOMETER 7-PRESSURE GAUGE 8-STEAM TRAP 9-GLOBE VALVE

Figure 6. Double steam seal barrier system

Seal Arrangements There are several basic seal arrangements for double gas seals. The most common design is the canister arrangement that utilizes back-toback inner and outer seals. Either contacting or noncontacting seal faces can be selected with this arrangement. When this arrangement is used, the process fluid must be monitored for the size and type of any solids present. Solids tend to get trapped and accumulate at the inner diameter of the inner seal and can prevent the seal from tracking properly. Some canister arrangements utilize a conventional contacting seal face without grooves for the inner seal. This design relies on the cooling effect produced by the process fluid

81

being pumped. It is not usually recommended where continuous dry running operation is likely. The major advantages of this design are its capability of operating as a single seal during pressure reversals and the low level of gas leakage into the process fluid it allows. One way of reducing the potential for seal hang-up, caused by the presence of solids or the polymerization of the process fluid, is an in-line seal arrangement used in conjunction with seal chamber flow modifications. A seal using conventional contacting seal faces without seal face grooves on the inner seal allows this design to operate with good single-seal performance, even during pressure reversal conditions. For some process fluids there may be an advantage in using a repeller type pump. The repeller evacuates the process fluid from the seal chamber during dynamic operation. This arrangement may be beneficial for a canister seal used with process fluids that contain solids or tend to polymerize. When a repeller type pump is used, a gas seal that does not rely on the process fluid for cooling should be applied since the removal of seal-generated heat is low. Heat removal can be accomplished by using either a noncontacting seal face canister arrangement or a contacting seal face canister arrangement that has a grooved inner seal face.

Process Fluid Considerations Dual gas seals applied to pumps are different from those applied to compressors, mixers, fans and blowers. In pumps, the inner seal is immersed in the process fluid at either the inside or outside diameter, depending on the seal arrangement chosen. Even though the barrier system gas pressure is greater than the process fluid pressure, there will be process fluid in a portion of the gap between the seal faces – due either to capillary action or to the wiping effect from shaft or sleeve runout (Figure 7). With process fluids containing solids that tend to polymerize, coke or salt out, the entrained solids must be removed by applying a flush over the seal faces, by providing a seal chamber that flushes the solids away, or by

82

using a repeller pump that keeps the process fluid away from the seal during dynamic operation. If a carbon material is used for one of the seal faces, entrained solids can erode the seal face to a point where excessive gas leakage will occur. In some applications hard seal faces such as silicon carbide mated against silicon carbide may be beneficial. Such hard seal faces may also be necessary when strong oxidizing process fluids are being pumped. With light process fluids that have a low boiling point, a double gas seal design should be selected that provides the lowest seal-generated heat to prevent vaporization of the process fluid. Although process fluid vaporization will not usually affect the performance of the seal, it could lead to pump vapor lock. BARRIER GAS PRESSURE =Po

optimum film strength in both static and dynamic operation. • A design that allows the O-ring to conform to the shaft regardless of inconsistencies in the O-ring surface caused by temperature change, chemical attack, manufacturing tolerances or obstructions on the shaft. • The use of direct-sintered silicon carbide to eliminate fatigue-type failures, stress risers and the potential for rotor disintegration. •A drive mechanism, designed specifically for the physical properties of silicon carbide, that eliminates a single-point drive pin in the rotor body. Drive is imparted through coil springs around the rotor to disperse drive loading stresses around the rotor O.D. • A means for cushioning and aligning the seal face rotor at its O.D. so that centering forces subtract from, rather than add to, centrifugal stresses developed during high-speed rotation of the seal.

Conclusion

PROCESS FLUID PRESSURE Pi

Figure 7.

Technology Advances Recent developments in dry seal technology are permitting more reliable applications for high-speed compressors. These developments also have extended the use of dry seal technology to lower-speed compressors, pumps, mixers and other equipment. The most recent development was unveiled with the announcement of a standard bore gas barrier seal designed specifically for small, conventionally sized seal chambers. This new double noncontacting seal face gas seal for ANSI, DIN and ISO standard bore seal chambers has a proprietary seal design that includes welded metal bellows spring loading. Operational capabilities include pressure to 300 psig (2070 kPa) and temperature to 500ºF (260ºC). Other dry seal technology advances include: • An improved seal face geometry that provides optimum hydrostatic and hydrodynamic load distribution around the seal ring, as well as The Pump Handbook Series

Successful field applications confirm that dry gas sealing technology can improve the operation of rotary equipment originally fitted with liquid lubricated mechanical seals. Optional contacting and noncontacting seal face designs, plus many alternate design arrangements, provide for a wide range of applications to match the pump and service. Dual pressurized dry gas lubricated mechanical seals are being used to retrofit pumping equipment in many critical applications. This technology helps plant managers meet environmental regulations for fugitive emissions and prevent process fluid contamination. It also extends the time between periodic maintenance, reduces horsepower requirements and lowers equipment, operating and maintenance costs.■ William V. Adams is Vice President of Technology for Durametallic Corporation. He has more than 25 years of experience in the sealing industry. This includes serving as chairman of the WG3 Emissions Task Force of STLE’s Seals Technical Committee and being a member of the API 682 Seals Standards Committee. Mr. Adams is also a member of the Editorial Board of Pumps and Systems.

MECHANICAL SEALS & SEALING SYSTEMS HANDBOOK

Shaft Sealing for Pulp & Paper Successful European strategies point the way to reliability. By Heinz P. Bloch

ulp and paper plants worldwide are under increasing pressure to reduce maintenance costs and achieve increased run lengths of their production facilities. These goals are to be achieved while paying attention to the urgent need to protect and preserve the environment. One primary impediment to the accomplishment of these objectives to date has been shaft sealing problems. Soft packing rings have served the industry for close to a century, but they require frequent replacement or adjustment. Moreover, soft packing must be lubricated by either the pumped product or an externally introduced flushing liquid. The result is either costly and environmentally unacceptable outward leakage, or expensive inward leakage. Mechanical seals appear to be the answer to the dilemma. However, while they have been used with great success in the HPI (hydrocarbon processing industry) since the late 1940s, their acceptance has been slow and fraught with setbacks in the PPI (pulp and paper industry). A great many of the early problems were clearly attributable to use of seals with springs located in the pumpage, or to the combination of inexpensive seals with inadequately selected seal materials. More recently, sealing problems have arisen when users have tried to economize by using conventional instead of stationary seals, or direct shaft-mounted instead of cartridge-type mechanical seals. On the other hand, remarkable successes have been documented by users who select the optimum seal geometry, seal type and seal material for a given application. A major mill

P

located in Scandinavia has mechanical seals installed in 334 centrifugal pumps. During a 7-year period these seals were replaced or repaired 1.65 times per pump. Thus, the average seal life was 4.2 years! Such life expectancies are achievable with optimally designed and selected seals. It is of extreme importance to recognize that the least expensive, off-the shelf seal will rarely (if ever) give this kind of service. Many inexpensive seals are too small, have the springs exposed to the pumpage, or incorporate less-than-desirable geometries and materials of construction. To find and accommodate better seals, the equipment owner must consider making certain changes in his or her selection and procurement policies. In fact, even physical changes to pump components may be needed and are sometimes easily justified. In this context, it is important to keep in mind that optimum seal geometries rarely fit in the conventional smallsize stuffing box. We should remember that this stuffing box was probably designed in the early part of this century for the purpose of accommodating soft packing. Stuffing box enlargement or the procurement of upgraded pump covers with considerably larger seal cavities have proven very helpful and should be considered for many pumps in pulp and paper plants.

Basic Application Considerations Contrary to the mechanical person’s intuition, the performance of mechanical seals is not heavily influenced by the medium or fluid being sealed. Instead, seal performance The Pump Handbook Series

depends primarily on temperature and pressure in the seal cavity since these define the state of the fluid. Or we could simply say that at one pressure-temperature combination the fluid between the seal faces is closer to vaporization than at another pressuretemperature combination. As a rule, dry running of seals must be avoided. If, therefore, we use single mechanical seals in a medium containing a certain amount of vapor or with the seal cavity at a pressure below atmospheric, the introduction of an external flush stream would seem appropriate. A possible exception would be clean pumpage. In this case a properly chosen carbon composition, as opposed to hard face, might tolerate occasional dry running. Seals with a narrow face width generate considerably less heat across the faces than do seals with wide faces. A narrow face is approximately 0.120 inches wide. A wide face measures 0.200 or 0.250 inches across. With a narrow face seal and pumpage temperatures at least 10ºC (18ºF) below initial boiling point at operating pressure, external flushing is not needed. Conversely, with wide faces or pumpage at temperatures close to boiling, a cool external flush would be appropriate. Note that jacketed stuffing boxes are considered an extremely poor choice and cannot be expected to be effective. Wear-resistant face material combinations, primarily SiC/SiC (in acid services) or tungsten carbide (WC/WC) in alkaline services are generally selected for mechanical seals in pulp and paper applications up to 120ºC (248ºF). Above this temperature, some manufacturers favor siliconimpregnated carbon against silicon

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solids. In this service category seals with geometries similar to Figure 2 would be typical. C. Corrosive media. For this service category, typical geometries would be as shown in Figure 3 for double, and Figure 4 for single, mechanical seals.

Source: Burgmann Seals

carbide (SiC/SiC).

bide (WC). However, in pulp and paper services with low solids contents, face materials belonging to the carbon, ceramic and other families may be quite suitable as well. Secondary sealing materials would generally be EP rubber or EP rubber with PTFE (Teflon®) covering for alkaline pumpage. Fluorocarbon rubber formulations (Viton®) or PTFE-covered fluorocarbon rubbers typically serve acidic media. For a few selected services, other secondary materials would give better service.

Seal Geometry Overview Double mechanical seals or tandem seals are always suitable for pulp and paper services. However, since they are more costly and may require elaborate support systems, they should be used only where single seals cannot perform adequately. Stationary seals, i.e., seals in which the flexing, spring or bellowsactivated face is stationary, have the rotating hard face closer to the pump impeller. This feature, together with a cone-shaped seal cavity (Figure 1), can represent a number of significant advantages. Longer seal life is often the result because stationary seals operating in a tapered seal cavity keep the pumpage in motion. This provides better heat removal and resists the settling-out of particulate matter on critical seal components. A first-cut approach to seal selection in the paper and pulp industries might be to separate the pumpage into three broad categories: A. Liquids containing a high volume percentage of solids. For this service category, seals with geometries shown in Figures 1 would be typical. B. Liquids containing a lower volume percentage of

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A. Collar B. Spring C. Elastomer Bellows D. Driving Band E. Disc

F. Retainer G. Primary Ring H. Mating Ring I. O-Ring

Figure 2. Single-acting, elastomeric seals suitable for liquids containing a lower volume percentage of solids Inner Seal Ring

Inner Shaft Packing

Flush Out

Outer Shaft Packing

Inner Stationary Insert

Outer Seal Ring

Inner Insert Mounting

Outer Insert Mounting Outer Stationary Insert

Spring Gland Ring Gland Gasket Flush In

Figure 3. Double mechanical seal for

Figure 4. Typical representation of single-acting, unbalanced, outsidemounted mechanical seal

Seal Material Overview Face materials should be wearresistant carbides. For acidic media this would be silicon carbide (SiC). For basic media it would be tungsten carThe Pump Handbook Series

Source: Durametallic Corp.

Figure 1. Stationary seals placed in cone-shaped housings are well-suited for “dead-ended” installations in solid containing pumps.

Source: John Crane Inc.

Seal Environment Recall our earlier note regarding the advantage of providing cone or taper-shaped seal cavities for seals in many pulp and paper services. This means the seal can be much closer to the impeller, and it enables the seal faces to be flooded by pumpage. Greatly enhanced heat dissipation and reduced compacting of solids will be experienced with this type of environment. Dramatic increases in seal life will be the result. Although relatively new in the United States, tapered or cone-shaped seal housings have seen decades of satisfactory service in thousands of pumps overseas. For pumpage temperatures and pressures at least 10ºC (18ºF) away from the initial boiling point, this seal housing configuration, as illustrated in Figure 1 should be seriously considered. If the pumpage temperature and pressure combination should disallow the dead-ended seal configuration that results from tapered seal housings, one of two alternative solutions should be considered. The first would be to use a single seal in conjunction with the conventional cylindrical stuffing box and throat bushing shown in Figure 5. In this case, a cool, pressurized seal flush would have to be injected. This means flush fluid would migrate into the pumpage, thus diluting it. The second approach would be to use a double or tandem seal configuration, as shown earlier in Figure 3.

Safety and Environmental Protection For external seals (Figure 4), the mill will have to provide suitable mechanisms for leakage run-off, collection or drainage. When seals are mounted internally as shown in Figure

5, leakage provisions should be threaded into the vent port, usually at the bottom location of the gland plate. If the leakage product is miscible with water, a continuous water or steam quench can be applied as depicted in Figure 6 and led off to a safe or recycle-designated area.

Steam Purge Bushing

Source: Durametallic Corp.

Steam In

Other PPI Machinery Sealing Applications Although not within the scope of this article, sealing applications for machinery other than centrifugal pumps deserve to be mentioned. Capable seal manufacturers are generally able to offer conventional or, in

Steam Out

Figure 6. Stationary, inside-mounted mechanical seal with steam quench provision

Source: Flexibox Inc.

CLEAN PRODUCT TO SEAL

Figure 5. Single mechanical seal with external flush provision

some cases, split mechanical seals for refiners, top entry mixers, side entry mixers, chippers, knotters, vertical screens, horizontal screens and certain positive displacement pumps. It is here where seals from the mining industry are judged to be of high potential value to pulp and paper mills. Again, Figure 1 illustrates how allowing the seal faces to be liberally flooded by pumpage helps keep seal face temperatures on the low side, and wide open seal cavities reduce the probability of particulate becoming compacted around seal components. Although there is not enough space here for us to provide a full listing of seal installations in the PPI, a number of U.S. and overseas seals similar to the designs shown in Figure

The Pump Handbook Series

1 have given excellent service worldwide. Redesign of pump covers to accommodate technically superior seals is paying rapid and sizable dividends!

Summary The evolution of sophisticated sealing technology in the petrochemical industry worldwide, and the success of mechanical seal applications in the European PPI, contrast sharply with the slow acceptance of mechanical seals in the North American PPI. Attempts simply to replace the traditional soft packing with inexpensive mechanical seals will lead to unnecessary experimentation and frustration. On the other hand, plants employing the European strategy will capture credits and opportunities that are certain to surprise even the optimists.■ Heinz Bloch is a consulting engineer in Montgomery, Texas who has more than 34 years of experience with process machinery. He retired from Exxon in 1986 and, since then, has specialized in machinery reliability improvement and maintenance cost reduction consulting on all six continents.

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MECHANICAL SEALS & SEALING SYSTEMS HANDBOOK

Reliability Through Fluid Sealing Management Having sealing trouble? Take a step back and remember to look at the problem from a system perspective. By Michael L. McCauley, Fluid Sealing Consultants Reliability is knowing you can count on your equipment and people to do what they are supposed to do. The more reliable a company is, the more the world will want its products or services. fluid process system has several sub-systems: pumps, valves, flanges and possibly the cylinders. Within these sub-systems, the largest opportunity to improve reliability is in the area of fluid sealing. Proven technology exists, but few plants fully embrace it. Each plant tends to use a small piece of fluid sealing technology and resists the use of others. Many plants need only to pull their existing technology, skills and knowledge together into a comprehensive program of fluid sealing specifications and procedures. Most times these decisions need to be made by plant or corporate level managers. Solutions arrived at, however, involve all essential departments in the plant.

A

Overview Failure to contain fluids at flanges, valves, pumps and cylinders is extremely costly for process plants. The most common reason for valve maintenance is leakage: 80% of pump repairs, 90% of seal failures (they do not wear out) and nearly 100% of flange maintenance problems are caused by leakage. A high percentage of bearing and motor failure is due to the intrusion of moisture and/or contaminants. Other costs due to leakage include increased housekeeping, product

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and energy loss, and damage to other components. Personnel safety can also be compromised. Recently great strides have been made by certain corporations in increasing the Mean Time Between Failure (MTBF) of fluid handling production equipment. The directions these companies have taken are similar. The key success factors are: • Active management support • Standardization on the best available fluid sealing technology • Well defined, clearly stated procedures • Emphasis on knowledge and skills training to support standards and procedures • Documentation of results

Active Support by Management Management support is critical to the success of any long-term fluid sealing program. The increased reliability of pumps, valves, flanges and the reduction of leakage and leakage-related failures must be a high priority for the company. Only if you have this support can you reduce production costs, increase production units and profitability and improve safety. The most compelling reason you must have management on your side, however, is to improve production equipment availability. This goal, and the process of obtaining it, must be explained clearly and repeatedly to personnel at all levels and in all departments. They need to understand it and be able to verbalize core issues when asked. An effective manager is an examThe Pump Handbook Series

ple of this process in action. He frequently goes to the lowest level personnel in the organization and asks them to explain why a particular program was implemented. If he gets the right answer, he sends a positive note to all the supervisors in that department. If he fails to get the proper response, he works his way up the chain of command until he does. He has then found the break in communications.

Standardize on the Best Fluid Sealing Technology Standardizing fluid sealing products has many excellent benefits. Cost reduction (reduced inventory, SMO savings, etc.) is one obvious one. Others include system reliability through effective training, reduction of installation errors and personnel confusion due to a multitude of materials, and improved troubleshooting effectiveness. Standardization on the best available fluid sealing technology can take a plant into an aggressively proactive fluid sealing program. The results can be dramatic and immediate. A five unit chemical plant, for example, had 161 individual mechanical seal parts in its storeroom; with standardization, the company reduced that to 11. By using the best available technology, decision makers also reduced the cost of repairs and doubled MTBF. The incidence of emergency callouts was reduced by 90%, and that indicates improved reliability.

Well Defined Procedures When an action is performed identically by several different people, you have a true procedure. When that action is not replicated exactly, you have results that are impossible to troubleshoot or predict. Real procedures, then, are written and followed. Procedures that are consistently followed yield consistent results. Ron Frisard, Packing and Gasket Engineer at A.W. Chesterton, and I teach a course designed to help clients obtain better flange gasket results. We demonstrate the importance of the correct parts (B-7 bolts, hardened washers, and 2h nuts), the right tools (calibrated torque wrench) and the right procedures (4-pass turnof-the-nut method) for gasket flange reliability. The average mechanic, given a regular wrench and doing it “his way,” will typically miss the 60% of yield goal by 30 to 80%. However, when the correct tools and hardware are used and the procedure is followed, the same mechanics will be within 5% of the target. In the area of pumps, a “Pump Repair Checklist” is a good example of a procedure that, if used consistently, will improve reliability and record keeping (See Figure 1).

Emphasis on Training If we want quality results, we must instruct our personnel in more than the rudimentary job activities. If a pump repair technician is to do more than replace parts, he needs the knowledge and skills to do a complete job, and he needs the written procedures to do the pump overhauls with consistency. Using a typical ANSI end suction centrifugal pump and mechanical seal as an example, let’s examine the knowledge and skills required by a qualified repairperson. Knowledge • Centrifugal pump operation • Reading and understanding a pump curve • NPSHR vs. NPSHA • Determining where a pump is operating (where the system curve intersects the pump H-Q curve) • Cavitation (all sources), evidence of cavitation • Basic corrosion and erosion • Sources and causes of vibration • Bearings (types, installation, lubrication, troubleshooting)

• When and how to cool bearings • Principles of balance • Basics of troubleshooting wear patterns • Basics of mechanical seals • Seal troubleshooting • Auxiliary seal piping plans (when and how to use them) • Basic knowledge of O-rings and how to troubleshoot them Skills • Use of basic hand tools • Pump assembly and disassembly • Alignment (laser and dial indicator) • Impeller clearances and how to set them properly • Seal assembly • Seal installation • Bearing installation • Use of a dial indicator • Use of precision measuring tools

Documentation of Results Each of the companies we work for makes investments and expects a reasonable rate of return. Programs under discussion here require an investment in time, tools and training. Managers willing to make these investments deserve a return on their capital. It can become an enormously successful investment, but it requires documentation of present conditions, costs and MTBF so that current life-cycle costs can be compared to anticipated future results.

Managing a Fluid Sealing Program For fluid sealing programs to work, systems must be more reliable. This is best accomplished by thinking on a “System Improvement Basis” and upgrading components through improved procedures, standards and specifications. A system is not just hardware; it is the hardware plus the procedures that are followed when personnel (operations, maintenance and engineering people) do their jobs. As an example, one subsystem that is in most plant systems is a bolted flange. This system is made up of two mating flanges, a gasket, bolts, washers, nuts, lubrication and the torque applied to the nuts. If the gasket leaks, one of these components is missing or incorrect. In the case of pumps, we have the pump casing, shaft, impeller and bearing housing; there are also the The Pump Handbook Series

sub-systems of bearings and their lubrication, and several sealing devices including gaskets, O-rings and the mechanical seal. Each of these can be affected by operational dynamic factors. Managing fluid sealing for improved reliability requires the setting of standards, specifications and procedures that will ensure that all of the components in the system are what we know they should be every time.

The Journey Each program I have been associated with has been unique. They all start in different places with different goals and resources. The first step is to make certain that the plans and goals of the plant are known, and that top management will actively support the new improvement efforts. There must be a clear understanding of the approach that will be used by management to measure performance. This will almost certainly differ from what the plant has used before. The work begins with understanding the equipment and fluids in the various systems, as well as the temperatures and pressures that are involved. All industry or governmental regulations and special safety precautions must be discussed at this point. Existing corporate or plant specifications and procedures for pumps, cylinders and valves should be reviewed. The equipment usually can be divided into four major categories: 1. Valves – AOV, MOV, manual, blocking and control 2. Flanges – including pipe flanges and all static gasket applications 3. Rotating equipment – including pumps, compressors, gear boxes, fans/blowers, etc. 4. Hydraulic and pneumatic cylinders The fluids are usually broken down into the following groups: 1. Water 2. Steam 3. Gases (air, nitrogen etc.) 4. Oil – lubricating 5. Hydraulic fluids 6. Chemicals

Building a Team At this point a team of key people is assembled. These should be the best from each of the following

87

areas: operations, engineering, maintenance, purchasing, management and key vendors (rotating equipment, sealing devices, pipes and valves). The team should have an outside facilitator or consultant who reports results and progress to management and keeps the team meetings focused and productive. The team starts by drafting any new standards, specifications and procedures needed to improve reliability. What is now being used? How was the choice of sealing devices made? What procedures are actually being followed? When and how were they created? No new, cutting edge technologies are to be considered here; only proven materials and procedures are to be incorporated by this team. Test/pilot operations are for another team, put in place later. The following is the beginning of typical standards, specifications and procedures. Valves • Packings: high quality braided Teflon®, in five-ring sets for all types of valves and all fluids below 500°F except flammable fluids; graphite five-ring sets for all flammable fluids above 500°F • Carbon bushings to fill deep stuffing boxes, as needed • All valves to be completely unpacked and inspected when repacking is needed • All gland bolts, nuts and washers to be lubricated with a premium nickel anti-seize compound Flanges • All flanges above 300# to be sealed with “spiral wound” CG gaskets • All bolts used on flanges to be B-7 grade stainless steel, except where plant standards state “carbon steel” bolts are to be used • Cut gasket styles used by application, per display chart in shop area • All bolts, nuts and washers lubricated with plant standard antiseize • Bolts torqued in a 4-pass procedure, using a calibrated wrench and turn-of-the-nut method Pumps Pumps shall be upgraded, where possible, to include the following components:

88

• Enlarged seal chamber to replace conventional stuffing box supplied for packing • Solid shafts for lower L3/D4 ratio to replace sleeved shafts supplied for packing • Bull’s-eye sight glass to replace automatic oilers • API auxiliary piping designs to be designated and installed as needed • Coupling alignments done to plant specification • All baseplates to be inspected and approved or repaired/replaced • Plant standard repair checklist to be followed and completed (see Figure 1) • Labyrinth seals or positive face seals to replace original oil seals Mechanical Seals • All seals to be balanced using nonfretting styles • All mechanical seals to be cartridge mounted unless equipment design precludes it • Split seals to be capable of either positive pressure or vacuum services • Single or double seals to be used with auxiliary piping, as specified The previous examples demonstrate that specifications and procedures do not need to be wordy. Keep in mind OSHA 1910.119 requirements where they apply. Once the specifications, standards and procedures are approved, and before implementation, the required new parts and tools must be brought in so they are available on “Day One” of any new program.

Information Availability If one is not already in place, each department should do an up-to-date equipment survey, by system. All detailed information available should be captured, and this survey should be made available to all plant personnel. Since keeping this survey information updated is critical, a centralized computer system should be used where at all possible.

Training Significant classroom/seminar and hands-on training must be done to ensure that hourly and supervisory personnel understand the overall program. Everyone must know how the new program will be implemented and how new procedures The Pump Handbook Series

will be applied. Training must be targeted to the needs of the people; a pre-test of personnel will help focus on individual needs. The training must outline the benefits of this new program as well as develop the knowledge, skills, information and familiarity with tools required. During this training, testing of skills and knowledge levels should be done; an immediate remedial training program should follow for those whose test scores are below the threshold. All first line supervisors should be required to attend the training and pass the test.

Key Vendors Vendors are critical to this process. They must be chosen carefully; you must become aligned with the best available products and customer service and the most knowledgeable representatives in your area. You must be certain of a vendor’s ability to deliver what he or she promises. Key vendors will be expected to participate in all aspects of the program and bring their value-added service to any agreement. During price negotiations, these services must be taken into consideration. Any vendors not making a fair profit with a specific customer will be forced to find a way to cut back on services or quality. They may be forced to abandon you as a customer. If they are fulfilling their end of the agreement, they have a right to a decent profit. If a key vendor cancels an account, the plant will surely suffer.

Implementation The improvements program should be started immediately after the training is completed so the momentum generated during training is not lost. Information on the new standards and procedures should be readily accessible to all involved plant personnel, and it should be coordinated with engineering, purchasing/stores, maintenance and operations. The program needs to be “front-loaded” with extra time given to it by all team members, so that a positive start is achieved. The program also needs the commitment of every team member to achieve long-term results.

Troubleshooting Troubleshooting should be done immediately, at the time of the over-

haul/rebuild, using key personnel and the new standards and procedures. The causes of all failures should be recorded and categorized for trend analysis. Troubleshooting must be part of the training so that machinists and repair people can do an adequate job at the time of equipment disassembly. A brief report should be written, and either the actual failed components or photographs of them should accompany the report, where practical for each equipment breakdown. These reports should go to the area managers immediately, and the project team should review each situation at a monthly meeting, determining what changes should be made, if any. Proposed changes should be presented to the area managers as suggested improvements for the next installation/ repair.

Summary Long term reliable performance of fluid sealing devices, and the equipment in which they operate, is achievable; this technology exists! It requires management’s support and a well-designed program to be successful. An “Improvements Program,” well conceived and supported by top decision makers, is very rewarding. It is not unusual for the cost reduction improvements to exceed the total purchases of sealing devices, and for the increased production, through higher reliability and availability, to be pure financial gain. Maintenance, in other words, can actually be a profit center. There are tremendous benefits for your organization when you keep that in mind. " Michael L. McCauley is owner of Fluid Sealing Consultants, Schererville, IN. He has 26 years of industrial experience with several organizations, calling on all major industries, developing programs and solutions to industrial equipment leakage. He has developed technical courses and conducts training seminars in the United States and internationally. He has a Bachelor of Science degree in marketing from the University of Illinois.

PUMP REPAIR CHECKLIST Date ________Pump I.D.# ____________Repairperson ________ Lock out – tag out # Proper use of all safety equipment # Ensure all forms of energy are relieved #Electrical #Pressure #Temperature # Drain and decontaminate pump as required # Completely clean pump with special attention to mating surfaces # Check stuffing box face: # Good gasket surface # 125 RMS or less? # True to shaft to within .005”? # Check shaft sleeve for burrs, scratches, or old set screw marks – eliminate any marks found. Make sure all sharp edges that seal must be installed over are chamfered # Sleeve finish to be 32 RMS or better # Sleeve O.D. to be + .000”, -.003” # Check bearing fit bore and shaft to recommended tolerance (Refer to bearing manual or vendor) # Pump sleeve must be sealed with new gasket, O-rings or TFE ring # Use all new O-rings of proper size and material # Replace bearings with correct new ones # Replace oil seals/or check labyrinth for proper condition # Check impeller for erosion, corrosion, cavitation or chipped impeller tips: # Repair/replace as needed, including dynamic balancing of new impeller on rotating assembly # Check that impeller is the proper size: measure diameter and record # Check wear ring – replace or correct as needed # Check: #Fit to housing #Fit to impeller #Running clearance Acid clean stuffing box cooling jacket, and pressure test, where # applicable # Replace bearing housing cooling jacket with cooling coil in oil sump (Consult with pump manufacturer’s engineering) # Dial indicate the shaft: Maximum Allowable Runout = .003” # Re-assemble pump and set proper impeller clearance (see pump manual) # Ensure mechanical seal is correct type and materials # Install seal following manufacturers recommended procedures # Re-install rotating unit/pump into system # Check rotation of motor prior to installation of coupling spacer # Align motor to pump # Manually rotate pump, assuring no rubbing or binding # Connect auxiliary piping, as directed, where appropriate # Ensure that all ports not in use are plugged, unless specified to be left open # Pressure test for static leaks, per plant procedure # Sign off, remove lock and tag, sign and date Figure 1. Pump Repair Checklist The Pump Handbook Series

89

MECHANICAL SEALS & SEALING SYSTEMS HANDBOOK

Choices in Emission Control for Rotating Equipment New technologies are proving their worth in real-world service. By Ken Lavelle, Flowserve Corporation, Fluid Sealing Division echanical seal manufacturers continue to advance seal technology to meet the strictest emissions control regulations for rotating equipment. In addition to reducing emissions, proper application of the latest sealing technology leads to increased reliability with corresponding increases in safety and lowered maintenance costs. Mechanical seal designs that directly seal a liquid product are well proven for reliability and emissions control in many applications. Single seal designs can be supplemented with liquid secondary seals operat-

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ing on a non-pressurized buffer fluid, dry running or lift-off gas type containment seals. These arrangements can provide near zero emissions control with high safety and reliability. Gas seal technology, originally developed for demanding gas compressor applications, continues to evolve as a sealing option. Double gas seal technology can provide zero emissions while increasing seal reliability in applications where fluid conditions may be inconsistent, such as batch operations in which pumps may tend to run dry or cavitate. Exciting new gas seal technology is extending the application range and

Figure 1. API 682 Type A low emissions single seal design

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The Pump Handbook Series

reliability of these devices, making it possible to use them in mixed liquid/gas services, and in other rotating equipment such as steam turbines and low speed mixers in chemical and pharmaceutical services.

Emissions The term emission is used in different ways in equipment sealing. It can be used to indicate any product that reaches the environment outside the equipment, or it may relate to a specific regulatory definition. Normally there are regulatory concerns only with products that can be hazardous to the environment, but all “emissions” have a cost in product and energy loss. Government regulation of mechanical seal emissions first appeared in the early 1980s. Specifically covered were releases of volatile organic compounds (VOC). The measurement technique defined—U.S. Environmental Protection Agency (EPA) Method 21—is still used today. Method 21 is a field measurement using a portable, explosion-proof organic vapor analyzer commonly referred to as a “sniffer.” This instrument measures the concentration of vapors in air of a sample taken one mm from the

Emissions (PPM)

Product Loss (kg/year)

100,000 10,000 1,000 100 10

1650 210 30 3 .5

Table 1. Product loss equivalents per seal

seal end plate. The measurement is expressed in terms of a parts per million concentration (ppm) (Ref. 1). In a typical hydrocarbon application such as the sealing of liquid propane under moderate pressures, emissions before 1980 regularly exceeded 100,000 ppm. Improved sealing technologies now typically reduce such emissions to less than 500 ppm with single seals, and down to zero ppm for dual seals. The use of terms such as ppm often leads to poor understanding of the actual amount of product loss. The EPA developed product loss equivalents with actual losses as shown in Table 1. These equivalents at best only give order of magnitude accuracy, but they can be used as reference points (Ref. 2). At 100,000 ppm in the early 1980s cost savings could readily be achieved by reducing product loss. In addition, as sealing technology was improved, seal life increased substantially, creating savings in equipment maintenance dollars as well. When sealing at levels of 100 ppm, the product loss per year is miniscule and for most fluids only represents a few dollars in cost. Reducing emissions to zero thus has a very small impact in savings due to product loss, but continued improvements in seal reliability remain a large opportunity. As an example, one large U.S. refiner reported an improvement in mean time between repair (MTBR) from 2 years to 6 years, a 40% reduction in overall seal maintenance costs and a 60% reduction in inventories from 1988-1994 (Ref. 3).

Low Emissions Liquid Seal Designs Low emissions seal technology began in the early 1980s in response to tightened emissions regulations. Initial design changes consisted of upgrades in materials and alterations in balance ratio to increase the closing forces of seal faces. While these improvements resulted in lower emissions at increased cost, high overall savings were achieved in reducing product loss and achieving longer seal life. During this time frame seal manufacturers also realized the need to optimize seal designs on test loops simulating actual pumping conditions, such as ethane and propane services. This testing, combined with modern finite element and fluid film analysis techniques, resulted in additional improvements, enabling users to meet most emissions regulations with simple single seal technology. These seals can withstand short process upsets without adverse effects on life or performance. Today low emissions technology is well proven and routinely achieving seal life of more than five years while meeting emissions requirements. The first industry standard to document these improvements and establish a seal selection methodology was API Standard 682, Shaft

Sealing Systems for Centrifugal and Rotary Pumps (References 4 and 5). Single mechanical seals provide reliable sealing for most VOC services under the following operating conditions: • Fluid specific gravity > .45 • Vapor pressure margin in the seal chamber > 25 psi • Flush fluid and pump operating conditions providing good basic lubrication of the faces In addition to API 682, several other published references clearly define this technology. Similar designs are available in cartridge seal packages for ANSI and DIN pumps. Improvements in seal chamber and pump design for modular pumps— such as enlarged tapered-bore chambers with flow modifiers to assist in removing solids from seal chambers, and bearing labyrinth seals—have been proven successful (Reference 6). Dual liquid seal designs utilizing unpressurized buffer fluids can show essentially zero emissions and high reliability. These designs are also well described in the literature and API 682. Dual seals using gas containment seals are gaining in usage, particularly for less severe services in which single seals would normally be specified but dual seals are required for regulatory compliance or for increased safety.

Figure 2. Series arrangement low emissions primary seal with secondary containment seal The Pump Handbook Series

91

Seal Vapor Balance Diameter

Face Materials

Carbon Face Projected 5 Wear Year Carbon Wear

Wear Nose Length

2.375”

Propane Carbon v. Silicon Carbide

<.0001”

1-.008”

.125”

3.625”

Propane Carbon v. Silicon Carbide

<.0001”

0-.008”

.125”

Table 2. Carbon wear test results (500 hours of testing @ 5 psi)

Seal Type Dry Running Lift Off

Normal Operating Pressure—Gas/Vapor

Liquid Containment

5 psi

750 psi

600 psi

600 psi

Table 3. Pressure ratings for standard containment seals

Dual liquid seal designs using pressurized barrier fluids are effective but rely on a system that can be relatively complicated to operate and maintain. Often in process services in the U.S. today, double gas seals are an attractive option to double liquid seals.

Gas/Liquid Containment Seal Designs Gas containment seals, used in series dual seal designs with low emissions liquid primary seals, are now well proven for emissions control and safety backup uses. These sealing devices were not included in the first edition of API 682, but a solid experience base is now available, and these designs will be included in the second edition being developed. Figure 2 shows a typical arrangement. There are two categories of containment seal design. Dry running seal faces have limited aerodynamic film support between the faces where the face materials are in dry sliding contact. Lift-off face designs use a face pattern that achieves full gas film support between the faces, making them essentially non-contacting. Contacting face designs may have face patterns such as deep hydropads or grooves that promote face cooling and make it possible for

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wear particles to migrate away from the contact area. Special carbon grades that have exceptionally low wear in dry gas operation are used, normally running against silicon carbide mating faces. Dry running seals have a lower operating pressure range in dry operation when acting as emissions control devices, but they can achieve far higher pressures when used for liquid containment with relatively low leakage in both operating modes. Lift-off designs can achieve higher operating pressures in gas operations, but they may have higher leakage rates in both gas and liquid modes.

Figure 3. Spiral groove face pattern The Pump Handbook Series

Dry running seals will normally have a measurable wear rate when running in gas. Wear rates for one manufacturer’s standard dry running seal design, which would have a minimum seal life of more than years, are shown in Table 2. Duty ratings vary by manufacturer and design, but typical ratings for process equipment containment seals are shown in Table 3. Various face patterns can be used in lift-off designs. Spiral grooves (Figure 3) have an aggressive pattern etched into the hard face. These grooves are typically on the order of 10 light bands deep, and achieving lift-off is dependent on the direction of rotation. Bidirectional etched patterns such as T grooves are also commonly used. Wavy face technology (Figure 4) has a bidirectional pattern with no etched grooves and a wave shape. The wave has high amplitude at the outer diameter to facilitate gas compression and liftoff forces while tapering to a sealing dam at the inner diameter. This wave shape has been highly resistant to contamination and can seal changing mixes of liquids and gases. For lift-off face patterns, the face pattern must generate adequate film thickness to support the sealing faces without damage while they are rotating. When the faces make the transition from gas or vapor to a more vis-

cous liquid, the closing force must be high enough to prevent the face separation from becoming too large, which can generate high amounts of leakage. Face pressure due to spring loads with these gas seal designs are normally on the order of 1/10 of the typical liquid seal face pressure, making resistance to hang-up, fouling and secondary seal drag very important.

When sealing volatile fluids with a primary liquid seal/secondary containment seal design, the secondary cavity will fill with vapor because even minor leakage from the primary seal flashes to a gas. Since a typical light hydrocarbon liquid can have an expansion rate of several hundred times in the liquid/gas transition, a miniscule

Figure 4. Wavy face design

Figure 5. Propane emissions at low pressure for wavy face seal

Balance Diameter

Pressure (psi)

Speed (rpm)

Fluid

Leakage (cc/min)

2.375”

300 300 300 600 300

3600

Diesel ISO 32 Oil ISO 32 Oil ISO 32 Oil ISO 32 Oil

2 2 8 6 15

3.625” 6.000”

3600 3600 1750

Table 4. Typical liquid leakage from wavy type lift-off containment seal The Pump Handbook Series

amount of liquid leakage produces a significant amount of vapor. The containment seal must be able to seal the vapors at low pressures to be emissions-tight. Leakage rates will vary between designs. Figure 5 is a plot of propane emissions for a 2.375” diameter wavy face containment seal design taken in ppm through the EPA Method 21 for pressures less than 25 psi. Leakage rates when sealing liquids will typically be higher for containment seals than for standard liquid seals. Both types of seals will normally produce emissions of 10,000 ppm or higher when sealing pressurized light hydrocarbon fluids, but there will be no visible vapor cloud. Non-volatile fluids may show light dripping depending on fluid viscosity and pressure. Typical tested leakage rates in liquid containment services are shown in Table 3. Piping plans for these containment seal designs are critical. A schematic of a typical piping arrangement is as shown in Figure 6. The following are important piping considerations: • For volatile liquids the secondary seal cavity is typically piped to a vapor recovery system. A .125” orifice restriction in the piping and a pressure alarm normally set for 10-25 psi can be used to detect increasing primary seal leakage. • A check valve is normally used in this line to prevent gases from the vapor recovery system from flowing back into the secondary seal cavity. Normal operating vapor recovery system pressures are below 5 psi, but the pressures can increase to 30 psi or more during plant upset conditions. The check valve cracking pressure, which will determine the normal operating pressure in the secondary sealing cavity, should be set for 2-5 psi. • When using containment seals in services with non-volatile fluids, hang- up of the secondary seal is a concern. In these cases the piping should include a low point drain.

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Dual Gas Seal Designs Dual gas seal designs, consisting of two gas seals pressurized by a barrier gas, are rapidly replacing conventional double pressurized liquid seals. They offer zero emissions sealing and a high resistance to offspec fluid or pumping conditions. A typical dual gas seal design is shown in Figure 7. Existing designs now have many years of operation behind them. These seals utilize an aerodynamic face pattern to generate lift-off forces as described previously for containment seals. Standard cartridge design dual pressurized gas seals are now available for a wide range of applications.

Normally nitrogen or air is used as a barrier gas, pressurized at 30-60 psi above the sealed equipment pressure. A typical barrier gas control panel is shown in Figure 8. A small amount of barrier gas is consumed in operation. A typical barrier gas consumption curve is shown in Figure 9. Since a control panel is used and barrier gas pressure must be maintained at least 25 psi over product pressure, the support system is normally more complex than required for non-pressurized dual containment seal designs,. Barrier gas must come from a reliable source, such as plant nitrogen or instrument air. Gas consumption can be on the order of 10-

New Technology Is Extending Limits for Gas Seals There have been several problem areas in gas seal technology. Hangup of the spring-loaded face has been one. Pusher seal designs use a dynamic gasket secondary seal, typically an elastomeric O-ring. Spring loads, as mentioned, tend to be low with gas seal designs. These O-rings can exhibit hang-up due to sliding friction, and this hangs up the seal. In addition, elastomers are sensitive to fluids and heat and may swell of up to 30% by volume. Hang-up of the seal faces can cause a rapid increase in leakage. I have observed small ANSI pump vapor lock due to high barrier gas leakage on the order of 8-10 SCFH into the pump head. The use of PTFE secondary seals can greatly reduce the sliding force and eliminate the fluid compatibility problems inherent to elastomers. A unique patented design that reduces elastomer drag and allows for swell is shown in Figure 10. This design consists of an elastomer O-ring that is lightly loaded by a garter spring. Over a thousand dual gas seals using PTFE or garter spring technology have been placed in service in the last 5 years. Solids—and fluids that crystallize or polymerize,—often present hangup problems in dual gas seals. Current seal designs may not recover stable operation after pressure reversals,

Figure 6. Typical containment seal piping

Figure 7. Typical dual gas seal design

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100 SCFH per day, depending on seal size, pressure and shaft speed. The use of gas sources such as nitrogen bottles is normally not recommended. A standard nitrogen bottle. which holds approximately 230 SCF of gas, may be depleted within a few days, or even more rapidly if there is a small fitting leak in the system. Pressure booster systems that provide high barrier gas pressure are an option, but the reliability of commercially available booster units that utilize a dry sliding seal may be lower than the pump/seal reliability, especially at higher pressures and gas volumes.

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The gas seal featured in Figure 12 is specially manufactured for steam turbines. It incorporates a metal bellows and wavy face design that is highly resistant to contamination and can operate in condensate, saturated, or dry steam.

Low Emissions Seals for Mixers

Figure 10. Low drag elastomer seal design

which can be induced by loss of barrier gas pressure or high sealing pressure operation. These problem areas are a subject of ongoing research involving exclusion devices and improved seal arrangements to further increase seal reliability. New dual gas seal designs using metal bellows seals eliminate the secondary gasket problems completely. As shown in Figure 11, these new designs can be made in compact packages that can be fitted to the old small chamber ASME pump designs.

Mixers are used in a wide range of services in the chemical, pharmaceutical and food industries, as well as in bio-process technology. In many of these services, it is more important to keep contaminants out of the process than to keep the process material contained. Rotating shaft seals for mixers have typically been traditional double seal designs with a barrier fluid, which provides positive product containment. These sealing designs have

been established for many years. In some processes even a small amount of barrier fluid leakage is not acceptable. This has led to widespread use of dry running type seal designs over the last decade. The basic technology is similar to the technology employed in dry running containment seals, using special dry running carbon grades running against silicon carbide faces. A single seal can be used in the vapor space of a top entry or side entry mixer, or a double seal with a pressurized gas barrier can be used. Normally barrier gas pressure is set at 30 psi above vessel pressure. Dry running seals of this type will show some measurable face wear, depending on operating conditions. In some services even a tiny amount of carbon contamination is unacceptable; in others the carbon face material may

Seals for Steam Turbines Emissions control can apply to nonhazardous products as well. A good example is gas seal technology being applied to the sealing of steam turbine drivers. Most turbine drivers today are sealed by rings or bushings, which can allow substantial leakage. This leakage has a cost in product loss, can result in safety hazards around equipment due to poor visibility in cold weather, and visible clouds of steam vapor above a plant can be reported as an environmental concern. Replacing a turbine ring seal with face type steam seals can reduce steam leakage by as much as 200 times. Reliable steam sealing requires a high resistance to contaminants such as pipe scale and chlorides, and high temperature capability. The ability to seal condensate, saturated or the dry steam conditions that can be experienced in start-up and dynamic operation, is critical to reliable seal operation.

Figure 8. Typical barrier gas control panel

Figure 9. Typical barrier gas consumption curve The Pump Handbook Series

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not have adequate fluid resistance to perform acceptably. Typical running speeds in mixers vary from 0 to 12 feet per second, limiting the use of lift-off face technology. New developments in lift-off technology have resulted in successful applications in mixer services, with both hard/soft (carbon vs. silicon carbide) and hard/hard (silicon carbide vs. silicon carbide) seal faces in double pressurized gas seals (Ref. 7). The hard/hard material combinations virtually eliminate process contamination and process emissions.

Application Recommendations for Low Emissions Sealing A wide range of low emissions sealing technology has been presented. Recommendations for specific

applications in the process industries are normally made on the basis of four factors: • Safety • Regulatory compliance • Reliability • Total cost to operate and maintain Examples of specific recommendations for low emission technology selections in typical refining and chemical services are as follows:

Refining Applications for API Pumps Single seals (API 682 Type A, B, C Arrangement 1) with floating carbon bushings are well proven for safety, reliability and low emissions performance in the majority of refining applications. .Where local emissions regulations require dual seals, single seals

Figure 11. Small cross section of a metal bellows dual gas seal design

with gas backup seals are normally recommended as the most cost effective option. Liquid/liquid dual seals (API 682 Type A, B, C Arrangement 2), properly applied, offer high reliability and resistance to upsets in pump operation or unstable vapor pressure. Double gas seals may offer advantages and be recommended in pumps where zero emissions are required, dry running conditions are frequent (tank farm or batch processes) or vapor pressure problems exist. Since many refining services have operating pressures that exceed plant nitrogen header pressures, the barrier gas supply should be carefully evaluated.

Chemical Applications for ASME Pumps Single seals are used in most chemical applications. Enlargedbore seal chambers are recommended for best operating life. Cartridge seals are usually recommended to provide a pre-tested package with minimal installation problems. Double gas seals are recommended in many areas, especially in batch services where the pumps can run dry. In many plants today, double gas seals are successfully replacing traditional double seals with pressurized liquid barrier fluids and increasing reliability. Low pressure services in most ANSI pumps/chemical process applications can normally use plant nitrogen headers for a barrier gas source, combined with a simple control panel. "

References 1. Determination of Volatile Organic Compound Leaks, U.S. Code of Federal Regulations, Title 40, Part 60, Appendix A, Reference Method 21. 2. Mechanical Seal Performance for Low Emissions of Volatile Organic Compounds, William E. Key, Ken Lavelle, George Wang, 13th International Conference on Fluid Sealing, 1992.

Figure 12. High temperature steam seal design

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The Pump Handbook Series

3. Major Refining Company Cuts Mechanical Seal Maintenance Costs with Standardized Reliability Program, V Ray Dodd, Robert C. Watkins, 1995 NPRA Maintenance Conference 4. API 682, Shaft Sealing Systems for Centrifugal and Rotary Pumps, American Petroleum Institute, First Edition, October 1994 5. Sealing Technology for VOC Control, Ken Lavelle, Bill Key, Pumps and Systems Magazine, April 1997. 6. Life Cycle Costs for Chemical

Process Pumps, Bob Urwin, Rich Blong, Chemical Engineering, January 1998. 7. Environmental Seals for Mixers, Hans Wilhelm Laarman, 1998 European Sealing Association Conference, Dusseldorf, Germany. Ken Lavelle is Director of Worldwide and Technical Services for Flowserve Corporation (formally BW/IP), in Temecula, CA. He has more than 20 years of experience in the mechanical sealing industry and is a member of the

The Pump Handbook Series

STLE (Society of Tribologists and Lubrication Engineers) Seal Committee and the SAE. Mr. Lavelle is past chairman of the STLE Seal Education Course and a member of the API 682 Task Force. He received his B.S. degree in Mechanical Engineering from the University of Illinois, Urbana, and holds a certificate in engineering management from California Institute of Technology. Mr. Lavelle is a member of the Editorial Advisory Board of Pumps and Systems Magazine.

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MECHANICAL SEALS & SEALING SYSTEMS HANDBOOK

Maximize Seal Flush Performance for Longer Seal Life Controlling the seal environment is critical, not optional. Choosing the right API plan means knowing the components and the requirements. Start here. By Mike Polk, Reliability Associates General Requirements The flush or injection to a mechanical seal is crucial to proper mechanical seal performance and operation. Many repetitive and/or premature seal failures can be traced to improper seal flush or quench systems. Often an adequate seal flush system is available but is not being used properly. For this reason it is essential that the flush system be designed so that it can be easily monitored and maintained. The use of gauges, rotometers, flowmeters, etc., will simplify the proper operation of a flush system. Although much of the material included in this article can be applied to seal flush systems in general, one area requires particular attention: hot oil seals operating at or near their auto-ignition points. These pumps represent one of the most critical populations of mechanical seals, not only in terms of reliability, but most especially for safety and environmental concerns. To better understand the requirements for a proper seal flush system, we must first appreciate what is happening to the mechanical seal faces during operation. The fluid in the stuffing box must be maintained at a pressure that is above atmospheric. This ensures a positive flow of lubri-

98

cating film across the seal faces to provide lubrication and cooling. As the lubricating film passes across the seal faces, it goes through two significant changes. First, the pressure on the fluid film is gradually reduced from full stuffing box pressure to atmospheric pressure. This is sometimes referred to as the pressure gradient across the seal faces. Second, the fluid film is heated as it passes through the seal faces. Face friction, liquid shear and face speeds are a few of the variables that affect the value of the heat generated. Normally, we would expect a 20–40°F temperature rise of the fluid between the seal faces. It is important to note that the fluid thickness normally varies from 50–100 millionths of an inch. Therefore, a very small volume of liquid is involved. Often the combination of pressure degradation and heat increase will result in the fluid film changing state from liquid to gas. This phenomenon is referred to as “flashing” across the seal face. Depending on how much and where this flashing occurs, the result can be loss of lubrication, even more adversely affecting face lubrication. In general, when the ratio of liquid phase to gas phase between the seal faces reaches 50% or less, the seal becomes unstable, The Pump Handbook Series

resulting in short seal life. If this condition cannot be rectified by a proper seal flush, an alternate seal design may be dictated. You may want to try a double/tandem seal, hydropads, or dry running designs—either spiral groove or wavy face. A seal flush at the proper temperature and flow rate can prevent or control flashing and provide an adequate lubricating film for the seal faces. There are two essential concerns in determining the proper pressure, temperature, and flow rate for a mechanical seal flush. The primary concern is to maintain the stuffing box pressure a minimum of 25 psi above the product vapor pressure. Also, the flush rate must be selected so that a proper flush flow rate is achieved. Stuffing Box Pressure When pump suction pressures exceed 250 psi, the stuffing box pressure should be maintained at least 10% of suction pressure above product vapor pressure at pumping conditions. Pumpage that is very light (> .5 specific gravity) or with a low viscosity (>10 SSU) might require even greater vapor suppression in the stuffing box. Vapor suppression is normally achieved by lowering the stuffing box temperature,

rates and can cause a mechanical malfunction of the seal. Metal bellows seals are particularly susceptible to problems caused by flow rates. Additional factors for consideration in selecting optimum flush flow rates include: 1. A relatively small loss of efficiency is incurred when using a product recirculation flush. 2. If a cool flush is used, whether internal or external, it may represent a heat loss in the process system. 3. If a processed fluid is used as an external flush, it may represent a rerun cost. All of these expenses can be significant in individual cases, to the extent that an alternate seal design may be dictated, however expensive the initial cost.

Flush Piping

Figure 1. Vapor pressure vs. temperature, light hydrocarbons

thereby reducing the vapor pressure in the box. However, vapor suppression can also be achieved by raising the stuffing box pressure. The latter is the more difficult option to achieve, as we shall see later. Consider the following from Figure 1, which shows the vapor pressure for various hydrocarbons. • pumpage = propylene • temperature = 75°F • stuffing box pressure = 160 psia • vapor pressure = 150 psia To achieve the required 25 psi stuffing box pressure above vapor pressure, we would have two options: 1. Elevate the stuffing box pressure to 150 psia + 25 psi (175 psia). 2. Lower the flush temperature to the point at which the vapor pres-

sure is 150–25 psia or 125 psia; that equals 62°F per the chart. Due to the steepness of the vapor pressure slopes for fluids such as hydrocarbons, it is often more practical to cool the flush medium than increase the stuffing box pressure. Flush Rate The flush rate must be selected so that a proper flush rate is maintained without upsetting the mechanical seal. In general, you can estimate the required flush flow to be .75 gpm per inch of seal diameter. A rate that is too low can cause inadequate lubrication of the seal faces, resulting in heat buildup, high face wear and flashing—all of which can lead to premature seal failure. If the flush rate is too high, turbulence can upset the operation of the seal faces. This results in high and erratic leakage The Pump Handbook Series

Designing the proper controls for a seal flush system is essential. The best systems are easy to maintain and provide an adequate lubricating film to the seal faces. As in all mechanical systems, the more devices required for proper operation, the lower the reliability. There are several basic systems. A brief description of each follows. Control Orifices These are the most popular option. There are several orifice designs, but the pipe type with a long cross section is preferred. To be effective, it must be located at least 12 inches ahead of the seal gland. A “Y” strainer should also be included upstream or ahead of the orifice. The orifice should be installed in the system piping so that it is not normally removed with the seal flush piping during pump overhaul. Union orifices should be avoided. Their performance is questionable, and they can easily be misplaced, even unknowingly, if the union is broken. This often occurs when the seal flush piping is dismantled for maintenance.

99

As previously noted, a rotometer or flowmeter should be included downstream of the orifice as a positive indication of seal flush flow. Pressure gauges are not reliable indicators of flush flow unless there is one on each side of the control orifice. Example Equipment: • 2” seal • 300 psi discharge pressure • 200 psi stuffing box pressure • .8 S.G. (specific gravity) 1. Recirculation flow 2.5” seal = 1.25 to 1.75 gpm 2. Head = (300 psi - 200 psi) x 2.31/.8 = 290 3. Enter right side of seal orifice chart at 290 head feet. Move left horizontally to seal size column (2 3⁄8” to 3 1⁄4”) that intersects the diagonal lines. An orifice 073” to .086” comes closest to providing required flow of 1.25 - 1.75 gpm.

Figure 2. Vapor pressure vs. temperature, common liquids

Figure 3. Orifice selection guide—orifice capacity: 68°F (20°C) water at 31 SSU (1.0 centistoke) viscosity (Courtesy Flowserve Corp.)

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The Pump Handbook Series

The flow across the seal flush orifice can be determined by using Figure 3. It is essential that all flow rates be calculated using feet of head, not pressure. Formulas for feet of head calculations are: • actual flow = (31/ssu) x orifice capacity • psi = (head feet/2.31) x specific gravity • head feet = (psi x 2.31)/specific gravity • kPa = (head meters/ .010212) x specific gravity • head meters = (kPa x .10212)/specific gravity To properly size the orifice, specific gravity at the flow temperature must be known. Extremely heavy products may require corrections for viscosity that are beyond this text. Given correct operating conditions and physical properties of the flush fluid, your seal manufacturer can supply accurate information. When the total head differential is too great for proper flow with a sin-

gle orifice, divide the head feet by the number of orifices to be used in series. Separate the orifices by a minimum of 6” when piping in series. The drill size in the seal gland should never be used as a control orifice. This method is unreliable for two reasons: 1. Turbulence can occur across the orifice, and because it is adjacent to the seal faces, can result in flashing at the faces. 2. Seal glands are typically made from soft materials and can easily erode at these locations. Seal Coolers These devices are one of a number of options used where vapor pressure vs. stuffing box pressure is of particular concern. Typically, a product recirculation flush is circulated through a seal cooler into the seal gland. This method can result in an energy loss due to cooling of the pumpage and involves the maintenance of the seal cooler. When a seal cooler is used to maintain the stuffing box below vapor pressure, a temperature indicator should always be included downstream of the cooler. This will enable accurate monitoring of temperatures and can alert operators to a potential problem before a seal failure occurs, allowing easy cleaning of fouled components. It is very important that seal coolers are cleaned at every seal change, or any time seal flush temperature exceeds allowable calculated limits supplied by the seal manufacturer, regardless of condition. Close Clearance Throat Bushings Bushings of this design are another effective means for controlling flow in the stuffing box. When properly sized and applied, close clearance bushings enable discharge pressure to be introduced into the stuffing box. This can be useful in maintaining vapor suppression if the seal design is capable of handling the increased pressures. Control is

Figure 4. Throat bushing sizing table (Courtesy Flowserve Corp.)

Figure 5. Choke tubing sizing The Pump Handbook Series

101

when designing the seal flush system. Additionally, increasing the stuffing box pressure significantly may require that you consider alternate seal designs.

Figure 6. API Plan 11 flush arrangement (Courtesy Flowserve Corp.)

maintained by the clearance between the bushing and the shaft, which is typically one-half normal throat bushing clearance. It is very important to remember that throat bushings are wear devices; as they wear, flush flow rates can increase dramatically. These increased flow rates can affect cooling efficiency and effectiveness. Seal design and stuffing box space may dictate the available bushing length, which can be a limiting factor. Figure 4 shows the throttle loss of feet (meters) of head per inch (meter) of sleeve length. Three examples using Figure 4 follow.

.5” long and all other conditions remained the same. Flow would be 3.2 gpm. As shown in Examples 2 and 3, the bushing clearance and/or length is critical to maintaining proper flows. Again, these flows must be corrected for viscosity. Figure 4 and the examples clearly define the problems associated with using throttle bushings as flow control devices. In general, if the throttle bushing length is less than 1.00” per 100’ TDH, an alternate method of flow control should be considered

Choke Tubes These also control seal flush rate and consist of a long length of small gauge tubing. The tubing provides the necessary head loss to arrive at the required flush flow rate. Figure 5 shows the necessary data to size the length of these lines for .25” OD tubing having an ID of .180” and .25” tubing with an ID of .364”. If there is no risk of losing flow due to high viscosities, this method provides the added benefit that some skin cooling will occur. Where a borderline cooling requirement exists, choke tubes have proven beneficial. The .25” OD tubing is normally wound around a large diameter mandrel to provide a compact piping arrangement.

API Flush Plans API Plan 11 (Figure 6) This plan is a product seal flush that originates at the discharge or a pressure point on the pump that is higher than stuffing box pressure. It is piped to the flush port of the seal flange or stuffing box tap. Plan 11

Example 1 • bushing length: 2” • shaft sleeve diameter: 2”: • total dynamic head across bushing: 100 TDH • clearance: .008” Therefore, flow would be .8 gpm. Example 2 Assume that the bushing wore to a clearance of .014” with all other conditions being the same. Flow would now be 4.7 gpm. Example 3 Assume that the bushing was only

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Figure 7. API Plan 13 flush arrangement (Courtesy Flowserve Corp.) The Pump Handbook Series

Figure 8. API Plan 21 flush arrangement (Courtesy Flowserve Corp.)

flushes are designed to lubricate and provide cooling for the seal. They can also be used to raise stuffing box pressure in applications where required vapor suppression margin is not present (i.e. suction pressure at or near vapor pressure of the liquid). The typical failure mode with Plan 11 flushes is blocked or plugged piping. For this reason the piping arrangement should be free of valves if possible to avoid inadvertent closing. Orifices can also become plugged and should be inspected each time piping is dismantled for maintenance. Checking surface temperatures with a temperature gauge is an effective means of monitoring Plan 11 flush performance. In hot services, if the line is plugged or obstructed, you will see a significant temperature change. The piping may be hot near the seal flange and at the origin, but temperatures will drop as you move away from these points. In cold services, the piping will be hotter at the seal flange flush port connection. Piping temperatures will cool as you move toward the seal flush origin point. The heat generated at the seal flange is generated by the seal and is an indication of no flow in the flush. Check for blockage or obstructions in the piping and connections.

API Plan 13 (Figure 7) This arrangement is a product seal flush that originates at the seal injection port of the seal flange and is piped to the pump suction or some point that has a lower pressure than the seal stuffing box. It is designed to lubricate and cool the seal and provides a vent to the seal chamber due to the reverse flow in the stuffing box. Plan 13 flushes can also be regulated to reduce stuffing box pressures when required.

API Plan 21 (Figure 8) This system is a product flush that originates at the pump discharge, or a point in the pump that is higher than stuffing box pressure, and piped through a heat exchanger to the flush port of the seal flange or stuffing box injection tap. In addition to lubricating and cooling the seal, the heat exchanger provides heat reduction when conditions in the stuffing box are at or near the vapor pressure of the pumped liquid. The goal is to reduce the flush temperature 100-150°F before it enters the stuffing box. In rare cases an exchanger can be used to heat the liquid if required. The exchanger should be equipped with a case drain to allow operators and maintenance personnel to drain any debris and to vent the exchanger’s case side. Attention should also be given to the exchanger’s water side. Draining debris and ensuring there is flow from the inlet to the return water supply should be a part of routine operator rounds. API Plan 23 (Figure 9) This flush is a closed loop product flush that circulates product from the seal flange through an exchanger.

Figure 9. API Plan 23 flush arrangement (Courtesy Flowserve Corp.) The Pump Handbook Series

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ing box. It is used where an external flush medium source is readily available and process conditions are very dirty. Plan 32 flushes are also used in very hot services when vapor temperature suppression is needed to avoid flashing at the seal faces. Plan 32 flushes must be checked regularly during operator rounds, so an accurate and reliable pressure indicator is required. As noted previously, using an external flush medium can result in additional processing expense.

Figure 10. API Plan 31 flush arrangement Courtesy Flowserve Corp.)

To develop head for circulation, a pumping ring is usually mounted on the seal. Some thermal siphoning will also occur. The primary purpose of a Plan 23 flush is to reduce the temperature of the flush while maintaining constant lubrication and cooling. It is often used when stuffing box conditions are at or near the vapor pressure of the pumped product. The exchange should reduce the flush temperature 150-200°F. The most common failure mode in this flush is an unvented system. To prevent this, operators should regularly vent the case or product side during routine operation checks. API Plan 31 (Figure 10) The Plan 31 flush uses product that originates at the pump discharge or a point on the pump that is higher than stuffing box pressure. It is piped through a cyclone separator that delivers clean process liquid to the seal injection port while particulates or solids cleaned from the flush are returned to the pump suction. When installed properly, the vortexing action creates a centrifuge effect that separates solids and drops them from the flush flow. The clean flush exits the top of the separator and is piped to the seal injection port. Because continuous flow is crucial to proper operation

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of the cyclone separator, no valves or orifices should be present in the flush piping. Another crucial design consideration is that the system should have no low points in the return line from the cyclone separator. Also ensure that it continually slopes to the pump suction. Low points can be gathering places for solids (similar to a “P” trap found in plumbing systems) and can result in a plugged line that will stop the separator’s cleaning action. Typical failures of Plan 31 flushes are caused by plugged separators or drain lines. Solids and erosion can cause grooving in the cyclone over time, which reduces the centrifuging effect. It is best to have separators with removable tops to enable inspection of the cyclone for grooving. If the top is not removable and a visual inspection impossible, the separator should be changed at each seal repair. Return lines from the cyclone must also be checked regularly to ensure continued flow. API Plan 32 (Figure 11) This seal flush uses an external source at a higher pressure than the stuffing box and is piped to the seal flush port or stuffing box tap. This plan provides clean flush medium to lubricate and cool the seal and maintain vapor suppression in the stuffThe Pump Handbook Series

API Plan 52 (Figure 12) This plan provides clean flush to secondary seals typically found in tandem seal arrangements where the primary (or inner) seal is flushed by another flush plan. The Plan 52 flush vents routine leakage from the primary seal to a flare or recovery system and provides lubrication and cooling to the secondary seal. These seals and flush systems are used to control leakage to the atmosphere for compliance with environmental regulations. They can be found in a large variety of services, most often where product vapor pressure is more than 14.7 psia at ambient temperature. The primary seals in all pumping applications leak an extremely small amount (< 500 ppm) of product across the seal faces during normal operation. A secondary seal is used to contain the leakage and allows it to mix with the barrier fluid. Over time this mixing slowly contaminates the system. To prevent the leakage from building pressure in the secondary seal chamber, a line is run from the barrier reservoir to the flare or recovery system. This enables disposal of the primary leakage in a safe and orderly manner. Seals with Plan 52 flushes often use pumping rings mounted on the secondary seal to provide head to circulate the flush. Thermal siphoning effects also provide flush circulation. Because the head developed for circulation is very low (< 10 ft TDH) careful consideration should be given

to piping arrangements. Use long radius piping sized to avoid head losses. Reservoirs should have an armored sightglass for operators to monitor fluid levels; bull’s-eye level indicators are not recommended. Because primary seal leakage contaminates the secondary seal flush, the seal reservoir should be equipped with a full diameter removable cover to enable easy cleaning and inspection. Reservoirs must be cleaned and the flush medium changed at each seal change. An important consideration for Plan 52 flushes is that you know the pressure gauge reading during normal operation conditions. This is the base or reference pressure level. Pressures above this reading should be monitored, since they might be an indication of primary seal leakage. The barrier fluid for the reservoir should also have good lubrication qualities and be compatible with the service. Some users add dye to the reservoir to make it easy to see, particularly in low light conditions. Plan 52 flushes are operatorintensive and require close monitoring. The flush normally maintains steady pressures; however, small fluctuations can occur due to flare header or process changes. These are usually acceptable. If a pressure increase occurs, check the fluid level in the reservoir to ensure that it is constant and stable. An increase in fluid level indicates above normal leakage rates in the primary seal. Normal primary seal leakage can cause fluctuations in levels if the leaked product vaporizes in the reservoir. Some products do not flash off in the reservoir and will cause a slight increase in levels over time. Flare header lines and recovery systems should always be checked to ensure that they are not blocked and are operating properly. Icing or sweating of lines can be an indication of leakage problems. Barrier fluids in Plan 52 flushes should be changed on a regular basis;

a six-month interval (as a minimum) is recommended. In some services, fugitive emission testing can cause a tandem seal application with a Plan 52 flush to be tagged for excessive leakage due to saturation of the barrier fluid over time. This can sometimes be remedied by simply changing the fluid. This situation highlights the need for routine preventive maintenance and fluid changes. API Plan 61 (Figure 13) These flushes are used in tandem seals that have dry running sec-

ondary seals and are found in services that require the primary (or inner) seal leakage to be contained and vented to a flare or recovery system. Plan 61 flushes can avoid the use of the barrier reservoir system that is used with Plans 52 and 53. With these seals, the primary seal is flushed with a standard flush plan. Because the primary seal leaks (as noted in the Plan 52 description), leakage must be contained and disposed of in an orderly manner to avoid environmental and safety concerns.

Figure 11. API Plan 32 flush arrangement (Courtesy Flowserve Corp.)

Figure 12. API Plan 52 flush arrangement (Courtesy Flowserve Corp.) The Pump Handbook Series

105

The Plan 61 arrangement will operate at the pressure of the flare or recovery system. An increase in pressure can indicate primary seal failure. Operators should make sure the valves do not become inadvertently closed or the system blocked. A check valve is required to avoid reverse flow into the seal chamber. API Plan 62 (Figure 14) This plan is designed to quench the atmospheric side of a mechanical seal through the quench port on

the seal flange and keep the seal free of product that has solidified after it has leaked across the seal from the product side. The most common sources of quench are steam, water and nitrogen. Heavy, hot hydrocarbons tend to coke or solidify at atmosphere pressures; caustic applications form crystals when their carriers evaporate. Typically steam or nitrogen will be injected into the quench port to reduce or eliminate oxygen and create a positive flow under the seal’s

Figure 13. API Plan 61 flush arrangement (Courtesy Flowserve Corp.)

Figure 14. API Plan 62 flush arrangement (Courtesy Flowserve Corp.)

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atmospheric side. This flow blows any product leakage from under the seal and prevents seal hang-up; the elimination of oxygen slows the coking process. When steam is used as the quench medium, care should be given to ensure that no water enters the quench port. In high temperature applications, water saturation can result in popping of the seal faces due to flashing of the water. In addition, grafoil gaskets can become damaged should they become water saturated and exposed to sudden high temperatures. The water will flash and expand, damaging the gasket. To prevent this from occurring, piping arrangements should never have the seal as the low point, and a low point drain with a steam trap should be provided. Piping in Plan 62 flushes must always rise to the seal from the low point drain. Nitrogen quench is also used in hydrocarbon services. These should have a reliable flowmeter to regulate the flow, because nitrogen consumption can be considerable and the cost may be difficult to justify. Water quench is usually found in caustic or similar services where evaporation of the product can cause crystals to form and damage seal faces and lock up the seal. Water quenches can become a housekeeping issue because they can result in a slow drip from the seal flange, much the same as packing. Plan 62 flushes should be checked by operators during daily rounds. Caution is critical: Never open a quench line using steam in high temperature (>400°F) services after the pump has been prepared for service or while it is in operation to avoid damage to the seal and protect the operator. Avoid over-pressuring steam quenches. This can result in condensation entering the bearing housing, contamination of the lubrication and damage to the bearings. Accurate and reliable pressure gauges are a must for operators to properly maintain correct quench pressures.

Failure usually results from valves becoming closed or plugging caused by debris. Should it become evident that the quench is not operating properly, shut the pump down in an orderly manner and establish quench flow before placing it back in service.

ensure that buffer gases are properly filtered. A coalescing filter with a rating of .1 micron and sized for a minimum flow capacity two to three times normal flow rate is recommended. Flowmeters should also be used to ensure adequate buffer flow, typically 10–50 scfm.

Optional Flush Plans Plans for non-contacting seal faces use a buffer gas to purge the seal faces. These seals use spiral or tee grooves etched in the seal faces, which develop head, form a dynamic barrier and pump product away from the faces. Face spacing when operating is usually 3-5 microns (micrometers). Should particulates or any solids migrate to the seal face area, damage can occur. A clean buffer gas such as nitrogen is injected into the seal injection port to purge the seal faces and provide cooling. Care should be given to

Summary In a typical process plant, centrifugal pump repairs can account for up to 80% of the rotating equipment maintenance costs. Proper design, operation and maintenance of seal flushes can have a major economic impact while reducing environmental and safety concerns.

Acknowledgments The author would like to acknowledge the help of Paul McMahan and Stuart Shoefstall of Flowserve Corporation for their guidance and input in this article. ■

The Pump Handbook Series

Mike Polk is Founder and Vice President of Reliability Associates, a rotating machinery consulting service company in Slidell, Louisiana. He retired from the Mobil Oil (formerly Tenneco Oil) Chalmette Refinery, in 1996 after 16 years as Staff Reliability Specialist and Machine Shop Supervisor. Mr. Polk has more than 30 years of experience with rotating equipment, including broad experience in the installation and repair of rotating machinery in oil production facilities, petrochemical plants and other industries. The author of several technical papers, he is affiliated with the Vibration Institute (past officer) and the Gulf South Compression Conference. He has a technical degree from Nashville Auto-Diesel College.

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MECHANICAL SEALS & SEALING SYSTEMS HANDBOOK

Estimating Heat Generation, Face Temperature and Flush Rate for Mechanical Seals By Gordon S. Buck, John Crane Inc.

umerous studies have shown that mechanical seal reliability is related to the state of the fluid between the seal faces. In turn, the fluid state depends on the seal face temperature. There is a strong interdependence between heat generation, seal face temperature and flush rate for mechanical seals that can be approximated by simple calculations. The effects of seal design, materials and fluid properties can also be illustrated using this model. Mechanical seals are widely used in centrifugal pumps to minimize or prevent leakage around the pump shaft. An example of a single cartridge seal meeting the specifications of API Standard 682 is shown in Figure 1. Although mechanical seals come in various types, configurations and arrangements, in this study, a single seal arrangement roughly similar to that shown in Figure 1 is assumed.

N

Heat Generation Mechanical seals generate heat as the primary ring and mating ring rub together. Heat is also generated through viscous shear of the fluid film between the faces and around the rotating components. For low viscosity fluids, most of the heat generated is from rubbing. In this

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case, the coefficient of rubbing friction for the primary ring and mating ring pair determines the heat generation. For viscous fluids, there may be no rubbing, but viscous shearing may generate about the same amount of heat as rubbing. An important distinction is that the wear rate is likely to be much greater when low viscosity fluids are between the seal faces. The heat generated by rubbing is: Hm = PmV Af f

(Eq. 1)

of friction. Unfortunately, there are many definitions and derivations of this coefficient. No matter what the source, definition or derivation, the coefficient of friction is not a constant. It ranges from around .03 to .3 and is frequently found to be around 0.1 for many applications. Naturally, the coefficient of friction is a function of the tribological material pair, but it also depends on the fluid being sealed. To make matters worse, it turns out that the coefficient also depends on the seal face load. Furthermore, the coefficient is reduced when the seal leaks. In spite of these limitations, the coefficient of friction is a useful means of comparing seal face materials, especially when tests are repeated under similar conditions.

For purposes of this article, the effects of viscous shear will be neglected; however, viscous shear can be significant. On the other hand, the normal sequence of events is that rubbing contact decreases as viscous shear increases. For this reason, Equation 1 can give good approximations even for viscous shear, especially if a representative value is used for the coefficient of friction. Mechanical seal calculations are considerably simplified by using a coefficient Figure 1. Cartridge mechanical seal for API 682 The Pump Handbook Series

Heat Transfer

Table 1. Coefficient of friction

Great care must be used when working with coefficient of friction data, especially with data from different sources. The coefficient is not measured directly; instead, it is computed based on measurements of heat generation or torque. A certain load, or basis of load, is assumed, and the coefficient of friction is computed. The tests might use mechanical seals or other friction devices such as a pin on disc tester. Some tests use pressurized seals; other tests might be in an unpressurized bath of fluid. Since the heat generation or torque is a result of both mechanical contact and viscous shear, the computed coefficient of friction includes both these effects. Table 1 shows the coefficient of friction for various face combinations. In mechanical seal engineering calculations, the product of pressure and speed, usually shown as PV, is widely used as a guideline for seal design and application. The standard calculation is described in Schoenherr (Ref. 1) as PV = [P (b - k) + Psp]V

Heat transfer plays an important role in seal performance. Both conduction and convection are significant in mechanical seals. Conduction is the process of heat transfer through solids. Convection is the transfer of heat from the solid to the surrounding liquid. Figure 3 illustrates the combined processes of conduction and convection. Because heat generation takes place in the sealing interface, the heat transfer process is first the process of conduction through the primary ring and mating ring. The thermal conductivity of these materials is very important. Materials like silicon carbide and tungsten carbide have relatively high thermal conductivities. Materials like alumina (ceramic) and carbon graphite have much lower thermal conductivities. See Table 2 for examples of thermal conductivity.

(Eq. 2)

The PV value can be important because it represents both wear and heat generation. Even so, it has only a very rough correlation with overall seal performance. In Equation 2, the pressure gradient factor, k, is taken as 1⁄2 according to Schoenherr. In actual service, the value for k may vary from 0 to 1. Nevertheless, it is convenient and the usual practice is to use k = 1⁄2 when dealing with PV calculations. Industry practice has been to limit the PV value based on seal face material combinations and to attempt to couple this limit with fluid properties. A typical limit is 500 kpsi ft/min for resin-filled carbons vs. silicon carbide for seals used with non-lubricating liquids in typical pump services. At another extreme, the same carbon used with alumina might be limited to 100 kpsi ft/min. Normal practice is to allow higher PV values for “lubricating” liquids. Equation 1 contains the quantity PmV to emphasize that the pressure term is the rubbing contact pressure; however, for purposes of this article, Equation 2 will be used to compute the PV product for use in Equation 1.

Figure 2. Heat generation is caused by rubbing and viscous shear

Figure 3. Conduction and convection heat transfer in a seal

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109

plex computations. Table 3 gives some representative values of the convective heat transfer coefficient. Equation 5 is correct for either the primary ring or the mating ring of the seal; however, it is not correct for the combination of the two. The complete heat transfer consideration is Table 2. Typical properties of materials

Convection takes place from the primary ring and mating ring to the surrounding fluid. Convection is usually evaluated as having three components: a convection (or film) coefficient, wetted area and temperature difference. The convection coefficient represents the combined effects of fluid properties, rotational speed, seal chamber design, flushing design and flush rate. Low viscosities and high speeds promote high convection coefficients. It is very important to have enough wetted area that heat transfer takes place without high temperature increases in the mating and primary rings. The heat transfer process can be represented mathematically. Because the sealing elements are physically small, there is a certain temptation to consider each element as a lumped mass—that is, at a uniform temperature. This approximation leads to very low estimates of the temperature at the seal faces and is not recommended. A more realistic approximation than the lumped mass approach considers the primary ring and the mating ring as separate elements. The individual elements can then be analyzed as one-dimensional fins. The fin is considered to have a heat flux at the base (seal face) and an insulated end. Because this is a onedimensional analysis, the temperature is considered to vary only along the length of the seal. This approach is illustrated in Figure 4 and recommended in Ref. 7. The seal face temperature that is computed with this model gives a rough average temperature.

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H = D hw ( The heat transfer rate through such a fin—using the classical arrangement of the variables from Ref. 8—is shown below. H=

hCλ A T tanh (mL) (Eq. 3)

For a seal shape, mL can be arranged as mL =

L w

hw λ (Eq. 4)

For a seal, the lumped mass approach would only be valid at low values of mL. For example, this would be true with extremely high thermal conductivity and/or very low convection coefficients. Making use of Equation 4 in Equation 3 yields H = D

hwλ (T - T∞) (Eq. 5)

with the qualification that L w

hw λ

≥2

Equation 5 is deceptively simple because it relies on the convection heat transfer coefficient. In practice, it is difficult to accurately predict this coefficient. Laboratory tests have shown that the values change not only with fluid, shaft size and speed but also with flush method, flush rate, seal chamber shape and many other parameters. Even simple estimates of the convective heat transfer coefficient can involve comThe Pump Handbook Series

λ1+ λ2)(T - T∞) (Eq. 6)

Although not strictly correct, the values in Table 3 could be extrapolated linearly for diameter and shaft speed. Significantly higher values could be obtained with good flush designs and rates.

Flush Rates For most seal designs, all the heat generated enters the fluid surrounding the seal. This means the fluid tends to increase in temperature. Unless it is cooled or continuously replaced, the fluid may become too hot for reliable seal performance. For this reason, most seals operate with continuous replacement of the surrounding fluid; this replacement process is called the seal flush. Figure 5 illustrates this process. There are many seal flush systems. The American Petroleum Institute has categorized and labeled the common systems; these systems are usually called API Flush Plans. The one most often used is Plan 11, which uses injection from pump discharge to flush the seal. Figure 5 would be an example of a Plan 11 flush if the source of the injection were the pump discharge. Under certain conditions, some seals can actually operate satisfactorily without a flush, but the usual practice is to recommend flushing. Flush requirements for seals are often given in terms of a minimum and a recommended flow rate. Applications without a flush are called dead ended seals. These usually involve nonvolatile fluids at low pressures and low speeds. For these

are allowed. Frequently, the minimum flush rate is relatively low—often less than 1 gpm. If the minimum flush rate cannot be provided, then the seal cannot be rated for that pressure and must be derated. Derating can be in terms of pressure or seal life. The amount of heat transferred to the flush is H = m Cp T

(Eq. 7)

Where m is the mass flow rate of the flush and T is the temperature rise. The minimum flush rate can be determined by combining Equations 1 and 5 and using an appropriate allowable temperature rise. However, the flush rate so determined is for the case of idealized mixing near the seal faces. Ideal mixing occurs in seals incorporating a good flush distribution system. Single point injection or withdrawal might also approach ideal mixing around the seal faces for small seals. On the other hand, the flush to a large low speed seal using single point injection may not be uniformly distributed around its circumference. Another consideration is the direction of the flush fluid flow. Flow toward the seal interface (injection) is more effective than flow away from the seal face. Axial flow paths are not very effective. Most of these effects are due to geometries that prevent flow near the seal interface; therefore, experience is the best guideline and may dictate greater or lesser flow rates. Field experience indicates and laboratory tests confirm that seal performance generally improves when the flush rate is greater than the minimum. In particular, heat transfer usually improves and the average temperature around the seal decreases with increased flush rate; as a result, face temperature and wear rate decrease. The recommended flush rate promotes these benefits.

Figure 4. Heat transfer through a fin

Figure 5. Concept of seal flush

seals, heat is transferred from the faces, through the liquid and into the metal surrounding the seal chamber. The transfer of heat from the liquid into the surrounding metal is called heat soak. The temperature of the liquid in the seal chamber rises until equilibrium is reached between heat generation and heat soak. The point of equilibrium depends on the fluid, pressure, seal design, pump design, etc. In comparison to flushed seals, dead ended seals are limited to low pressures. A certain minimum flush rate is necessary to obtain the performance rating given by the manufacturer. This minimum is usually determined by an energy balance computation. Heat generated between the seal faces is assumed to be absorbed by the flush through ideal mixing. This raises the temperature of the flush. Typically, increases of 15°F for water and low volatility hydrocarbons, 30°F for lube oils and 5°F for volatile hydrocarbons

Table 3. Example of convective heat transfer coefficient for 4” OD cylinder at 3600 rpm

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111

The recommended flush rate should be based on experience with similar applications. Considerations include performance goals and fluid properties as well as the design and interaction of the seal chamber, gland, flush plan and seal. In the absence of specific experience, a simple rule of thumb is the larger of 1 gpm/in of seal size or the minimum flush rate. Questions are sometimes asked about the maximum flush rate. Although increasing the flush rate beyond the recommended value may produce further improvements, by definition this effect is rapidly diminishing. Unless close clearances and high velocities are involved, there are probably no detrimental effects from very high flush rates.

Example Consider the seal shown in Figure 6 in 100°F water service at 100 psig and 3,600 rpm. The balance ratio is 122%; the face area is 1.865 in2. The mean diameter of the face is 2.375 in., so the mean velocity at 3,600 rpm is 2,238 ft/min. Assume that the unit spring pressure is 26 psi. From Equation 2, the PV value is PV = [100(1.22 - 0.5) + 26]2238 = 219,324 psi ft/min (Eq. 8) Then in Equation 1, the heat load is 219324(1.865).18 60 = 5678 Btu/hr 778 (Eq. 9) Equation 7 can be used to estimate the minimum flush rate by allowing a 15°F temperature rise in the water: Hm =

m=

5678 Btu/hr = 378 lbm/hr = 0.76 gpm (1 Btu/lbm F)(15 F) (Eq. 10)

If the seal were flushed with 0.76 gpm of water at 92.5°F, then the exit temperature of the water would be 107.5°F, and the average operating temperature of the water surrounding the seal would be 100°F. The next step is to calculate the average face temperature. The thermal conductivities of the carbon and alumina were given as 5 and 15 Btu/hr ft °F respectively. Table 3 gives the convective heat transfer coefficient for a 4” cylinder in water as 4,800 Btu/hr ft °F; for the 2.375” mean diameter of this seal, an estimated value of 2,850 will be used. The temperature difference between the seal face and the flush is then (T - T∞) =

5678 2.375  12

= 149 F .25 4800 12

(

5+

15) (Eq. 11)

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At this point, some judgments can be made about the performance of the seal. First, the computed average face temperature is 249°F. This means that the water between the seal faces will be flashing to steam toward the ID of the seal face where the pressure approaches atmospheric. Also, whereas cool water is not a particularly good lubricant, hot water is even worse. This seal might not work well in this service. What can be done to improve the situation? Suppose that the flush rate is increased to 2 gpm. With the increased flush rate, the temperature rise in the flush becomes 5.7°F, and the average temperature around the seal becomes (92.5 + 5.7/2) = 95.3°F. So the increased flush rate only decreases the seal face temperature to (95.3 + 149) = 244°F. The seal faces are still too hot. As mentioned previously, experience indicates that the maximum allowable “PV” for carbon vs. alumina is 100 kpsi ft/min. The PV for this example is 219 kpsi ft/min. Also noted previously was that silicon carbide could be used up to 500 kpsi ft/min. Suppose that the alumina mating ring were replaced with a silicon carbide ring. Table 1 shows that the coefficient of friction for carbon vs. silicon carbide is 0.10—much less than for alumina. Table 3 gives the thermal conductivity of silicon carbide as 80 Btu/hr ft°F. Because the heat generation is directly proportional to the coefficient of friction, the heat generated with carbon vs. silicon carbide faces is easily computed to be (.1/.18)5,678 = 3,154 Btu/hr. Similarly, the new minimum flush rate is (3,154/5,678)0.76 = 0.42 gpm. The seal face temperature with carbon vs. silicon carbide is 3154 (T - T∞) = = 45 F 2.375 .25 4800 ( 5 + 80)  12 12 (Eq. 12) Therefore, the expected face temperature using carbon vs. silicon carbide is only 145°F—a considerable improvement over the initial combination of carbon vs. alumina. This computed improvement is due to the reduced coefficient of friction and increased thermal conductivity of silicon carbide over alumina. Seal performance could be improved further by making design changes in the seal components. For example, if a balanced seal with a balance ratio of, say, 75%, were used instead of the unbalanced seal, then the heat generation would be reduced to only 1,641 Btu/hr. The face temperature would be only 123°F.

Summary Using only handbook-type equations, it is possible to get quantitative estimates of seal performance that provide useful guidelines for the application of mechanical

The Pump Handbook Series

seals. Such calculations illustrate the importance of heat generation, heat transfer and flush rate. By working out a few such problems, the engineer can gain new appreciation for specific parameters such as balance ratio, coefficient of friction, thermal conductivity, wetted area and flush rate. ■

Acknowledgments The author wishes to thank John Crane, Inc. for supporting this work. As always, Ralph Gabriel and Jim Netzel were helpful in guiding this work and suggesting improvements.

References 1. Schoenherr, K. S., “Design Terminology for Mechanical End Face Seals,” Society of Automotive Engineers Transactions, Vol. 74, Paper Number 650301 (1966). 2. Schoenherr, K. S., “Life and Wear of Mechanical Seals” American Society of Metals Wear Conference (1969). 3. Buck, G. S., “A Methodology for Design and Application of Mechanical Seals,” American Society of Lubrication Engineers (1979). 4. Lebeck, A. O., Principles and Design of Mechanical Face Seals, New York: Wiley-Interscience (1991). 5. API Standard 682, “Shaft Sealing Systems for Centrifugal and Rotary Pumps,” First Edition, American Petroleum Institute (1994). 6. Gabriel, R P. and Niamathullah, S. K., “Design and Testing of Seals to Meet API 682 Requirements,” Proceedings of the Thirteenth International Pump Users Symposium, Texas A&M University, College Station, Texas (1996). 7. Massaro, A. J., “The Mating Pairs” Concept for Mechanical Face Seals, Society of Tribologists and Lubrication Engineers, 42nd Annual Meeting, May (1987).

Figure 6. Primary and mating rings

Gordon Buck is Chief Engineer of Field Operations for John Crane Inc. He has been with the company for 12 years. In his previous positions he worked at both refineries and chemical plants, as well as in pump and mechanical seal sales. Mr. Buck has more than 34 years of experience and has published a number of technical papers, software programs and books. Nomenclature Af —seal face area Ah—heat transfer area for convection b—geometric balance ratio, dimensionless C—circumference Dm—mean diameter, inch f—coefficient of friction H—heat generation rate for the seal h—average convective heat transfer coefficient k—pressure gradient factor, dimensionless N—shaft speed P—pressure PV—a parameter combining seal face contact pressure and velocity, psi ft/min ∆P—differential pressure, psi Psp—unit spring load on the seal face ∆P—pressure differential = P1 - P2 The Pump Handbook Series

H—heat transfer rate for the seal T—temperature T∞—bulk temperature of the liquid ∆T—differential temperature V—velocity at the mean face diameter, ft/min w—face width Greek Symbols λ—thermal conductivity Subscripts 1—sealing element #1 (the primary ring) 2—sealing element #2 (the mating ring) f—seal face coefficients

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MECHANICAL SEALS & SEALING SYSTEMS HANDBOOK

Mechanical Seal Failure Analysis Don't let mechanical seals be the weak link in your pumping system. Every failure leaves clues. Read the signs and prevent future damage. By Ralph Merullo, Education and Development Manager, Chesterton Global Training, A.W. Chesterton Co.

echanical sealing devices are often the weak link of fluid handling rotating equipment. Proper material specification and environmental controls will lead to long mechanical seal life. Failed mechanical sealing devices often will leave clues that can be used to increase the life of the replacement seal installed in the equipment. The basic principle of a mechanical sealing device combined with the observations of the damage to the seal help us to better understand the causes and therefore the solutions to our mechanical seal failures.

M

Mechanical Seal Basics A mechanical seal is a device that houses two highly polished components (known as faces). One face rotates, the other is stationary. The polished seal faces are sealed to their driving mechanisms by secondary seals, which may include many shapes of elastomeric components, gaskets, spring energized seals or adhesives. The polished seal faces are pressed together by a spring mechanism to provide adequate force to affect a seal. The force acting between the seal faces increases in direct proportion to product pressure. Facts must be harvested from the system and equipment to determine the cause of the failure to avoid being

114

victimized by the "best seal since sliced bread" salesmen. Credible vendors will pursue diligent analysis, which begins with a collection of facts to include: • liquid description/concentration • liquid temperature • shaft speed/size • pressure in the seal chamber • environmental control /seal design Additional information may be required. Common questions include: 1. What is the ratio of the shaft diameter to the shaft length? 2. Is a control valve used upstream/ downstream of the pump? 3. Did a trained technician install the seal? 4. Does the expectation exceed the limits of the design? 5. Does the liquid have any unusual vapor/pressure properties?

Mechanical Seal Examination Examine and document damage to each component of the mechanical seal, but do not attempt to explain the failure until all observations of the seal components have been recorded.

Mechanical Seal Disassembly Prior to disassembling the mechanical seal, obtain a cross-sectional drawing and place it above the surface that will be used to disassemble the seal. Lay out paper towels or kim-wipes under the mechanical seal. A light background helps to highlight small pieces, which may fall from the seal during disassembly. Disassemble the mechanical seal by laying out the components so that the inboard components closest to the impeller are placed to the left of the components farthest from the impeller. This step is critical to identifying the effect of each component on its adjacent components.

Documenting Observations

Figure 1. Seal face fractures

The Pump Handbook Series

Documenting observations is best described by the acronym FEM, which stands for Faces, Elastomers and Secondary Seals, and Metal Components. A mechanical seal is a group of seals all working together. If any one seal

fails in a single mechanical seal, then the entire mechanical seal fails. Mechanical seal faces are highly polished to establish a surface that is flat within .000023" from the highest peak to the lowest valley.

Visible leakage occurs when the peak to valley distance exceeds .000106". If the polished surface is damaged beyond this limit, the seal will drip. A series of drips will form a stream. Shaft rotation causes the stream to cover the driver with a spray until operations or maintenance covers the equipment with some sort of deflecting shroud.

What is a Wear Track? An automobile or truck traveling down a dirt or gravel road will eventually wear two ruts in the road surface. The wear is indicated by the tire's contact with the ground. The narrow

Faces

Photo 1. Scored seal face

Photo 2. Crystallized scoring

Typical observations of damaged mechanical seal faces and their causes are indicated in Photos 1-7. Photos 1-3 show different types of scoring, which occurs when particles become embedded in the soft face. The product then evaporates between the faces, causing poor lubrication. Crystals form between the faces, further damaging seal face flatness by causing scratches. Photo 4 shows heat checking, which results from poor lubrication that causes localized expansion, which is relieved by cracks. Photo 5 illustrates the effect of erosion, which occurs when abrasive particles penetrate the spaces between the seal faces. Chipping is shown in Photo 6. This occurs when product evaporates between the seal faces. Evaporation causes expansion and the seal faces separate. Seal springs push the faces back together, causing chipping. This damage can happen on either the inside or outside diameter. The seal faces in Photo 7 have been pitted. This damage is also caused by poor lubrication and/or low thermal conductivity, which both cause localized heating. Strain occurs from thermal expansion and is released by cracks. Material is wiped away by face contact. Figure 1 shows a seal face that has been fractured. High startup torque in viscous service can cause unprotected ceramics to fracture in this way. Tension due to very high inside diameter pressure can also cause this kind of damage.

Photo 4. Radial cracks/heat checking

Photo 5. Erosion

Photo 6. Chipping on inside diameter

Photo 3. Coked product scoring The Pump Handbook Series

115

Photo 7. Pitted resin-based carbon/silicon carbide

seal face will have a similar effect on the wide seal face. The seal designer attempts to choose materials that will not form a rut, however, discoloration of the wide seal face is often the result. The discoloration is known as the seal’s wear track. The wear track is a record of the contact between the narrow seal face and the wide seal face. Examples of wear track damage are shown in Photos 8-10. In Photo 8, the wear track is wider than the narrow seal face. The shaft is orbiting the true rotating axis at the seal faces or a non-rotating narrow face was installed off-center. The seal faces in Photo 9 are damaged because of poor installation or the shaft orbiting the true axis of rotation. Photo 10 is an example of uneven wear track, which is related to mechanical or hydraulic seal face distortion.

Photo 8. Wear track wider than narrow seal face

Elastomers There are many types of elastomers used in mechanical seals. This article will discuss the use of O-rings, which are widely applied in mechanical seals. An O-ring is a round elastomeric ring that is made with a round cross section. The O-ring is used to form a seal between two adjoining surfaces. These surfaces may or may not be contacting. Very often one of the surfaces is sliding along the axis of shaft rotation (Figure 2).

Photo 9. Mis-centered wear track

Fig 2. O-ring damage

Photo 10. Uneven wear track

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The O-ring forms a seal between the two adjoining surfaces by being placed in a groove that is slightly smaller than its molded shape. The O-ring's materials attempt to expand against the adjoining surfaces, forming a seal. The seal fails when the O-ring loses some or all of its elasticity or some or all of its molded dimension. Typical O-ring failures are shown in Photos 11-14. The O-rings in Photo 11 have suffered compression. Heat or chemical attack destroyed the polymeric bonds or elastomeric filler materials. Photo 12 shows two O-rings that have suffered extrusion or nibbling. The elastomer has been subjected to mechanical damage caused by high pressure or excessive clearances. Photo 13 is of an O-ring that went through explosive decompression. The elastomer absorbed chemical, and the chemical expanded, forming bubbles. The bubbles then blew out the lowpressure side of the elastomer. The damage to the O-ring in Photo 14 is obvious; it's been twisted. This damage was caused by excessive relative motion combined with chemical or thermal attack.

Photo 11. Compression set O-ring

Photo 12. Extruded or nibbed O-ring

Metal Components Metal components in mechanical seals exist to achieve one of three purposes: Support Seal Faces or Elastomers Seal faces are often "shrunk fit" into a metal housing. The housing is usually heated to 500°F to cause slight thermal expansion. The seal face material is pressed into the housing, then the assembly is allowed to cool. The housing shrinks around the seal face, causing a "hoop" stress. The seal face is then polished to the required flatness. The flatness is distorted if the housing is exposed to high temperature. The distortion occurs due to removal of the "hoop" stress caused as the housing cools. This flaw has prompted many manufacturers to offer "monolithic" seal face designs.

Photo 13. Explosive decompression

Photo 14. Twisting The Pump Handbook Series

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Monolithic designs do not use a housing around the seal face; the seal face is slotted to provide a means for driving or preventing rotation. Elastomers are housed in grooves as discussed earlier. The critical element of the groove is the finish. The groove must be smooth (32 RMS finish) to effect a seal. Five pits smaller than the head of a pin clustered on the contact surface caused by chemical attack can cause the elastomer to leak.

Photo 15. Scoring on the outside diameter of the seal face

Photo 16. Corrosion

Photo 17. Flattened set screws

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Cause or Prevent Seal Face Rotation Mechanical seals typically enable a shaft to rotate without allowing a fluid to leak out. This requires one seal face to remain stationary while the other rotates. Typically a drive pin, tab, clip or screw is used to hold the seal face in position and to drive the rotating housing and seal face. Provide Initial Closing Force To Keep the Seal Faces Contacting Mechanical seals are designed to prevent visible leakage from the seal faces. Designers specify a spring mechanism to provide adequate, but not excessive force to the seal faces. Typical spring mechanisms include single coil spring, multiple coil spring, wave spring and metal bellows. Springs typically become damaged or cracked when exposed to excessive axial motion. Improper spring material selection can lead to premature cracking. Locating the spring mechanism away from product to prevent clogging leads to enhanced reliability. Metal components are also vulnerable to damage. Examples are shown in Photos 15-17. The parts in Photo 15 have been scored, either by the rotating or stationary components. The part in Photo 16 has suffered chemical attack, which is often made worse by high temperatures. Photo 17 shows damaged drive screws. The shaft sleeve material was significantly harder than the set screws. The Pump Handbook Series

Explaining the Damage Once the components are examined, the analyst seeks to explain the damage, in conjunction with the operating conditions and liquid properties. Thin or Thick Fluids Mechanical seals generally work well in 20°C water. Liquids that are less dense than water may be poor film formers, causing high frictional heat at the seal faces. Liquids that are much more dense than water may not conduct frictional heat away from the seal's faces. An additional danger of high density fluids is increased torque at the seal faces due to higher viscosities. Fluids That Crystallize Liquids that harden or crystallize can damage the flatness of the seal faces, causing the seal to leak. The solution is to control the liquid temperature around the seal to prevent hardening/crystallization. A second approach is to provide an alternate fluid (other than the sealed liquid) to the seal faces to prevent damage. High Temperatures Elevated temperatures can make some liquids become more aggressive, causing swelling or corrosion of mechanical seal components. Mechanical seal cooling is generally worthwhile after a mechanical seal failure in an aggressive acid, alkali or solvent where temperatures exceed 100°C. Seal Environment Several publications outline strategies to control the environment around a mechanical seal. The most notable are API Standard 682 "Shaft Sealing Systems for Centrifugal and Rotary Pumps" and ANSI/ASME B73.1M "Specification for Horizontal End Suction Centrifugal Pumps for Chemical Process." Both outline environmental control "plans" that stipulate piping arrangements to achieve higher reliability when using

a mechanical seal under specific conditions. Questions that the Analyst Must Answer 1. What caused the damage to the mechanical seal components? 2. To prevent future failures, should the seal materials, design and/or environmental control plan be changed? 3. Are the changes acceptable to the equipment and system currently sealed?

Conclusion Damage to the faces, elastomers and metal components are signs that provide direction. Use the operating conditions and the damage to the seal

components to generate conclusions and strategies to enhance reliability. Presenting a conclusion while searching for the evidence to support it will lead to frustration and a temptation to purchase a solution instead of addressing the true cause of failure. ■

References 1. John C. Dahlheimer. Mechanical Face Seal Handbook, Chilton Company, 1972. 2. R.M. Austin, B.S. Nau, N. Guy and D. Reddy., The Seal Users Handbook, BHRA Fluid Engineering 3. Marco Hanzon, "Mechanical Seal Failure Analysis Handbook", A.W. Chesterton Co., 1997.

The Pump Handbook Series

Ralph Merullo holds degrees in Mechanical and Manufacturing Engineering. Prior to joining Chesterton, Mr. Merullo worked with the U.S. Army as a Program Manager at the Army's research laboratory in Natick, Massachusetts. He has been with Chesterton for ten years. During that time he has coordinated or created and delivered many successful seminars and training programs at both the National and International Sales Meetings as well as at customer sites around the world.

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Materials for Seal Faces A primer on the different composites, metals and other materials that make up your mechanical seals By Gordon S. Buck, Chief Engineer - Field Operations, John Crane Inc. election of the proper seal face materials is essential for the successful operation of the mechanical seal. In fact, it could be argued that selection of materials is the most important decision a seal designer makes. In evaluating materials for seal faces, both the properties of the individual materials and the combination of the tribological pair must be considered. In general, dissimilar materials are used for seal faces. These materials are frequently thought of as the "soft face" and the "hard face," although sometimes two "hard faces" are used. Mechanical seal design would be considerably simplified if the "perfect" seal face material could be found. With such a material, the designer would not be concerned about balance ratio, face widths, heat generation, flushing, corrosion, etc. Therefore, a tremendous incentive to develop improved seal face materials exists. Even though a perfect seal face material is not likely, the ideal can be described based on our experiences and problems with existing materials. This ideal material would: 1. Be wear resistant 2. Promote low leakage 3. Promote low friction 4. Have high strength 5. Embody good thermal properties

S

120

In Equations 1 through 3, the 6. Be corrosion resistant 7. Be easy to manufacture into seal nomenclature is: shapes (1-v) 8. Have a low cost R1 = S

aE

Wear Resistance Seal face wear can be divided into five types: 1. Adhesive 2. Abrasive 3. Corrosive 4. Erosive 5. Surface wear from thermal stress or fatigue.

R2 =

R3=

(1-v) Sk aE

(1-v) k S( ) aE pc

For the most part, wear resistance is usually considered to be directly related to hardness. Adhesive wear also is a function of the tribological pair of materials. Corrosive wear is related to the chemical resistance of the material at the operating temperature.

a v p E S k c

Thermal Shock

The first factor, given by Equation 1, is probably the easiest one to visualize. R1 has the units of temperature (°F in this case) and represents the maximum sudden temperature change that can occur without a fracture. The second factor, R2, includes the effect of thermal conductivity and is frequently listed as a property of materials. It has the units of Btu/hr ft and represents the effect of gross heat conduction on the surface temperature.

When materials are subject to large temperature gradients, the resulting stresses can cause cracks or even complete fractures to develop. Material properties that promote resistance to thermal shock are high strength, low expansion, low modulus of elasticity and high thermal conductivity. One way to evaluate the effect of these parameters is to consider a "thermal shock" parameter. There are three methods of evaluating this parameter: The Pump Handbook Series

coefficient of expansion, 1/°F Poisson's ratio Density, lbm/cu ft Modulus of elasticity, psi Flexural strength, psi Thermal conductivity, Btu/hr ft °F Specific heat, Btu/lbm °F

The third factor, R3, includes the ability of the material to absorb energy by virtue of its specific heat and density. It has the units of sq ft °F/hr. For all three of these factors, a larger number represents an improvement in resistance to thermal shock. Because its effect is small, Poisson's ratio is not considered in some references to thermal shock. Table 1 shows why materials such as 316 SS and alumina are not good seal face materials where thermal shock is a major concern. In Table 1, the listing for silicon carbide is representative of both alpha-sintered and reaction-bonded silicon carbide. The exact properties of the various silicon carbides vary considerably from manufacturer to manufacturer.

elasticity. Materials with a low modulus, such as carbon, are more easily made into compliant shapes than materials such as tungsten carbide.

Mechanical seal design would be considerably simplified if the "perfect" seal face

Leakage Leakage is probably more a result of the seal design rather than a property of the material, but good face materials can certainly promote low leakage seal designs. In most seals, the actual face separation is strongly related to the surface finish of the materials. Therefore, materials having and maintaining smooth surfaces generally leak less than those with rough surfaces. Leakage also is related to the compliance, or ability of the seal faces to conform to each other. Compliance is generally thought of as a function of the seal shape; however, it is strongly influenced by the modulus of

material could be found. Friction Mechanical seal calculations are considerably simplified through the use of a coefficient of friction. Unfortunately, this coefficient of friction is not a constant and ranges from around .03 to .3. Naturally, the coefficient of friction is a function of the tribological material pair, but it also depends on

the fluid being sealed. To make matters worse, it turns out that the coefficient of friction also depends on the seal face load and is reduced when the seal leaks. In spite of these limitations, the coefficient of friction is a useful means of comparing seal face materials, especially when tests are done under similar conditions. Table 2 shows coefficients of friction for various face combinations. As shown in Table 2, there is a considerable variation in coefficient of friction for various materials. Even when specific material formulations are tested, the coefficient of friction depends on the fluid being sealed, the seal load and aspects of the seal design such as face distortion.

Strength A good mechanical seal material must not only be strong enough to resist the stresses of normal operation, it also must be strong enough to survive the manufacturing process, storage and the rigors of installation. The strength, hardness and rigidity of carbon graphite-based materials is generally an order of magnitude less than that of metals and ceramics such as steel, tungsten carbide or silicon carbide. This means that more design effort is normally directed toward the component manufactured from carbon graphite. The primary reason for the use of carbon graphite in mechanical seals is this material’s self-lubricating qualities, not its strength.

Material

R1

R2

R3

Cast Iron

300

9700

186

316 SS

250

2360

39

Alumina (85%)

200

1300

35

Tungsten Carbide (6% Co)

860

49800

1070

Silicon Carbide

400

33800

1100

Carbon Graphite (Resin)

640

5100

230

Table 1. Thermal Shock Factors The Pump Handbook Series

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Material #1

Material #2

Carbon

Cast Iron Alumina Stellite Tungsten Carbide Silicon Carbide (R) Silicon Carbide (A) Silicon Carbide and Graphite (Composite) Bronze Tungsten Carbide Silicon Carbide Silicon Carbide Silicon Carbide and Graphite (Composite)

Tungsten Carbide

Silicon Carbide

Coefficient

Range

.12 .16 .16 .10 .06 .07 .05

.07 to .15 .07 to .40 .1 to .3 .05 to .17 .02 to .12 .02 to .12 .02 to .1

.12 .16 .10 .08 .06

.1 to .2 .1 to .4 .05 to .15 .05 to .15 .04 to .10

Table 2. Coefficients Friction (based on net load, in water)

Metal-filled carbon graphites generally show improvements in strength, hardness and modulus of elasticity over resin-filled carbon graphites. Tungsten carbide is at the other extreme from carbon graphite. Tungsten carbide has a very high compressive and tensile strength, is very hard and has a high modulus of elasticity. Silicon carbides are even harder than tungsten carbides but are much more brittle and greater care must be taken during installation and removal. These difficulties in handling have caused many users to prefer tungsten carbide in spite of the low frictional characteristics of silicon carbide.

Thermal Properties The thermal aspects of mechanical seals are a major factor in seal performance and reliability. Two of the major material properties are thermal conductivity and thermal expansion. The thermal shock characteristics of materials have already been

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discussed. Although thermal conductivity enters into the thermal shock parameters R2 and R3 directly, its effect on seal face temperature is probably more important. Carbon graphite materials generally have a thermal conductivity of around 5 to 8 Btu/hr ft °F; metal-filled carbons are somewhat higher. In contrast, tungsten carbides and silicon carbides have thermal conductivities ranging from 40 to 100 Btu/hr ft °F. In a typical seal with carbon vs. tungsten carbide or silicon carbide faces, the majority of the heat transfer takes place through the non-carbon element. Stainless steels, Stellite and alumina have much lower thermal conductivities than tungsten carbide, and seals using these materials will run considerably hotter than one using tungsten carbide or silicon carbide. The thermal expansion of seal face materials is related to both the seal face temperature and the coefficient of expansion of the material. In order to The Pump Handbook Series

minimize the effects of face temperature on distortion, a low coefficient of expansion is desired. The coefficient of expansion of carbon graphites, tungsten carbides and silicon carbides is similar. This is fortunate and enables some degree of substitution in seal face materials within the same design family. Alumina is higher and stainless steels still higher. Any differences in coefficient of expansion become especially important when a seal is manufactured by shrink-fitting components made from different materials. In this case, if the operating temperature is sufficiently different from the manufacturing temperature, the seal faces may become distorted. In an extreme case, the components may become loose.

Corrosion Resistance Corrosion of carbon graphites is usually more related to the binder than the carbon graphite. Metal-filled carbons are especially susceptible to corrosion,

but a suitable resin-filled carbon can usually be found for most services. Carbon graphites are not recommended for aqua regia, oleum or perchloric acid. Resins in common use are attacked by lithium hydroxide, potassium hydroxide, sodium metophosphate, anhydrous ammonia, sodium diphosphate and sodium cyanide.

generally better in corrosion resistance than nickel-bonded tungsten carbide. The "free silicon" in reaction-bonded silicon carbide can be attacked by strong oxidizing chemicals. Alpha sintered silicon carbide has no free silicon; it is considered to be the most corrosionresistant of all the seal face materials.

Ease of Manufacture

The cost of seal components is generally related to the hardness and chemical resistance of the material.

Alumina has good corrosion resistance, and high purity alumina is very good. Before the introduction of silicon carbide, alumina was the preferred corrosion resistant material in many mechanical seal services. The two most common variations of tungsten carbide are cobalt-bound and nickel-bound. Nickel-bound tungsten carbide is the more corrosion resistant, although the cobalt-bound tungsten carbide is more than adequate for most services. Neither is as good as alumina. The chemical resistance of silicon carbide is excellent. The two most common variations of silicon carbide are reaction-bonded and alpha-sintered. Of the two, the alpha-sintered is the more corrosion resistant, but even reaction-bonded silicon carbide is very resistant to chemical attack. Both are

Many of the desirable material qualities for a seal face are not so desirable during the manufacturing process of that component. In particular, the hardness and high strength of many materials make manufacturing very difficult. A common approach is to mold the "green" material into a near finished shape before completing the manufacturing process. Carbon graphites are typically molded to a rough shape before being impregnated with resin or metal binder. Some simple shapes with small cross sections may be machined from cylindrical stock. The final shape is machined. Faces are always lapped. Because of their hardness, tungsten carbide and silicon carbide pieces always must be ground to their final dimensions. Faces are always lapped. Seal components made of very hard materials such as tungsten carbide and silicon carbide are frequently repairable. The repair process consists of chemical and mechanical cleaning and relapping. Caution must be used to ensure that dimensional tolerances are maintained. Softer materials, such as carbon graphites, are frequently not reused, especially if they have been in service for an extended period of time. These softer components generally have more extensive face damage than the hard component and are also less expensive to replace. In the case of carbon graphites, there also may be a concern about chemical attack of the binder.

Cost

and the labor involved in changing out the seal parts. For this reason, most seal users prefer to utilize the best available materials in their mechanical seals. Currently, the most popular material combination is a premium resin filled carbon graphite versus silicon carbide. In relative terms, material costs can be arranged as follows: • Machinable carbon graphite • Carbon graphite (impregnated after molding) • Cast iron • Leaded bronze • Ni-resist • Tool steel • Alumina • Dairy bronze (nickel bronze) • Stellite • Tungsten carbide (cobalt binder) • Tungsten carbide (nickel binder) • Silicon carbide (reaction bonded) • Silicon carbide (alpha sintered) • Silicon carbide/graphite composite The order of the preceding list also is influenced by the shape of the component part and the stocking and pricing policy of the manufacturer. In many cases, there is very little difference in the price of the various tungsten carbides and silicon carbides. The additional cost of tungsten carbides and silicon carbides is somewhat offset by the fact that components made from these materials frequently can be repaired. ■ Gordon S. Buck is Chief Engineer of Field Operations for John Crane Inc. He has been with the company for 13 years. In his previous positions he worked in both refineries and chemical plants, as well as in pump and mechanical seal sales. Mr. Buck has more than 35 years of experience in this industry, and has published a number of technical papers, software programs and books. Contact him at [email protected].

The cost of seal components is generally related to the hardness and chemiThis article was excerpted from the cal resistance of the material. This cost official conference proceedings of is normally a small fraction of the total PumpUsers Expo 2000. cost of removing the pump from service The Pump Handbook Series

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Packing in the 21st Century

It’s come a long way, baby!

Packing is an expendable sealing device. It is the oldest, and still one of the most common seals in the industrial world. Packing does not necessarily stop fluid from leaking, but it controls fluid by stopping or reducing leakage from containment to atmosphere. Modern mechanical compression packing is usually braided or compressed into a square or shaped cross section and used to keep fluids inside pumps, valves, agitators, compressors and other equipment. Some other terms for packing are mechanical packing, compression packing, pump packing, stuffing box packing, or square rope. Packing is used for control of fluid loss, which is essential to the successful operation of equipment in fluid handling. The common places for packing use are where pump and other rotating shafts, reciprocating rods, or valve stems breach equipment containment. Equipment using packing can be found in petrochemical, pharmaceutical, chemical, steel mill, pulp and

paper, wastewater treatment, mining, utilities, and nuclear industries, and many others. It's impossible to list all the fluids sealed; water, steam, caustic, oil, soaps, sewage, diesel, kerosene, alum and various chemicals can all be sealed with packing

The Mechanics of Packing Fluid pressure exists and requires containment where the shaft enters the vessel. A mechanical arrangement called the packing box or stuffing box is attached to the vessel. Packing is inserted between the inside diameter of the box and the outer margin of the shaft. A gland fitted around the shaft exerts force to form the packing. As stud nuts or bolts are tightened against the gland, the gland forces the packing to compress against the bore of the box and conform to the shaft. If the shaft is static, there is usually no leakage but, if the shaft turns or reciprocates, the seal will likely need to leak slightly. Leakage of fluid along the shaft helps cool the shaft as heat-creating

Bob Mathews, President, Uptime Resources Figure 1. Water leaking past the gland and down the sleeve

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PS0501PG13

C M Y K

motion is applied. A drop per second to a drop every three seconds is the ideal atmospheric leakage in most centrifugal pumps. The condition of the shaft sleeve and the stuffing-box bore has to be good to achieve this rate of leakage (Figure 1).

Selecting Packing With the number of manufacturers and with all the types of packing available, selection may seem difficult. Take time to ask all the right questions for your personal safety, plant uptime, packing cost, equipment repair, and environmental concerns. At least, ask the following: 1. On what type equipment is the packing going to be used? 2. What is the fluid to be sealed? 3. What is the pH of the fluid? 4. Is a flush required? What kind of flush? How clean or hot, etc.? 5. What pressure is to be sealed? Is it line, suction, discharge or combination pressure? 6. What is the operating temperature? 7. Is there thermo cycle? 8. What is the shaft/sleeve speed? 9. What is the shaft/sleeve diameter? 10.What is the stuffing box cross section? 11.What is the stuffing box depth? 12.What is the distance to the injection port from stuffing box face? 13.How many rings of packing are needed? 14.In what position do you put the rings? 15.Is a lantern ring required? What material, width, position and clearances are required for the lantern ring? 16.What color do you need? The equipment itself provides important information in selecting the type packing that will be required. End suction centrifugal pumps will require one set of packing. The rpm must be converted to fpm. There will be suction head and discharge head and more information to gather depending on the exact set up.

Figure 2. Calculation of packing cross section

Center-hung, double-suction centrifugal pumps will require two sets of packing. Again, rpm must be converted to fpm. Only suction head is present at the stuffing boxes. The fluid pumped is likely gland water, but there might be abrasives to consider. A knife gate valve is manually operated, so speed is not an issue. Movement is reciprocating, and you will have line pressure and processfluid concerns. A butterfly valve stem rotates only one quarter of a turn. The movement is not a critical factor. The process fluid is the only real consideration. Know the fluid and all the elements involved before choosing a packing. For example, if a valve thermo cycles, you will add Bellville washers to liveload the gland. If the process has fibers or solids, you might want to flush them away from the packing at the throat bushing area. This will keep the contaminants from getting between the sleeve and the packing. pH values are always important to the decision-making process. The decision The Pump Handbook Series

of what type packing to use just can't be made safely without this information. 0 to 6 pH indicates an acid, with 0 being the most acidic. A pH of 7 means the material is neutral. 8 to 14 pH indicates a caustic material, with 14 being the most alkaline. When looking at which packing will perform in a certain pH, remember that the temperature and/or dilution can change the pH. When selecting packing, you will also want to consider packing density. In dynamic applications, about 70% of the wear is on the outer two packing rings nearest the gland follower. Higher density packing helps transfer the gland load to the bottom of the stuffing box. Variations in temperature, dilution, contamination and pressure can all affect the materials and configuration chosen for packing. Find out all the information you can before making your selection.

Types of Packing There are many braided or constructed types of packing to choose from, and when you add in the variety

125

of fibers like TFE, PTFE, Nomex, Kevlar, carbon, graphite and others, the numbers multiply and add to the confusion. Throw in hybrid constructions and hybrid fibers, and the choices increase even more. This all sounds difficult, but stay with it. Packing costs far less than maintenance labor and lost production time. In no time, you can sort out what you need for an application. Most companies that manufacture packing have information sheets to help you select the correct packing for the job. Don't let anyone tell you that “the black packing”, or “the white stuff”, or that “yellow-rope packing” can fix leaks. One style of black packing could have a pH range of 0 to 14 and another black packing could range from 3 to 12. Going by color alone can get you into trouble. Color is only important when fiber-color contamination upsets the process (e.g., black packing is not the packing for the pump or agitator on a paper machine that produces white paper).

Fitting Packing The packing cross section can be determined from the stuffing box and the packing sleeve. Measure the stuffing box ID and the shaft sleeve OD, and subtract. Divide the difference by two to get the cross section (Figure 2). Packing is made in many different sizes. The size needed for an application is most likely on a shelf ready to be shipped. If the size you need isn't available, the packing can be made for your application in metric or inch sizes. How do we make the packing fit in the stuffing box? The answer, too often, is, "With a hammer." No way! This is not acceptable. Never beat packing with a hammer! Packing is made to fit the stuffing box, so use what fits. If the stuffing box is worn out, fix it or replace it. If the shaft is off-center, center it. If the stuffing box bore is pitted or rusted, clean it up, coat it, sleeve it, line it or do something to correct the situation. Don't beat the packing. Hammering the packing breaks the fibers, and causes the packing shape to become out of square. This will cause the gland to cock on the sleeve. Consequently, the more you torque the gland nuts, the greater the chance the gland will contact the sleeve and cut into it, wearing the gland at the same time. When the pump starts, the hammered fibers can wash out and make for a short-lived packing set. Look for the washout at startup; it will be the color of the packing. When looking at wear on shaft sleeves and stuffing box bore try not to exceed 0.045" for cross sections of 1/4"– 9/16", and 0.062" for 5/8" and larger. Replacing or repairing a badly worn sleeve can triple the packing-set service life. If there is no time to properly repair that off-centered shaft or fix the stuffing box, and if the packing shape must be altered, use a rolling pin. Another temporary fix is

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injectable packing. This author has witnessed such packings working well and still running beyond three years in some applications. Evaluate these with your supplier before you try them.

Installing Packing Good installation skills are important to the life of the equipment. Seat each ring correctly in the stuffing box, one at a time. Tools used to install packing vary in size and shape; most are custom-made and user-friendly. PVC pipe pieces, split down the middle, work well. The author uses a piece of 1/4" key stock, 12" long with two bends of about 45° (one 2" from one end and the other 3 1/2" from the other end). This tool works for many sizes of stuffing boxes, and the design can easily be modified it to fit many more. After ring installation, it is always good to inspect the joint with a mirror and flashlight before installing the next ring. The ring joints are to be staggered so they do not line up to create a leak path. Failure to adhere to this principle results in leakage control problems, which greatly reduce packingset and equipment life. Divide the number of rings into 360 to determine the number of degrees to space the ring ends apart. For example, 360 ÷ 2 rings=180°. When using more than 4 rings, a 90° stagger will work just fine. Note that a pump with the packing arrangement 2L3 is not a 5-ring set when considering the ring stagger. It is two sets, one with two rings and another set with three rings. In the 2L3 packing arrangement, the lantern ring divides the packing sets. The two inboard rings are to be staggered 180° and the three outboard rings are to be staggered 120° apart. There may be an occasion when the flush is not needed in the 2L3 packing arrangement. If this is the case, don't add more packing. The normal arrangement will seal with five rings. If there is a seal water inlet and packing is to be used in place of a lantern ring, the inlet should be plugged so the packing will not extrude in the inlet. If packing is allowed to extrude into an opening, or a wear groove, the packing may hang up and, as a result, the inner rings will not be loaded or evenly loaded. This can allow the forward rings to spin in the stuffing box and, if the rings cock in the stuffing box, the gland can also cock. This can cut into the sleeve. The two most common kinds of ring ends used to form the joints of a complete packing ring are called skive- (or angle-cut) and butt-cut (or square-end cut) ends. It is very important to remember that, regardless of the kind of end used on a ring, the joint formed must close well to minimize leakage. Skive-cut ends are preferred whenever they can be used on a particular style of packing, because they form a sealing lap that closes the ring joint under pressure. The type of end used on a particular style of packing depends on whether the ends tend to fray easily before or during installation When making a proper skive cut to a packing ring, first

The Pump Handbook Series

cut a 45°angle to an end of a length of packing. Wrap the packing around the mandrel, making certain not to stretch the packing. Next mark the packing. There are two perpendicular lines on the wrapped packing, one at each end of the 45° angle at the start end of the packing ring. Cut diagonally through the marks with the packing wrapped on the mandrel. Never cut packing on the shaft/sleeve; the knife can burr the coating. The author uses a shaving- sharp knife with a safety guard. Packing cutters can be used to make neat clean cuts, but be certain that the cutter you use is one that produces ends that close up. Butt cuts are easy to cut on mandrels, and are necessary on types of packing that fray or unbraid when skive cut. These packings are usually non-impregnated graphite filament, plastic core, or metallic packing. A proper butt is not simply a cross-sectional cut through the packing. Because the outside of the ring is longer than the inside, a square cross-sectional cut will generate a "V" shaped gap in the ring joint. Kevlar is an exception to the normal cutting methods. The edge of the knife should be file-made and have tiny teeth to cut the stronger-than-steel fibers. The edge will be similar to a small multi tooth or mini hacksaw blade.

Lantern Rings Lantern ring holes and grooves enable the seal water to flow along the shaft sleeve. The lantern ring is too often placed where we think it should go and not where it aligns with the flush port. Keep in mind that the lantern ring will move with gland adjustments. All packing installation procedures need to assure that the flush port position is confirmed for the lantern ring placement. During installation, be sure the lantern ring is free of foreign materials. Fit the lantern ring to the stuffingbox bore and not the shaft sleeve. About 0.030" diametric clearance works for most packing applications. This clearance permits the flush to lubricate

between the packing and the sleeve. When bending loads are placed on centrifugal pump shafts, this clearance keeps the lantern ring from contacting the sleeve. Metal lantern rings will damage the shaft sleeve and Teflon will burn if the speed is very fast. Steel or cast iron lantern rings will rust in the stuffing box and hang up. These can take time to remove during repacking jobs.

Running with Dry Packing Occasionally, you may hear of a pump started with no medium, or with the suction dry. A packed pump, dependent on the product as the coolant/lubricant, can burn up without liquid in the pump. Also, some plants run a line to the lantern ring and out the other side. This should be done with care. Use flowmeters to guarantee that the correct pressure and flow are getting to the packing. Most packings don't need this unnecessary, costly water on the floor, unless the pump is pumping a hot product and heat reduction is required. Insufficient fluid in the packing glazes the inner packing surface, and a leak occurs as operating pressure is reached. We have all burned the end of a nylon string or rope to keep the ends from fraying, and we know that the end gets hard. When most modern manmade packings burn in the stuffing box, they do the same thing, and burned packing doesn't seal very well. Also, please don't breathe the fumes when the packing is burning! Level controls and other signal devices will prevent many of these problems. There is waterless packing that works well enough when applied correctly, but anyone adjusting it must understand it. This packing is tightened, possibly 1/4". The nuts are backed off for a few minutes and then hand tightened during installation. Check the manufacturers instructions for the exact settings. These arrangements will run hotter than flushed packing sets, normally 120° to 180°F. A 1/4 to 1/2 turn too much on an adjustment nut of a boiler feed pump, and the heat can melt the shaft.

The Pump Handbook Series

Break-In Period/Startup It sounds simple, but millions of dollars are wasted each year because packing is not adjusted properly. If packing is not adjusted carefully, burnout can occur, causing production delays.

It sounds simple, but millions of dollars are wasted each year because packing is not adjusted properly.

Lack of technical maintenance during the break-in period accounts for most of the startup problems in packed equipment. Proper selections and installation techniques for packing are extremely important. Naturally, if the packing material does not suit the application or the packing is installed incorrectly, startup problems are inevitable. At startup, and for several hours afterward, packed stuffing boxes must have the attention of qualified personnel. To overcome the deformation and the break-in difficulties of square packing, some millwrights use grease to help out. Grease on the outside of packing, in a smooth stuffing box, will encourage the rings to spin. Grease will also attract dirt and other grit, creating a sandpaper effect on the shaft sleeve. Other lubricants are being used and have similar or worse effects. Never-seize is sometimes used. This compound is lubricant with metal particles. Modern packing is already lubricated, and metal

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Figure 3. Sleeve damage from dry packing (left) and dirty flush (right)

is not needed. The author suggests that it never be used. WD-40 and other spray lubricants are also being applied and could chemically attack the packing and/or its lubricants. Sprays will always wash away some of the original breakin lubricants, increasing the chances of burning at the startup.

Dealing with Problems When the centrifugal pump is operated at its preferred operation zone or pump curve, with all valves open, all forces within the volute are in equilibrium. The only radial load on the pump is the weight of the shaft. With good pump rebuilds, no pipe strain, good alignment, and no invasion of contamination to the bearing housing, the pump, packing, and bearings will operate well for years. When problems arise in pumps, they normally have one of three sources: operations, design, or maintenanceinduced problems. When starting a centrifugal pump with the suction supply valve above the pump, make sure the suction valve and any recir-

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culation or cooling lines are open. Fully close or partially open the discharge valve as dictated by the startup procedures. Watch the pressure gauges. If discharge pressure is not quickly obtained, stop the driver, check the suction, troubleshoot the system, and then attempt to restart. Slowly open the discharge valve until the desired flow is obtained. Monitor vibration levels, bearing temperature, and excessive noise. Vibration can result from lots of causes, including a bent shaft. A bent shaft is physically bent or distorted. Dial indicating can readily reveal this, and bent shafts rarely get out of the shop. A bending shaft, however, will be straight when checked with the dial indicator but bends when operated. Closing the discharge valve on a centrifugal pump will force the impeller away from the cutwater, causing a bending on the shaft. The bending leaves an opening between the shaft and the packing, which can cause a leak in the upper cutwater. When the discharge is opened again, the leak occurs 180° The Pump Handbook Series

from the first leak. Check for these wear patterns during pump rebuilds, and have the operators advise maintenance when these conditions exist. The sleeve must be centered to the bore for the best possible packing seal. The sleeve and stuffing box condition will affect the seal integrity. In Figure 3, the color of the left sleeve clearly indicates that the sleeve reached a very high temperature. The sleeve tells a story; read the story and you understand what happened. This will give you the ability to stop recurrence. The sleeve on the right would have been headed in the same direction as the left one, if it were in the same application. The one on the left was caused by the Kevlar packing fiber being run too dry. The one on the right shows abrasive fiber imbedded in the packing and compressed to the sleeve. The sleeve on the right was most likely in a pump that was flooded with product before the flush was turned on, or the flush water was dirty.

Packing should be sacrificial. Replace it when it is still soft, and the job will be easy. Look for the leak at the shaft for a packing leak. If the leak is around the gland, it is a stuffing box bore problem. The stuffing box doesn't get milled and wear in this area is a big problem, but the packing gets the blame (Figure 1). Don't let this happen. The outboard rings do most of the sealing. The two

closest to the gland take 70% of the compression. Only five rings are needed to seal. However, each additional ring does throttle some pressure. These additional rings are now being replaced with bushings in many applications. A smooth stuffingbox bore is important to performance. With any pumping application, but especially in the case of crystallizing products, it's important to remember that air is a big problem. The pump gets shut down and, as the liquid escapes and air enters the system, product dries and sticks parts together. This condition can cause your packing to stick to the sleeve and spin the packing in the stuffing box, or it can glaze the packing at the sleeve as the pump is restarted. Use packing sets. They eliminate waste, speed installation, reduce misapplication, reduce inventory, and help track cost. Most packing manufacturers can provide you with pre-fitted or die-formed sets, produced in a "grit free" clean facility. Use flowmeters and gauges. They're the only way to know what you are doing. The flowmeter is designed to monitor, measure and control seal flush liquid in a stuffing box. When flow and pressure are optimally adjusted, it saves water and energy consumption. The flowmeter can be used in other areas. For example, quenching is used to smother the leakage, and quench water for the packing gland reduces the heat transmitted from the shaft to the bearings. Seal water for packing controlled at 10 to 15 psi above stuffing box pressure and at 0.5 to 2 gph is common. Always remember that final adjustments should be made only when the seal water system is at its lowest. This will ensure constant seal water to the packing. If seal water is lost and packing is burned, replace the packing.

Scheduling Maintenance/ Installation for Packing

rope fibers, from old, worn ropes. The Romans tapped water from an aquifer and piped it to the fountains of the city. They used bell-and-socket pipe, and they used packing fibers, twisted and held with crimped lead, to seal the pipes. In the early days of ships, sailors would have to pump the ship's bilge. In the process, they were constantly wetted with seawater. They would wrap, cram, or tie old rope fiber and cloth around the pump shaft. Over time, they eventually developed a containment area in which to stuff the materials. This containment area is now called the “stuffing box.” History is interesting, but this is the 21st century. To meet the demands of the modern world, the packing indusThe Real Cost try and modern packing have of Good packing advanced unbelievably. Jump on the Look at your packing cost (not price). train or get left behind. ■ An engineer friend revealed some eyeopening numbers. This may or may not Acknowledgment: The author apprebe your case, but rusted rings are ciates the assistance of Richard Cowen in sometimes not removed and can enable the preparation of this article. excessive leakage to the process. In this engineer’s case, the annual cost to Bob Matthews is the president of remove one gpm of water from the Uptime Resource, a reliability consulting process at a particular stage was $7,600. firm in Baytown, TX. Mr. Matthews has There was about 20 gpm too much more than 30 years of experience in flush water flowing into the process on mechanical applications, 10 of which were some pumps, costing $152,000 per year in hands-on pump repairs. He is a frequent per pump. Problems on too many contributor to Pumps & Systems, and has pumps can waste millions per year. been one of the most popular presenters at This smart engineer, and others like recent PumpUsers Expo technical conferhim, realize that packing is not simple, ences. For information on the applications and not small potatoes. Some large referenced in this article, contact him at companies can save more than [email protected] or by phone at $100,000,000.00 per year in energy, (281) 421-4918. water, and environmental costs with (Editor's Note: This article was upgraded packing programs. Is packing price important? Yes. But adapted from the Official Proceedings you must consider the quality of the of PumpUsers Expo 2000.) packing, the cost of downtime, maintenance, packing life, and all your operational needs. Paying twice the price might be a good deal.

perform a packing task can be a nightmare. The time to pack a pump can be the same time it takes to fill a tooth: 30 minutes to two days, depending on the damage and the accessibility. The author has seen two men work hard for two days to carve one set of packing out of a stuffing box. On the other hand, some workers wash out their packing with water in an instant. Don't get discouraged with time inconsistency involving packing work, even on the same pump at a different time. It is a job worth doing, and worth doing right. Packing should be sacrificial. Replace it when it is still soft, and the job will be easy. Take the time to study your equipment. Develop a predictive/preventive program to replace packings.

Packing: It’s come a long way and has a long

How long does it take to install way to go. The first packing was made of used packing? Scheduling maintenance to

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Built to Last: Working with a New Mechanical-Seal Standard

Attention to components from the outset yields reliability in a refinery’s systems.

James P. Netzel, Chief Engineer, John Crane, Inc. David Redpath, Senior Rotating Machinery Engineer, BP Amoco Oil Neil M. Wallace, Technical Director, John Crane EMA

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In terms of reliability, a pump must be viewed as a series system, where the failure of one of the components means failure of the system as a whole. If the seal, bearing, coupling or shaft is inoperable, then the pump is inoperable and must be repaired. This article describes the startup history of the first refinery built in accordance with the pump seal requirements of API Standard 682 (1994). A sustained, cooperative effort to improve pump seal reliability, and the resulting improvement on pump (and, indeed, plant) reliability is highlighted. The mission of API 682 (1994) is to create a specification for seals that have a good probability of meeting emission regulations and have a life of at least three years. To meet the requirements of API 682 (1994), testing would be done on a simulated refinery pump operation. This would include operating at continuous duty, pump shutoffs, fluid vaporization or low flow and running the seals without a flush. The test conditions were as follows: • Fluid sealed: Propane • Pressure: 250 psig • Temperature: 90ºF • Speed: 3600 rpm Two- and four-inch seal sizes were tested. Test time was 100 hours. The results were excellent and have helped to define seal design, materials and flush arrangements. In the field, seals of the API specification are exceeding the minimum life of three years. For the first time, end-users are beginning to see seal life established, not on some maximum pressure-velocity value, but essentially based on operating conditions. This means that today’s test information

can be used to begin setting life limits on seals based on their operating environment. This test work does not include vibration from poor piping, pump installation and other external causes. Therefore, when the Mean Time Between Failure (MTBF) for a seal or pump is substantially shorter than its estimated MTBF, the installation must be reviewed to completely eliminate other factors that would reduce seal life.

A New Refinery Several years ago, a refinery in Thailand was built to meet the requirements of API Standard 682 (1994). The experience from this refinery demonstrates how operator and vendor, working together, can quickly and effectively resolve problems and improve performance. Every advantage was taken to incorporate the latest technology to reduce operating costs and minimize environmental impact of the new refinery. Where applicable, mechanical seal selections were based on API 682 (1994). Construction was completed in March 1996.

A Focus on Seals During plant commissioning and early operation, mechanical seal “failures” were higher than expected and the company formed a task force to address the problem. The team included representatives from Operations, Maintenance, Integrated Machinery Inspection (IMI) and the seal vendor. The first task was to establish a true picture of the situation. MTBF was found to be around 30 months with 15 “bad actors” identified. An initial target MTBF of five years was set with an ultimate objective

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of eight years’ (12 month rolling sample) “pacesetter” performance. Each month, the team would meet to review every seal replacement of the previous month. Operations’ input at this meeting was significant. Often, they were able to provide details of the pump operation or product handled that influenced what changes, if any, were required. It was only following agreement among the team members at these meetings that changes to materials, configurations or pumps, etc., were made. Through the team, failure modes have been identified and operator training undertaken to improve performance. Plant performance and improvements have been reviewed around three measurement bases: • Seal replacements by failure type • MTBF/MTBR (Mean Time Between Repair) • Repair costs

Cases The greatest improvement has been in the reduction of operations-related failures. In the early days of operation, seal failure due to dry running occurred from: • Incorrect valve operation • Strainers being blocked by debris in the pipe-work • Problems with the circulation flow in coolers (viscous plugging) The flow to coolers was an operational problem, but also related to design. The seals were designed to run with a Plan 23 cooler. This worked well under normal operating conditions even though there was some sludge in the product, but when the pump was on standby, the product left in the cooler became highly viscous. When the pump was restarted, the pumping ring (API Plan 23) did not have sufficient head capacity to drive the viscous plug from the cooler. Consequently, cooling was minimal, resulting in a temperature increase in the seal chamber, vaporization at the seal faces and seal failure. Using Plan 21 taken off first stage discharge and the same cooler, the inlet temperature was increased but is still low enough for the duty. The Plan 21 has enough impetus to drive the plug from the cooler on startup. Note that the

selection of Plan 23 was driven by API 682 (1994) without consideration of the viscosity of the product at cooling water temperatures. (Editor’s Note: Appendix B of API 682 specifies consideration of “High freeze point and viscous products” when selecting Plan 23. A pending revision also notes that Plan 21 “has sufficient delta P to allow good flow rates.” ref. www.api.org/tf682, see appendix B.)

For the first time, end-users are beginning to see seal life established, not on some maximum pressure-velocity value, but essentially based on operating conditions. In another case of operational failure, related in this case to plant design, a dryrunning secondary seal was piped from the top of the gland plate via an orifice plate, and check and isolator valves, to the flare header, located approximately 50 ft above the seal. Normal product leakage caused the seal to become permanently pressurized to approximately 1.3 bar (19 psi). Under these conditions, the seal became hot, and coke was formed, which led to hang-up. In this case, the team removed the secondary seal and replaced it with a floating carbon bushing and a steam quench piped to the drain system. A final example of operations-related failures concerned a double seal leaking barrier oil, smelling of H2S, from the outboard seal. Seal chamber pressures were found to be as designed. Even increasing barrier oil pressure did not stop the oil from being contaminated by the product. Pressure control for the system was being regulated, not by the pressure control valve, but by a pressure relief valve, which meant that the barrier pressure was constantly dropping below that of the seal chamber for milliseconds before the pressure was restored. The compressor to which this seal was installed had a constant supply of seal The Pump Handbook Series

water piped to suction that was at a higher pressure than the seal chamber. By using this water, piped through the seal chamber and orifice to suction, the unit not only was made more reliable, but large savings also resulted from reduced power consumption and elimination of barrier oil. However, not all problems were operational. For example, by training the technicians, a problem of silicon-carbide faces being broken during fitting has been resolved and is no longer a problem. The team also introduced some flexibility into the plant specification. By relaxing strict adherence to API 682 (1994), they were able to introduce PTFE O-rings for some applications where TFE/P copolymer or perfluoroelastomer O-rings had failed. The original perfluoroelastomer O-rings exhibited problems of severe swelling in some seals. The replacement fluoroelastomer is performing satisfactorily. The original O-ring selection was driven by the presence of sulfur and H2S on the data sheet, diverting attention from the otherwise preferred material. One last item is not really a seal problem, but one of process design and miscommunication. Perhaps more than any other, this item highlights the benefit of bringing together expertise from operator and vendor. The seal was a single bellows with carbon versus silicon-carbide faces on API Plan 32. Seal lives could be as short as five to six hours. The supply pressure of the Plan 32 was around 7 to 8 bar (102 to 117 psi) but the pump seal chamber pressure was found to be around 17 bar (250 psi). This resulted in the seal running on slurry rather than the Plan 32 clean injection. The pump impeller was drilled with balance holes to get the seal chamber closer to the suction pressure. A floating carbon bushing was also fitted in the bottom of the seal chamber to slightly increase the pressure and to reduce the usage of the Plan 32 flush. This was moderately successful, with the seal achieving lives of around four months. It was then found that the bronze baffle sleeve under the bellows was becoming worn by radial movement of the bellows assembly, which was attributed to wet steam (in realty, hot water). A steam trap gave little improvement due to the low 131

usage rates. The Plan 62 was changed to a nitrogen quench, which stopped the wear on the baffle sleeve. Wear on the carbon face was still a concern, with lives still being relatively short. When the seal was inspected, traces of catalyst were being found in the seal chamber. To make the seal more tolerant to catalyst, the carbon face was changed to tungsten carbide, which further improved performance and increased service life to over 12 months. It has been noted that there is still a constant catalyst presence in the Plan 32 flush system, and filtration is being installed to remove this. The last seals inspected had a brown glazed deposit on the tungsten face, which could be from the product or the catalyst. The option of using medium-pressure steam for its cleaning and cooling properties is being considered. Modification of the seal faces, from carbon to silicon carbide, is also being considered as the deposits are only found on the tungsten face.

Figure 1. Seal replacement by failure type as percent of annual replacements

Results – Reliability Figure 1 gives a clear insight into the drivers that are now influencing seal replacement. During the first year, operational reasons accounted for over half of seal replacements at the refinery. This figure had fallen to around 10 percent at the last data issue. Conversely, a situation now exists where around 30 percent of seals are replaced before their useful life is complete. This suggests that attention needs to be directed to other parts of the equipment. Operational experience and design improvements have led to a condition where operations-related failures are being supplanted by non-failure replacements. Perhaps the most frequently used measure for plant reliability is MTBF, and this has been plotted for the life of the plant (Figure 2). The initial MTBF figure of 28 years is meaningless and is a function of the calculation method requiring time for stabilization but, clearly, within four months, the value has settled to around three years. While this “meets API objectives,” it is very low when compared with experience from modern plants. Eighteen months after startup, two developments can be seen in the graph. First, the plant MTBF starts to steadily increase, finally exceeding the target (pacesetter) value three years after startup. Second, while replacements “settle” (albeit with some natural fluctuations), failures have diverged and constitute only about 60 to 70 percent of all replacements. This suggests that seal MTBF is no longer the overwhelming influence on pump MTBF, and that other components are starting to affect overall reliability. This differential between replacements (which includes planned maintenance) and failures can be expected to increase as MTBF continues to rise. The curves for MTBF and MTBR from April 1998 show this more clearly (Figure 3). While the general form of the MTBR curve is similar to that for MTBF (naturally, as it is highly influenced by it), there is an increasing divergence between the two. This status reflects that routine maintenance and failure of other components are starting to influence the curves more than actual seal failures, and reflects the success of the program. Previous papers on reliability have demonstrated the variation of MTBF between different units of a plant. This can also be seen 132

Figure 2. Mechanical seal MTBF

Figure 3. MTBF and MTBR curves

in Figure 4. The lowest (12-month) MTBF is found in the continuous catalyst regeneration (CCR) unit, which includes some light hydrocarbon duties (although, as there are a total of only 14 pumps, the results could be distorted by a relatively high number of seals that required change to O-ring materials). Three other units showed MTBFs close to or below four years. The crude distillation unit (CDU) includes high temperature applications, as does the residual fluid cat cracker unit (RFCCU, which also includes light hydrocarbon applications). By contrast, the hydrotreater unit (HTU) covers a wide range of applications and seal types. Units that involve pumping of light hydrocarbons or hightemperature fluids generally exhibit the lowest MTBF figures, despite the fact that these are often given most attention at the specification stage. This gives rise to two questions.

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comes from comparison with other plants. Wallace, et al. (1999 and 2000), reported MTBF that was achieved and (for 10 plants where data were available) the cost per seal installed. Comparing the MTBF for the Thai refinery with those plants showed that, in two years, it had gone from being in the lowest 15 percent to being in the top 6 percent of performers (with evidence of continuing improvements anticipated). Equally significant, and perhaps more important to the refinery, CPSI had gone from being very poor to on-a-par with the best, within the same period. Whereas the annual spending on seal repairs was almost $700,000 during the first year of operation, this had dropped to below $230,000 within two years, an average savings of $234,000 per annum.

Figure 4. Mechanical seal MTBF by plant

Summary

Figure 5. Monthly mechanical seal repair cost (Data courtesy of Star Petroleum Refining Co. Ltd.)

• How bad could they be if they did not receive this attention? This should serve as a reminder not to become complacent just because MTBFs are increasing. • How good can general applications become if they are given the same level of attention?

Results – Costs While MTBF is a measure used extensively through the process industries, cost per seal installed (CPSI) is possibly of greater importance to the plant operator. Duty for duty, it is likely that a dual seal will give higher MTBF than a single seal (though this is not necessarily an automatic fact). Dual seals are, however, considerably more complex than single seals and, therefore, more costly to operate/maintain/repair. Actual seal repair costs fell considerably over the first three years of refinery operation, with average monthly repair costs at less than 25 percent of the first year costs. The monthly cost (Figure 5) reflected both seal repairs and non-failure replacements and, like the repair curve, showed a reducing (but fluctuating) trend. From early 1998, these fluctuations tended to disguise the overall trend for costs, so the graph also includes both a polynomial that “smoothes out” the curve and annual averages for each year. Interestingly, while MTBF exceeded plant pacesetter targets, the CPSI remained approximately 10 percent above target, suggesting that the ongoing cost reduction indicated by the polynomial would be confirmed as more data became available. A good measure of success of an operator/vendor partnership

The Thai plant has been heralded as the first grassroots refinery built to the API 682 standard (1994). While this has undoubtedly helped the plant to quickly achieve world-class levels of efficiency, it is clear that it is not an automatic guarantee of success. Commitment from both operator and vendor, and a close working relationship between the two, have been shown to provide major benefits, repaying the cost of implementing the scheme over and over again. When an operator sees the damage done to a silicon-carbide face as a result of running a pump with a blocked strainer, he can reduce the risk of it happening again. When there are copper particles in a seal on liquified petroleum gas (LPG), the operator can offer a possible source. When a face is broken after two minutes, the maintenance personnel will identify if it had been difficult to install or had to be done very quickly (and, perhaps, that this is how it got broken). They also start to understand operations, and will tell the operators when they pass something that is not right, saving a potential failure. The reliability people will say the seal is out for pump-bearing problems, not seal failure, so the seal vendor does not spend hours looking at seal parts trying to find a cause for a nonexistent failure. By working closely with all three groups, the seal vendor learns more about plant operations and problems, can gather essential data there and then, and can offer more effective solutions for the future. ■ James P. (Jim) Netzel is Chief Engineer at John Crane Inc., in Morton Grove, IL. He joined John Crane in 1963 and has more than 35 years of experience in the design and application of mechanical seals and systems. Mr. Netzel’s accomplishments include five patents on various seal designs, and he has contributed numerous technical papers and articles published through STLE, ASME, BHRA, AISE, and various trade publications. He also has written chapters for the Pump Handbook and the Centrifugal Pump Handbook. Mr. Netzel is a Fellow of the Society of Tribologists and Lubrication Engineers (STLE) and on STLE’s Board of Directors. He is past Chairman of the ASME/STLE International Tribology Conference and past Chairman of the Seals Technical Committee of STLE. Mr. Netzel has a B.S. degree in Mechanical Engineering, from the University of Illinois.

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David Redpath is a Senior Rotating Machinery Engineer for BP Amoco Oil, Refining Technology Group, at Sunbury on Thames, Middlesex, England. He provides technical and reliability improvement support for BP Amoco refineries worldwide. Mr. Redpath has 32 years’ experience in the specification, selection, testing, operation and troubleshooting of rotating equipment in refining and oil production, including 22 years with BP Oil. He is a Chartered Engineer and a member of the Institution of Mechanical Engineers, where he has served as a member of the Fluid Machinery Committee. Mr. Redpath has an Honors degree in Mechanical Engineering from the University of Liverpool. Neil M. Wallace is the Technical Director of John Crane EMA (Europe, Middle East, and Africa). Based in Manchester and Slough, England, he is responsible for technical matters in John Crane EMA and Asia Pacific. He previously worked with Renold Limited and Flexibox International, which he joined in 1974. Mr. Wallace has extensive experience in the field of mechanical seals and power transmission couplings, and has presented many

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technical papers around the world. He is a Fellow of the Institution of Mechanical Engineers and a Chartered Engineer; Chairman of the British Standards Working Group on Mechanical Seals; past Chairman of the Mechanical Seals Division of the European Sealing Association; and a member of the API 682 Task Force. Mr. Wallace has B.Sc. degree from Manchester University (1965).

Proceedings of the Eighth International Process Plant Reliability Conference, Houston, Texas.

Acknowledgements

The authors would like to thank the management of the Alliance Refinery Company PU - North Refinery in Thailand for permission to publish data from their plant on seal reliability. They would also References like to thank Peter Bowden of John Crane API Standard 682, 1994, “Shaft Sealing EMA-AP for his help in developing the Systems for Centrifugal and Rotary information on this refinery. Pumps,” American Petroleum Institute, Washington, D.C. (Editor’s Note: This article is an excerpt Wallace, N. M., Redpath, D., and Netzel, from a paper entitled, “Toward Reduced J. P., 2000, “Toward Reduced Pump Operat- Pump Operating Costs, Part 2 – Avoiding ing Costs,” Proceedings of the Seventeenth Premature Failures,” which was presented International Pump Users Symposium, at the 18th International Pump Users Turbomachinery Laboratory, Texas A&M Symposium. It is reproduced with permisUniversity, College Station, Texas, pp. 171- sion of the Turbomachinery Laboratory (http://turbolab.tamu.edu). From Proceed186. Wallace, N. M., David, T. J., and Bowden, ings of the 18th International Pump Users P. E., 1999, “Quantifying Reliability Symposium Proceedings, Turbomachinery Improvements Through Effective Pump Laboratory, Texas A&M, College Station, Seal and Coupling Management,” Texas, pp. 135-146, copyright 2001.)

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✓ R E L I A B I L I T Y- D R I V E N P U M P M A I N T E N A N C E H A N D B O O K

Table of Contents A Systems Approach to Pump Reliability...................................................................1 Improving Pump Reliability at Thermal Power Plants .............................................3 Improving Performance of a Water Transport Pump................................................6 Early Maintenance Staff Involvement ......................................................................10 Team Approach Produces Winning O&M Formula - Part 1....................................12 Team Approach Produces Winning O&M Formula - Part 2....................................15 Team Approach Produces Winning O&M Formula - Part 3....................................18 Reducing Acoustic Emissions ....................................................................................22 Addressing Pump Vibrations - Part 1 .......................................................................25 Addressing Pump Vibrations - Part 2 .......................................................................27 Lubrication Systems for Rotating Equipment ..........................................................29 Lubrication Oil Viscosity Classification… Don’t Get Confused...............................34 Probe Installation Tips ..............................................................................................35 Bearing Basics ............................................................................................................38 Troubleshooting Boiler Feed Pumps.........................................................................42 Torsional Analysis in Couplings Selection ................................................................43 Not All Grouts are Created Equal .............................................................................47 Bearing Lubrication Trends and Tips .......................................................................49 A Multistage Vertical Pump in Light Hydrocarbon Service - Part 1........................53 A Multistage Vertical Pump in Light Hydrocarbon Service - Part 2........................56 What Does It Cost? ....................................................................................................60 Keys to Improve Pump Record-Keeping ...................................................................63 Reliability Improvements to a 13 Stage Charge Pump ............................................68 Reliability Improvements to a High Pressure Charge Pump...................................71 Reliability-Driven Pump Maintenance .....................................................................76 Fluid and Material Basics..........................................................................................85 Baseplate Design Affects Reliability .........................................................................87 Exploring Bearing Lubrication Options....................................................................90 How to Extend Pump Bearing Life...........................................................................95

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Table of Contents Mounting and Clearances for Casing Wear Rings....................................................99 Factors in Pump Suction Piping .............................................................................102 Alignment Movement Methods...............................................................................105 Foundation Tips.......................................................................................................109 Evaluating Pump Monitoring Options....................................................................110 Applying Predictive Maintenance Measures...........................................................112 Vibration Analysis Yields Good Vibes ....................................................................116 Getting the Full Benefit from High-Efficiency Motors ...........................................121 Coupling Strategies ..................................................................................................127 Adjustable Speed Drive Offers Pump Flexibility ...................................................132 Solid State Controls Tune Pump Operation ...........................................................135 Pumps for Haz/Rad Wastewater Service ................................................................138 Alignment Tolerances for Spacer Couplings...........................................................143 Couplings Alignment - Let’s Talk Tolerances .........................................................147 Machinery Alignment vs. Coupling Misalignment ................................................149 Build a Better Foundation to Reduce Costly Downtime .......................................151 Flow Sensing For Pump Protection ........................................................................154 Reliability Profile: Champion International ..........................................................157 Tips and Trends in Sanitary Pumping....................................................................160 Computer-Based Pump Reliability.........................................................................164 Proper Repairs Avert Failures: Part 1 - Centrifugal Pump Gasket & Impeller .......................................................176 Part 2 - Single Stage Horizontal Pumps ..................................................................181 Part 3 - Multistage Horizontal Pumps ....................................................................183 Part 4 - Vertical Pumps ...........................................................................................187 Evaluating Motor and Drive Solutions....................................................................191 Slurry Pump Wear Factors ......................................................................................194 Tools That Do Our Work for Us - Or Do They? ............................................204 Users Q&A: Canned Motor Pumps ...............................................................206 Winning Maintenance Strategies at Thorn Creek.........................................209

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Table of Contents Innovations in Multiphase Hydrocarbon Operations ...................................211 Key Indicators - The Measure of Reliability .................................................217 New Coatings Make Their Mark ..................................................................220 Rim and Face Alignment Procedures for Direct Motor Driven Equipment..224 “Best-of-Class” Lubrication for Pumps and Drivers .....................................227 Universal Drive Shaft Maintenance..............................................................230 Optimizing High Energy Pump Operation ...................................................232 Managing Vertical Pumps in a Changing World...........................................236 Improving Progressing Cavity Pump Reliability...........................................241 Water-Lubricated Fluid Film Bearings Can Be Trouble-Free........................244 Oil Lubrication for Process Pumps and Related Equipment........................248 Injectable Packing Compounds - An Alternative Sealing Technology ..........260 Multistage Pumps in the CPI .......................................................................263 The Role of Grouting in Standardized Installation Practices .......................270 Variable Frequency Drives for a Vacuum Pump System ..............................274 Turnaround Time! ........................................................................................276 The Critical “First Hour” Startup of a Turbine-Driven Boiler Feed Pump ..278 Methodologies for Calculating MTBF – Part I..............................................282 Methodologies for Calculating MTBF – Part II ............................................287 Packing & Rotating Equipment ....................................................................291 Which Alignment Method Is Right for You? ................................................294 Troubleshooting Liquid Ring Pumps ............................................................296 Replacement of the Boiler Feedpumps at Janschwalde Power Station ........298 The Nuts and Bolts of Vibration Signature Analysis ....................................305 Energy Savings Pay for Reliability Improvements at Chevron......................311 Teamwork Key to Pump Reliability Upgrades at Conoco ............................315 Maximizing Reliability in Sealless Pump Operation ....................................318 Laser Shaft Alignment of a “Bark Hog” ......................................................323 Balancing Pump Impellers ..........................................................................325 Revitalizing Vertical Lineshaft Turbines ........................................................329 Sealing Lime Slurry at Alberta-Pacific ..........................................................333

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Table of Contents AlliedSignal GeismarWorks............................................................................334 Well Pump Applications for Mine Dewatering ..............................................337 Hazardous Fluid Pump and Sealing Systems: Reliability Driven Improvements ........341 Classifying Chemicals to Ensure Effective Sealing ........................................344 Reliability: Big Opportunity or Big Distraction? ............................................347 Controlling the Seal Environment—A Key to Seal Reliability ........................353 Winning MTBR Strategies ......................................................................................................356 Reliability Starting with Procurement: Maintenance’s Role in the Purchase Cycle....361 Overhauling 36 Vertical Process Pumps and Motors under Adverse Conditions......367 Selecting Motors & Drives for Optimum System Reliability ................................................371 Understanding and Using Shaft-to-Shaft Alignment Measurement Systems ............379 OEM Pump Materials and Their Relationship to Quality and Reliability ........386 Piping-to-Pump Alignment: Getting It Right! ......................................................393 Alignment Monitoring Report........................................................................................399 Construction Impacts on Pumps and Systems ............................................................403 Predictive Maintenance Programs: Building a Complete Package ............................412 Best Practice: High Temperature Slurry Applications ................................................421 The Importance of a Bearings Inspection Program ....................................................425 Coupling and Alignment Strategies ..............................................................................431 Building a Better Foundation ........................................................................................436 Successful Submersible Operation Part 2: Inspection and Maintenance ..................444 Using VFD Technology in a Water Distribution System..............................................448 Strategies to Reduce Pump Repairs at Petro-Canada ..................................................455 Finding and Solving Vibration Problems......................................................................461 Spike Energy™ Measurement ........................................................................................467 In-Plant Perspective: Eli Lilly and Company................................................................477 In-Plant Perspective: Dupont ........................................................................................483 Building Reliability and Safety into Roatating Equipment ........................................489 In-Plant Perspective: KoSa..............................................................................................496

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Table of Contents In-Plant Perpective: Thorn Creek Basin Sanitary District ..........................................503 Paper Mill Stock Pump Improvements ........................................................................507 Root Cause Failure Analysis ..........................................................................................510 Continuous Monitoring of Sealless Pumps ..................................................................517 Increased Reliability at a Small Refinery ....................................................................526 Water Contamination of Equipment-Lubricating Oil ................................................................532 Pro-Active Maintenance for Pumps...............................................................................................538 Bearing Protection Devices and Equipment Reliability: Part I – Constant-Level Lubricators..........................................................................545 Bearing Protection Devices and Equipment Reliability: Part II – What is Really Justified?.............................................................................548 Maintenance Outsourcing – Critical Issues..................................................................................553 Pump Reliability – Hydraulic Selection to Minimize the Unscheduled Maintenance Portion of Life-Cycle Cost...........................................557

All materials © 2002 Pumps & Systems, LLC. No part of this publication may be reproduced without the written consent of the publisher. The publisher does not warrant, either expressly or by implication, the factual accuracy of the articles or descriptions herein, nor does the publisher warrant the accuracy of any views or opinions offered by the authors of said articles or descriptions.

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R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

A Systems Approach to Pump Reliability BY GREGORY ZIMMERMAN ump six has gone down again. That’s the third time in two years.” If your first reaction to such a statement is to ask what’s wrong with this pump, you may find yourself asking that same question when pump six goes out for the fourth time. The problem may appear to be a failure of the seal, bearing, or some other component. Data on repairs typically show that 80% of pump failures are manifested as seal failures. In some cases seals are not selected properly, or they do not meet the stated design specifications. If the shortcoming is in the seal selection or design, the new American Petroleum Institute (API) Standard 682, Shaft Sealing Systems for Centrifugal and Rotary Pumps, should help (see “API Improves Seal Selection and Testing,” page 29). But often, even though a problem shows up as a component failure, it actually arises elsewhere in the pumping system. Unless the root cause of the problem is found and fixed, any replacement pump will suffer the same fate as its predecessor. To find that root cause, you’ll need to take a close look at the procedures you use to design, specify, install and operate your system. The issue isn’t just the pump—it’s the whole pumping system. It’s also important to realize the “system” in which the problem lies may not be the pumping system but may be your company’s procedures instead.

“P

PILOT ERROR Companies that collect pump repair data repeatedly find that the leading cause of pump failure is incorrect pump selection. The pump is not being used in conditions for which it was designed. In other words, it’s operating away from the best location on the curve, it’s subjected to piping strains, it’s not properly aligned, or the user can be blamed in some other way for an error of omission or commission.

The same thing—pilot error— often applies to aircraft. Several years ago, investigators were troubled by a number of crashes for a new fighter plane. Investigations showed that nearly all crashes could be attributed to pilot error. But should planes crash just because pilots aren’t perfect? Shouldn’t the aerospace engineers understand a pilot’s limitations and design the aircraft to be at least a little forgiving? Similarly, pump manufacturers are starting to build in some forgiveness. They’re listening to their customers, trying to understand how pumps are installed and operated in the “real world.” Many manufacturers take advantage of user meetings, such as Texas A&M University’s International Pump User’s Symposium, held annually in Houston. The manufacturers sit in on the discussion groups, listening to users’ concerns and later designing pumps to accommodate them. Some pump manufacturers have more formal procedures in place, such as quality function deployment methodologies. Manufacturers bring in key personnel from the customer’s plant to explain to pump designers what factors are most important in the operation, how the product will actually be used, and, perhaps most importantly, why the product is being misapplied. The manufacturers are beginning to understand why customers misapply pumps. And they’re finding out that often it’s not out of ignorance that mistakes are made but instead because users must sometimes operate under unreasonable constraints. For example, everyone knows that piping strain is bad news for pumps. So systems should be designed to eliminate piping strain. Yet piping strain is a fact of life in many installations. What does the manufacturer do? Instead of blaming the users for poor design, some suppliers now furnish flange adapters that can absorb piping strains.

The Pump Handbook Series

Another difficulty is alignment problems. Again, everyone knows that the pump and driver must be in near-perfect alignment. But that’s hard to do in the field, especially given reduced numbers of field staff and reduced training budgets for the staff that are still around. To make it easier for their customers to meet alignment requirements, some manufacturers have designed their ANSI pumps so that the motor’s proper position can be locked in. The motor can be removed and then locked back into place without realignment. By simplifying installation procedures, manufacturers are trying to minimize the chance for “pilot error.” As one manufacturer says, “Our designs haven’t gotten to the bulletproof stage yet, but they’re certainly more robust, more forgiving, than they used to be.” Of course, no amount of design wisdom can overcome the basic laws of physics and chemistry. Pump users must still be sure they’ve selected the right materials, the appropriate type of pump, the right seals and couplings. And users will never be totally relieved of their maintenance responsibilities. Education is still needed in pump selection and operations. Manufacturers are perhaps in the best position to provide that education. Those who do are often surprised by the willingness of users to learn more about how to choose and operate pumps. In designing pumps that take more of user’s needs into account and in providing more information, pump manufacturers are building closer ties with their customers. And for their part, users are beginning to understand the value of working closely with their suppliers instead of simply choosing the lowest bidder for each project. The level of trust, the feeling of partnership between the vendors and users (and likewise between the pump manufacturers and the seal or motor manufacturers), is increasing in the

1

industry. Some users have even begun to “outsource” their repairs back to their manufacturers. Such an arrangement lets the users concentrate on their core business (i.e., chemical processing) rather than having to become pump repair experts.

COMMUNICATION WITHIN USER COMPANIES Before partnerships between vendors and users can be effective, however, work groups within user companies must become partners. One problem that will continue to hamper progress within user companies is the separation of engineering, purchasing, operations, maintenance, and other departments affected by pump purchasing decisions. Persons responsible for keeping the capital budget as low as possible buy the cheapest pumps they can find. The high operations and maintenance budget that results from this approach is often regarded as someone else’s problem. This overall difficulty will continue to plague companies as long as they consider capital costs, maintenance, and operations costs as being “owned” by separate internal entities. Even pumping people don’t always communicate and cooperate as much as they should. The result can be over-specified pumps. John Joseph, supervisor of central shops for Amoco’s Texas City plant explains: “The size of API pumps stated in requests for bids is often 50% to 100% too large.” Joseph attributes the problem to too much overcompensation. “I’ve been encouraging vendors to take initiative in replying to customer’s requests for bids,” he says. “Vendors should get the user engineers and process people together to talk about what the pump is really going to be called on to do.” In such a meeting the process engineers show on a formal process diagram what they need to pump to do, what the minimum and maximum flow rates will be and for what periods of time the pump will operate under normal conditions. (For example, “We’ll need to be at this point on the curve for 5% of the time.”) Then they lay out the startup and shutdown

2

requirements, how long and how often the process will be at particular temperatures, whether a flushing liquid is used, what it’s specific gravity is, and so on. “Once everyone agrees to that, they can eliminate all the layers of fudge factors,” Joseph says. “And if the pump size is dictated by off-spec conditions that will be in force for long periods of time, that must be accounted for and agreed to rather than hidden in some unquantifiable fudge factor.” Joseph suggests that by digging into the real requirements of the application, API pump users are often surprised that they can sometimes scale back two or three pump sizes and still meet all their requirements for normal as well as unusual conditions. A short meeting can yield big dividends. But the reality in many user companies is that no one takes the lead in bringing the right people together. That’s why Joseph is encouraging vendors to initiate the process. Limitations in the pump selection process aren’t just the user’s fault, however. According to Joseph, vendors can also help by changing their approach to presenting pumps. His first request is that vendors narrow their reported pump curves. “Start at the Best Efficiency Point (BEP) flow rate, go to 50% of that and mark through everything to the left of that on the curve,” he requests, “because most centrifugal pumps cannot operate reliably and efficiently in this low flow area.” Some pump designs are far more restrictive than that. Secondly, he requests that manufacturers publish NPSHR curves that add back the 3% head loss used to generate today’s NPSH R curves. “That changes a typical NPSHR curve into one that gives the user’s engineers the real story,” Joseph says. “Then, instead of relying on fudge factors, the users can bracket the BEP. With proper system design and this improved method of pump selection, cavitation and recirculation won’t be a problem, and the user can purchase a pump that’s the right size and provides the most efficiency and least maintenance cost for the capital

The Pump Handbook Series

dollar. The pump size selected is usually smaller, which reduces the costs of piping, foundation, and motor. Energy costs are lower, and because the pump is operating at or near the BEP flow rate, reliability is higher and maintenance costs lower. ”Users of high suction specific speed pumps know that the sweet spot on the curve is narrow,” Joseph says. ”High suction specific speed implies low tolerance to off-BEP-point operations. Users know that these pumps must be kept in a narrow operating range around the BEP flow rate at all times.” The bottom line is that there’s no simple, quick fix to pump reliability problems. Only rarely will a new seal permanently solve the problem with a down pump. Sometimes, using a system approach, a redesign of the pump hydraulics fixes or improves the problem. The most effective method is to address the problem from the first conception of the system. And everyone involved in the whole system must work as a member of a team. Good communication within user companies and between vendors and engineering contractors sounds easy but often requires people to overcome ancient barriers. Until someone takes the lead, we’ll always be wondering why pump six has now gone down for the fourth time in two years. ■ Gregory Zimmerman is Associate Editor for Pumps and Systems.

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R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

Improving Pump Reliability at Thermal Power Plants BY ALBERT SIMON, DR. JOHANN GÜLICH, DR. ULRICH BOLLETER, DR. DUSAN FLORJANIC, AND TOM MCCLOSKEY

FORECASTING CAVITATION

Model machine used to determine cavitation behavior.

During the investigation we looked for ways to forecast cavitation damage as a function of the suc-

The Pump Handbook Series

FIGURE 1 1.2

d3=1.025 (Gap B =2.5%, Gap A = 2%)

1.1

Ue Gap A

0.8

D2

0.7 0.6 0.5

100

0.4

80

0.3

60

0.2

Gap B

0.9

d3=1.043 (Gap B = 4.3%, Gap A = 2%)

d3=1.013 (Gap B = 1.3%, Gap A = 0.9%)

40

D3

1.0

d1=D1 D2 d3=D3 D2

D1

An important part of our investigation was determining how different parameters affected the shape of the feed pump. In addition to a high pressure coefficient and good efficiency, the pump needs a stable characteristic, which means properties that increase as the flow quantity decreases. We looked at: • specific speed of rotation “ηq” • gap between impeller blade and diffuser vane (gap B) • gap between the shrouds of impeller and diffuser (gap A) • overlap “Ue” between the shrouds of impellers and diffuser • inlet diameter “d1” of impeller • blade shape and distribution of blade loading. We built several model machines to observe the flow and to measure the system’s parameters. A model machine was built for stroboscopic flow observations to be carried out on the impeller, the diffuser and the return flow channel. These investigations included the influence of different combinations of d3 and gap A on the shape of the characteristic. Figure 1 shows the strong interaction between these parameters, particularly in the part-load range using the specific speed ηq ~ 30 as an example.

Efficiency curve η (%)

D

HYDRAULICS

p m tic pu ris ed cte Fe ara ch

uring the 1970’s the number of coal-fired power plants with a 400 mw rating in the United States decreased markedly. The Electric Power Research Institute (EPRI) and Sultzer Brothers Pump Division investigated the problem and discovered that feed pumps contributed to this decline. In order to make the power plants more efficient, and to attempt to isolate and solve the efficiency deficit, the authors looked at a variety of potential pump problems: • hydraulics • cavitation • rotor dynamics • hydraulic and mechanical interactions of a feed pump with the system surrounding it.

Best efficiency Efficiency

0.1 0 0 0.02 0.04 0.06 0.08 0.10 0.12 0.14 0.16 Volumn efficiency ϕ

Characteristic optimization by varying distances between impeller blade and diffuser vane (gap B) and the distance between the impeller shrouds and diffuser (gap A). tion head (net energy level in the inlet cross-section of the pump), of the impeller construction materials, and of the properties of the fluid being pumped. Determination of the rate of erosion by using the bubble length is possible knowing the bubble field configuration as a function of the suction head for different discharge rates, based on stroboscopic observations. Based on long duration tests on the model machine using variations of suction head, delivery head, discharge rate, temperature of medium being pumped and of impeller constructional material and field results, we determined the rate of erosion at the point of strongest cavitation attack. We could forecast the approximate rate of erosion. Since the bubble field size is practically the same for cold or hot water up to a temperature of 180°C, cold water can be used to determine the bubble field working in feed pump operating conditions.

3

In the case of feedwater, the influences of gas content and the temperature cancel out each other to such an extent that the cavitation intensities for cold, aerated and hot, deaerated feedwater are almost the same. We could only determine the erosion rate, however, when the bubbles were present on the surface of the impeller blade, since that’s the only time bubble field length can be measured. If a stroboscopic observation shows clouds of bubbles, we can determine where the bubbles implode by paint coating removal test. The paint coating test can also be used to determine cavitation damage. Using cavitation noise to determine the erosion rate is made easier because all types of cavitation, include recirculation, have been established. No special equipment is needed, such as Plexiglass, for bubble observation. Listening to cavitation noise is a good preventative maintenance tool for rapidly determining the approximate erosion rate in installations. However, this does not determine the exact location of bubble implosion. By using the paint test method and measuring cavitation noise, we saw an economic hydraulic layout of feed pumps with a provision of a prolonged or specific operating life of the first stage impeller.

FIGURE 2

4

Discharge line

/ZI/ /Zs/

Feedwater tank Booster pump Main pump 3λ/4 ~ q

/ZB/

Control valve

λ/4

FUFH

Simplified feedwater system, with schematic of main feed line, including various natural frequency acoustic standing waves of mass flow fluctuation for a closed control valve.

ks Hydraulic and mechanical forces occur on the rotor of a pump stage, which are classified into excitation forces (FH are hydraulic radial forces, FU is mechanical unbalance) and interaction forces [ZS] are the suction labyrinth forces, [ZI] at the impeller and [ZB] at the interstage bush. Ks is the stiffness. Excitation forces are not dependent on rotor vibrations.

Hydraulic unbalance created by geometric deviations in the impeller is radial force that occurs at a frequency corresponding to the speed of rotation. It is generally significantly greater than the mechanical unbalance. Vibrations in all gaps and spaces where liquid flows generate the hydraulic interaction forces between the rotor and stator. In the classic rotor dynamic model they are represented by the direct and coupled coefficients of stiffness (k, kc), of damping (c, cc) and of mass (M, mc).

in the inlet and at the diffuser entry. This influences on bearing loading and the dynamic coefficient of the bearings. And these in turn affect the rotor dynamic behavior of the pump.

FIGURE 3

ROTOR DYNAMICS

Vibration frequency = rotating frequency 100% discharge rate 25% discharge rate

1500 Radial force (N)

The rotor dynamic behavior of a multi-stage feed pump with high stage head is determined by the hydraulic and mechanical forces which are applied to the rotor. There is a difference between the so-called excitation forces that are present if the rotor does not vibrate at all and the interaction forces between rotor and stator, which arise when a rotor vibrates laterally. The different forces and the corresponding zones are shown in Figure 2. The hydraulic excitation forces consist primarily of the static radial thrust, the hydraulic unbalance and broad band radial forces, particularly in the part-load range. Static radial thrust, which can also be thought of as time-independent hydraulic excitation, occurs due to flow asymmetries

FIGURE 4

Vibration frequency = 0.1 X rotating frequency Peak (2260)

1500

1000

1000 Hydraulic 500

Mechanical Q=6.3 Q=2.5

0 ks Zs Z1 Interaction at 100 µm parallel displacement

ZB ∑

Effective value

500

Peak Effective value

0 ks Zs Z1

Excitation (unbalance) Q = balance quality level (ISO 1940)

Interaction at 100 µm parallel displacement

ZB ∑ Broad band excitation at vibration frequency/ rotating frequency = >0-0.20

Comparison between the radial interaction forces acting at the impeller and the excitation forces

The Pump Handbook Series

A comparison of all hydraulic excitation and interaction forces acting on an impeller are shown in Figure 3. A vibration amplitude of 100 µm was assumed for the quantitative determination of typical interaction forces. Knowing the forces, and the methods of increasing damping, designers can create reliable machines that have no tendencies to develop rotor dynamic faults. In a pumping system, pressure and mass flow fluctuations propagate acoustic waves. Where these waves are reflected — for example, from a closed valve — standing waves are formed with a minimum and maximum pressure for pressure and mass flow fluctuations. Both the formation and damping of standing waves are dependent primarily on the positions of the pumps, valves and other components in the system. Acoustic resonances in a pumping systems — arising from a coincidence between excitation frequency and emanating from a pump, valve or other component and an acoustic natural frequency of the system — have often only a small damping. Thus, pressure pulsations starting

out from pumps, valves or other components are amplified and lead to unacceptable vibrations in pipelines. It is then necessary to design as large a damping system as possible. However, because pipelines make only a slight contribution to the damping of pressure pulsations, the damping must be achieved through other components, such as pumps, valves or throttling sections. We built an acoustic model where each element of the system is represented by a transfer matrix which links pressure and mass flow fluctuations between inlet and outlet. Figure 4 shows a simplified feedwater system. The feed pump incorporated a slightly unstable characteristic. With the valve closed, the natural frequency of the system (13 hz) is associated with negative damping, i.e. vibrations are excited in the system (Figure 4). In this instance, the second natural frequency corresponds to the 3/4 wavelength, which has a mass flow fluctuation ~q maximum in the region around the pump. When the valve is half way open, all waveforms having positive damping, since the valve now represents a

The Pump Handbook Series

resistance which pulls vibration from the system. When the length of pipeline between the pump and valve is shortened and the valve is closed, the damping of the second natural frequency becomes positive, since the maximum of the mass flow fluctuation q is taken away from the pump. Thus, it is possible to design pump systems that are not responsive to exciting pressure pulses emanating from system components. If you need more information on improving your feed Pumps’ performance, EPRI’s “Feedwater Pump Technical Specification Guidelines” is a good resource. ■ Albert Simon, Dr. Johann Gülich, Dr. Ulrich Bolleter, Dr. Dusan Florjanic (retired) led this research effort and are with Sulzer Brothers Ltd. Pump Division, Winterthur, Switzerland. Tom McCloskey is Manager of Turbomachinery in the Generation and Storage Division of the Electric Power Research Institute. He serves on the Editorial Advisory Board for Pumps and Systems.

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R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

Improving Performance of a Water Transport Pump BY STEPHAN S. FLORJANCIC AND ALAN D. CLOTHER he six pumping stations of the Rio Colorado Tijuana Aqueduct were operating under unfavorable conditions. Specifically, pump impellers were wearing prematurely due to cavitation erosion. The root cause of the problem was low Net Positive Suction Head (NPSH). To reduce construction costs, the suction reservoirs of the pipeline system were built at relatively slight elevations above their pumping stations. This problem could not be solved at its source, however, because construction changes (e.g., raising the reservoirs or rerouting the suction piping) were cost prohibitive. The only option for extending the life expectancy of the impellers was to improve the pump design. The goal was to achieve at least twice

T

the original life expectancy. The constraints were the less-than-optimal NPSH conditions and certain inherent operating restrictions that would normally be considered undesirable for new pipeline layouts.

The bubbles then cavitate (implode) in regions of higher static head. The implosions can cause extreme stress on the impeller material, actually wearing it away. This is called cavitation erosion.

NSPH AND CAVITATION

THE SYSTEM

NPSH is defined as the total head, above vapor pressure, of the fluid at the suction nozzle. NSPH thus measures the margin against vaporization for the fluid entering the pump. NPSH requirements increase with flow rates above the Best Efficiency Point (BEP) due to high local velocities in the fluid. Thus, pumps running above BEP require greater suction head to maintain a sufficient NPSH margin. At low levels of available NPSH, vapor bubbles form within the lowest static head regions of the impeller.

The pumps were 24-inch, horizontal, single-stage double-suction pumps (Table 1). The system was designed for three such pumps to be operated in parallel at each station. Multiple pumps running in parallel operate very close to BEP. But the pipeline was often operated with only one pump running at each station. Because lower total flow reduced system resistance, the single pump operation forced the pump to operate at a flow considerably greater than BEP (Figure 1). The higher flow rate increased the required NPSH and led to excessive cavitation erosion. We inspected a typical impeller subjected to this treatment over 3,000 hours of singlepump operation and 13,000 hours of dual-pump operation. Maximum depth of erosion was 8 mm (0.315 in). This depth represented 75% of the thickness of the vane—the impeller had reached the end of its useful life. The measured depth and width of damage was comparable to estimated erosion rates based on measurements of cavitation noise.

FIGURE 1

RIO COLORADO 24x24x25C HSB, STATION 2 System and Pump Curves 1,2 or 3 Pumps Operating in Parallel Head [m] 170 3 Pumps 160 System Resistance

1 Pumps 150 2 Pumps

DESIGN CHANGES 140

Best Efficiency

130

120 0

0.5

1

1.5 25%

2

2.5

3

3.5

4

Flow (m°3/s)

5

Changes in the impeller and casing (Figures 3 and 4) involved:

System and pump curves.

6

4.5

Using state-of-the art methods, we developed a new impeller design. Flow conditions were modified and the design head reduced to prevent unnecessarily high flows during single-pump operation. Although some of the changes do not represent the best solution for new pipeline layouts, they were reasonable compromises for extending impeller life in this situation.

The Pump Handbook Series

face finish compared to normal sand casting.

FIGURE 2 Cavitation Bubble Distribution and Weight Loss per Unit Time as a Function of Cavitation Coefficient at Constant Speed

• Cutting back and profiling the inlet splitters to avoid vortex shedding into the impeller eye. • Profiling the volute lips.

PREDICTING DAMAGE L

Head L

60 57 50

L

L

0%

3%

40 25 x - Weight loss measurements

∆G T

(mg/h)

20 15 10 “theoretically”

5 0

0,3 0,4 0,5 0,6 0,7 0,8 0,9 1,0 1,1 1,2 1,3 1,4 1,5 Cavitation coefficient σu 1 Cavitation erosion delay given by finite cavitation resistance of the material Distribution of cavitation bubbles and weight loss. • Increasing the surface area of the vane at the inlet. The vanes were extended farther toward the impeller hub, thereby reducing loads on the vane.

which reduced excitation of vane pass frequencies.

• Reducing the vane inlet angle by an average of 3.5 degrees, which reduced the mismatch of flow and vane angle.

• Narrowing the width of the impeller discharge by 4% and reducing the angle of the discharge vane by 1 degree. This change steepened the pump’s characteristic curve and limited flow rates during single-pump operation.

• Changing the leading edge of the vane to an elliptical profile, which made the vane less sensitive to variations in flow angle.

• Upgrading the impeller material from 316 SS to CA6NM, which increased resistance to cavitation erosion.

• Extending the center web to the full outer diameter to accommodate vane stagger of the two sides,

• Casting the impeller using ceramic core techniques. This yielded better tolerances and improved surThe Pump Handbook Series

Insufficient NPSH available causes loss of head. Required NPSH levels are usually set for 1 to 3% head loss, a condition relatively easy to measure and a point at which pump performance begins to deteriorate. However, NPSH values do not directly define cavitation erosion and damage because cavitation starts well before the pump head deteriorates (Figure 2). Unfortunately, the reservoir levels in the Rio Colorado Pipeline do not provide a sufficient NPSH margin to prevent cavitation damage. There are two ways to predict cavitation: a special pump can be constructed with a plexiglass window so analysts can observe the creation of bubbles on the impeller, or the process can be detected indirectly via measurements of noise. For this a piezoelectric pressure probe is installed in the pump suction chamber. For this project the construction of a special model pump was not economically feasible, so we relied on noise measurements. This requires a careful setup to reduce background noise because cavitation is not the only source of noise. Cavitation noises can be discerned by monitoring head loss—the noise increases distinctly immediately before head loss occurs. It peaks at about 1% head loss and then drops. At greater head losses, cavitation bubbles dampen the noise.

TESTING We conducted suppression tests for varying flows on the test bed, comparing NPSH characteristics for alternative configurations and evaluating the effects of the design changes. Cavitation noise and vibrations were measured on the test bed and on-site. Test bed measurements allowed for a wide variation of flow and NPSH. Onsite measurements represented the low-variability in flow and NPSH seen in actual field conditions.

7

ed the prominent vane pass frequency peak in the original design. The new pump had excellent overall vibration behavior and was much smoother-running than the original pump.

FIGURE 3 SITE-STATION #2, PUMP #1, NEW-NEW STA2-PMP 1N SITE-STATION #2, PUMP #4, OLD-OLD STA2-PMP 40

Old Design, OB, Axial

EVALUATING IMPELLER LIFE

PK VELOCITY IN IN/SEC

STA2-PMP 40-POA 14-MAR-91 16:02 PLOT SPAN 0.16

Old Design, OB, Vertical STA2-PMP 40-POV 14-MAR-91 16:01

New Design, OB, Vertical

0

STA2-PMP 40-POA 13-MAR-91 14:14

New Design, OB, Axial 0

100

200

300 400 FREQUENCY IN HZ

500

600

STA2-PMP 40-POV 13-MAR-91 14:14

Vibration spectra for old and new design. Hydraulic tests showed the anticipated increase in slope of the pump characteristic curve, the increased BEP of the new design and the limitation on maximum obtainable flow. The new design features a reduced NPSH-3% level at rated flow and a 20 to 30% decrease in cavitation noise at design flow. At higher flows (which will not be attained at the site), the NPSH-3% is greater with the new design due to optimization of the impeller at a lower flow. (The shockless flow for the old design was above BEP but was close to BEP for the new

design.) The new design clearly increased the NPSH margin for singlepump and parallel operation. Cavitation erosion was reduced substantially. In addition, the modifications in the case and impeller staggering also reduced pressure pulsations, which reduced pump vibrations (Figure 3). Vibration levels were generally low. The 1 x rpm and 2 x rpm vibrations indicate small residual imbalance and minor misalignment. The new vane configuration led to a small 2 x vane pass frequency but virtually eliminat-

TABLE 1. DESCRIPTION OF PUMP Type/Size Design Speed Old Impeller New Impeller Old Flow Old Head New Flow New Head

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HSB 24 x 24 x 25C (disch., suc. nozzle, imp. diameter [in.]) Horizontal, Single Stage, Double Suction, Between Bearings 1,780 RPM D2 = 22.7 in.; D1 = 14 in.; 7 vanes no stagger D2 = 22.7 in.; D1 = 14 in.; 7 vanes staggered 24,200 gpm at BEP (87.5% efficiency) 460 ft. at BEP 20,800 gpm at BEP (89% efficiency) 440 ft. at BEP

The Pump Handbook Series

Calculations done according to published methods and based on field measurements showed that the new design allowed a guarantee of a 25,000-hour life for the impeller under threepump operation and a 13,000hour life under single-pump operation, where life-expectancy refers to erosion depth no greater than 75% of the vane thickness. The formula for predicting erosion shows how the gains in impeller life were achieved: Erosion rate = 9.4 x 109 x (Fcor/Fmat) x (Noise level/Tensile strength)2.9

The constants in the formula are empirical values based on a large amount of test data. Fcor, the corrosion factor, was 1 for both impeller materials used. But Fmat, the materials factor, was 1 for the ferritic stainless steel impeller and 1.7 for the austenitic CA6NM. Thus, the greater cavitation erosion resistance of CA6NM, expressed by its 1.5-fold greater tensile strength compared to stainless steel, is partially offset by the material factor. The change in material properties alone increases the impeller life when subject to cavitation to about 190% of the original value. The less intensive cavitation was immediately apparent in the 20 to 30% reduction in cavitation noise. Introducing this change into the above formula shows that the improved hydraulic design alone also increased impeller life to about 190% of the original value. Note that new materials becoming available may allow an even greater increase.

INSPECTION OF THE PROTOTYPE IMPELLER ON-SITE The prototype impeller was inspected after 3,000 hours of operation on site. Based on earlier measurements of cavitation noise, we

extrapolated the calculated remaining life of the impeller and checked the guarantee given. Data logged by the users indicated the length of time the pump was operated under various conditions (Table 2) and enabled us to establish the maximum wear rate. The impeller showed no loss of metal. Variations in vane thickness, as measured by an ultrasonic probe, were well within the 10 mil tolerances of the casting technique. The area close to the vane inlet appeared to be polished, indicating the region where cavitation persisted. Available NPSH was not sufficient to prevent cavitation entirely, but the new design clearly reduced the intensity of cavitation and eliminated premature impeller wear. The results of the on-site inspection were better than those predicted from the test bed analysis. The new impeller will exceed its guaranteed lifetime. ■

Modified pump at the Rio Colorado Tijuana Aqueduct.

TABLE 2. PRO RATA LIVE EVALUATION Operating Units 1 Pump 2 Pumps 3 Pumps Total

Operating Hours 985 1,955 0 2,940

Guaranteed Hours 13,000 18,000 25,000 N/A

Percentage Used 7.6% 10.9% 0.0% 18.5%

Predicted (mils) 28 41 0 69

Erosion (mm)* 0.72 1.03 0.0 1.7

Stafan S. Florjancic, Ph.D., is senior field engineer, Sulzer Bingham Pump Company, Portland, OR. Alan D. Clother is a hydraulic and applications engineer, Sulzer Bingham.

*Vane thickness is 0.5 in., end of life is reached when 75% of vane thickness is eroded.

The Pump Handbook Series

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R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

Early Maintenance Staff Involvement BY PUMPS AND SYSTEMS STAFF aintenance expertise is invaluable. When a maintenance job needs to be done, you want it done right and done quickly. But if you call on your maintenance staff only after the fact—only after something needs attention— you’re missing an opportunity to improve your operations even further. To take full advantage of the knowledge of your maintenance staff, you should involve them before the production process starts and call on them in the selection of equipment used in the process. The Oregon Overlays Division of Simpson Timber has been successful with that approach. As Rich Cooper, maintenance coordinator for the Portland-based company explains, “When we choose a pump, we want to know not only that it will be able to do

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the job, but also that we will be able to keep it up and running.” The divison makes phenolic resin-impregnated paper used for overlays such as on wafer board, plywood, signs and easy-release concrete forms. The production process involves pumping formaldehyde, phenol, poly-vinyl-acetate, methanol, caustics, resins, glues and abrasives. “It can get tricky fast,” says Cooper. Cooper plays a major role in selecting pumps for the company. “The process engineer will come to me with the application specs — the required flow, the viscosity, temperatures and so on,” Cooper explains. “I’ll look at those parameters and at the elbows and valves and pipe restrictions and figure out the head we need. Then I discuss the application with manufacturers to find which of their

products meet those requirements. Next I look at the design of the pump from an installation and maintenance perspective. I make sure we’ll be able to get it in and out and that it won’t be too hard to rebuild. And I consult with our other maintenance people to be sure they haven’t had any problems with that type of pump.” This approach has headed off potential problems. For example, an air pump used in one application worked well until the company cut a trench around it to contain any possible spillage. Then the pump developed pulsation problems. On the recommendation of maintenance staff, the company switched to a mechanical pump. The change solved the problem. In another case, an air pump had recurring problems with freeze-ups. The maintenance staff

Rich Cooper discusses pump maintenance with John Wheeler, millwright.

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The Pump Handbook Series

solved the problem — again, by choosing a different pump. The maintenance staff meets every Tuesday to discuss problems. By comparing notes on maintenance activities, they can also spot trouble before it happens. “Our maintenance meetings are an open forum,” says Cooper. “Everyone is free to put in their ideas. Everyone has expertise to lend to the decisions.” Cooper himself brings a good deal of experience to the process. Before he took on the job of maintenance coordinator, he was a millwright for seven years. Part of his job was to rebuild pumps. So he knows first hand how they operate, what causes them to breakdown and how hard they can be to get back on line.

SHARING EXPERTISE IS A WAY OF DOING BUSINESS The division runs up to 100 processes with a total of about 100 employees. “Although we’ve grown our staff, we still get a lot done without a lot of people,” says Cooper. “We count a great deal on our suppliers. We have some very good ones that can pinpoint a solution.” But most of all, the company relies on the experience of each employee. Early maintenance involvement is just one example. Empowerment and sharing of knowledge might be the latest buzzwords of management gurus, but it’s been the philosophy of the company for a long time, according to Cooper. “The attitude around here has always been that we’re all on the same team,” he explains. “Production staff will do some minor maintenance. They’ll alert maintenance staff to unusual noises or other information that will help us. They don’t just demand that we get over there and fix a problem.” All employees are expected to share information and ask questions. Cooper says that no one feels as though he or she must always come across as the all-knowing expert in order to protect their job. “Of course we take pride in our individual areas of expertise,” he says. “But then we know that no one is an

expert in everything. We know that there’s someone in the plant who’s an expert in one particular area and that we’re free to call on that knowledge.” To promote this attitude, the company has an integrated Plant Services Department instead of splitting maintenance, quality assurance and engineering into segregated departments. Thus no barriers exist between these groups. “We always feel free to talk to other people in Plant Services,” says Cooper. “We go to meetings of the QA and Engineering groups and we have them in our meetings. The plant manager comes to our maintenance meetings about once a month.” Maintenance personnel are also invited to production meetings, and vice versa. “We’re all learning about the effect our activities have on the next department,” says Cooper. “Then if we have a choice of how to do a job, we can be sure to pick the one that works best for us and for the other departments. We’d never know that unless we were always in communication.” “All of us feel free to make suggestions because we know the suggestions will truly be considered,” says Cooper. “We are allowed to exercise initiative and make choices. And when we see our ideas put into action, we have more motivation to do it again.” For example, Cooper thought that a problem in a packing system could be solved easily by using an adapter he had seen in a brochure. Cooper’s supervisor responded with a “sounds good to me,” instead of nit-picking-it-to-death as managers in some companies would. “Our managers are good at acknowledging the expertise of our staff,” says Cooper. No one proclaims himself to be the sole owner of knowledge either. “For example, our engineer is an expert,” says Cooper, “but he also knows to ask us about proposed changes to a system. He relies on our hands-on expertise.” Cooper anticipates that his participation in pump selection will become even more important as

The Pump Handbook Series

the plant begins to use newer technology to meet environmental and safety concerns. By looking at cutaway drawings and at vendor-recommended maintenance procedures (e.g., skills, spare parts and tools required), Cooper will be able to assess the difficulty maintenance will have repairing pumps. In keeping with the teamwork philosophy, Cooper will be sure to call on the expertise of the current pump repairman. “He’s seen some new things since I left that area,” says Cooper. Cooper recommends that all plants involve maintenance staff in the selection of equipment. “It works well for us and if a plant works hard at making sure ideas get shared, it can’t help but go forward.” He believes that the company’s success during the recession is due in large part to this attitude. “We all learned that we can make money even in lean times if we work together on the same team,” he says.

FOR MORE INFORMATION The management of the Oregon Overlays Division of Simpson Timber is starting to formalize the company’s long-standing philosophy of information sharing and early maintenance involvement around the Total Productive Maintenance concepts of S. Nakajima. Nakajima’s books explain the advantages of early maintenance involvement and suggest ways of putting the principles into practice. ■ Introduction to Total Productive Maintenance by S. Nakajima, published by Production Press (1988) and TPM Development Program: Implementing Total Productive Maintenance by S. Nakajima, published by Production Press (1989). Contact your local bookstore for ordering information.

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R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

Team Approach Produces Winning O&M Formula-Part 1 Major gains in pump procurement, installation and reliability at BASF plant. BY: KAREN SELF, STEPHEN R. TREICHLER AND MOHAMMAD A. ZAMIN

eamwork and communication, essential for a quality process, are often overlooked in large capital projects. Yet critical communications among engineer, machinery specialist, plant operations, maintenance and reliability personnel (i.e., among planner, buyer and user) occur too late in a project, resulting in poor rapport, spotty installation, increased start-up problems and higher overall costs. End result? An unhappy customer. Dependable initial equipment design and selection, proper installation, and consistent operational support and training are the cornerstones of smooth pump start-up, operation, maintenance and reliability in any new process plant. When plant engineering, production and maintenance personnel work as a team these key factors are effectively integrated for overall quality. Various disciplines envision a successful project in differing ways and these differences can reinforce communications problems. For example, to engineering, a successful project meets schedule and budget. While operations values this, too, they also require low long-term operating costs and reliable equipment. The accompanying sidebar lists problems that can plague pump procurement, start-up and long-term reliability. Better, more timely communication and teamwork can help solve many of these problems.

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tion, procurement and installation procedures. Results from 18 months of operation demonstrate the success of this approach. Our team’s primary goal was to develop a program to help reduce start-up problems and extend the plant’s MTBR (mean time between repair). We hoped to improve machinery reliability–today’s key to cutting plant operating costs and remaining competitive. In particular, we directed our reliability group to help maintenance and production identify possi-

ble causes of failure and suggest and implement measures to improve MTBR. At the on set of the project, the pump team, consisting of the machinery engineer from corporate engineering, the reliability engineer from the site reliability group and the maintenance mechanical group leader from the plant, formed to identify plant needs and select reliable pumps to meet project goals. The manufacturing representative from production and the lead project engi-

PHOTO COUTESY OF BASF CORP.

THE OLD WAY— ENGINEERING VS. PRODUCTION

FORMING THE TEAM During a recent $35 million expansion project, the BASF plant in Freeport, TX applied a long-term team reliability approach to pump selection and installation. This grassroots wastewater treatment project presented an excellent opportunity to apply a team concept for specifica-

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BASF’s Freeport, TX wastewater treatment plant

The Pump Handbook Series

neer kept abreast of the team’s progress and participated in key meetings and decisions. Figure 1 outlines the team chronology. Initial meetings stressed team goals, communication guidelines and each team member’s responsibility. The first plant meeting covered the following general issues: • document distribution and review •

plant philosophy regarding reliability

•

vendor list approval

•

previous plant experience with vendors

•

vendor support and pump manufacturers’ inspection and testing procedures

•

preferred vendors for auxiliary items, such as seals, couplings and drivers

•

plant involvement in installation and pre-start-up activities

•

parts and vendor stocking programs

MAINTENANCE AND OPERATIONS REQUIREMENTS Commonly, plants rely primarily on company standards along with established project piping and instrumentation diagrams and flowsheets when developing mechanical specifications. Often, larger engineering offices apply only the company’s technical standards for equipment specifications, overlooking (or not even receiving) plant or project specific requirements. This practice inevitably leads to costly problems and delays during installation. Purchase specification A professionally engineered project is based on a well defined specification, which should: • state design flow conditions for the pumping equipment •

describe the project philosophy

•

include plant-specific requirements

•

indicate severe duty considerations (erosives, corrosives, extreme temperatures, etc.)

•

define possible short-term excursions, such as operating under reduced flow at start-up, turndown operation, emergency conditions, etc.

Even though our wastewater treatment plant project was well defined, some process parameters were being evaluated at the same time the pumps were being specified and selected. Consequently, continuous and timely communication within the team and with vendors was vital. The first required document, an inquiry prepared by the machinery engineer, called for comprehensive information from many disciplines: Process and operations personnel provided data covering the complete operating range, quantity and type of corrosives and erosives as well as acceptable metallurgy. Reliability provided information on plant preferences, such as preferred seal suppliers, coupling suppliers, enhanced pump design features for longer life, as well as pump selection criteria. Maintenance recommended seals and installation procedures and supplied historical data including previous experience with various makes of pumps. Electrical provided motor data, approved motor suppliers and variable frequency drive considerations. Materials reviewed the metallurgy and recommended paint systems for the motors, pumps, guards and baseplates. Additional requirements for baseplates, inspection and testing, along with the company’s technical standard for pumps, were also documented in the equipment specification. The approved vendor list was screened to identify suppliers with a strong plant service history. All team members reviewed the final inquiry specifications. Technical Bid Evaluation The technical evaluation on all bids involved an examination of facThe Pump Handbook Series

tors such as bearing life, shaft stiffness, impeller sizing, and design flow vs. best efficiency point (BEP). In addition to costs, equipment options, power debits and maintenance experience were also evaluated. The team carefully reviewed proposals against the inquiry documents and evaluated all exceptions taken by vendors. The purchasing department then contacted vendors to reconcile discrepancies. The team applied a power debit against the less efficient pumps, accounting for the additional power used over the unit’s life. Current and predicted interest rates and the cost of power contributed to the power debit calculation. Most companies consider the average life of rotating equipment as nine to twelve years. However, this life span assumption leads to a very high power debit figure, which may skew the selection toward the more efficient pumps—regardless of other factors, such as cost or mechanical reliability. So our team used a threeyear pay out period for the project and included the estimated dollar value in the specifications. Consequently, the power debit was important in the selection of only 8% of the pumps. Because our team held the bid list to pump vendors with a favorable plant reliability history, we decided not to include any repair cost estimates in the evaluation—eliminating any subjective opinions that could bias the bid evaluation. However, differing component designs and standards of quality control among vendors may correlate with differing MTBR even in pumps built to the same standards. If a plant has a comprehensive maintenance database, these differences can be easily quantified. After preliminary evaluations plant maintenance and reliability personnel reviewed the vendor proposal, technical assessment and purchase recommendations. Even though maintaining confidential commercial terms can be difficult during wide internal reviews, cautionary notices on the bid packages prevented breaches of trust. Since bids fell within an acceptable range, the final selections were based more on technical

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factors than cost. As a result of this evaluation process, we decided to purchase the majority of our pumps from two suppliers. Four additional suppliers provided a small number of specialized pumps. Next month we’ll discuss procedures for quality pump design after an order is placed, and present vendor case histories illustrating the importance of the team approach. ■ Editor’s Note: This article is based on a paper presented by the authors at the 1994 International Pump Users Symposium. Karen Self is the Reliability Engineer for BASF at the Freeport site. Stephen R. Treichler is a Mechanical Group Leader at the same facility. Mohammad A. Zamin is a Machinery Engineer at BASF Corporation’s Houston Engineering Office.

Potential Problems Impacting Pump Procurement, Start-Up and Long-Term Reliability

PURCHASE SPECIFICATION A professionally engineered project is based on a well defined specification, which should: •

state design flow conditions for the pumping equipment

•

describe the project philosophy

•

include plant-specific requirements

•

•

•

Design criteria for all operating conditions, such as startup, partial load, or emergency operations, not specified.

indicate severe duty considerations (erosives, corrosives, extreme temperatures, etc.)

•

Maintenance or Production requirements not fully considered.

define possible short-term excursions, such as operating under reduced flow at startup, turndown operation, emergency conditions, etc.

•

Technical standards outdated or incomplete.

•

Plant’s local requirements not communicated to the machinery engineer.

•

Plant preferences for vendors with exceptional field maintenance service, vendor stocking programs, site spare parts compatibility, or maintenance technicians familiar with equipment overlooked.

•

Failure to communicate all requirements to pump vendor.

•

Communication breakdown between vendor’s local sales office and vendor’s engineering/fabrication office.

•

Lack of communication between the pump vendor and its sub vendor.

•

Inadequate pump vendor and sub vendor quality control.

•

Insufficient and poorly documented inspection and testing at manufacturer’s facility.

•

Damage during shipment or installation.

•

Insufficient inspection upon arrival at job site.

•

Installation instructions missing, insufficient or ignored.

•

Poor field follow-up.

•

Lack of experienced personnel.

•

Missing or late review of documents.

TECHNICAL BID EVALUATION

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•

bearing life

•

shaft stiffness

•

impeller sizing

•

design flow vs. best efficiency point (BEP)

•

costs

•

equipment options

•

power debits

•

maintenance experience

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R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

Team Approach Produces Winning O&M Formula-Part 2 BY: KAREN SELF, STEPHEN R. TREICHLER AND MOHAMMAD A. ZAMIN ump start-up, operation, maintenance and reliability depend on teamwork and communication. This month’s article addresses issues occurring after an order has been placed with a pump vendor. Follow-up is key—as illustrated by the case histories presented.

P

PUMP MANUFACTURER QUALITY EFFORT Once a purchase requisition was delivered to a pump supplier, the design effort shifted to the manufacturer. At this stage in the project, the requisitioner focused primarily on the drawing and delivery schedule and conformance to quality standards by the manufacturer. All agreements made during the pump selection process were included in the revised data sheets which were issued to the vendor as a contract document. If the purchase requisition is comprehensive and the pump is supplied by an established vendor, the pump should, ideally, meet all the requirements of the customer. However, the process of design and fabrication involves many organizational subunits of the supplier and its subvendors, and lack of communication or commitment among these groups may result in nonconformance. If a vendor did not perform according to the purchase contract, our previous philosophy would dictate returning the pump along with a request that the deficiencies be corrected. This philosophy assumes that all defects would be found in field inspection, and that there would be sufficient time to make the repairs. Our approach for this project was based on the concept of partnerships. Both the supplier and the purchaser were part of a team, working toward the common goal, to obtain a pump that met the specification. Instead of promoting an adversarial relationship, this strategy focused on the question, How can we help each other meet the mutual goal?

PURCHASE ORDER REVIEW MEETING

Ensuring quality in the end product depends largely on communication of the agreed purchase requirements to the vendor’s design engineering staff. A post-award meeting between the customer’s requisitioning engineer and the supplier’s design engineer pays big dividends toward ensuring the completeness of this communication. Ordinarily, the vendor’s sales office (or distributor) communicates the requirements of the purchase specification to their design engineering office. Some vendor sales representatives will send the entire specification package, including all technical standards, as received from the customer. Other vendors require their sales staff to sift through the customer’s specifications and transmit only the deviations from their standard design. But many factors may result in a failure to fully communicate the needs of the customer to the engineer responsible for finalizing the design of the pump. For example, technical standards can be confusing and time consuming; requisitions may not be clearly written; and bid revisions, options, and changes made prior to ordering may cause further confusion. Moreover, sales personnel may not be aware of all the optional features of their pumps. A post-award meeting, occurring one to two weeks after the order is placed, enables the participants to review the order line by line, clarify needs and requirements and resolve any uncertainties. Attendance of the requisitioning engineer and the vendor’s design engineer is mandatory for the success of the meeting, but sales personnel, inspectors and purchasing agents might also benefit from being there. Face-to-face meetings provide the basis for more cooperative interaction among all parties, and they enhance the visibility of the order in the venThe Pump Handbook Series

dor’s schedule. However, post-award meetings do add costs in dollars and time. For very small orders or projects on a tight budget, a telephone conference call may be a better option. Post-award meetings were not conducted for this project due to time constraints as well as information indicating high quality among the selected vendors. This decision resulted in some surprises during the performance testing and shop inspection stages. These surprises are discussed later in the CASE HISTORIES section.

BENCHMARKING AND MEASUREMENT Benchmarking is an important part of quality control. This process includes monitoring the degree to which a manufacturer adheres to its own design tolerances. In the past, we have inspected newly-delivered pumps at random, completely disassembling them and measuring components against tolerances. Occasionally, a pump from a supplier failed this test. The problem with this approach to benchmarking tolerances is that the vendor’s warranty could be voided if a vendor representative is not present to witness the disassembly, measurements, and reassembly. This approach also adds to cost and schedule. Suppliers for this project were given the measurement responsibility. First, they were directed to mark clearances and tolerances of mating parts on the drawings. This data would facilitate repairs in the field. Then, our technicians spent a few days at one of the manufacturing plants prior to the assembly stage. They picked parts at random and checked them or observed vendor measurements. A significant number of out-of-tolerance parts were discovered and corrected before assembly. In the opinion of this team, the cost of the technician’s time and travel was justified by the potential problems eliminated.

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WITNESSED INSPECTION In addition to obtaining material certification for all alloy components and documentation for the hydrotests of pressure-containing parts, we implemented witnessed performance tests, attended by an outside inspector, for all pumps. While these tests added to the cost of the equipment, they provided assurance that the equipment would perform to specification. The requisitioning engineer, the unit maintenance engineer, and the operations engineer visited the two primary suppliers to see the first few performance tests themselves. These visits served to highlight the importance the company assigned to quality and to instruct the outside inspector with regard to our particular specifications.

FIGURE 1. PUMP TEAM CHRONOLOGY Purchase orders to pump vendors

Corrective actions by vendors

Fabrication

Team interface with pump vendors • Review of pump drawings • Shop pre-assembly inspections • Witnessed performance tests

Delivery

• “Ownership” shifts to site maintenance - oil change - inspections - weekly checks

CASE HISTORIES The bids for pump fabrication and delivery were awarded to four vendors. The problems arising with each vendor underscore the importance of applying a comprehensive team approach during selection and procurement. • Case History One Vendor one supplied 24 centrifugal pumps of various sizes. This supplier was willing to provide tolerances and fits critical for repair and assembly of the pumps. Plant technicians visited the vendor’s shops prior to assembly and found some shaft and seal tolerances outside the vendor’s own specifications. These deviations were immediately corrected. During the first few performance tests, the pumps operated as designed. Vibration and temperature rise were acceptable. Shaft runnouts, checked on three of the pumps, were within tolerance. However, a number of flaws were also evident to the team, including low spots in flange gasket surfaces, motor alignment holes machined part way through, poor paint application, base plate welding contrary to specification, and inadequate coupling protection of rotating components. These defects were repaired expediently by the supplier. However, when the pumps were delivered, additional problems were discovered. For example, epoxy paint had been sprayed in the bellow of the

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outer seals, and the seal glands were not properly oriented. The supplier incurred the expenses for additional inspection and subsequent repair. • Case History Two Vendor two supplied 13 centrifugal pumps in a variety of sizes. This supplier provided critical tolerances rather reluctantly. Part dimensions were checked, but not documented, prior to assembly. This supplier’s local sales staff had been directed to excerpt from our specifications a list of options and deviations from their standard design. This list was the only document sent to the factory engineering office. As a result, our technical standards and purchase specifications were not available to the vendor’s design team, and a number of our requirements, such as shaft turnout and type and duration of testing, were not communicated to the factory. Of the three pumps disassembled for runout checks, only one was within our specification. Another exceeded the supplier’s specifications as well. This pump, we discovered, had a defective bearing carrier. Because this vendor had no form of assembly documentation to verify that critical checks were conducted, all pumps were then checked for runout. These remaining pumps demonstrated tolerances within the supplier’s specification but not within our own. Upon delivery, the base The Pump Handbook Series

plates of four pumps were found to be stitch welded as opposed to our specification requiring continuous welds, and corrosion was evident in the joints. These defects were corrected by the supplier. Base plate mounting pads were not parallel, and excessive shimming was needed under most pump and motor feet for alignment. • Case History Three Two self-priming centrifugal pumps were purchased from a third vendor. This vendor did not, as a rule, deviate from their design standard. The order was placed through a local sales distributor. And approval drawings were received from and returned to the distributor. Due to turnover in the sales staff, these drawings were not returned to the pump supplier. While our expediter was told that everything was on schedule, the supplier had not followed up when approval drawings were not returned in time. We had, in the meantime, overlooked the missing certified prints. When these pumps were delivered, much later than required, the base plate paint did not meet our specification. • Case History Four Twelve small gear pumps were purchased from the fourth vendor for sampling service in the plant’s analyzer systems. Our process requirements for these pumps called for low flow

with no pulsation. Although the product was specified as clean, small strainers were installed at the suction of each pump. In addition the pumps had been constructed with a metal gear against a soft plastic gear. As a result of these defects, pump performance was very poor. A large amount of trash in the product stream caused frequent choking of the strainers, and the small amount of dirt that passed through the strainers caused premature wear of gears and shafts. The average life of these pumps was only a few months. The misleading assump-

tions about the quantity and type of erosive in the product stream were the primary cause of faulty pump selection in this case. However, poor communication among our process, maintenance and design engineering groups, and insufficient importance assigned to the selection of these particular pumps because of their relative low cost added to the problem. Next month’s concluding discussion will focus on the installation process, as well as the results of the team approach, including highlights of our successes and some areas for

The Pump Handbook Series

improvement. ■ Editor’s Note: This article is based on a paper presented by the authors at the 1994 International Pump Users Symposium. Karen Self is the Reliability Engineer for BASF at the Freeport site. Stephen R. Treichler is a Mechanical Group Leader at the same facility. Mohammad A. Zamin is a Machinery Engineer at BASF Corporation’s Houston Engineering Office.

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R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

Team Approach Produces Winning O&M Formula-Part 3 Traditional differences are replaced with common goals. BY: KAREN SELF, STEPHEN R. TREICHLER AND MOHAMMAD A. ZAMIN uring a recent $35 million wastewater treatment plant project, BASF implemented a team approach to their pump operation and maintenance program with the impetus to improve machinery reliability. This approach to pump specification and procurement procedures was highlighted in parts one and two of this article. As shown by this month’s concluding section, the team approach was continued throughout the course of construction and installation. And, this strategy is credited for the decisive success of this project.

D

ON-SITE INSTALLATION One key to the success of the project was timing. Clearly, decisions about how equipment will be installed should not be made after it has been delivered. Therefore, the mechanical installation procedure was negotiated early-on in the project. During these negotiations, engineering and reliability facets of installation were addressed so that the goals of both sides could be met. One of the first agreements regarded ”ownership” of the equipment. During past projects, construction personnel tended to set new equipment aside until the schedule called for installation, and maintenance and production personnel were not consulted regarding the storage of this equipment. In contrast, for this project the maintenance members of the team were notified upon delivery, at which point each piece of equipment was thoroughly inspected, the oil was changed, and the reservoirs were filled. If some discrepancy with specification was discovered by the technician, the construction supervisor and the pump team were notified. In some cases the pumps were disassembled for inspection. In a few cases a pump vendor was required to make corrective repairs. During stor-

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age each piece of equipment was wrapped to protect it from adverse atmospheric conditions. However, the shafts were made easily accessible to make it easier to spin them on a scheduled basis. This simple ”ownership” agreement accomplished two goals. First, maintenance was on board during the early construction phase of the project, taking responsibility for what would ultimately be theirs. Second, construction did not have to worry about storage and maintenance tasks that might interfere with timely startup and mechanical completion. The complete installation procedure is outlined in Table 1. To aid in installation, an inspection check sheet (Figure 1) was developed for the maintenance technicians assigned to the project. These people provided quality control assistance for both construction and maintenance. Each section on the check sheet reflected decisions made early-on in the project. This check allowed early detection of problems, making corrections easier and cheaper. In addition, baseline information on each piece of equipment was recorded to facilitate future troubleshooting.

TRAINING Training was also essential to the project’s success. Production developed detailed job analyses. Each step in the operation of the equipment and the process was outlined with quality and safety factors highlighted. Maintenance also reviewed the analyses to ensure that important reliability factors were included.

PROJECT RESULTS Some plant start-up experts say that one should anticipate a 20-25% failure rate of rotating equipment during the combined start-up process and first six months of operation. The Wastewater Treatment Plant (WTP) project installed more than 50 mediThe Pump Handbook Series

um to large centrifugal pumps. The start-up was smooth, encountering no delays due to pump or mechanical problems. Less than two thirds of the budget for mechanical start-up was consumed. And this money was spent primarily on field checks and inspections prior to start-up. During the first year of operation, equipment reliability has been outstanding. Since start-up, only six seal failures have occurred. Two of these resulted from pump cavitation caused by a siphon created in one of the process streams because of poor pipe design. One failure resulted from a broken spring in the seal and two others from operational error. The last failure was due to pump cavitation that developed when a slug of foreign matter got caught in the pump. No motor or bearing failures have occurred since commissioning. Post start-up vibration monitoring shows little or no deterioration in pump conditions. Total costs on pump repair to-date have been less than $20,000. In contrast, another production facility constructed during this same time frame which did not apply the team approach toward procurement and installation was plagued with numerous mechanical problems prior to and after startup. As a result, an inordinate amount of field work and revision were required throughout the unit. The start-up budget was depleted, and more than $40,000 was spent in the first month alone on pump repairs. A comparison of typical equipment failures on these two projects is given in Table 2. The negligible increase in our routine maintenance budget can also be attributed to the team strategy. The WTP site maintenance budget currently averages 3.6% of total installed cost. Even though the expansion effectively doubled the equipment count in the plant, the budget of routine maintenance increased by

TABLE 1. INSTALLATION PROCEDURE 1.

Remove pump and driver from baseplate. Clean pump and motor mounting surfaces. Inspect baseplate for sufficient vent holes to eliminate the possibility of trapping air between foundation and baseplate.

2.

Chip top surface of concrete pedestal to remove laitance. Blow with air to remove dust or moisture. Wrap anchor bolts in polyurethane or other soft material so the grout does not stick to them.

3.

Level pump and motor mounting pads to 0.002 in/ft. in both horizontal directions. Use cylindrical stainless steel bearing pads under the jackscrews. Flanges must be within 0.0625 in (1/16 in) using machinist level (Starrett No. 98 or equivalent). Pump discharge flange must be level within 0.005 in/ft, both lengthwise and crosswise.

4.

Form for epoxy grouting.

5.

Mix epoxy carefully per grout manufacturer's specifications. a) Proper mixing temperature is important. b) Ensure sufficient quantity for a single pour.

6.

After epoxy has set and cured, remove forms and clean up. Check baseplate for voids. Carefully fill voids, do not over-pressure.

7.

Mount pump and motor and rough align. a) Alignment: Couplings must be aligned (using reverse indication method) to within 0.003 in. Number of individual shims under motor feet must be kept to a minimum. Do not place shims under the pump feet. b) Piping: Piping must be supported independently and cause minimal stress on pump casing. Stress will be measured by placing dial indicators on pump coupling and loosening and retightening flange bolts. Maximum allowable deflection will be 0.003 in.

8.

Final align. Use reverse indicator method.

9.

Tighten suction and discharge flanges of the pump.

10. Recheck alignment. 11. Install flush piping and seal pots. Check markings on gland to match flush connections. 12. Connect motor and bump to check directions. NOTE: If installed equipment is not to be used for an extended period of time, maintain lubrication and shaft rotation check schedules. only 1.7% of the total installed cost. Of this, only 0.8% was budgeted for mechanical maintenance.

OLD TRADITIONS VERSUS NEW REALITIES Traditionally, engineering personnel have not taken a proactive approach in consulting maintenance or production with regard to their concerns for a particular project.

Conversely, production and maintenance people tend to postpone direct involvement until too late in the project. This lack of interactive involvement can be overcome with frequent visits by all parties to the site prior to and during construction. These visits can be augmented with frequent meetings of plant personnel at the engineering offices. Site personnel are The Pump Handbook Series

at times reluctant to break away from their respective plants and other responsibilities to participate in the early design stages. Unfortunately, this is when their input is most needed. Engineering and plant management must support the active interaction of engineering, maintenance and production and not view these meetings as a waste of time and money. Cost has historically been the dominant component in the pump selection process. However, we believe the selection process should be biased toward technical merit, even though all pumps selected are not the cheapest. This bias can be promoted by a good inquiry package as described in part one of this presentation. Other factors to weigh during selection are past vendor service records with the plant and technician familiarity with a particular vendor or product. These factors were considered during the WTP project, and the pump selection was accepted in its entirety on the first pass. Generic inquiry packages serve no useful purpose. Extra design and/or construction specifications must be included in the inquiry. Otherwise, the manufacturer or construction contractor has no choice but to charge for add-ons. These additional charges erode contingency money unnecessarily. Furthermore, a mediocre inquiry package may lack sufficient technical information to back up a decision to purchase a more expensive piece of equipment. A large project organization may be particularly prone to poor inquiry packages. Because of the number of people who must review the package, it may become weak from over-editing, which can tend to lead to deletion of critical design parameters. Another stumbling block to project success is the early entrenchment of project members who are resistant toward change. It is crucial to obtain early consensus with regard to project goals among all individuals involved in pump selection and installation. In the past, our company had foregone the post-award meetings with the vendors. However, we reconsidered this practice because our

19

AREAS FOR IMPROVEMENT

FIGURE 1

No project is without flaws. For the WTP project we identified the following areas for improvement:

PUMP INSPECTION/COMMISSIONING CHECK SHEET PUMP ID: PUMP DESCRIPTION: PUMP DATA

KEY DATES Date Inspector

Pump Service Manufacturer Model/Size / Impeller: Size: Mtl Motor: HP RPM Mfg Mechanical Seal (Single Double Manufacturer Model

Design Flow Head Sp.Grav

gpm

Clearance Frame None Flush Plan

):

INDIVIDUAL INSPECTION

Delivery Notification Initial Inspection Installation Alignment Fit-up Start-up & Acceptance Eq File Complete

NOTES

Visual inspection - look for shipping damage, flaws, defects Compare to vendor drawings - materials, dimensions, quality On-site storage - oil, N2 purge (?), location Obtain pump manuals, drawings for equipment file Add pump to stored equipment preventive maintenance schedule Initiate equipment checksheets, documentation, etc.

NOTES

Witness equipment relocation, installation Witness alignment - obtain alignment sheet for files Witness grouting (epoxy grout only) Witness fit-up for pipe stresses Witness final alignment (if required) Ground wire installed

•

Review of piping design Piping design review should include consideration of process flow and mechanical equipment requirements. ”Buy-in” from all involved parties Significant amounts of time, money and energy can be wasted correcting problems that occurred because of a lack of consensus. The earlier the ”buy-in,” the better. •

PRE START-UP

NOTES

Seal flush installation - open, clear Check lubrication - change oil if required Check rotation Review I/E installation - stop/starts, alarms, shutdowns Pressurize seal to check for initial seal leaks

START-UP

NOTES

Participate in initial start-up Monitor pump, seal. and motor performance Obtain vibration signatures for equipment file Add pump to production lube and preventive maintenance schedule Complete equipment file documentation: Equipment drawings Certified drawings on components Spare parts list Input equipment into changes

20

• Attention paid to smaller pumps There are no ”small, insignificant” pumps. All pumps need to be treated with the same attention to detail, regardless of size.

•

INSTALLATION

experience has shown that, depending on the vendor, the information received by its manufacturing unit may not clearly convey the specifications originally submitted to the vendor sales representative. Some companies require that only the deviations from standard manufacturing items be sent to their manufacturing facility. Others want the entire pack-

• Storage of equipment Experience has shown that static equipment requires more protection from the elements than that which is in-service. An enclosed storage building would more effectively protect the equipment awaiting installation.

age sent to them. The former situation creates information filters through which less and less of the original request remains intact. Removing these filters through post-award meetings (face-to-face for larger parcels, conference calls for smaller ones) has proven to be invaluable.

The Pump Handbook Series

Mechanical installation bid package If specific installation procedures (e.g., methods of alignment or grout procedures) are desired but are not included in the package (possibly due to over-editing), the bid package must be considered incomplete and re-evaluated. •

Post award meeting Post award meetings with the vendor’s manufacturing representative to discuss equipment specifications on an item-by-item basis is extremely important. For some pieces of equipment or for smaller orders, a conference call may substitute for a face-to-face meeting. •

Maintenance database A maintenance database should be established to determine the maintenance cost associated with poor installation. The same database could be expanded in order to evaluate the maintenance costs associated with the different suppliers. In this way, a maintenance debit could be applied to pumps that have a poor maintenance track record.

TABLE 2. FAILURE STATISTICS

WTP PLANT B

MECH. SEALS 6 26

EQUIP. BEARINGS 0 8

SUCCESSES OF THE TEAM APPROACH The benefits of a team approach are so numerous that any group undertaking a project, regardless of size, should not hesitate to incorporate this approach from the beginning. Highlights of these benefits include: • improved ”ownership” by all engineering members throughout the project •

early involvement of plant personnel (both production and maintenance)

•

excellent transition of responsibility from engineering to the plant

•

networking and sharing of information, both intra-company and inter-company.

MOTOR BEARINGS 0 3

MOTOR WINDINGS 0 2

CONCLUSION In a team approach the traditional differences in the various member objectives are replaced with a common objective to obtain the best total value for the company. Team members strive for a positive interaction to make the right decisions the first time. This effort requires members to communicate well with one another, addressing important project factors at appropriate times during the execution of the project. The early assembly of the team members was a key factor in the success of our project. The representatives from the engineering, production, maintenance and reliability groups were dedicated and knowledgeable about pump applications in a process environment. This dedication helped ensure the necessary attention to details required for a quality job. ■

The Pump Handbook Series

Editor’s Note: This article is based on a paper presented by the authors at the 1994 International Pump Users Symposium. Karen Self is the Reliability Engineer for BASF at the Freeport site. She joined BASF in 1989 after working for BeCon Construction Co. Ms. Self has a B.S. in Mechanical Engineering from the University of Oklahoma and is a member of ASME and the Vibration Institute. Stephen R. Treichler is a Mechanical Group Leader at the same facility. He joined BASF in 1985 after 10 years as a section leader at Gilbert/Commonwealth. Mr. Treichler has a B.S. in Mechanical Engineering from Michigan State University and a B.A. from Trinity University in San Antonio. Mohammad A. Zamin is a Machinery Engineer at BASF Corporation’s Houston Engineering Office. He is responsible for specifying and selecting rotating equipment for BASF’s chemical plants in the Gulf Coast area. He joined BASF in 1990 after 12 years at Exxon Chemicals and 7 years at Dawood Hercules in Pakistan.

21

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R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

Reducing Acoustic Emissions BY WILLIAM MABE ith all the hoopla and concern over the Clean Air Act and the problem of volatile pump emissions, it is easy to forget another important pollutant source: acoustic emissions or, in everyday language, noise. The same technology that is helping the pump manufacturer eliminate seal leakage has also gone a long way towards reducing objectionable noise. To understand noise, we’ll review the fundamentals of sound. Then we’ll look at the remarkable improvements in noise reduction that have occurred through product development, most recently in the sealless canned motor pump.

W

FUNDAMENTALS OF SOUND

The human ear is a highly sensitive organ that responds to sounds of an enormous acoustic range. To avoid large unwieldy numbers, decibel scales based on logarithmic ratios have evolved to represent sound levels. Figure 1 shows the relative magnitude of noise from common sounds in decibels (db). Sound power is the fundamental property of a sound source. Complete

FIGURE 1. EVRYDAY NOISE IN DB 140 dB Threshold of pain 130 AIRPLANE 120

100 JACKHAMMER 90 80 AUTOMOBILE 70 60 OFFICE 50

10

HOME

20

0 Threshold of hearing

22

Subtraction It’s surprising that when you subtract 99 db from a system totalling 100 db, it leaves a balance of 93.1 db. Keep in mind that subtracting a noise source that is narrowly separated from the total noise has questionable validity because instrumentation resolution is at best about 1/2 db and typically suffers from continuous drift. As you might expect, small sources subtracted from a large total sound source leave a balance approximating that of the total — the total noise does not change appreciable: 100 - 100 = 0 100 - 99 = 93.1 100 - 90 = 99.54 100 - 80 = 99.96 100 - 90 - 90 = 99.0 100 - 95 - 95 = 95.65

Averaging Similar noise averaged logarithYou can logarithmically add, mically produce results close to those subtract and average individual obtained by simple arithmetic averagsound levels. Summations are made ing. Logarithmic averaging sources of by combining the antilogs of individsubstantially differing db levels biases ual sound levels, taking the logarithm the result toward the higher source of the sum times 10, according to the level. following general formula: ∑db = 10 LOG ∑ (ANTILOG db )= TABLE 1. SOUND POWER & SOUND POWER LEVELS 10 10 LOG ∑ (10 db )

40

10

ual sources, regardless of the individual sound levels. Adding sources which differ substantially yields a total which is nearly equal to that of the larger source. The smaller source does not appreciably affect the total. For example: 100 + 100 = 103 90 + 90 = 93 100 + 99 = 102.5 100 + 90 = 100.4 100 + 90 + 90 = 100.8 100 + 100 + 100 = 104.8

DOING THE MATH ON NOISE

110

30

source definition consists of frequency and directivity information in addition to sound power. Instrumentation to directly measure sound power, however, does not exist. Sound measurement resorts to sensing of indirectly related sound pressure, which is readily done with microphone-type instrumentation. Sound pressure levels are highly dependent upon the acoustic character of the measurement and upon the surroundings, i.e., measurement distance, directivity of the source, room reverberation, sound reflections, and potential standing waves. A given power source will produce sound pressure levels which vary substantially depending on the acoustic environment in which the measurements are made. Users, of course, are interested in the sound which will result when a machine is placed in their particular environment and must have sound power (PWL) information in order to make sound estimates pertaining to their installation. On the other hand, OSHA noise limitations are based on readily measured sound pressure levels (SPL). So you can see the need for translation between sound power and sound pressure. The relationship for a typical semi-anechoic test chamber is shown below: PWL = SPL + 14 (db)

NATURE

Addition Two identical sound sources added together yield a total sound level which is 3 db above the individThe Pump Handbook Series

PWL

W (Watts)

100 90 80 70

.01 .001 .0001 .00001

Reduction Below PWL = 100 0% 90% 99% 99.9%

FIGURE 2

TABLE 2. SOUND LEVEL REDUCTION & PERCEPTION TO EAR

RELATIVE RESPONSE (dp)

+10

BLADE PASS RANGE

0 -10 -20

db Reduction

Subjective Effect

Power Reduction

3 5 10

Just Perceptible Clearly Perceptible Half as Loud

50% 68% 90%

one millionth of a pound per square inch of sound 10 10 10 10 10 pressure. The percentage HZ. dbA Weighting Scale column (Table 1) also sheds light on the relationship between db reduction and 100 + 100 = 100 sound power levels. Asking for 20 db 100 + 99 = 99.53 noise reduction is equivalent to ask100 + 90 = 97.4 ing for 99% sound power reduction. 100 + 80 = 97.0 You also need to consider the effect on the people working in an NOISE REDUCTION area where you are attempting to Now that we’ve got all the math reduce noise. Shutting down one of out of the way, let’s see how this two identically noisy machines obviapplies in the real world. For examously reduces sound power to half ple, two identical sources added the original noise power level, but it together increase the noise level by 3 is just as perceptible to the ear (Table db as we saw in the logarithmic addi2). A 10 db reduction in machine tion, while 10 sources added together noise is interpreted by the ear as increase the total noise by 10 db. So if being about half as loud, so 20 db a battery of 10 identical pumps pronoise reduction or 99% sound power duces a total of 100 db, shutting reduction is perceived by the ear as down 9 of the 10 units only cuts the being 1/4 as loud. noise down to 90 db. Of, course, you want to avoid To gain a better appreciation for damaging the ears of the people who noise, you need to know about the work in the plant. OSHA prescribes trivial amount of power associated exposure limits based on the “A” with some common machine sounds. weighting scale (dbA) which reflects Precious little energy escapes into the ear sensitivity as a function of sound air in the form of noise, even at noise frequency (Figure 2). OSHA limits levels deemed grossly unsatisfactory worker exposure to 90 dbA for 8from the human standpoint. Note on hour daily exposure, but often pump Table 1 a relatively loud noise of 90 users ask for 85 dbA, thus allowing db represents only 1/1000 of a watt of the use of three units simultaneously energy. In fact, the threshold of hearat 85 dbA. Or this leaves a 5 dbA ing (0 db) is considerably less than cushion for an application in the pump user's FIGURE 3. NOISE LEVELS IN SEALLESS PUMPS acoustic environment. -30 -40

SOUND LEVEL (db)

100

2

VIP-801

80

3

4

20 GPM 260 FT. 9 KW

5

dbA 59.5

WHAT CAN YOU DO ABOUT NOISE?

Noise in pumps comes from primarily two sources — the impeller blades and the motor fan. Traditionally, the way most pump users

59

60

49

47

50

52

55 43

40

43

39

20

0

0 31

63

125 250 500 1K 2K 4K FREQUENCY, HZ

8K 16K

20K

The Pump Handbook Series

deal with noisy pumps is to put them in the doghouse. These sound absorbing enclosures reduce noise by 10 db, or cut the sound in half (again, sound is measured logarithmically). But what if you need to reduce your noise even more? New design and technology can help. Pump noise that impacts the “A” weighting the most, generally comes from blade pass energy. Blade pass frequency is given by: Where Z = number of blades, and N = rpm. Through the typical speed range and impeller number of blades encountered, the predominant frequencies generally range through the curve peak region in Figure 2. The amount of overall noise is reduced by altering the blade pass frequency in Table 3 for a gear driven centrifugal pump. The more impeller blades your pump has, the lower the noise from the pump. On this pump we compared an 8-vane impeller to a 24-vane impeller. Note the dramatic 11 db improvement with the increased number of vanes. Unfortunately, most impellers in service today have 6 to 8 vanes. Using more vanes does not appreciably decrease efficiency in the pump and is usually not much more expensive. These impellers are more difficult to cast, though, and they can sometimes aggravate cavitation problems in the pump. Canned motor sealless pumps offer another opportunity to reduce noise levels in installed equipment. These pumps are quieter because there is no motor fan, which can contribute up to 10 db. If we didn’t have to put motor fans on pumps, our plants would be significantly quieter (Figure 3). Note that the overall noise level is about 60 dbA, not much more

23

TABLE 3. NOISE VS. BLADE PASS ENERGY Frequency Range (hz) 8-Blade 24-Blade

31 69 60

63 89 78

125 77 77

250 75 69

500 74 69

1K 74 66

2K 74 65

4K 77 63

8K 84 72

16K 80 68

dbA 86 75

than the background level in a typical semi-anechoic noise test chamber. You could also reduce noise by 3 db if you used a 24-blade impeller in this pump. With all the possible improvements in pump noise reduction, soon you’ll not be able to hear if your pump is running! ■ William Mabe is Manager of Engineering Services at Sundstrand Fluid Handling in Arvada, Colorado. He also serves on the editorial advisory board for Pumps and Systems.

24

The Pump Handbook Series

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R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

Addressing Pump Vibrations, Part I BY W.E. (ED) NELSON xcessive pump vibrations cause premature failures of pumping systems. Unless the root cause of the vibration is eliminated, replacing failed parts will buy time but the failure will recur. Only with a good understanding of the symptoms of hydraulic instabilities will rotating equipment engineers and operations and maintenance staff be able to detect impending failures— and fix problems before their systems fail. Although complete engineering guidelines are not available for determining the root causes of many pump problems, a few basic considerations can be a good beginning to a predictive maintenance program.

E

REQUIRES SYSTEMS APPROACH Pump vibrations underscore the fact that a pump is only part of an overall system. Your search for causes of vibration and your remedies to those problems must encompass the entire system. For the purpose of problem solving, a typical pumping system can be divided into eight components: • foundation • driver • mechanical power transmission • the driver pump • suction piping and valves • discharge piping and valves • instrumentation for controlling pump flow • alignment anchoring devices. Any one of the components can cause or accentuate vibrations. For example, excitations from the vibrations of the driver (motor, steam turbine, gearing) can be transmitted to other components. Excitations can arise in the coupling area, especially due to misalignment of the driver or eccentrically bored coupling hubs. Distance between shaft ends (DBSE) in excess of the axial flexing limits of the coupling can trigger vibrations. Adverse flow conditions such as cavitation, intake vortex or suction recirculation can cause flow disturbances. Unfavorable dynamic behavior of piping can arise because of loads

from dynamic, static or thermal causes including resonance excitation. Many of these problems involve the mechanical nature of the pump. But hydraulic problems are just as common. While pump handbooks and manufacturer’s guides provide useful advice about mechanical aspects, they don’t often discuss hydraulic aspects. When they do, it’s usually in the form of technical data, tables of pressure drops and conversion charts. Such information provides little practical value for selecting, operating and maintaining centrifugal pumps. Here are three brief, practical guidelines that will help you address hydraulic vibration problems: 1. Select and operate centrifugal pumps at or near the manufacturer’s design rated conditions of head and flow. This is usually at the point of best efficiency (BEP). Pump impeller vane angles and the size and shape of the internal liquid flow passages can only be designed for one point of optimum operation. For any other flow conditions, these angles and liquid channels are either too large or too small. 2 Do not operate at excess capacity. Any pump operated at a flow significantly greater than BEP and at a lower head will surge and vibrate, causing bearing and shaft problems and excessive power consumption. 3. Do not operate at reduced capacity. When operating at flow significantly less than BEP and at a higher head, the now incorrect vane angles will cause eddy flows within impeller and casing and between the wear rings. The radial thrust on the rotor will increase, causing higher shaft stresses, increased shaft deflection and potential bearing and mechanical seal problems. Coupled with increased radial vibration and shaft axial movement, continued operation in this mode will accelerate the deterioration of the mechanical and hydraulic performance of the pump.

VIBRATION MONITORING The systems approach involves The Pump Handbook Series

the consideration of part of the system as well as the hydraulic and mechanical aspects of design and construction. The hydraulic and mechanical aspects are interrelated. They cannot be separately evaluated or corrected without disastrous results. Also, all three phases of predictive maintenance — surveillance/monitoring, analysis/diagnosis and correction/remedy — must receive equal emphasis. Many monitoring programs get bogged down in the second step because of problems in interpreting the data. And many times the third step consists of a recommendation to replace the journal bearings, which treats the symptoms but not the root cause of the vibration trouble. The objective is to correct the problem, not just collect data and establish records. Knowledge of the pump’s hydraulic design features and its mechanical condition is necessary to accurately evaluate any vibration data. In pumps (unlike turbines and compressors), hydraulic effects are important. The pumped fluid has a large mass compared to the rotor and the fluid transmits any pressure pulsations throughout the system undiminished. The rotor and casing interact at off-design flow and can cause confusion in analysis of the vibration patterns of the rotor. Other system interactions can complicate the vibration analysis. Multistage pumps can produce confusing patterns of vibration forces occurring in the wear rings and the thrust balancing drum has a great influence because of its large area and high pressure differential. Tangential forces acting on the rotor can lead to self-excited vibrations. A spectral analysis of the vibration shows peaks related to specific causes of vibration (Figure 1).

RECIRCULATION EFFECTS Liquid flow in the internal channels of the impeller and casing is a complex phenomenon, especially at off-design conditions. Flow is not uni-

25

0.5

1.0

Vane Passing

Misaligned

Hydraulic Unbalance Misalignment

Dynamic Unbalanced

1

Oil Whip

2

Suction Recirculation

3

Oil Whirl

Shaft vibration amplitudes

FIGURE 1. SPECTRAL ANALYSIS OF VIBRATIONS

Ed Nelson is a consultant to the turbomachinery and rotating equipment industries. He serves on the Pumps and Systems Editorial Advisory Board.

2.0

3.0

Motor vibration frequency/machine speed

form. Instead, it is characterized by instability and violent changes including stall, back-flow, eddy-type circulation, turbulence and cavitation. These instabilities (i.e., recirculation flows) are always present in a pump and are exacerbated by offdesign flow rates. Severe recirculation damages pump components: recirculation at the impeller eye damages the eye (the inlet areas of the casing); recirculation at the impeller tips alters the outside diameter of the impeller; recirculation around impeller shrouds damages thrust bearings. Recirculation also erodes impellers and causes failure of mechanical seals and bearings. When recirculating, the liquid has multiple opportunities to restrike the vane surfaces before being discharged. The resulting noise subsides with increased suction head make it easy to confuse recirculation with bubble cavitation. This is unfortunate because recirculation can be more destructive than cavitation. The tendency toward recirculation flow strongly influences the minimum flow under which a pump can operate. Traditional pump curves assume uniform flow and thus can be misleading and even invalid. Watch out for recirculation problems with: • low NPSHR or high suction specific speed impellers. Because of highly undesirable flow characteristics, the only reliable way to predict trouble in these conditions is to calculate suction specific speed. Suction specific speed is defined as: N SS = [(rpm) * (gpm/eye) 1/12 ] / (NPSHR)3/4.

26

tion. Cavitation that occurs during recirculation is of mechanical origin and is not necessarily related to temperature. If application permits little or no turndown, you may need to maintain a very high NPSHA or limit your operations to near BEP to prevent problems with recirculation. ■

Failure to understand the problems of recirculation, minimum flow and hydraulic stability can lead to severe maintenance and operating troubles. One study showed that over a five-year period pumps with high suction specific speeds (i.e., greater than 11,000) failed at twice the rate of pumps with lower specific speeds. For cold water and general service applications, suction specific speeds should be 8500 or lower; for boiler feed and condensate applications and general hydrocarbon service, suction specific speeds should be 8500 to 11,000. Pumps designed for suction specific speeds in excess of 12,000 are generally for special applications only. • double suction pumps. Dimensional inaccuracies inherent in sad casting of impellers leads to both mechanical and hydraulic imbalances. The imbalances are magnified by the double suction configuration. (By the way, hydraulic imbalances cannot be eliminated by mechanical balancing of the impeller.) • energy levels above 650 ft of head and 250 ft per stage. These are the lower levels for really serious recirculation problems but smaller pumps can also be affected. • water or other narrow boiling range liquids. Pure liquids such as water are homogeneous and vaporization can occur instantaneously. In addition, water has a high vapor-to-liquid volume ratio. Mixed chemical or petroleum liquids composed of fractions that vaporize at different temperatures and pressures will have much less violent cavitation. But don’t assume that these latter liquids aren’t susceptible to cavitaThe Pump Handbook Series

The Top VibrationRelated Pumping Problems 1. Improper suction conditions (e.g., NPSHR > 12,000). 2. Recirculation caused by pump not properly designed for application 3. Vane pass frequency vibrations caused by improper Gap B. 4. Pump not accurately aligned for actual heat rise of pump and driver and pipe strain. 5. No attention to coupling fitting, balance or keys. 6. Improper selection of seals, fitting and flush/cooling. 7. Improper lubrication (e.g., too cold, water ingestion and condensation). 8. Improper selection of baseplate and grouting procedures. 9. Improper maintenance procedures in shop; bearing selection/ fitting wrong. These problems can be expensive. The cost of pumps exceed that of other rotating equipment five to one. By Charles Jackson, P.E., Turbomachinery Consultant, Texas City, Texas..

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R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

Addressing Pump Vibrations, Part II BY W.E. (ED) NELSON perating a pump at reduced capacity is another action to avoid if you want to limit vibration. What is reduced capacity? No fixed rules are available to answer that question. In fact, some disagreement exists as to why minimum flow is required. API 610 has used two definitions of minimum flow. One is continuous thermal flow — “the lowest flow at which the pump can operate and still maintain the pumped liquid temperature below that at which NPSHA equals NPSHR.” The other is continuous stable flow — “the lowest flow at which the pump can operate without exceeding the noise and vibration limits imposed by this standard.” Continuous thermal flow is almost meaningless as a guideline for preventing reduced capacity problems. Suction and discharge recirculation, impeller erosion, seals and bearings damage and cavitation occur long before thermal flow limitations

come into play (Figure 1). Minimum continuous flow is the more useful definition. Greater minimum flow rates are necessary for meeting continuous flow requirements than for thermal flow requirements. Other practical considerations that influence minimum flow rates include: • Time. Is the low-flow condition continuous, a momentary emergency condition or something in between? • Design for low NPSH (high specific suction speed). These pumps, which are prone to recirculation cavitation in the impeller inlet, require higher minimum flows. • Nature of the liquid. Water can vaporize almost instantaneously, so minimum flow must be higher.

O

B.E.P.

DISCH. RECIRC.

SUCTION RECIRC.

REDUCED IMPELLER LIFE

LOW BRG. AND SEAL LIFE

LOW FLOW CAVITATION

HIGH TEMP. RISE

HEAD

FIGURE 1

FLOW Head flow characteristics (S. Gopalakrishnan, 5th International Pump Users Symposium)

The Pump Handbook Series

•

Double suction pumps. Since these pumps are more susceptible to recirculation, they require minimum flows of 60% to 80% of BEP. These additional guidelines will help you avoid damage caused by low flow rates: •

Assume that no pump is designed to operate for longer than about 15 minutes below 50% of BEP flow. • Provide minimum flow bypass systems for high-energy pumps. • For impeller patterns with suction specific speeds greater than about 11,000, insure a minimum flow of 60% to 70% of BEP. • Realize that pumps with suction specific speeds greater than 20,000 may require a minimum flow of 100% of BEP. • Be sure that pump liquid channels (Gap A,Gap B, discharge impeller-diffuser overlap) have adequate overlap and positioning. Any single percentage number for minimum flow is arbitrary. These guidelines can provide smoother operations for pumps that operate off design, but they may not eliminate all risk.

OTHER HYDRAULIC VIBRATIONS One major cause of vibrations is vane passing, a hydraulically induced vibration at a frequency determined by the number of impeller vanes, the number of stationary vanes and the pump rpm. This type of vibration is created by the momentary disturbance of the wake of the liquid exiting the impeller liquid channels by the stationary diffuser or volute vane tips. The larger Gap B, the more the unguided flow path can smooth out before it contacts a diffuser vane or volute tongue. This increases the

27

possibility of the flow reaching a more favorable inlet path into the diffuser or volute, especially at design capacities. Determination of the vane passing frequency sounds easy, but it can be confusing. The most common pump design has an impeller with an odd number of vanes and a double volute in the casing. The vane passing frequency in this pump is the rotational speed times the number of vanes.

Even combinations of impeller vane and volute or diffuse vanes can induce very high forces that can fracture vane tips, impeller shrouds and volute tongues. Diffuser pumps have a larger number of diffuser vanes and a narrower Gap B than volute pumps — and a different interaction of the rotor and casing. Vane passing frequency in diffuser pumps does not correspond to the number of vanes in the

Accurate Computations for Preventing Cavitation in Reciprocating Pumps Cavitation is a major cause of damage in reciprocating pumps. It occurs in pumping systems when the negative peak of the dynamic pressure wave, added to the steady state pressure, approaches the vapor pressure of the liquid. Acceleration head calculations, the computations typically used to calculate the additional suction head required to prevent cavitation, assume quasi-static conditions of flow. These calculations are inadequate because they ignore the dynamic acoustical response characteristics of the fluid. To be accurate, calculations of pulsation levels in pumping systems must consider the dynamic flow by taking into account all the parameters which significantly influence the system, including the pump fluid end, the pump valves and the associated suction piping. To obtain more accurate computations for suction head requirements, we simulated pump systems using a computer program that predicted pulsations and onset of cavitation. Pumps and suction piping form a complex acoustical system with many natural acoustic frequencies that can be excited by the flow modulations generated by the pump. A reciprocating pump generates pulsations at integer multiples of the pump speed with the highest amplitude components normally at the plunger frequency and its harmonics. These harmonics can be amplified by the acoustic frequencies of the system. Amplification factors are typically 10 to 40 for pulsation resonances. Thus, when excitation frequencies and an acoustic natural frequency of the system coincide, the pulsation can become amplified and result in severe cavitation in the pump manifold and suction piping. Our model calculates the pulsation amplitudes at any point in a piping system and evaluates the potential for cavitation in the system. It lets designers evaluate the effectiveness of any given accumulator or modification of the piping. Designers can examine the complex wave and spectral analysis over the speed range and accurately assess NPSH before installing a pumping system. For example, in one study we tested the effect of numerous configurations of accumulators on cavitation. Accumulators can reduce suction pulsations to more reasonable levels, but we found that selection of these devices without regard to system acoustics can actually worsen the situation. We also found that adding an orifice at the location of the accumulator can reduce pulsation amplitudes and the potential for cavitation in the manifold. The orifice size is critical and should be optimized using proven simulation techniques. For full details of the simulation, refer to our paper “The effects of pulsations on cavitation in reciprocating pumps,” presented at the Energy-Sources Technology Conference and Exhibition, Houston, TX, January 22 - 25, 1989. J.C. Wachtel, J.D. Tison and S.M. Price, Engineering Dynamics, Inc., 16117 University Oak, San Antonio, TX 78249.

28

The Pump Handbook Series

impeller or casing. In multistage volute pumps, the degree of positioning of the volutes is severely limited by case design. It is necessary to cut the keyways in the impellers randomly to assure that vanes on adjacent impellers are not aligned and do not pass volute tongues simultaneously. Manufacturers don’t always position keyways randomly. Instead, they may leave impellers lined up. The result is high vane passing frequency vibrations. Be sure to check the alignment of impellers in witness testing, reassembly of a new pump or replacement of impellers during maintenance. Blunt vane tips can also cause problems by generating hydraulic disturbances in the impeller exit wake area, even with correct Gap B. Careful sharpening of the impeller trailing edge can reduce this cause of vibration (see “Ask the Experts,” Pumps and Systems, January 1993). Vortexing is another source of hydraulic vibrations. Vortexes can occur when the NSPHA falls below NPSHR. They can be very damaging because they effectively reduce NPSHA. Minimal suction head is likely to result in air entrainment, which reduces pump capacity further. As little as two percent air or gas in the liquid can reduce pump capacity by 10 percent. This information about hydraulic vibrations should help you address your vibration problems. Remember, though, that hydraulic vibrations are only part of the vibration puzzle. Mechanical vibrations are also important (especially in variable speed pumps and vertical submerged pumps). But mechanical vibrations seem easier to quantify and are generally covered thoroughly in pump handbooks and manufacturers information. With those resources and a clearer understanding of hydraulic vibrations, you’ll be much better prepared to identify the causes of vibration-induced problems and to eliminate them, not just patch over the symptoms. ■ Ed Nelson is a consultant to the turbomachinery and rotating equipment industries. He serves on the Pumps and Systems Editorial Advisory Board.

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R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

Lubrication Systems for Rotating Equipment BY SCOTT HANSEN AND SAM HOPKINS

A

•

heat load (Btu/h)

•

pressure required at lubrication system outlet

•

inlet temperature of the cooling media

•

oil temperature at lube system outlet

•

location of the lubrication system

•

available utilities

•

applicable specifications

A number of industry specifications are available to help operators select a lubrication system. All API specifications relating to rotating equipment have sections on lubrication. The most comprehensive API specification is API-614, which covers lubrication systems for special services. Many customers incorporate some API specifications. The following list provides available API specifications for rotating equipment, particularly those that require a selfsupporting force fed lubrication system: API 610 Centrifugal Pumps for General Refinery Services API 611 General-Purpose Steam Turbines for Refinery Services

API 612 Special-Purpose Steam Turbines for Refinery Services API 613 Special-Purpose Gear Units for Refinery Services API 614 Lubrication, Shaft Sealing & Control Oil Systems for Special-Purpose Applications API 616 Combustion Gas Turbines for General Refinery Services API 617 Centrifugal Compressors for General Refinery Services API 618 Reciprocating Compressors for General Refinery Services API 619 Rotary-Type Positive Displacement Compressors for General Refinery Services API 672 Packaged, Integrally Geared Centrifugal Plant and Instrument Air Compressors for General Refinery Services A lubrication system should include the following at the very least: Reservoir - This is the “storage tank” that contains enough oil relative to retention time. Main Shaft Driven Pump - This can be any style pump that is used continuously to force the oil to flow around the system. Auxiliary Motor Driven Pump This pump is normally identical to the main shaft driven pump, but is only used for start ups and emergency situations. Relief Valve - This valve can be used for two different functions: 1) as a safety relief valve, which will open with a build up in line pressure, and 2) to reduce the flow in the line to the required level (normally not a good design practice). Cooler - This unit reduces the temperature of the oil to the required

level. The cooling media are usually water or air. In some cases the cooler has been used to heat the oil during a cold start-up situation by using steam as the medium. Filter - This is to ensure no foreign particles can enter the main equipment’s bearings. The filter will take out particles smaller than what is visible by the human eye. Pressure Switches - These are used as a “safety” precaution, and usually involve alarm type situations in which the system falls below the required pressure level or rises above the preset level. See an example of pressure set points shown later in this text. Pressure Indicator - This is a visual check to monitor pressure reduction as oil flows through the system (coolers, filters etc.). Temperature Indicator - This works the same as the pressure indi-

PHOTO COURTESY OF LUBE SYSTEMS, INC.

lubrication system is the lifeblood of any piece of rotating equipment. It is the most important system for keeping machinery operating effectively for extended periods. By supplying cooled, clean oil to the surfaces, a lubrication system eliminates friction and heat build-up, and reduces excess wear. When selecting a lubrication system for large rotating equipment, you have a number of options. However; sometimes it can be very difficult to specify a system that is economically feasible, reliable, easily maintained and long lasting. The following information is necessary to size a lubrication system correctly: • oil flow rate to bearings/rotating shaft

Lubrication system for API 610 horizontal pumps

The Pump Handbook Series

29

Below is a description of each main component in API-610 and information on how to specify the correct size, and key considerations needed for sizing.

FIGURE 1. TYPICAL PRESSURIZED LUBE-OIL SYSTEM PSLL

PSL B

PSL A

PI

RESERVOIRS

Driver

PDI Duplex filter

Pump T1

T1

T1 F1

T1 F1

F1

Shaftdriven oil pump

T1

Heat exchange

P1

P1

T1

Motor-driven auxillary oil pump Oil reservoir

F1 T1 PI PDI PSL A PSL B PSL L

Flow Indicator Temperature Indicator Pressure Indicator Pressure Differential Low-pressure switch (auxilliary pump start) Low-pressure switch (alarm) Low-pressure switch (trip)

Instrument (letters indicate function) Gate valve Relief valve Line strainer Pressure control valve Check valve Block and bleed valve Reflex-type level indicator

cator except it monitors temperature through the system. Interconnecting pipe. When specifying a non API standard system, the specifications should be written around the components stated above and additional consideration given to the reservoir, materials of construction, location of pumps, cooling media and console location.

Adequate capacity is the most important factor when sizing a reservoir. The most common mistake is to undersize this component. The API610 specification requires 3 minutes retention time. Note: Retention time and total capacity do not mean the same thing. To size for “total capacity,” the following factors must be considered and then added together to arrive at the final number (Figure 2): 1. capacity for the rundown level 2.

capacity for the maximum operating level

3.

capacity for the minimum operating level (retention level)

4. capacity for the suction low level When a system’s capacity requires a low flow, the required volume of the reservoir will be small. This makes it difficult to design for proper access for manufacture, cleaning and maintenance. With a small reservoir, other equipment can not fit on top. In this situation, capacity does not become the important factor, but rather the access

Let’s focus on API-610, 7th edition, and how this specification can be applied to a lubrication system design. In particular, section 2.10 discusses lubrication. Figure D.6 shows the P&ID for this lubrication system as shown in Figure 1. Figure D.6 details the bare minimum components required to meet the API specifications. This specification requires all pressure retaining parts to be carbon steel, all piping to be stainless steel (including the reservoir) and all piping to be fabricated to ANSI B31.3.

30

PHOTO COURTESY OF LUBE SYSTEMS, INC.

API-610, 7TH EDITION

API 614 unit

The Pump Handbook Series

FIGURE 2. RESERVOIR LEVELS AND OIL LEVEL GLASS DETAILS To suit glass length (1 in Minimum) Rundown level

A

Maximum operating level

B C

Submerged positive displacement pump Submerged centrifugal pump

E

to the reservoir with all equipment in place. A baffle, separating suction and return sides, and proper sizing will control foaming. The longer the oil is in the tank, the less chance entrained air will remain in the oil. This will reduce the chance for any air reaching the suction line of the pumps. A larger reservoir is always a safer design approach. Reservoirs have many accessories—all require attention and any one can cause a problem if overlooked. These include: • level indicators and switches • temperature indicators • heaters • fill strainers • foot valves • baffles • sloping bottom • drain connection • return connection • suction connection • access holes

Since the 10% for valve set and 10% for the accumulation is lower than 25 psig, we will use the straight 25 psig. Total pump pressure required = 60.0 psig

Suction level

D Alternative pump suction arrangements

Manufacturer's Standard Glass Length

Minumum operating level

49.50 x 1.1 = 54.45 psig

Pump Suction Level 1

outlet flow by using a back pressure regulator. Example Required Outlet flow: 3.5 gpm 3.5 X 1.15 = 4.02 gpm or 3.5 x 1.20 = 4.20 gpm When required flows are low, a positive displacement pump (3 screw or gear) is the better selection. Where required flow rates are over 100 gpm, a centrifugal pump should be considered. The excess flow will be returned to the reservoir via a back pressure regulating valve. This excess flow allows for control of the system during various flow changes while in operation. Pressure is the next consideration: System outlet pressure : (assumed)

The above calculation does not 2 in. minimum take into consideration the oil pressure being higher than the cooling water pressure. It is safer to design a system with the oil pressure being higher than the cooling water to avoid any water entering the lubrication oil system if a leak in the cooler (heat exchanger) develops. The next calculation takes into consideration the cooling water pressure. Cooling water pressure: 90 psig (assumed) Required pressure of oil: 10 psig Cooler differential pressure: 5 psid Total: 105 psig

FIGURE 3. FILTERING SYSTEM FLOW Flow Clean

15 psig

Differential pressure - dirty filter: 25 psid Cooler differential pressure: 5 psid

Flow Dirty

Flow Dirty

Total : 45 psig

PUMPS To determine the required flow rate from the auxiliary/main oil pump, the total outlet flow should be multiplied by a factor of 1.15 to 1.2. This allows for control of the

Rule: Plus 10% or 25 psig, whichever is greater Relief Valve Set : 1.10 x 45 = 49.5 psig Relief Valve accumulation additional: 10% The Pump Handbook Series

Filter Element

31

Rule: plus 10% or 25 psig whichever is greater Relief valve set 1.10 x 105 - 115.5 psig Relief valve accumulation additional 10% 115.5 x 1.1 = 127.05 psig Since the 10% for valve set and 10% for the accumulation is greater than 25 psig, we will use the calculated figure. Total pump pressure required = 127.05 psig Note: Due to the cooling water pressure, the filter differential pressure and the outlet pressure are not a consideration in this calculation. The schematic (API-610, Figure D6) indicates the main and auxiliary pump is external from the reservoir, but allows for a vertical cast iron pump to be submerged in the oil reservoir. There are advantages and disadvantages of external and vertical pumps (Tables 1 and 2).

MOTOR The required horsepower should be selected based on the pressure of the relief valve setting, plus the accumulation of an oil viscosity at 70°F (650 ssu). Remember, a little more horsepower will not be harmful, but not enough could be drastic.

COOLER One cooler is required. Basic tube and shell coolers are allowed by specification, including carbon steel, admiralty tubes, cast iron bonnets, carbon steel tube sheet, and nonremovable bundle (fix tube sheet design). The cooling water flows through the tubes and the oil through the shell. A fix bundle means the tubes must be reamed to be cleaned. If the cooler is fouled or plugged and cannot be cleaned, then it must be replaced. Generally the tubes are 1/4 inch or 3/8 inch outside diameter. Other tube sizes can be specified. Placement of the vents and drains needs to be considered when locating this type of equipment.

FILTER DUPLEX All carbon steel construction with a built in transfer valve is the basic design. Filtration is generally

32

10 micron nominal. Flow through the element is from outside to inside, leaving all contaminates entrapped on the outside of the element (Figure 3). The element can be manufactured in many types of media, but the most common is pleated paper, which allows for disposal after service. Clean differential pressure sizing should not be higher than 5 psid at 150 ssu. Differential pressure design for disposable (replaceable) elements is 20 to 25 psid at a viscosity of 150 ssu. Collapse pressure is 75 psid. Cold start up, meaning high viscosity of the oil, can cause damage to the filtering element, deformation or failure. It is necessary to understand the problems that can be caused due to cold start-up. Consideration is required for the venting, draining and cross-over (equalizing) connections. The size of the pressure vessels (filters and coolers) depends on whether the component meets ASME VIII, Div. 1.

PRESSURE SWITCHES Three switches are required per specification. Their functions are to determine any high or low pressures in the system (Table 3). If a low pressure registers, the auxiliary pump will be initiated, and if there is a high in the system the auxiliary pump will shut down. Each switch housing consists of two independently adjustable, single pole, double throw contacts. This allows for six set points. The auxiliary pump should have a manual “shut off” and should not be done automatically. If the auxiliary pump comes on, this means there is a problem in the system that should be addressed immediately. Manual shutdown of the auxiliary pump is recommended to avoid automatic cycling.

PIPING AND VALVES All piping and valves upstream of the oil filter should be stainless steel, socket weld. Flanges must be used to break the piping. All piping downstream of the filter should be with butt-weld connections. All welding should to be performed by a certified code welder. Piping arrangements must be designed for ease of maintenance. All piping must be true The Pump Handbook Series

and “plumb.” Flange breaks should be used to allow any component to be removed for maintenance.

GENERAL CONSIDERATIONS Location of the lubrication system in relation to the main rotating equipment is very important. Many systems are placed directly on the same baseplate as the pump and drive system. This allows for all interconnecting piping to be established, and incorporated, in the design for manufacture in the factory, which minimizes any extra piping to and from the shaft driven main oil pump. Always: 1. design with the minimum pipe length to the shaft driven suction line to ensure the lowest suction loss possible. 2.

use a foot valve in the suction line to ensure a “prime” situation.

3.

check the suction lift capacity of the shaft driven pump.

4.

prime the main pump suction line prior to start up, to assist suction lift.

The rotating equipment dictates the return line height into the reservoir. If the return line has a low centerline the reservoir has to be made wider and longer to maintain the same required volume. This is an additional point to consider when designing the overall dimensions of the reservoir. All accessory equipment is mounted on or around the oil reservoir. Once the layout has been determined, the design is incorporated into the main rotating equipment baseplate. For accessibility, consideration should be given to the filters, pumps, cooler, switches and instrumentation. Another alternative in incorporating the lube system with main rotating equipment baseplate is to offer a free standing console. From a maintenance stand point this is a preferred method. However, the option requires additional space and interconnecting piping. When locating the free standing console, consider the NPSH required to the shaft driven pump.

RECOMMENDED CHANGES TO API-610, 7TH EDITION

TABLE 1. VERTICAL CAST IRON PUMP ADVANTAGES: ▲ Less cost ▲ No mechanical seal required ▲ Less space required DISADVANTAGES: ▼ Suction strainer in reservoir can’t be cleaned

without removing the pump from the reservoir

1. Add gate valve to pump suction line auxiliary oil pump if flooded suction, since removing the pump will mean draining the complete reservoir. 2.

▼ Can’t do a positive visible check on the

coupling engagement ▼ When servicing a vertical pump, the reservoir

access hole must be covered so as not to allow foreign matter (dust, dirt, water, etc.) in the reservoir

TABLE 2. HORIZONTAL STEEL CASE OIL PUMP ADVANTAGES: ▲ “Y” strainer is external of the reservoir to ease maintenance for cleaning ▲ Can do a positive check visible on coupling engagement DISADVANTAGES: ▼ More expensive ▼ Mechanical seal required ▼ More space required ▼ When serving this style pump, Figure 1 does not require a gate valve on the suction side of the pump. This should be an option because if repair is required on the pump, the complete reservoir will have to be drained

Relocate back pressure regulator to upstream of the cooler, this will allow the excess oil flow to return to the reservoir prior to filtration and cooling, and thus, enable smaller sizing of the piping, cooler and filter. It will also reduce the transient condition, and excessive back pressure during the pump run down (shaft driven main and motor driven auxiliary).

3.

Always specify a heater. A heater is obviously useful in cold conditions, however, it is also useful in a warm environment to evaporate condensation.

4.

All piping should be welded. Screwed piping is subject to leaking, which often occurs during shipment due to traveling vibration.

CONCLUSION The last few years have brought a new emphasis on improved industrial productivity, requiring maximum performance from pumping systems. But placing a premium on increased machinery operation requires advanced lubrication fluids and systems. Today, a number of specialized manufacturers of lubrication systems have extensive experience serving customers in the process industries. They can help operators meet manufacturing demands by ensuring that the quality of your lubrication system matches the quality of your industrial processes. ■ Scott Hansen, President of Lube Systems, Inc., Newton, NJ., has been involved with lubrication issues for twenty years. Sam Hopkins, International Sales Manager of Lube Systems, Inc., has worked in fluid dynamics for over 15 years. He holds a mechanical engineering H.N.C from Reading College of Higher Technology in the United Kingdom.

TABLE 3. SWITCH SETPOINTS WITH FUNCTIONS 1. Pressure Signal “Low” (PSL)

10 psid falling Start Aux. Oil Pump

2. Spare 3. Pressure Switch “Low” “Low” (PSLL)

8 psid falling Shut Down System

4. Spare 5. Pressure Switch “High” (PSH) 6. Pressure Switch “High” (PSH)

18 psig rising Aux. Pump Alarm 15 psig rising Permissive Start of Main Equipment

The Pump Handbook Series

33

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R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

Lubrication Oil Viscosity Classifications... Don’t Get Confused BY JERRY T. SHELBY election of the proper lubricant for pumps, or any equipment, depends on many factors. Speed, load temperature and ambient conditions must be taken into account when choosing the right lubricant. In most cases, the equipment manufacturer will have already made most of the critical decisions concerning the best lubricant for their equipment, providing a recommended list of several “approved” products. This list will, of course, indicate the oil company name and the product designation. However, over the past few years, some changes in the lubrication industry have lead to product nomenclature that can be confusing to those unfamiliar with these changes. Today’s lubricants are formulated from selected base stocks and a wide array of additives to enhance performance. Although the products may be conventional mineral oils (or synthetic oils), the most important aspect of their performance with respect to equipment is viscosity, the oil’s resistance to flow. The higher the viscosity, the thicker the oil. In an effort to standardize the way a product’s viscosity is designated, the lubricant industry has generally moved toward the ISO (International Standards Organization) viscosity system or grade numbers. Before the ISO system was introduced in the U.S., almost all manufacturers of lubrication oils gave each of their products a unique name along with a number. The number usually referred to the oil’s viscosity in Seconds Saybolt Universal (ssu) measured at 100°F or 210°F. To confuse the competition—which, of course, only confused users—the significant zero was sometimes left off the number. That is, instead of 150, it would be 15. Some companies further confused the issue by developing their own name and numbering system, which had no relation to the rest of the industry with respect to viscosity. As an example of this, consider the following oils, all essentially equivalent: Citgo Pacemaker® 15, Arco Duro® 150, Mobil DTE® Light, Texaco Regal® A, Exxon Teresstic®43, and Sunvis®

S

34

916. This made it practically impossido not update their manuals and lubrible for the average user to choose the cant recommendations as often as they correct viscosity without contacting should. So recommended products the oil company. Enter the ISO, in an may reflect the old name and numberattempt to end the product selection ing system and not the current ISO sysand viscosity confusion at the field tem. If, for example, the old manual level. called for a Phillips Magnus® 150, Based on the kinematic viscosity which had a viscosity of 150 ssu at (centistokes or cSt) at 40°C, the ISO 100°F, and the user ordered a Magnus system designates 18 grades from 2 to 150 from Phillips today, he would 1500 (Table 1). The number is the receive an ISO 150 oil, which is approxmidpoint of a viscosity range. For imately 750 ssu at 100°F, or five times example, an ISO 32 viscosity grade as heavy as the intended oil. Needless can range from 28.8 cSt to 35.2 cSt at to say, lubrication problems would 40°C. This gives the oil companies most certainly develop if the ISO vissome flexibility when formulating cosity oil were used where the ssu visproducts, since it would be impossicosity oil was intended. ble to have similar products made by Don’t just take the recommenddifferent companies meet an exact ed oil from a plate or manual. It may viscosity. However, all ISO 32 oils be old or dated. When in question, must fall into the range. If you have call the equipment manufacturer or trouble remembering which ISO visyour oil supplier to be sure that the cosity relates to which ssu viscosity, type of oil, and most important, the just multiply the ISO viscosity by 5, viscosity, is what was recommended for a close approximation of the old by the manufacturer for the specific ssu viscosity. application in the first place. Recognizing this as an effective Otherwise, expensive problems could method to end the product viscosity/ result in equipment downtime and name confusion, most companies failure. ■ renamed their products with ISO desJerry T. Shelby, is president of ignations. This assures users get the Lubrication Consultants, Inc. proper viscosity from company to company and product to product by matching the numbers. Of course, the names are differTABLE 1. VISCOSITY CONVERSION CHART ent, but within product groups, ISO Viscosity Kinematic Viscosity Saybolt Viscosity hydraulic oils, gear oils, turGrade cSt @ 40°C (104°F) ssu @ 104°F bine oils, etc., matching ISO 2 1.98-2.42 32 numbers will produce equiva3 2.88-3.52 36 lent viscosities. However, not all oil companies adopted the 5 4.14-5.06 40 ISO system. Most notable of 7 6.12-7.48 50 those that did not is Mobil Oil, 10 9.00-11.0 60 an industry leader. For what15 13.5-16.5 75 ever reason, the company 22 19.8-24.2 105 retained its unique product 32 28.8-35.2 150 name and numbering system 46 41.4-50.6 215 and, unless one knows the 68 61.2-74.8 315 Mobil format, it is difficult to 100 90.0-110 465 relate the Mobil product viscosities to the ISO viscosities of 150 135-165 700 similar products. 220 198-242 1000 Although the ISO system 320 288-352 1500 has eliminated most of the con460 414-506 2150 fusion that existed from one 680 612-748 3150 company’s products to another, 1000 900-1100 4650 there is still a caveat for the user 1500 1350-1650 7000 to consider. Some manufacturers The Pump Handbook Series

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R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

Probe Installation Tips BY LARRY COVINO roper proximity probe installation is critical to effective machinery monitoring, as improper installation may lead to an inaccurate indication of machine condition. When installing proximity probe systems, a user should consider the following: • Ensure the inter-compatibility of the transducer system components. These components must also be suitable for the application and operating environment.

P •

•

•

•

•

•

•

Tag and document system components to assist with future identification and maintenance. Check components electrically before and after installation to assure system integrity. Develop and maintain a probe orientation plan to assist with future servicing and troubleshooting of machine faults. Ensure the probe target area exhibits uniform conductivity and magnetic permeability and is free from residual magnetism and surface defects, such as scratches, dents, rust and corrosion. Identify the shaft material. If the material differs from the standard set for your proximity probe system, recalibration may be necessary. Check proximity probe target area clearance. Insufficient side clearance and/or available shaft surface will inhibit the accuracy of the signal by altering the scale factor and linear range. Figure 1 illustrates proper probe tip side clearance for a typical 8mm proximity probe. Provide adequate spacing between proximity probes to avoid signal interference known as cross-coupling. Figure 2 displays the minimum tip separation needed to prevent cross-coupling on an 8mm probe system.

FIGURE 1 8.9 mm Min. (0.35 in)

Mounting Surface

8.9 mm Min. (0.35 in)

Mounting Surface

17.8 mm Min (0.70 in)

8.9 mm Min. (0.35 in)

Side Clearance and Rear Clearance

Counterbore

Proper probe tip side clearance for 8 mm transducers

FIGURE 2 Radial Vibration Mounting

Radial Vibration Mounting

15.2 mm (0.6 in) Min

Radial Vibration Mounting 15.2 mm (0.6 in) Min

Shaft

35.6 mm (1.4 in) Min

40.6 mm (1.6 in) Min

Thrust Position Mounting

Shaft

Minimum tip separation needed to prevent cross-coupling on an 8 mm transducer

BENEFITS OF EXTERNAL MOUNTING The advantages of mounting probes externally include ease of probe adjustment and removal withThe Pump Handbook Series

out machine disassembly, and easily accessible, externally located connectors. Using external housing assemblies, as shown in photo 1, provides

35

FIGURE 3 O° TDC CL

n

r

Bearing strong back

Ve

Aluminum mounting block

Bearing upper half Flexible conduit

Safety tie wires

n tio ra ib lV ta °R on 45 riz Ho

io

at br Vi L l a ° tic 45

Safety tie wires

Bearing Cover

Rotor Connector protector Lower machine case Bearing babbitt

Junction box

Cable seal

Note: As viewed from driver end.

Bearing clearance

SPECIAL CONSIDERATIONS INTERNAL MOUNTING

Typical internal probe installation complete protection for probes and cables, reducing the risk of physical damage. Probes should be mounted through the casing when it is integral to the bearing. This will assure proper transfer of rotor-related vibration information and accurate radial position measurements to identify the rotor position within the bearings. In any case, measurements should be referenced to the rotor’s point of constraint (i.e., the bearings). This location will provide for proper machine protection and enhanced diagnostic information. Case or bearing mounted seismic transducers should be added to enhance, not replace, the use of the proximity probes.

SPECIAL CONSIDERATIONS EXTERNAL MOUNTING •

36

Unsupported probe sleeve lengths should be no more than 12” for 3600 rpm machines to prevent the sleeve from resonating and prompting false signals.

ative to the bearing. Finally, if there is significant pressure under the casing, externally-mounted probes are not suitable. In these instances, the probes should be mounted internally. A high pressure cable seal should be used as required with internally mounted probes. A typical installation is diagrammed in figure 3. On larger machines the outer housings are not integral components to the radial bearings. Consequently, probes should be mounted internally and referenced to the rotor’s point of constraint (i.e., the bearing). Probes used to measure radial vibration should be within 3” of the radial bearing. Trust position probes should be mounted within 12” of the trust bearing.

In general, make the probe sleeve as short as possible. •

In order to install external housings, the machine case must be drilled and tapped. Care should be taken when machining to ensure the probe tip will be normal to the shaft circumference after installation.

•

Sufficient clearance external to the machine is essential to allow for easy installation of the probe and sleeve.

• Ideally, when installing probes internally, they should be mounted in blocks bolted to the bearing retainer within 2” of the shaft to avoid

WHEN IS INTERNAL MOUNTING APPROPRIATE? On some machines, it is physically impossible to mount the probe externally. For example, in some configurations, the casings are not integral with the bearings and provide relatively poor vibration information. In addition, thermal expansion of the casing may change the position of the mounting location rel-

The Pump Handbook Series

External probe housing assembly

probe resonance and aid installation. •

Probe brackets should be sturdy in design to avoid resonance or failure. Bracket resonance should be 10 times machine running speed.

•

Brackets should also be positively secured using the tie wire or tab-locking washers to prevent the bracket or probe from coming unfastened.

•

Windage due to shaft rotation and/or oil spray may cause damage to cables if they are not properly restrained. So probe leads should be securely fastened inside the machine.

•

When routing the cable, the bend radius should never be smaller than specified, as a tight radius could alter the cable’s electrical characteristic and will damage the cable.

•

Cables should exit the machine below the horizontal split but well above the oil level. And, if possible, the cable should exit at a location that will allow the probe/extension cable connection to be external to the machine. External connections will facilitate calibration checks and troubleshooting.

•

Cable exit holes should be sealed with appropriate glands to prevent leakage of oil and other contaminants to the proximity housing. Never use an armored probe or extension cable to exit through a machine casing, as oil will leak out between the cable and the armored sleeve.

•

•

Apply the recommended connector protectors or the heat shrink tubing supplied for your systems to protect the probe to extension cable connection from contaminants and provide isolation from electrical ground. Never use electrician’s tape to protect the cable connection, as oil and heat may cause the tape to deteriorate and melt.

difficult otherwise to replace probes once the machine is operating.

Larry Covino is the Design and Installation Service Manager for Bently Nevada Corporation in Buffalo, NY.

See the accompanying discussion of Torque Measurement for a typical application of proximity probes.

Torque Measurement A rotating machine does not turn at constant angular velocity. Even under a steady state condition, angular and torsional vibrations, due to varying loads, driving forces or oscillations within the shaft itself, cause slight variations in the speed and twist of the shaft, respectively. High levels of angular or torsional vibration may seriously shorten the life of a machine. Shaft stresses and potential cyclic stresses in the gear teeth must be properly documented to provide accurate information about the life expectancy of a rotating system, based on the number of starts and stops of the motor. Torsional or angular vibration can be measured in several ways. One method employs two proximity or optical transducers to observe a reference mounted on the shaft. The reference is usually a precision manufactured gear or a reflective tape bearing evenly

spaced marks. The transducers are mounted 180° apart, in-line with the reference gear, to observe the time interval from one tooth to the next. If the shaft is moving at a constant velocity, there is no variation in the toothto-tooth timing and no angular vibration. If there is angular vibration, the tooth-to-tooth timing will change. The transducers will detect these changes and send the signal to a monitoring instrument. Knowledge gained from torque measurements may lead to fewer machinery failures and lower energy costs. Such approaches are particularly applicable in: • test stand/shop testing • on-site commissioning • on-line condition monitoring • performance optimization • research and development testing

End View

Top View

Transducers

Shaft

Gear

When installing probes internally, it is often prudent to install redundant probes as it may be The Pump Handbook Series

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Bearing Basics BY RAY RHOE ntifriction bearings, which can utilize either balls or rollers, are used to transfer radial and axial loads between the rotating and stationary pump and motor assemblies during operation. Even under the best of installation, maintenance, and operating conditions, bearing failures can and will occur. The purpose of this article is to provide a working-level discussion of bearings, the types of failures, and how bearings should be installed and maintained for optimum life expectancy. Due to space limitations, we cannot address all the different sizes and types of bearings available, or all the constraints currently utilized in design. However, because electric motors are used more often to drive centrifugal pumps, our discussion will be based on bearings typically used in quality motors. These bearings usually include a single radial bearing and a matched set of duplex angular contact bearings (DACBs). Together, these bearings must: • allow the unit to operate satisfactorily over long periods of time with minimum friction and maintenance • maintain critical tolerances between rotating and stationary assemblies to prevent contact and wear • transmit all variable radial and axial loads in all operating conditions, which include reverse rotation, startup, shutdown, maximum flow, and maximum discharge pressure Each bearing has a specific purpose. The radial bearing, which is located at one end of the motor, only transfers radial loads such as minor unbalanced rotor loads—and the weight of the rotor itself in the case of horizontally oriented components. The DACBs

A

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Photo 1. Typical radial bearings

must transfer radial loads at the other end of the motor, and they must transfer all axial loads. Photo 1 shows several typical radial bearings, and Photo 2 shows DACBs.

DIFFERENT BEARING CONFIGURATIONS Radial bearings may be provided with either 0, 1, or 2 seals or shields that are effectively used to prevent entry of foreign material into the bearings. If the bearing is equipped with one seal or shield, the installer should determine which end of the motor the seal or shield should face. Failure to install radial bearings properly in the correct orientation may result in the blockage of grease or lubricant to the bearings during routine maintenance. The orientation of DACBs is more complex, DACBs must be installed in one of four configurations, as determined by design: 1. face-to-face 2. back-to-back 3. tandem: faces toward the pump 4. tandem: faces away from the pump The “face” of the DACB is that side that has the narrow lip on the

The Pump Handbook Series

outer race. The “back” of the bearing has the wider lip on the outer race and usually has various symbols and designators on it. Photo 2 shows two pairs of DACBs. The pair on the left is positioned faceto-face while the pair on the right is back-to-back. Note that the lip on the outer races of the first pair is narrower than on the second pair. This distinguishing characteristic provides an easy identification of which side is the face or back. In tandem, the narrow lip of one bearing is placed next to the wide lip on the other. In other words, all bearing faces point either toward the pump or away from it. To facilitate the installation of DACBs, the bearing faces should be marked with a black indelible marker showing where the burnished alignment marks (BAMs) are on the back. This is because the four BAMs, two on each bearing, must be aligned with their counterparts, and not all BAMs are visible during installation. For example, when the first bearing is installed in a face-to-face configuration, the BAMs are on the back

ing cannot be hammered into position or removed and reused because it will be destroyed internally by these actions.

2. DACBs

Photo 2. Two pairs of DACBs, with the pair on the left positioned face-to-face, the pair on the right back-to-back

side, hidden from the installer. Marking the face of each bearing allows the installer to see where the BAMs are, so that all four BAMs may be aligned in the same relative position, such as 12 o’clock.

BEARING PRELOAD Under certain operating conditions (hydraulic forces, gravity, and movement of the pump and motor foundation such as on a seagoing vessel), the rotor may be loaded in either direction. If this occurs, the balls in a DACB with no preload could become unloaded. When this happens, the balls tend to slide against the races (ball skid) rather than roll. This sliding could result in permanent damage to the bearings after about five minutes. To prevent ball skid, bearing manufacturers provide bearings that have a predetermined clearance between either the inner or outer races. Face-to-face bearings have this clearance between the outer races. When the bearings are clamped together at installation (the outer races are clamped together), the balls are pressed between the inner and outer races, causing the preload. Back-to-back bearings have the clearance on the inner races, which are usually clamped together with a bearing locknut. By increasing the clearance between races, the preload can be increased from zero to a heavy load. This way, when conditions

cause the rotor loads to change direction or be eliminated, the bearing balls will still be loaded and ball skid should not occur. One disadvantage of using preloaded bearings is that bearing life will be reduced due to the increased loading. Preloaded bearings should not be used unless design conditions require them. If uncertain about the need for preload, users should contact manufacturers.

BEARING INSTALLATION Once the proper bearings have been obtained and the correct orientation determined, installation is relatively simple. The shaft and especially the shaft shoulder should be cleaned and any welding or grinding operations secured. The bearings must be installed in a clean environment, and the shaft must be free of nicks and burrs that may interfere with installation.

1. RADIAL BEARINGS To install radial bearings, they should be heated in a portable oven to 180–200°F. Then, using clean gloves and remembering the correct orientation, quickly slide each bearing over the shaft and firmly onto the shaft shoulder. Do not drop or slap them into position. Experience indicates that you have about 10 seconds after removing the bearing from the oven before it cools and seizes the shaft. If it seizes the shaft out of location, remove it and scrap it. The bear-

The Pump Handbook Series

Installation of DACBs follows the some procedure, except that additional care must be taken to position the bearings properly, line up the burnished alignment marks, and not erase the indelible marks added on each bearing face. After the first bearing has been installed, rotate the rotor (if necessary) so the alignment mark on the inner race is at 12 o’clock, then rotate the outer race so it too is at 12 o’clock. Before proceeding with the second bearing, mentally walk through the procedure. Remember which direction the face goes and that the burnished alignment marks must be in the same position as the first bearing marks. Also remember you have about 10 seconds before the bearing seizes the shaft. The purpose of aligning the four burnished alignment marks is to minimize off-loading (fight) and radial runout loads that will occur if the true centers of the bearings are not lined up. Minor imperfections will always occur, and they must be minimized. Failure to align the marks will result in the bearings loading each other. DACBs come only in matched pairs—they must be used together. To verify that a pair is matched, check the serial number on the bearing halves—they should be the same, or properly designated, such as using bearing “A” and bearing “B.”

NEW BEARING RUN-IN After new bearings have been installed, they should be run in while monitoring their temperature, noise, and vibration. Run-in is often called the “heat run” or “bearing stabilization test.” To perform this test, first rotate the pump and driver by hand to check for rubbing or binding. If none occurs, operate the

39

installed and the balls ride on the ball ridge located on the outer race. Evidence of reverse loading appears as “equator” bands around the balls. Contamination (failure to follow Rule 3): Contamination of bearings almost always occurs during installation, but can also occur when liquids or other constituents from the pump leak or are present in the surrounding environment. If contamination is found in a new bearing before installation, the bearing should be carefully cleaned and repacked. Evidence of solid contamination in used bearings usually appears as very small, flat dents in the races and balls. Photo 3. Radial bearing disassembly unit at the design rating point and record bearing temperatures every 15 min. Bearing temperatures should increase sharply and then slowly decline to their normal operating temperature, usually 20–60°F above ambient. During the heat run care should be taken to ensure that the temperature does not exceed the value specified by the manufacturer. If it does, the unit should be secured and allowed to cool to within 20°F of ambient, or for 2 hours. The unit may then be restarted and the test repeated as necessary until the bearing temperature peaks and begins to decline. If, after repeated attempts, bearing temperatures do not show signs of stabilization, too much grease may be present. The bearing should be inspected and corrective actions taken as necessary. Now let’s cover some basic rules to follow when working with bearings: 1. Never reuse a bearing that has been removed using a gear puller, even if it is new. The bearing has been internally destroyed in the removal process. See “True Brinelling” under “Bearing Failures.”

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2. DACBs must be installed in the correct orientation. If not, they may experience reverse loading and fail. See “Reverse Loading” under “Bearing Failures.” 3. Bearings must be installed in a clean environment. Contamination is a leading cause of premature failures. 4. Do not pack the bearing and bearing cups full of grease. Excessive grease will cause overheating and ball skid. See “Excessive Lubrication” under “Bearing Failure.”

BEARING FAILURES Failure to follow these four basic rules will result in premature bearing failures. These and other failures will occur for the following reasons: True Brinelling (failure to follow Rule 1): This type of bearing failure occurs when removing bearings with a gear puller. The force required to remove a bearing from a shrink-fit application is great enough that when it is transferred through the balls to the inner races, the balls are pressed into the inner and outer races forming permanent indentations. Reverse Loading (failure to follow Rule 2): Reverse loading occurs when DACBs are improperly The Pump Handbook Series

Excessive Lubrication (failure to follow Rule 4): Too much grease in a bearing may cause the balls to “plow” their way through the grease, resulting in increased friction and heat. If the bearings and bearing caps are packed full of grease, ball skid could occur. When it does, the balls do not roll, but actually slide against the races. Experience shows that the bearings may be permanently damaged after more than five minutes of ball skid. Finding packed bearings and bearing caps is a good indication that too much grease caused the bearing to fail. Bearing manufacturers usually recommend that bearings have 25-50% of their free volume filled with grease. Excessive Heat: Failure to provide adequate heat transfer paths, or operating the component at excessive loads or speeds may result in high operating temperatures. Evidence of excessive temperature usually appears as silver/gold/brown/blue discoloration of the metal parts. False Brinelling: False brinelling occurs when excessive vibrations cause wear and breakdown of the grease film between the balls and the races. This condition may be accompanied by signs of corrosion. A good example of how false brinelling could occur would be when a horizontally positioned component is shipped

Zero-leakage magnetic liquid seal developed to retrofit process pumps across the country and not cushioned from a rough road surface. The load of the rotor is passed through the bearing balls, which wear away or indent the races. Evidence of false brinelling looks similar to true brinelling, but may be accompanied by signs of corrosion where the grease film has not been maintained. Correction simply involves protecting the unit from excessive vibration and using specially formulated greases where past experience demonstrates the need. Fatigue Failure: Even when all operating, installation, and maintenance conditions are perfect, bearings will still fail. In this case, the bearings have simply reached the end of their useful life, and any additional use results in metal being removed from the individual components. Evidence of fatigue failure appears as pits.

BEARING DISASSEMBLY FOR INSPECTION Now that we know what to look for in failed bearings, let’s see

how we disassemble a bearing for inspection. Before disassembling any bearing, however, turn it by hand and check it for rough performance. Note its general condition, the grease (and quantity thereof), and whether there is any contamination. If solid contamination is present, the particles should be collected using a clean filter bag as follows: 1. Partially fill a clean bucket or container with clean diesel fuel or kerosene. 2. Insert a clean filter bag into the kerosene container. This ensures that no contamination from the container or the kerosene gets into the filter bag. 3. Using a clean brush, wash the grease and contamination out of the bearing. The grease will dissolve and any contamination will be collected in the filter bag for future evaluation.

RADIAL BEARING DISASSEMBLY After removing the grease and any contamination, you should disassemble radial bearings by removing any seals or shields, which are often held in place by snap rings. Then, to remove a metal retainer, drill through the rivets and remove both retainer halves. Then the bearing should again be flushed (in a different location) to remove any metal shavings that may have fallen between the balls and races when drilling out the rivets. If the bearing does not freely turn by hand, some metal particles are still trapped between the balls and races. Next, place the bearing on the floor as shown in Photo 3 with the balls packed tightly together on the top. Insert a rod or bar through the inner race and press down, hard if necessary. Note: If the balls are not packed tightly together, disassembly will not occur.

The Pump Handbook Series

DACB DISASSEMBLY To disassemble DACBs, support the face of the outer race and press down against the inner race. The back of the bearing must be on top.

HANDLING, TRANSPORTATION, AND STORAGE Common sense applies in handling, storing, or transporting precision bearings. They should not be dropped or banged. They should be transported by hand in cushioned containers, or on the seat of vehicle—not in a bike rack. They should be stored in a cool, clean, dry environment. Because nothing lasts forever, including bearing grease, bearings should not be stored for more than a few years. After this, the grease degrades and the bearings may become corroded. At best, an old bearing may have to be cleaned and repacked, using the correct type and amount of grease.

MAINTENANCE Routine maintenance of bearings usually involves periodic regreasing (followed by a heat run) and monitoring bearing vibrations, which will gradually increase over long periods of time . To maintain pumps and drivers that are secured for long periods of time, simply turn the rotor 10–15 revolutions every three months by hand. This will ensure than an adequate grease film exists to prevent corrosion of the bearing. If this action is not taken, the bearings may begin to corrode due to a breakdown in the grease film. ■ Ray W. Rhoe, PE, has a BSCE from The Citadel and 15 years’ experience with pumps, testing, and hydraulic design.

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Troubleshooting Boiler Feed Pumps BY GARY E. GLIDDEN y favorite work order is,”The boiler feed pump seems to be laying down!” And, my favorite comeback is, ”All feed pumps lay down, condensate pumps stand up.”

M

LOW FLOW TROUBLESHOOTING After determining whether flow suddenly or gradually decreased, take stock of the following key trouble areas when your boiler feed pump “lays down,” or delivers low flow. 1. Start by checking the suction. Some of these questions may sound simple but are often overlooked by operators. •

Is there enough water in the feed tank?

•

Could there be a restriction in the suction line?

•

Are all the suction valves open? (A valve that is partly open can show proper suction pressure but still starve the pump.)

•

Are you sure the plug has not pulled off the valve stem?

•

Does the pump show any vibration on the inboard bearing?

•

Are there any loose or obstructive objects in the feed tank or in the suction eye of the first stage impeller?

2.

Next, examine the discharge. A bad recirculation valve, either stuck open or leaking, may lose a lot of water. Also, a discharge valve not fully open will restrict flow, as will a stuck check valve.

3.

Inspect the warm up lines. On many pumps, these lines tie the discharge to the pump’s suction. If the valves are left open or leak, they may act as a recirculation valve.

4.

Calibrate the flowmeter. Often times the flowmeter, a frequent culprit in low flow situations, is mistakenly checked last.

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5.

Check gaskets and fits. A blowout of the gasket that seals the high pressure discharge from the low pressure suction will result in a large loss of water. In addition, some pumps have only a metal-tometal fit at this junction, which may sustain a slow increase of flow loss as a small ”cut” turns into a ”big washout.” Fits and gaskets on pump internals can cause plenty of headaches.

Results of a ”hard rub”

TROUBLESOME INTERNAL CONTACT In my experience, I’ve found that few boiler feed pumps simply “wear out.” If nothing contacts inside a feed pump, nothing wears. The real issue is not whether pump parts rub but when they rub. Internal pump conditions during hot standby or start-up are not always understood by operators. Any time a pump sits still, the shaft will sag a little. Sagging is accentuated if the pump is not evenly warmed up. Nozzle loading can also compound shaft sag as well as pump case bowing. Firing up a nozzleloaded pump will result in decreased life span of the wear rings, at the very least. In a worst case scenario the case rings could weld to the impeller rings, completely incapacitating the pump. In addition, if the pump is an unbalanced unit which has a balancing drum or disc, the balance line “leak off” should be charted and monitored. A gradual increase in “leak off” is normal, but a sudden increase, especially after a “trip” or start-up, signals that the rotating drum has contacted stationary faces. In a lower-pressure, lower-speed pump, say 2800 psi at 3600 gpm, this contact might take a while to show up. However, when stationary faces on a high-pressure, high-speed pump – for example, 5600 rpm at 4200 psi – rub, you can expect a work order reading, “Pump froze up – will not let the turbine go on turning gear,” in reference to some $80,000 – $100,000 – worth of impellers being welded together in an ugly mess. The Pump Handbook Series

A capscrew sucked up into the first impeller suction eye. This started a vibration that caused the case wear ring to break loose and also be sucked into the impeller.

SUMMARY Although there are many reasons for operational troubles and failures, a thorough knowledge of your pumping system is crucial in diagnosing problems before they lead to irreparable damage. ■

ABOUT THE AUTHOR: Gary E. Glidden is Crew Leader of the Houston Lighting & Power Company and member of Pumps and Systems’ User Advisory Team.

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Torsional Analysis in Couplings Selection BY JIM MAHAN ecent developments in higher-efficiency diesel engines are allowing in-line three-, four-, and six-cylinder diesels to do work that once required V-8 (or larger) powerplants. With the availability of these smaller, lighter-weight diesel engines, more pumping applications are converting from electric or gasoline power to take advantage of lower diesel operating costs. This trend seems most evident with centrifugal pumps used in such applications as portable flood clean-up or construction-site dewatering (“trash” pumps), deep-well irrigation and booster pumps in agricultural service, and large volume circulating pumps used in commercial fish and shrimp farming. While engine downsizing may be a blessing in terms of initial cost reduction and operating economy, it can easily aggravate the relationship between engine and driven equipment in terms of torsional vibration and the resonance that can amplify this vibration to destructive levels. As a result, the dollars saved in fuel economy might well be lost in repair expenses and downtime costs due to repeated replacement of broken crankshafts, spline shafts, couplings, bearings, gears, and seals. The purpose of this article is to examine the sources of torsional vibration problems, and to review the important role that careful coupling selection can play in minimizing these problems by damping vibrations and tuning the engine-driven system to shift destructive resonance speeds away from the engine’s operating rpm range.

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WHY TUNING IS NEEDED Normal diesel operation creates torque pulses that momentarily accelerate or decelerate crankshaft rotation. This is caused by the firing order, the firing angles, and the number of cylinders in the engine. Due to the diesel’s high compression, its pulses are larger than those of other

Photo 1. The worn splines on this hydraulic pump drive shaft are a typical indication of torsional vibration problems.

engines. In addition, “pressure spikes” occur because ignition doesn’t always occur precisely when it should and the fuel doesn’t burn completely on each ignition stroke. These torque pulses, which can be as high as ten times the engine’s normal operating torque, can add substantially to the total amount of torque transmitted through the system and to the wear and damage that result from this transmission (Photo 1). Torque pulses ripple through the system as torsional vibrations that cannot be seen or felt like the familiar linear (up-and-down, side-to-side) type of vibration. Consequently, the damage they cause is often mistakenly attributed to some other cause, such as shaft-to-shaft misalignment, improperly specified components, or faulty parts.

will remain fairly constant (which is not always the case), the time required for twisting and untwisting will be determined primarily by the inertial relationships between engine and driven equipment, and by the spring rate (stiffness) of various elements in the drive train. This twistuntwist time increment, expressed as a ratio to one minute, identifies the natural frequency of the system. In simple two-mass systems (engine and single load, connected by a single shaft), natural frequency can be closely approximated by the following formula, in which CTDYN is the dynamic torsional stiffness (spring rate) of whichever connecting device is torsionally softest—usually the coupling—JA is the engine’s inertia and JL is the load’s inertia (stiffness and inertia data are readily available from all equipment and component OEMs).

SYSTEMS HAVE NATURAL FREQUENCIES The stop–start forces applied by torque pulses exert a twisting pressure on the drive shaft and other drive train elements, which simply react as all springs do; rather than absorbing the twist energy, they give it back by untwisting. Assuming that the magnitude of the torque pulses The Pump Handbook Series

Natural Frequency (CPM) = 60 2π

√

(J +J ) CTDYN n A L (JAxJL)

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FIGURE 1 CPLG STIFF @ 885000 in-lbs/ rad

16

8 CYL (2)

EFFECT: ENGINE & LOAD INERTIA (JA & JL) ON NATURAL FREQUENCY

15

6 CYL (2)

14 13 13

IT

12

PM

FREQUENCY (000)

11 'S.

10 CT

E

9

EFF

L CY

, LES

ON

C EX

EQ

FR

&R

C

CY

8 CYL (4)

6 CYL (4)

8

JA VERY LIGHT 4 CYL (4)

7 6

JA LIGHT

5 JA MEDIUM JA HEAVY

4 3 2 1 0

0

2

4

6

8 10 JL DRIVEN EQUIP

12

14

16

18

20

0 0.2

0.6

1

1.4

1.8

2.2

2.6

3

Effect of engine and load inertia on natural frequency. Precise determination is possible with a more detailed calculation known as a Holzer Analysis, in which known frequencies are brought into the system on a trial-and-error basis and their resultant energy balances calculated until a frequency is found for which the summation of energy gained or lost through all torsional deflections balances to zero. In such analyses, frequency computation becomes very time consuming and complex and is best left to experienced consultants. Generally, when smaller engines replace larger ones (assuming all other components remain essentially the same), the reduced mass and inertia at the driving end tends to lower the natural frequency of the system.

RESONANCE AMPLIFIES VIBRATION All diesel engines “excite” their driven systems at specific frequencies that vary in direct proportion to engine rpm. At those rpm levels where the engine’s excitation frequency and the driven system’s natural frequency coincide, a condition of resonance emerges. Each torque pulse applies its twist energy at

44

about the same time that the shaft has untwisted to release energy from the previous pulse. This greatly amplifies the twist/untwist motion. Those rpm levels at which resonance occurs are termed critical speeds, for obvious reasons. For the standard four-cycle inline diesel engine system, the speeds at which resonance occurs (critical speeds) are calculated as the natural frequency divided by N/2 for the primary resonance speed, and by N for the secondary resonance speed, where N represents the number of cylinders. Critical speeds for systems using smaller engines move toward the operating rpm range because the engine’s excitation frequency drops with the move from multi-cylinder or two-cycle engines to four-cylinder, four-cycle engines. Diesel-driven centrifugal pumps generally operate at 1,200 or 1,800 rpm (approximating the typical pump design rpms of 1,180 and 1,750). The trend is to operate at the highest optimum speed so as to allow use of smaller, more compact driven equipment. This results in a situation where faster operating speeds, lower The Pump Handbook Series

excitation-frequency slopes, and higher natural-frequency curves combine to put critical rpm values closer to the required operating speed ranges (Figure 1). The traditional system, using large-inertia, multi-cylinder engines combined with big, low-speed equipment, would have relatively high natural frequencies with relatively low operating rpms. Accordingly, critical rpm values—occuring where excitation frequency meets system natural frequency—usually would appear farther away from the operating speed range. Engine rpm can pass through the critical speed ranges with no problems if it is done quickly. The trouble starts when engine rpm remains in the critical zone for longer periods, allowing natural frequency and excitation frequency to resonate at the operating rpm level. In this situation, careful selection of couplings is needed to move the natural frequency away from destructive resonance ranges.

CRITICAL WINDOWS FOR RPM Torsional vibration resonance typically occurs across an rpm range from 0.7 to 1.4 times the critical rpm, establishing a window in which engine excitation forces can damage system parts. For example, if resonance or critical rpm peaks at 1,500 rpm, the system will see torsional vibration problems anywhere between 1,050 rpm (0.7) and 2,100 rpm (1.4). To prevent such problems, new systems designed with lightweight three-, four-, or six-cylinder engines must have a complete torsional vibration review to identify the system’s natural frequencies and resultant critical rpm ranges. If this determines that critical rpm ranges exist near idle or operating speeds, it becomes necessary to conduct a complete multimass torsional vibration calculation, which usually requires hiring a specialized consultant. It is also worth noting that as engines become smaller and lighter in relation to their driven equipment, the magnitude of their torsional vibration pulses tends to increase and

MASS FACTOR: MAGNITUDE OF TORQUE PULSE

cause they usually involve a rightponent, connected by a variangle gear drive, U-joint shafting, and ety of torque-transmitting elea very long, torsionally soft down-hole ments that offer a wide range pump shaft. of standard torsional stiffness Multiple impellers usually are values, from highly elastic stacked in series at the bottom of the rubber to an almost rigid well to increase pressure, which plastic. This wide variation in results in higher unit inertia at the far stiffness lets you tune your end of the train. In this kind of applisystem’s natural frequency. cation, the coupling is only one of four Each coupling adds its or five spring elements, which means own spring rate (dynamic torthe system has four or five natural fresional stiffness) to the system, quencies subject to engine excitation. altering the system’s natural Here, the coupling must be positioned frequency. In this way, stiffer at the engine flywheel in order to procouplings, usually used for tect the entire drive shaft, but comrelatively low-inertia loads mon practice often leads to couplings such as hydraulic pumps, improperly positioned elsewhere in will move the natural frethe system (Photo 2). quency of the system When coupling selection for upward. Softer elements, these types of applications is based on usually appropriate for hightorsional analysis in addition to the er-inertia loads such as larger usual mechanical and torque-load compressors or multiple specs, it will maximize the service life pumps driven through splitof your equipment and greatly reduce ter boxes, will move the natdowntime and repair costs associated ural frequency downward as with premature fatigue and failure of well as absorb shock or damp key parts (Photo 3). vibration energy. Proper coupling selection also End suction pumps or assures you that if the system is operself-primers used as boosters ated at the wrong speed, or if a fault or dewatering pumps fall into occurs elsewhere in the system, the the category of medium-inercoupling will serve as a fuse and tia equipment, which typicalPhoto 2. Thin-profile engine-flywheel-type break before something more expenly requires a soft coupling. coupling with pump installed on pump sive does. As a result, premature These pumps are usually anamounting plate housing (upper half of coupling failures can be taken more lyzed as simple two-mass sysplate is cut away to show coupling). tems, but the many FIGURE 2 pump sizes and the match-up to do more damage to drive train com13 a wide variety of ponents. This can be seen in Figure 2, 12 LIGHT ENGINE engines and horsewhich depicts how pulse magnitude 11 power requirefalls off sharply as the ratio of engine ments call for very 10 inertia to load inertia increases (i.e. soft couplings. By lighter drivers absorb less torque 9 using very soft coupulse energy than heavier drivers and 8 plings, the full pass more twisting force along the 7 range of pump sizes drive shaft). 6 and engine alternaObviously, engine rpm standards, tives can be covthe number of cylinders, and the iner5 ered with a single tia values of both the engine and its 4 type of standard driven equipment are not easily 3 coupling rather changed. Fortunately, however, the 2 than requiring a resonance of the system can be 1 custom selection for changed to prevent it from occuring HEAVY ENGINE each installation. within the operating range. 0 0 0.4 0.8 1.2 1.6 2 2.4 2.8 3.2 Deep-well irriJA / JL TUNING THE SYSTEM gation pump appliCouplings designed for enginecations tend to be Torque pulse magnitude (mass factor) vs. ratio of driven systems basically comprise a more complex be- engine-to-load inertia (JA/JL) driving component and a driven comThe Pump Handbook Series

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reliably as a signal that something else is troubling the system—perhaps cylinder misfiring problems, poor fuel or filters, ambient temperatures or airborne contaminants, unrealized torque overloads, misalignment, or improperly specified parts. Never assume that the couplings already in place in your system have been selected within the disciplines described above. As one of the least expensive elements in any engine/ pump system, couplings too often are added as an afterthought rather than factored into the design up front. If your pumping system seems to be chronically suffering the kinds of wear and breakage symptoms that could result from torsional vibration and critical frequency problems, it’s very possible that replacing the original equipment coupling with one selected via torsional analysis can put those problems behind you forever. Experienced representatives of broad-line coupling suppliers can help you prepare a torsional analysis and select the proper couplings. ■

Photo 3. A high-torque rubber-block-type coupling shown in operation between a diesel engine and generator. The jaw-type design is mounted directly to the flywheel and has a torque capacity of up to 177,000 in./lbs.

Jim Mahan, currently Centaflex Product Manager for Lovejoy, Inc., has 30 years of sales, marketing, and engineering experience in centrifugal pumps and power transmission. He holds a BSME degree from Worcester Polytechnic Institute and an MBA from Illinois Benedictine College.

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The Pump Handbook Series

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Not All Grouts Are Created Equal BY RICHARD D. MYERS wo common causes of excessive vibration and resulting pump failure are out-of-balance rotating components and shaft misalignment. Unfortunately, the manifestation of the first problem may be confused with the effects of the second. Even with sophisticated vibration spectrum analysis equipment, misalignment—caused by grouting problems—may be improperly diagnosed as an unbalanced part or a bearing wear problem. When alignment is correctly identified as the culprit, the underlying reason for the problem may be deterioration of the cementitious grout.

T

THE IMPACT OF MISALIGNED SHAFTS Angular misalignment between shafts connected by a flexible coupling introduces forces that can produce torsional and lateral vibrations—two indicators of a basic misalignment problem. Misaligned shafts can cause premature packing or mechanical seal wear, resulting in excessive leakage. More stringent restrictions on fugitive emissions to comply with clean air and water regulations enforced by the Environmental Protection Agency include provisions for substantial fines—and plant shutdowns. In addition, these new regulations require the use of more exotic mechanical seals for many installations. This

Preparing the baseplate for installation and grouting. increases the financial impact of premature failures. With every revolution of a misaligned shaft, severe stresses alternating between tensile and compressive loads are exerted on bearings and their housings. This will cause progressive opening of bearing housing races (bearing materials are harder than their housings) to the point where they will eventually lose their ability to retain the bearings, bearings will wear out, and shafts may bend or break. Before a catastrophic breakdown occurs (i.e., a broken shaft or coupling failure), a misaligned pump-

FIGURE 1. PUMP FAILURE ANALYSIS IN THE PETROCHEMICAL INDUSTRY Failure Type Mechanical Seals Bearing Distress

Percentage of all Failures 34.5 20.2

Pump failure analysis in the petrochemical industry

FIGURE 2. VIBRATION LEVELS BEFORE AND AFTER PROPER GROUTING Vertical Plane Before Grouting After Grouting Pump 1 35.61 1.94 Pump 2 4.98 0.44

Horizontal Plane Before Grouting After Grouting 5.42 2.26 5.65 0.48

Vibration levels before and after proper grouting The Pump Handbook Series

ing system will probably fail due to bearing distress or serious damage to an expensive mechanical seal. In the petrochemical industry these are the two most common causes of pump failures (Table 1). Note: bearings are milled to tolerances of 12 millionths of an inch, and the average cost of a mechanical seal is $4,000/year.

SOLUTION Pumps are typically mounted on baseplates with machined mounting surfaces. Baseplates are, in turn, secured to concrete foundations with epoxy or cementitious grout. Ideally, vibration is transmitted through the baseplate to the foundation and down through the subsoil where it can be absorbed. As indicated by the example in Table 2 of two particularly severe applications for ash-slurry pumps, overall vibration levels are reduced significantly with proper grouting. After grouting, a tenfold reduction in vertical vibration levels is common. Properly installed, an epoxy grout will bond to a concrete foundation with a tensile strength of 2,000 psi—significantly greater than that of concrete. The bonded baseplate is effectively transformed into a mono-

47

mortar mixers to be cleaned with water, eliminating the need for harsh chemicals and flammable solvents on the job site. Other features to look for include high compressive, tensile, and shear strengths, as well as deeppour capabilities and non-shrink properties.

CONCLUSION

The epoxy grout cures in the form for 24 hours before removing forms and final torquing of the anchor bolts. lithic structure including the concrete foundation. The single-block monolith significantly dampens vibration resulting from resonance with the natural frequency of the baseplate. Another important reason to use an epoxy grout instead of a weaker cementitious grout is that the epoxy will seal the concrete foundation, preventing damage from moisture and lubricants. (Because there is a poor bond when cementitious grout is used, leaking or spilled liquids can seep between the underside of the baseplate and the foundation.) With approximately eight times the vibration damping effectiveness of concrete and 30 times that of steel, an epoxy grout can extend mean running time between failures (MTBF) three- or four-fold. However, to take full advantage of the excellent vibration damping characteristics of epoxy grout, intimate contact must be maintained with the pump bedplate. This is easily accomplished by following installation procedures. The end result is a significant increase in MTBF, significantly reduced maintenance expenses, as well as a substantial increase in the life of mechanical seals and bearings.

while curing. This eliminates cast-in stresses that can cause cracks and flaws in the essential monolithic structure—a problem common to high exothermic grouts. In addition, the grout should provide a low coefficient of linear expansion (CTE). The value should be close to the CTEs of the metal baseplate and the underlying concrete foundation. And a grout with low dusting aggregate will provide a safer, cleaner working environment. It can also simplify installation and clean-up procedures. Currently available products allow tools and

CHOOSING THE RIGHT GROUT When selecting an epoxy grout, a product should be chosen that exhibits a gentle exothermic reaction

48

The completed installation.

The Pump Handbook Series

Epoxy grouting of pump baseplates to their foundations will keep pumps in service longer with fewer unscheduled outages and fewer expensive repairs or retrofits due to misalignment or deterioration of an inferior grout. Richard D. Myers is the Escoweld Marketing Manager for ITW Philadelphia Resins. He did his undergraduate studies at Lycoming College, Williamsport, Pennsylvania and his graduate work in biochemistry at New York University. His experience includes approximately 30 years of marketing industrial coatings and grouts.

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R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

Bearing Lubrication Trends and Tips BY DON EHLERT earing lubrication in process pumps is performed by various methods, from traditional oil baths or sumps to pure or dry sump oil mist lubrication. Other less common methods include grease, circulating oil, and purge or wet sump oil mist. Though pumps are the workhorses of many industries (refining, chemical/petrochemical, and pulp and paper), for some reason bearing lubrication is often not considered a high priority. However, concern for pump reliability is growing as a result of record keeping, increasing repair costs, and professional papers that continue to give lubrication issues attention. Process pumps move product from one point to another through a piping system that may stretch several hundred yards or several hundred miles. Pump operation is normally remotely activated, 24 hours a day continuous, and performed in all types of weather with minimum operator involvement. These pumps endure tremendous abuse from the product being pumped and their environment. A pump is built to certain specifications so it can withstand what is going through it and what is happening around it; but the bearings that allow the pump to run are made to handle loading conditions only, not ambient conditions. Bearings, usually out of sight and out of mind, are the heart of the pump. If a bearing is gone, the pump does not run and no product is pumped. An upset condition then occurs, which may cause any number of events, most of which are inconvenient and costly. If bearings are the heart of a pump, then lubricant is the life blood; but too many process plants pay little or no attention to bearing lubrication.

B

BEARING HOUSINGS Water is not the only culprit that shortens bearing life or contaminates oil. Thermal cycling caused by temperature changes tends to draw air-

Conventional oil sump application. borne contaminants into the bearing housing. Damaged or worn seals allow product to enter the housing and create severe problems. After more than 18 years of working around pumps, I am still amazed at how contaminated oil can be while a pump is still running. I have pulled the plug on bearing housings expecting slightly thick and dirty oil to drain out–only to be surprised when sludge oozed out. This is not common, but it does happen. I’ve also seen a pump come into the shop with charred paint on the outside and fresh, clean oil in the bearing housing. Needless to say the rolling elements of the bearing are beyond recognition. Some are welded together. Others are hollow with a hole in one side. Plant maintenance departments have gone to great lengths to improve MTBF by installing special housing seals to keep contaminants out and by using synthetic oils with improved lubricating properties. Some shops have elaborate repair procedures that The Pump Handbook Series

document all tolerances to make sure everything fits perfectly before assembly, and that require special seals and synthetic oil. I have heard of repair shops that guarantee several years of operation without an oil change when specified seals and oil are used. Everything these shops are doing helps; but other plants are achieving long MTBF simply by initiating good lubrication programs.

THE SURVEY To learn more about the everyday life of pumps, I took an informal survey, asking general questions about the number of pumps, standard method of lubrication, oil change interval, mean time between failure (MTBF), and average pump repair cost. I received replies from 18 plants or individual operating units from across the U.S. and Canada. These included refineries, paper mills, and chemical plants from Texas, Louisiana, California, Washington, Montana, and western Canada.

49

FIGURE 1 B 12 MONTHS ENDING 7/86

A

WITH GREASE (440)

20

50

WITH GREASE (156)

40 30 20 10

WITH OIL MIST (400)

0

MOTOR BEARING FAILURES

MOTOR BEARING FAILURES

3 1/2 YEARS

10

WITH OIL MIST (400)

vals. Two plants use a quarterly schedule, four have a six month schedule, and five are on annual schedules. Two plants recognize the importance of changing oil but indicated that it was seldom done if ever. Of course, the pumps on pure oil mist did not require oil changes. One plant with all pumps on oil bath indicated a monthly interval, and the units had an average MTBF of less than two years. The plant with the longest MTBF (six years) is on an annual schedule, and the plant with a three-year MTBF is on a six month schedule.

SEALS

0

Lip seals are used in about 60% of the pumps in the survey, across all industries that replied. Labyrinth seals are used in Effect of oil mist on electric motor bearing life. (A) Compares failures at two plants run by the same company in different states. (B) Compares fail- about 38% of the pumps. Oddly enough, people at every plant ures in two process units in one plant. knew what type of bearing housing seals were used, and many gave exact numbers or These plants have a total of are 1.4 and 5.5 years. Out of the ten percentages when asked for quanti20,460 pumps. MTBF ranges from a plants using oil mist, only two have ties. low of 200 days to a high of six years. an MTBF of less than one year. These It is possible that replacement Repair costs are as low as $1,000 and two facilities have small, modularcost is a major factor in as high as $7,500, with an average type oil mist systems determining which seal to repair cost of $3,500. Three plants did without alarm features. use, and lip seals tend to not give a repair cost, and one of I was disappointed be the most economical. these plants did not provide an that more plants using All plants using pure oil MTBF, which I assume is less than oil mist did not respond Thermal mist except one use lip one year. to the survey. I believe seals predominantly Lubrication methods for these that the totals would cycling because a slight positive pumps break down like this: 14,106 have been much differpressure is maintained in pumps use traditional oil baths, 5,388 ent if they had. Several caused by the bearing housing and pumps use pure oil mist, 508 use plants that recently conthere is no oil level to contemperature purge oil mist, and the remaining 458 verted from oil sump to taminate. use other methods, including grease oil mist have reported changes tends and circulating oil. longer MTBFs and a LUBRICANT Oil sump has the lowest MTBF, considerable reduction to draw Fourteen plants use 200 days, as well as the highest, six in pump repair costs. mineral oil exclusively. airborne years; however, five plants that are Many plants I’ve dealt One plant uses synthetic 100% oil sump have an MTBF of less with over the years have contaminants throughout on pure and than one year. These five have 7,215 had a drop in the numpurge oil mist applicapumps, and their average repair cost ber of mechanical seal into the tions and has a two-year is approximately $3,000 per pump. failures after installation MTBF. Another uses The minimum of $1,000 was used for of oil mist lubrication. bearing synthetic throughout in the one plant that did not show an OIL CHANGE INTERVAL housing. oil bath, and this plant average repair cost. How often should had the shortest MTBF Two plants with a total populaoil be changed? This is a of all plants. Two plants tion of 850 pumps are 100% oil mist matter of opinion. Of use both types of oil, lubricated. Of the 850 pumps, 92% the 18 plants surveyed, only 11 had and their MBTFs are three and four use pure oil mist. The rest use purge routinely scheduled oil change interoil mist. MTBF for these two plants PLANTS

50

OPERATING UNIT

The Pump Handbook Series

Dry sump/pure mist application (API 610 7th Edition).

years. Both of these plants are predominantly oil mist.

OIL MIST Oil mist is a dispersion of tiny oil droplets, 1 to 3 µm in size, suspended in an air stream moving at about 24 ft/second. These droplets are transported through a piping system to the point of lubrication, where they pass through an orificed fitting known as a reclassifier that controls the amount of lubricant being applied. The controlled mist, which looks like smoke, not only lubricates bearings, but it uses less lubricant than a traditional oil sump. A recent study funded by the State of Texas determined that pure oil mist lubrication could extend bearing life by as much as six times over the traditional oil sump. It should be noted that oil mist has an oil-to-air ratio of 200,000 to 1. This is well below the lean limits of flammability, so it will not burn or support combustion. Oil mist is often referred to as a vapor, but it is not. It is an aerosol, and because typical R&O turbine oils are used in these systems, oil mist is not a volatile organic compound (VOC). Opinions on oil mist lubrication are varied. Those who like it outnumber those who don’t. Those who’ve had a bad experience with oil mist were probably involved in its misapplication or a poor system installation.

The installation of the distribution system is the most critical part of a completed oil mist system, with proper sizing of the reclassifiers a close second in importance. Some plants have experienced so many failures after the installation of oil mist systems that they have given serious consideration to taking the systems out and never using oil mist again.

The oil mist distribution system normally consists of schedule 40 galvanized pipe with malleable iron 150# screwed fittings. This is where problems with oil mist systems usually start. Anybody can screw pipe together, and often just anybody is hired to do just that. Oil mist will leak through very small gaps, and if proper procedures are not followed, every screwed joint will leak. Another problem is traps or low spots in the pipe. If not properly sloped, the pipe will fill with oil and every pump on the system will fail. If you are considering an oil mist system, do not cut cost or corners on system installations. It will catch up with you and eventually cost you more because of pump repairs and down time.

CONCLUSION Obviously, there are no clear cut trends here. Pump life and performance vary widely, depending on product pumped, environment, and length of time in service. But some of the things happening out there in the field make a case for paying more attention to pump lubrication. I’ve often heard that lubrication accounts for about 30% of all pump

Wet sump/purge mist application.

The Pump Handbook Series

51

this cost to lubrication related failures, we are still looking at $6,493,500 being spent annually on lubrication-related repairs, and this averages out to $1.3 million per plant each year. This amount could cover installation costs for 8 to 12 oil mist systems, and these systems could eliminate up to 90% or more of lubrication related failures. This could save these plants $1 million a year, on average, and provide a return on investment in less than 1.5 years. Heinz Bloch reached the same conclusion in a paper titled “Large Scale Application of Pure Oil Mist” (Ref. 1).

FIGURE 2 $K 100

83.9

82.1

74.1

80 60

56.2

40

27.5

25.9

25.6

23.3

20 0 DCU

FCC

AVU

3 UNIT AVG.

REFERENCE BEFORE MIST

AFTER MIST

Annual pump bearing repair costs before and after installation of oil mist lubrication for three different process units within the same plant. failures. If this is the case, there is a tremendous number of needless pump failures. From this survey of just 18 plants, only five had an MBTF of three years or longer, and seven had an MTBF of less than one year. This is the most interesting point in this survey. Let’s consider the five

plants using 100% oil sump that have an MTBF of less than one year. Pumps in these plants totaled 7,215, with an average repair cost of $3,000. If we do some math, we are looking at five plants spending $21,645,000 annually on pump repairs. If we attribute only 30% of

1. H. Bloch. Large Scale Application of Pure Oil Mist. ASME Paper 80-C2/LUB-25. Presented at the ASME Conference, August 1988. ■ Donald C. Ehlert is currently a Senior Sales Representative for Lubrication Systems Company, where he has worked for the past 15 years. Mr. Ehlert’s experience includes system design, layout, installation, and maintenance of lubricating systems. He has been involved with lubrication equipment and technology for 24 years.

TABLE 1. LUBRICATION SURVEY RESULTS Plant 1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17 18 Totals

52

Pump Quantity and Type of Lubrication Population Oil Sump Pure Mist Purge Mist Other 3,000 3,000 175 50 100 25 200 20 120 60 650 585 65 900 900 1,200 1,152 48 800 800 300 300 600 600 2,500 2,500 1,230 922 246 62 40 25 15 350 150 150 50 300 268 30 2 515 515 4,700 2,820 1,410 235 235 200 200 2,800 84 2,632 56 28 20,406 14,106 5,388 508 458

MTBF <1 yr. 2 yr. <1 yr. 1.4 yr. <1 yr. 2.1 yr. 6 yr. <1 yr. 3.5 yr. <1 yr. 1.3 yr. <1 yr. 3 yr. 4 yr. <1 yr. 2 yr. 5.5 yr. 2.6 yr

The Pump Handbook Series

Average Repair Cost $5,000 ? $1,000 $4,500 $1,200 $3,500 $3,000 $4,500 $1,000 ? $1,400 $5,000 ? $3,000 $3,750 $7,500 $1,300 $6,500 $3,500 Average

Types of Bearing Housing Seals Lip Labyrinth Other x x x x x x x x x x x x x x x x x x x x x x x x x x x x x x

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R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

A Multistage Vertical Pump in Light Hydrocarbon Service BY KENNETH R. LAPLANT AND CHARLES R. RUTAN critical piece of equipment in an Olefin plant is the ethylene product pump. Design and operation of this pump is difficult because ethylene is a high vapor pressure fluid with poor lubricity. The chances for periodic plant upsets and human error, evolution of processing requirements and the need to stay within budgets make the task even more challenging. Designs of these pumps are continually refined as users and manufacturers strive to increase MTBF and incorporate new technologies for processing. This article discusses the design and operation of two vertical 4 x 8 x 10-1⁄2 double case pumps for ethylene service (Figure 1). The pumps have a 20 stage opposed configuration, originally designed with 20 impellers, and could produce 9300 ft tdh at 550 gpm from a source of liquid ethylene at -37º F and 325 psig. The pumps take suction at the bottom, 10 ft below grade, then pump the liquid up through 10 stages where it passes through a cross-over channel to the top of the pump. The flow then continues down through the final 10 stages to the center of the inner pump case where the liquid is channeled into the outer case. Then the liquid travels up to the discharge nozzle through the annulus formed by the inner and outer cases. The pumps are set in a 10 ft diameter can anchored in a 3 ft-thick reinforced concrete slab. The outer pump case and associated suction and drain piping are 24 inches from the bottom of the outer can. To provide insulation, pearlite was blown into the void between the pump case and outer can. To provide a level surface for the outer pump flange, a flat steel plate was epoxy-grouted into the foundation. A 1000 hp vertical, 4000-volt, 3600 rpm induction motor drives the pumps. The motor has a high thrust load capacity. The coupling is a rigid spool design and the motor coupling

A

hub has been modified from a slip fit to an interference fit. The pumps can produce 2300 psig discharge pressure at normal conditions. But downstream system losses were less than expected so the pumps were destaged by two impellers. After consulting with the manufacturer, the best location from which to remove impellers was chosen with special concern given to pump shaft thrust and rotor dynamics. The destaging decreased energy requirements and improved the stability of the system by bringing the pump closer to its BEP.

ing the suction will begin to boil and the pumps will vapor lock which causes severe damage to both pumps. Consequently, operations under minimum flow must be controlled carefully. A larger exchanger operating off one of the process refrigeration systems could handle the load and may prove economically justifiable.

FIGURE 1

INSTALLATION In cold, light hydrocarbon service, the piping system must be designed to minimize pipe strain at the pump flanges and eliminate any possibility of gas pocketing. The design must also address the potential for low temperatures. Liquid ethylene can reach temperatures as low as -147º F when it vaporizes as it is released from a high pressure to a low pressure area. In this installation, the suction tank stands 20 ft above grade. It has an 84-inch internal diameter and is 18’ 8” tall (including heads — seamto-seam height is 17’ 2”). The liquid ethylene in the tank is at equilibrium and is boiling. The tank was situated in such a way as to minimize the number of piping bends, thereby minimizing head loss. Accessibility for maintenance, namely head clearance, was another factor. At design conditions, NPSHA is 19.9 ft and pump NPHSR is 16 ft. All horizontal suction piping has a slope of 1” in 10 ft to prevent pocketing of vapors. The offset disk block valve used for the main suction block valve did not provide adequate isolation, so a second valve was installed with a bleed connection between the valves. Heat builds up in the ethylene when the pumps operate on minimum flow recycle to the suction tank. Within an hour the liquid enterThe Pump Handbook Series

Piping/pump outline.

Vents are located: •

at the plan 13 (reverse) seal flush at the seal gland

•

at the seal chamber, below the seal

•

at pump discharge

•

downstream of the check valve before the discharge block valves

•

between the pump discharge and the discharge block valve

The vent between the pump discharge and discharge block valve is a 1⁄2” stainless steel tubing vent open to the atmosphere 20 ft above the pump. This vent is used only as a momentary test vent before startup. All other vent lines are tied

53

through a manifold to the liquid drain system and to the vapor space of the suction vessel.

OPERATION Before startup, the system between the suction and discharge block valves is purged for a fixed time with nitrogen to atmosphere. Next, the vent valves are lined up to the flare system and the pump is flushed with methanol to remove moisture. The methanol is then purged from the system with nitrogen and ethylene is slowly introduced to the pump system to cool it. Care is taken to ensure that the ethylene does not thermally shock or distort any system component or cause any leaks. After about 20 minutes the liquid drain lines are closed and the vent valves opened to the suction vessel. The system is allowed to remain in this condition for a minimum of 8 hours so the pump and piping reach equilibrium with the suction tank and fill with liquid. At this point, the pump is ready for startup and all vent valves (except seal flush) are closed. The plan 13 flush, which is at the high point of the pump case, vents all ethylene vapors generated from insulation losses. The pump is then kept in ready mode; before actual start-up, the 1⁄2” vent valve is used to check for vapor. Operations personnel must respond immediately to pump shutdown. If the suction valve is not closed before the pump comes to a stop on coast down, energy stored in the liquid between the discharge check valve and the first stage impeller will cause the pump to rotate in reverse at speeds up to 3600 rpm.Under these conditions, pumpage vaporizes. Without liquid pumpage, heat builds up, bushings aren’t lubricated and no hydraulic dampening occurs. The result is damage to the internal parts of the pump and the need for an overhaul. When maintenance is due, the pump is isolated and the ethylene is purged with warm (ambient) nitrogen through the flare system.

DRIVER SYSTEM The thrust bearings and upper and lower radial bearings are in an oil bath equipped with cooling coils.

54

Thrust bearings are a Kingsbury tiltpad design. Radial bearings are a sleeve design with oil grooves cut axially and circumferentially. The upper journal is 8.75” in diameter, 0.50” thick and 3.63” long. The upper shaft journal would not hold its shape and was found to be 0.003” to 0.005” outof-round. A diametrical bearing clearance of 0.008 ± 0.00025” is required for stable motor operation. Insufficient clearance will cause the bearings to overheat and melt the babbitt. Excess clearance will cause oil whirl in the rotor and damaging vibrations. Several modifications were made to the system to improve its reliability. One change involved the disassembly process. Each time the driver was disassembled, the upper section of the shaft would bend. The cause was found to be the stepdown of the shaft which was designed to accommodate the oil bath’s stand pipe. When the shaft was raised and lowered during disassembly, the weight of the armature on the stressed end of the shaft caused the shaft to bend 0.005”. The old lifting procedure was to use an eye bolt and shackle to lift the shaft, which was then lowered on a wooden cradle. The shaft no longer bends because a new procedure uses two additional nylon straps — one placed around the armature windings and another just below the windings. The top is now lowered slowly while the two other points are lifted. The oil cooling coils were also modified. The original coils, which were Monell tubing failed after one year of service because of under deposit corrosion. The coils were replaced with copper refrigeration coils. The new coils transfer heat better and improve oil and bearing temperatures, and the copper is compatible with the cooling tower water. The coupling hub was also changed to improve alignment tolerances. The original coupling was a key-driven slip fit with a segmented ring keeper. Tight alignment tolerances were difficult to achieve and to hold. The coupling was modified by turning down one motor shaft to remove the keyway and provide a square shoulder for the coupling to stop against. A coupling hub was The Pump Handbook Series

machined with a 0.0015” total interference fit to the shaft. The other motor shaft was to be machined in the same manner six months later. But when the motor hub was removed, the shaft was found to be cracked. A new hub was therefore machined from ASTM 4140, using a specification similar to that for a turbine shaft. The coupling end was ground to a 1⁄2” per 1 ft taper and a 410 stainless steel band was shrunk on the hub as a vibration probe target.

FIGURE 2

Vibration monitor probes.

INSTALLING AND USING THE VIBRATION PROBES Two proximity probes were mounted to the top of the motor to reference the axial position of the motor shaft. Two radial probes were mounted on the outboard (top) of the motor, just above the shaft bearing journal. In addition, two radial probes were mounted below the pump coupling hub on the seal gland to read the vibration of the pump shaft (Figure 2). A tri-accelerometer mounted to the bottom of the motor was replaced with proximity probes because the accelerometer did not provide a true representation of the shaft vibration. The two probes mounted on the

pump seal gland were also inaccurate because they were often at or near a node of the pump shaft vibration. These readings only detected misalignment of the pump and driver and internal rubbing caused by excessive clearances of the upper pump bushing. To improve vibration monitoring, another set of radial probes were installed, referencing the motor shaft at its coupling hub. Because the motor is rigidly coupled to the pump, the new probes provide true readings for vibration of the rotor system. Vibration levels are monitored with the motor running solo (i.e., uncoupled) whenever the pump and motor are uncoupled for maintenance work. This monitoring can detect bearing problems, especially oil whirl, in the motor. Normal levels of vibration are about 1 mil pk-pk (peak-topeak). After the pump is coupled to the motor and the system is ready for operation, the pump is started on minimum flow recirculation to verify the work performed. Technicians know that amplitudes below 5 mils pk-pk indicate that they did an excellent job.The best amplitude achieved on a coupled system is 3 mils pk-pk.

When the maintenance task is complete, pump start-up is monitored with an oscilloscope, spectrum analyzers and a digital vector filter. Expected vibration levels at the motor shaft coupling hub are 5 to 7 mils pk-pk. With the old coupling design, 10 mils pk-pk was considered very good. The pump will run indefinitely at levels of 12 mils pk-pk. Vibrations above this level call for a pump overhaul. If levels reach 18 mils pk-pk, the pump is shut down and repaired to avoid major damage. Vibration levels are typically 1 to 2 mils pk-pk higher when the pump is on minimum flow recycle than when it is fully loaded. The increase is attributable to flow instability when the pump operates back on its curve. The minimum flow was sized for 280 gpm (38 percent of BEP) but this rate was inadequate for recycle runs in excess of one hour. The minimum flow is now set to 400 gpm (55 percent of BEP). The vibration alarm level is 12 mil; the danger level is 18 mil. Radial vibration alarms are audible. The only conditions that trigger a shutdown are low level in the suction

The Pump Handbook Series

tank and thrust problems (i.e., axial position of the motor rotor). Since access to the coupling area is restricted by the vibration probe brackets and connections to the probes, two small bayonet-type electrical heaters have been installed to turn on automatically upon pump shutdown. The heaters, mounted in a two-piece aluminum clamp, prevent ice from forming around the top side of the seal gland, onto the pump shaft and coupling. Next month, part 2 of this article will review the pump parts used in this multi-stage vertical pumping application. It will also discuss the couplings and alignment problems and solutions. ■ Ken LaPlant is a rotating machinery consultant. His company is Machinery Resources, Inc, Houston, TX. He has a BSME from the University of Vermont and is a registered Professional Engineer in the state of Texas. Charlie Rutan is a mechanical area maintenance manager for OxyChem Petrochemicals in Alvin, TX. He has a BS degree from Texas Tech.

55

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R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

A Multistage Vertical Pump in Light Hydrocarbon Service, Part II BY KENNETH R. LAPLANT AND CHARLES R. RUTAN ood pump design requires good communications between manufacturers and users. Designers must understand the process in which pumps will be used, including the habits of the operations and maintenance personnel. Modifications should be made with the input of the manufacturers. This case clearly illustrates the need for good communications. The original construction used 18-8 stainless steel parts extensively. Engineers at the plant objected, but the pump company and designers at the customer’s corporate headquarters felt that 18-8 SS was a good choice. The plant engineers were right: on startup both pumps seized. Teardown revealed that the interstage bushings and most of impeller wear rings had been severely rubbed. Many of the stationary rings were 188 SS with flash chrome applied to the bore. Likewise, most of the rotating rings were 18-8 SS that had an “electrolyzed” hardening applied to the outside diameter.

G

Plant engineers knew from experience that only bronze or a suitable nonmetallic material was reliable in ethylene or ethane service. The pumps were quickly retrofitted with SAE 660 leaded bronze surfaces to mate to 18-8 SS parts. One drawback to this scheme is that copper acetylides can form if acetylene occurs in ethylene streams. But this phenomenon is highly unlikely in the nearly pure ethylene process at this plant. Another advantage of bronze in this application is its ability to absorb momentary impacts that occur during startup, shutdown, and occasional upsets. Nonmetallics may be a superior material for wear parts in such service and should be seriously considered by users facing similar conditions. Dynamic balance of assembled rotors is imperative for long-term, trouble-free operation of a vertical multistage pump. Users should specify a stacked dynamic balance and request that the balanced rotor not be disassembled. This requires split case rings in many locations, although

solid rings can be used in others. The split rings are suspended away from the mating ring with tape during balancing. The split rings and the bushings must be designed to match the integrity of a solid bushing and not allow fasteners and alignment pins to loosen and fall into the pump during operation. Unfortunately, the pumps in this case could not be dynamically balanced. Only the impellers were balanced, and they were balanced individually before assembly. The balance sequence for the rotor starts with an individual balance of all impellers to the API 610 7th Edition 4W/N tolerance. The stacked balance operation then proceeds, working from the middle out until the tolerance is met. With a 500 lb rotor, this pump requires a finished tolerance of 7.9 gram-inches per plane. Materials must be selected carefully. Shaft and impeller materials that survive cold service often tend to gall during removal. Under such circumstances the impellers and sleeves can be difficult to remove. Removal can take a long time and require application of heat, which can damage other parts in the pump. Users should consult with a qualified metallurgist to select the best materials. Another option is to hard-coat the impeller and sleeve fit surfaces to lessen the chance of damage during disassembly. The pump has a detailed shaft manufacturing procedure that assures a straight and stable finished shaft. The fit surfaces are hard-coated with chrome oxide, which decreases the probability of damage to the shafts.

COUPLINGS The standard coupling for vertical multistage pumps is the rigid hub and spool with a lift nut to position the pump rotor. This coupling has a good track

Rotor in balance stand

56

The Pump Handbook Series

Fully assembled coupling record; but if not manufactured and maintained properly, it can cause alignment problems. After striving for straight pump and motor shafts, users must ensure that the coupling joins the two into one straight rotor system. Otherwise, the system will be kinked. The coupling must be machined very carefully during repairs and manufacture. Analyses show that perpendicularity, concentricity, and parallelism of mating surfaces are of primary concern in this pump. Also, hub-toshaft fits and radial clearances of rabbet fits on the coupling must be as tight as practical. Thus, design, manufacture, and maintenance require high precision. Rust is a major concern in selection of coupling materials for these pumps because they are installed in the Texas Gulf Coast area. Addressing the problem of rust turned out to be more complex than it first appeared to be. Resistance to rust was not the only consideration: strength, stability,

and resistance to galling were also important. ASTM 4140 remained the preferred material for the motor and pump hub, but 17-4 PH was selected for the spool. The pump hub must slip on the pump shaft with minimal clearance. This necessitated a material that will not gall on the 15-5 PH shaft. The lift nut is still made from 4140 but it is now coated with 8812 tungsten carbide–cobalt on the mating faces and outside diameter where it registers on the pump hub and spool. The coating is rust-free and hard. The hardness permits grinding to exact tolerances and resists wear from handling. It also provides a nongalling surface for the 4140 pump hub and the 17-4 PH spool it mates with, as well as the 15-5 PH threads on the pump shaft. To make the radial fits for the four major parts of the coupling as tight as practical, the motor hub was resized to have a shrink fit to the The Pump Handbook Series

motor shaft. This change eliminated slop in that connection. A compromise is necessary between the need for the pump hub to slide on the pump shaft and the need for minimal slop. A maximum diametrical clearance of 0.0005” is used in this position. The three registers on the flanged parts of the coupling require the same compromise, but with a wider clearance to accommodate short registers and large diameters. A diametrical clearance of 0.001”/0.002” is used in that position. This clearance creates minimal runout problems at assembly because there is no key at these fits. Manufacture of the coupling components is straightforward but requires attention to detail. The limit for perpendicularity is 0.0003”, for parallelism 0.0002”, and for concentricity 0.0002”. Grind reliefs and chamfers in the design aid manufacture and coupling to these tolerances.

57

identically assembly.

TABLE 1 Motor Coupling Hub Flange Top Spool Flange Bottom Spool Flange Pump Coupling Flange Probe Target Pump Shaft

0.0010” TIRR 0.0010” TIRR 0.0015” TIRR 0.0015” TIRR 0.0020” TIRR 0.0020” TIRR

Specifications for final alignment A detailed procedure was developed for assembly of the coupling in the shop. Requirements of assembly include: • keeping surfaces immaculate • watching for and avoiding surface scratches and digs • avoiding excess and uneven heat application • installing the half keys for balancing during assembly • ensuring that bolts do not bind • ensuring that the lift nut does not affect runout in any orientation • ensuring that bolt and nut weights match within 0.1 grams— not matched to the coupling A set of precision mandrels is used in assembling the coupling for checkout and balance. The mandrels are coated with chrome oxide in the fit areas to ease removal of the hubs and reduce scuffing and galling. The mandrels for both pump and motor hubs allow for a 0.0005” to 0.0020” shrink fit. A diagram and detailed checklist guides each phase of the assembly. An individual balance of the four major components precedes assembly. All components are singleplane balanced according to ISO G2.5 using magnetic mounting plates instead of mandrels (mandrels would introduce inaccuracies into the balancing operation). The lift nut complicates balancing. Its balance must not be altered during assembly balance. It must be free of match marks and it must be capable of being located in more than one orientation. In other words, while the lift nut is present in the assembly balance, no weight correction can be made on it. After the assembly balance, match marks are placed to orient the two hubs and the spool

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for

ALIGNMENT

future

0.001”/0.002” is allowed at the register fit between the cover and pump. These clearances and machining tolerances on the cover yield an actual radial runout of 0.0015”/0.0025” on the cover bore in relation to the pump bore. The parallelism of the two mating surfaces or faces of the upper cover is more critical than these radial runouts. A parallelism between the two mating surfaces of 0.0005” was specified, resulting in a finished relationship to the pump bore within 0.001”. These tolerances eliminate the need for shimming or other questionable techniques to obtain acceptable alignment. With the upper cover mounted to the pump, the motor mounting bracket and motor are installed, with utmost attention given to cleanliness of the fit surfaces. Motor-to-pump alignment is then performed by mounting a dial indicator to the motor coupling hub and indicating off the upper cover bore (0.001” TIRR) and the upper cover face (0.001” TIRF) on a 4” radius (Figure 1). Proper alignment required that the seal chamber be installed after the motor. Two modifications made this step possible. First, a compact design lightened the seal chamber.

Alignment of multistage vertical pumps is crucial for continued reliability. Close alignment starts at the foundation. readings The mounting plate of this pump is grouted level within 0.002” per foot. From this point, the other components are stacked, resulting in a perfectly vertical pump shaft. Precision machine work at the factory and conscientious care of these mating surfaces by technicians in the field is imperative to achieve perfect alignment. Mounting of the upper cover or head to the pump can affect alignment dramatically. This step is critical to ensure that the rotor will run on the center line of the case (which was properly line bored at the factory). If the upper cover is not properly cared for or is not a precision component, it can change the reference to center line. Because the cover must be installed on the pump case after the rotor is installed and before the pump is installed in the field, the inner bore and upper face on the cover become primary reference points for all alignment operations. Thus, the cover cannot be aligned in the field. It must register perfectly to the FIGURE 1 pump case as installed. With this in mind, MOTOR users expect a nearly perfect relationship between the cover and the pump. In reality, though, economics and the capabilities of a precision machine shop determine the tolerances. To determine the relationship between the cover and the pump, the cover was mounted on the pump case with the rotor removed to access reference points in the case bore. The axial and radial runouts of the upper cover, pump case register, and cover/case assembly were checked at a precision facility. A Motor to cover alignment diametrical clearance of The Pump Handbook Series

.001 TIR F MAX

UPPER COVER .001 TIR R MAX

This modification permitted installation without the need for an overhead lifting device. The second modification was to shorten the pump shafts to increase the DBSE (distance between shaft ends) to permit installation of seal housing and seals after installation of the motor. A meticulous procedure is now used to install the coupling. The procedure was developed by plant staff with little help from the manufacturer. In fact, our experience was that while pump manufacturers stress good alignment, they provide little help to the customer in developing good procedures applicable to the plant environment. Also, factory alignment for test stand purposes relies on equipment not often found in the field and is usually performed by “old timers” who carry the procedures around in their heads. The rigid coupling design used on this pump has eight mating surfaces. All must be in near perfect condition for acceptable, long-lasting alignment. The alignment procedure thus requires that all parts be kept very clean and protected and pass a

detailed visual inspection. If any doubt exists, a detailed mechanical inspection is to be made. Installation begins with a shrink-fit of the motor hub to the motor shaft. Sometimes (e.g., after an overhaul) the motor hub is installed at a repair facility where runouts and fits can be checked and corrected as necessary. Hub runouts are checked after any installation, whether in the field or in a shop. Indicators are placed on the shaft and hub during runout check to determine whether the motor has shifted during rotation of the shaft (the motor is susceptible to shifting due to the use of sleeve radial bearings and tilt-pad thrust bearings). Next, the pump hub and lift nut are positioned on the pump shaft. The coupling spool is then installed, with special care taken to align the balance match marks. The bolts are then evenly tightened to prescribed torque and another set of runout readings is taken. The total rotor float is measured at this time and the lift is set within tolerance. Finally, the pump hub is connected to the spool

The Pump Handbook Series

while aligning match marks between hub and spool, and another set of runout readings is taken on the pump hub and shaft. Assuming that the previous alignment and coupling checks are accurate, out-of-tolerance runout readings at this time usually indicate a bent shaft. If acceptable readings cannot be obtained, the root cause of the problem must be determined and solved. Shimming or selective torquing of the coupling is not permitted. A final set of readings is taken and must be within specs (Table 1). Design and operation of this critical pump is complex. By following the procedures developed over time through direct experience with this pump in this service, the plant is now able to achieve high reliability and minimized costs. Users in similar situations should find in this example many ideas for improving their operations. ■ Ken Laplant, P.E., is a rotating equipment consultant for Machinery Resource, Inc., Houston, TX. Charlie Rutan is principle engineer in mechanical maintenance for Oxy Petrochemicals, Alvin, TX.

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What Does It Cost ? BY GARY GLIDDEN ntil recently the main consideration when we were building a new power plant or replacing old equipment was up-front costs. In other words, the lowest bidder always got the job. This approach was dictated partly by public utility commission rules, but also by the attitude among plant engineers that “a pump is a pump.” The cheaper the better. As a result, we ended up with twenty different models of boiler feed pumps from five different manufacturers. Other pumps (e.g., condensate, condenser cooling water, and auxiliary cooling water) purchased the same way made the mix even more confusing. To keep downtime as brief as possible, we had to stock

U

an entire inventory of spare parts for each model of pump. Every OEMrecommended spare part was kept on hand. Although confusing and expensive to support, this diversity of pumps had one advantage: it made the plant a pump testing center. Maintenance records showed clearly which model of pump offered best performance and life in which service. For example, in the 1960s we put two sister units into service. Each had a 100% capacity boiler feed pump driven by the main turbine. The designers and purchasers liked the setup. It did not require very much auxiliary equipment and involved no additional expense for a drive unit.

A main-shaft-driven boiler feed pump.

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The Pump Handbook Series

But we encountered many problems. First, our work force was divided into specialized groups: a crew for the main turbine, a crew for the boilers, and a crew that worked on all components between the turbine and boilers. Because the main boiler feed pump sat on the turbine deck, the boiler crew and the turbine crew constantly fought over use of the overhead crane. (The turbine crew usually won.) Another problem was that if either boiler feed pump went down the whole unit went down. (Remember, each pump ran at 100% capacity). Other units had two or three 50% capacity pumps. If one of those pumps went down, the other

TABLE 1–REPAIR SHOP MAINTENANCE COSTS Inspection Main Boiler Feed Pump #1,2,3 Main Boiler Feed Pump #4 Start-up Boiler Feed Pump #1,2 Boiler Feed Booster Pump #1,2 Circulating Water Pump Condensate Pump Heater Drip Pump Cooling Water Pump Screen Wash Pump Compressor

Labor Hours Material $ Labor Hours Material $ Labor Hours Material $ Labor Hours Material $ Labor Hours Material $ Labor Hours Material $ Labor Hours Material $ Labor Hours Material $ Labor Hours Material $ Labor Hours Material $

70 160 70 60

125 100 80 100 75

Average Repair

Major Repair Rebuild

350 $30,000 500 $30,000 350 $30,000 160 $3,000 1,000 $11,000 300 $3,000 275 $4,000 175 $1,500 225 $1,800 175 $3,000

700 $120,000 700 $160,000 700 $160,000 500 $30,000 1,500 $22,000 5,000 $8,600 500 $8,000 250 $4,000 350 $6,000 200 $9,000

This is a five-year average of labor and materials costs for three types of repairs. The data is used to plan outages and figure man hours and materials costs for future repairs.

units could keep operating at least at reduced loads. But ours couldn’t. And the situation was even more complex. The two pumps rotated in opposite directions because no one thought to provide a reversing gear for the fluid drive coupling on one pump. Often, as we were rebuilding the pump, someone would ask, “Are you sure you put in the right element?” Invariably, this question would come after the element was in place and the pump head was on. Spare parts were a nightmare, too. Some parts would interchange but others wouldn’t. Worst of all was the pumps’ nasty habit of self destructing once each year. Major holidays (especially Christmas Eve) seemed to be their favorite time to come apart. Mechanics became quite superstitious about it. If a pump hadn’t broken down in eight or nine months, no mechanic dared to tempt fate by mentioning that pump. If anyone commented about how well the pump seemed to be working, people would throw coffee cups at him. These two pumps cost our company a great deal of money in down time and repair costs. And the overtime wages they generated paid for many a house and car.

By the 1980s, these two units were scheduled for “low load conversion,” which would help maintain high loads during the day yet bring the units down to 60 Megawatts at night. This was our chance to replace the two pumps from hell. But plant managers didn’t see it that way. They just put up the “there’s nothing wrong with the old pumps” stone wall. So we had to get our little ducks in a row to show what was wrong with the old pumps. We needed hard numbers to show how much they had cost over nearly 20 years of operation. Our preferred alternative was to install two 50% pumps, each driven by its own turbine. But this design was rejected because the extra turbine would add too much exhaust heat to the hot well. We then reasoned that if we couldn’t get a new system, at least we could get different pumps. That seemed like an easy sell—the facts spoke for themselves. Apparently, however, they didn’t speak loudly enough. We didn’t get approval for different pumps. If nothing else, we wanted to install a modification on the fluid drive coupling of one pump that would let the two pumps turn in the same direction. Again, no sell. Too much money. The Pump Handbook Series

So we simply put in new pumps of the same design. We’re still stuck with two pumps that rotate in opposite directions. And we still have to buy two spare elements and all those extra spare parts that don’t interchange. But there is a bright side. The new pumps don’t break down as often. In fact, the single outage over the past ten years was due to failure of the boiler feed pumps. The moral of the story is to consider long-term costs, not just up-front costs. For example, we could have saved money if we had consolidated the spare parts inventory. As you consider long-term costs, don’t forget your smaller, less expensive equipment. It is just as crucial. For example, one of our 750 Megawatt units had three small ammonia pumps (purchased according to lowest bid). These were designed simply to maintain the proper pH in the condensate water. We always had problems with these pumps. A different brand of pump installed at the three other units in the plant caused no such problems. Yet it took a major incident to convince management to replace our ammonia pumps. We had taken one of the problem pumps out of service for repairs. That night the watch supervisor called me at home to report a problem with maintaining the right pH in the condensate. I called out two mechanics and went back to bed. A few hours later the watch supervisor called me back. With panic in his voice, he told me that I had better do something because they were about to lose the unit. About the time I pulled through the plant gate, I heard the safety alarms go off. It took us until the next night to get all three pumps running and the unit back on line. What does it cost for a 750 MW power unit to be down for 20 hours in the heat of summer? A lot more than the price of three new ammonia pumps. Management saw the light on this one. About three months later, we had three new pumps. These two instances occurred before we had computer systems to track labor and material costs. In the case of the two boiler feed pumps, we spent several months digging

61

TABLE 2–MAINTENANCE COST COMPARISON Craft

Number of Work Orders

Labor Hours

Aux Turb Total

2 1 3

404.0 6.0 410.0

Aux CMSH Elec I&C Mech Total

1 3 1 5 2 12

496.0 1,062.5 17.0 29.5 72.0 1,677.0

Labor Material Cost Cost #1B Boiler Feed Pump $8,901.19 $40,344.40 $136.29 $0.00 $9,037.48 $40,344.40 #2A Boiler Feed Pump $11,229.81 $22,358.07 $23,604.06 $100,464.16 $330.55 $1,181.50 $581.54 $1,407.47 $1,666.15 $0.10 $37,412.11 $125,411.30

Other Cost

Total Cost

$0.00 $0.00 $0.00

$49,245.59 $136.29 $49,381.88

$0.00 $2,093.12 $0.00 $0.00 $0.00 $2,093.12

$33,587.88 $126,161.34 $1,512.05 $1,989.01 $1,666.25 $164,916.53

A summary of two reports showing maintenance costs over five years for two pumps of the same size and type.

through files and old work orders to come up with our facts and figures. Now, five minutes on the computer is all it would take. If we had such a system back then, we could have saved a lot of money by getting an answer to the question “How much does it really cost?” Yet computers only provide the facts and figures. Someone still needs to know to look at long-term costs— someone still needs to suspect that spending more money up front might save money over the long term. ■

Gary E. Glidden has been employed by a gulf coast power company for 23 years. For the past 13 years, he has been Foreman in the Energy Production Department. Most of that time was spent in field inspection, monitoring, troubleshooting, and repair of all types of rotating equipment, especially boiler feed pumps. Currently, Mr. Glidden is Crew Leader of the Pump Division in the Central Repair Shop.

TABLE 3–SPARE PARTS RETURNED TO INVENTORY Crew

Labor Labor Material Other Total Hours Cost Cost Cost Cost 4A 4 $139.03 $0.00 $0.00 $139.03 4A 43 $1,377.97 $6,574.56– $0.00 $5,196.59– Crew Subtotal 47 $1,517.00 $6,574.56– $0.00 $5,057.56– 5A 3 $90.95 $313.82 $0.00 $404.77 Crew Subtotal 3 $90.95 $313.82 $0.00 $404.77 8A 23 $631.42 $55.66 $0.00 $687.08 Crew Subtotal 23 $631.42 $55.66 $0.00 $687.08 Total for Craft: 73 $2,239.37 $6,205.08– $0.00 $3,965.71– A summary of a report by one maintenance group showing spare parts repaired and returned to inventory.

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R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

Keys to Improve Pump Record-Keeping BY JULIEN LE BLEU

Olin Chemicals designs ”user friendly” overhaul manuals and data sheets. ecord keeping has always been important. The corporations that do a good job of it also seem to manage their businesses more effectively. The need to make our equipment operate longer, with improved reliability, requires finding a better way to capture the information needed in an easily understood manner. With the implementation of OSHA 1910, it makes even more sense to improve our methods of recording data taken from pumps. With this in mind, work is progressing at our Olin Chemicals plant in Lake Charles to make the overhaul manuals and data sheets more ”user friendly” and complete. If personnel can access and record maintenance data in a record based on illustrations of the pump, they are more apt to use the information and provide accurate documentation. An individual using the data sheet should have all of his information in one place, if possible. In addition, this information should be near the position for recording the data points. Figures 1 and 2 represent the best method we have found to record the data in a manner that answers the questions the pump repairer has during an overhaul. Figure 1 (a and b) is a ”build-up and tear-down” set for a magnetically driven sealless pump. Figure 2 (a and b) is a ”build-up and tear-down” set for a sealed pump. We use a picture of the equipment with an information box for recording the required information, the acceptable limits and an indication of where, in the pump, to look for the information. We try to minimize the amount of writing required, so many of the answers can be provided by circling yes or no. We hope to incorporate these drawings into a windows-type com-

R

puter environment that will use a scanned-in image instead of a drawing. Data from the information boxes will be automatically transferred into a database for analysis and manipulation. Also, limits checking will be incorporated into the data entry, so obviously bad data can not be entered. The data checking will reduce the number of typing errors as well. This format lends itself to our ultimate goal of becoming a paperless plant. Overhaul manuals correlate with the new data sheets and are tailored to the specific pump or piece of equipment. The improvements we have made include: 1. Larger type for some workers to see without their glasses. 2.

A split page format (Figure 3) that allows someone familiar with the work to move down the left side of the page from step to step. Those less familiar with the material or process can read additional explanations on the right two thirds of the page.

3.

The information solicited by the data sheets (Figures 1 and 2) is referred to in the text of the overhaul manual. As a result, confusion about where to put the information or when to record it is minimal.

4.

5.

To eliminate constant turning of pages, all cut away or exploded view drawings are on 11 1/2” by 17” pages, and may be folded out past the edges of the text for easy reference. Within the same pump model there are differences such as metallurgy, seals, and modifications to make each pump operate in its process. These are all captured on the leading page of each overhaul manual (Figure 4). Therefore, the pump will be rebuilt correctly even if the person doing the overhaul has not worked on the particular unit for a long time. The Pump Handbook Series

6.

Data that is sometimes confusing in the text of the original equipment manufacturer’s manual is clarified. As an example, one pump manufacturer has an impeller setting guide that states, start with .015” clearance and add .002” for each 50 degree increase in operating temperature above 100°F. We changed that ambiguous text to a graph representation depicting impeller clearance versus operating temperature. In-stead of calculating the required clearances and possibly making a mistake, the graph makes it easy to find the correct clearance by knowing the operating temperature that neces- sitates setting of the impeller.

7.

Because work is performed on a word processor that accepts scanned-in images, the information may be updated even if the pump is later modified. Moreover, access to the documentation may be limited by the system, so only authorized personnel are able to make changes in the information.

The use of the word processor also lends itself to the elimination of paper and the need for additional typing by importing the word processing files into the network system.

ABOUT THE AUTHOR: Julien Le Bleu is Senior Associate Maintenance Engineer at Olin Chemicals, Lake Charles, LA. He has more than 20 years experience in the field of rotating equipment, including 13 with Olin. Mr. Le Bleu is responsible for all of the rotating equipment in the company’s largest chemical plant.

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The Pump Handbook Series

FIGURE 1A FIGURE 1B

FIGURE 2B FIGURE 2A

The Pump Handbook Series

65

FIGURE 3 WORTHINGTON DM-1010 25-50 NM

OVERHAUL PROCEDURE

5. Slide the drive end bearing sleeve (74B) onto the inner rotor shaft (6).

The sleeve should be firmly slid into place, and the slot in the end of the sleeve should be aligned with the anti-rotation pin. If the slot and pin do not align properly, the drive end sleeve bearing must be removed and realigned

6. Slide the spacer sleeve (78) onto the shaft (6).

7. Slide the pump end tolerance ring (74C) onto the shaft (6).

8. Slide the pump end bearing sleeve (74A) onto the shaft (6).

The bearing sleeve end slot should be facing toward the pump side, and in a position which will align itself with the pump end thrust washer pin.

9. Measure the bushing clearance (Should be 0.0015”, See “A” on the Build Up Report.)

Record the value in the proper place on the Build Up Report.

10. Insert the inner rotor assembly through the drive side of the bearing carrier (215).

Ensure that bearing parts are not damaged.

11. Install the impeller key (32) and the pump end thrust washer (86B).

The thrust washer contains an integral pin which must be inserted into the drive end slot. The pump end thrust washer is kept from rotating by the impeller key.

*sample form

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The Pump Handbook Series

FIGURE 4 Title: WORTHINGTON DM-1010 PUMP OVERHAUL PROCEDURE Equipment #: _______________________________________________________________________________________ Property Index # (PI):________________________________________________________________________________ Serial #: _______________________________________________ Pump Model:________________________________ Pump Size: ____________________________________________ Revision #: __________________________________

SPECIAL OR SPECIFIC FEATURES OF THIS PUMP: 1. _________________________________________________________________________________________________ __________________________________________________________________________________________________ 2. _________________________________________________________________________________________________ __________________________________________________________________________________________________ 3. _________________________________________________________________________________________________ __________________________________________________________________________________________________ 4. _________________________________________________________________________________________________ __________________________________________________________________________________________________

NOTES:___________________________________________________________________________________________ __________________________________________________________________________________________________ __________________________________________________________________________________________________ __________________________________________________________________________________________________ __________________________________________________________________________________________________ __________________________________________________________________________________________________ __________________________________________________________________________________________________

*sample form

The Pump Handbook Series

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R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

Reliability Improvements to a 13 Stage Charge Pump BY DAVID A. JOHNS eliability of a high pressure charge pump has been a problem for more than 20 years. Unplanned unit outages and sometimes fires accompany pump failures. This article details the joint effort of the OEM and the user to solve the problems associated with design and installation. The end result is a pump that has run for three years without incident, decrease in performance, and increase in vibration. The lessons learned from this case history can be applied to pumps by any manufacturer to resolve chronic reliability problems. The Hydrocracker Unit at the Norco Manufacturing Complex depends on a single (unspared) pump to charge oil to the unit (Table 1). Since installation in 1965, this pump has had a record of poor reliability with several seal failures resulting in fires. Previous attempts to improve reliability have made some progress, but the pump has never been able to run from turnaround to turnaround. The majority of the seal failures have been on the coupling end, and most of the failures have occurred during pump startup. Several options to improve the pump’s reliability were considered: • Replace the pump with a new barrel pump.

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•

Add a spare pump.

•

Install a volute type pump.

•

Modify the pump with vendorrecommended hardware improvements.

MODIFICATION SCOPE Since this pump was un-spared, all of the upgrade work had to be performed during a scheduled unit turnaround. To minimize the downtime required, the pump’s spare inner case and rotor assembly were sent to the manufacturer’s repair facility for modification prior to actual shutdown. In addition, changes to be made on the pump case and bearings were closely coordinated with the manufacturer’s shop in advance of shutdown. The manufacturer’s recommended and installed hardware improvements included: • metal-to-metal head to case fit •

larger diameter shaft, stepped at each impeller fit, with shrink fit impellers located by split rings and double keys for each rotor component

•

tapered piston

•

Kingsbury LEG thrust bearing

locknut

balance

•

spiral groove wear rings and bushings

•

remachined bearing housing horizontal split line to remove gasket

•

interface fit bundle (covers)

The user’s specified changes included: • metal bellows mechanical seal •

integral impeller wear rings

•

hard coat rotating wear rings and bushings

•

provisions for thermal growth control

•

reduction of pipe strain

SHOP/BUNDLE MODIFICATIONS Upgrade of the spare bundle was performed in the manufacturer’s shop to lessen the amount of time required for the pump upgrade during the actual turnaround. Extensive repairs were required to the existing bundle before the upgrade work began. New impellers were needed due to the

After reviewing the options, the decision was made to modify the existing pump because of the following advantages: • It was the lowest cost option. •

No piping or pump baseplate modifications were required.

•

Manufacturer-recommended changes had a good track record at other facilities.

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Impeller with counter rotating spiral grooves

The Pump Handbook Series

large amount of welding required to incorporate the integral impeller wear rings. The intermediate covers required extensive weld repair to return all dimensions to original specifications. This was because previous bundle repairs included true cutting, which adversely affected all of the bundle’s axial dimensions. Once these repairs were complete, the real bundle upgrade work began.

position was established by a shaft shoulder on one end with impellers and sleeves stacked from the shoulder. This type of rotor construction has several disadvantages: • Galling probability during rotor assembly increases with the length of same diameter shaft that impellers and sleeves have to pass over.

METAL-TO-METAL HEAD TO CASE FIT

•

Axial stack-up errors are likely, because everything is referenced from one point, the shaft shoulder.

•

Difficulty maintaining rotor straightness, which depends on the squareness of the impeller hub and spacer sleeve faces when the components butt together.

The discharge head serves many functions. It seals the pump’s discharge end and closes the pump’s casing. The head also locates the bundle axially and radially. Additionally, it positions and supports the thrust end bearing housing, the stuffing box, the balance piston seal, discharge diffuser and the gland plate. The head to case joint was a leakage area. The original pump assembly specifications called for clearance between the pump head and casing. The pump designers, in an effort to maintain rotor-to-case alignment, also wanted the head parallel to the case within 0.002 in. The machinist had to tighten 24, 2.0-in studs holding a 500-lb chunk of steel and keep the surfaces parallel within 2.0 mils, which is like winding a Rolex with an impact wrench. This was a challenging task under ideal laboratory conditions, and nearly impossible under field conditions. The metal-to-metal case fit has several advantages: • easier and faster assembly •

minimizes chance of leak

•

easier to obtain proper bolt torque

The modification required machining of the discharge head, the mating surface of the casing, and the inner discharge head. All other machined faces and bores were welded and remachined so that the entire case was brought back to factory specifications.

STEPPED SHAFT WITH DOUBLE KEYS AND SPLIT RINGS

The original shaft was the same diameter from the first to the last impeller, with each impeller driven by a single key. Proper axial impeller

The new design shaft has several features to ease assembly and increase stack-up accuracy. Starting from the first stage, the shaft diameter under each impeller is progressively smaller. Therefore, the impeller has clearance on all fits except the mounting fit. Each shrunkfit impeller is individually located by split rings that fit into machined grooves on the shaft. The impeller is machined to retain the rings. The sleeves are fabricated as part of the impeller instead of each being a separate item. The manufacturer performed a rotordynamics analysis of the new design to predict stability. The model predicted a very stable rotor, which has been verified after three years of operation.

SPIRAL GROOVE WEAR RINGS AND BUSHINGS

Standard wear rings attached to

the impeller by tack welds or set screws were originally supplied. These separate impeller wear rings had a tendency to crack and come off the impeller if suddenly heated, such as during mechanical contact or with the pump running dry. These loose metal parts would then cause further damage in the pump. Modifications were made to minimize the potential damage to the pump due to wear ring touch off or failure. Both the impeller eye and shaft sleeve wear rings were made integral with the impeller. This eliminated the possibility of the ring loosening or breaking. The impeller had sufficient material to allow recutting and making case rings undersize. Wear ring renewal after one recutting required weld buildup of the sealing areas and remachining. The rings and mating bushes were made with spiral grooves on the surface instead of the typical smooth surface (Photo 1). The grooves in the rotating part were in the opposite direction of the grooves in the stationary part. The counter spiral grooves had two benefits: 1. If a gall developed, the grooves would provide a path for the material to exit the seal area, thus minimizing damage to the wear rings. 2.

They provided a better seal, thus increasing efficiency.

The rotating wear rings and balance piston were coated with tungsten carbide to create a greater surface hardness differential between the rotating and stationary parts. Due to the long flexible rotor, there is a high probability that mechanical contact will occur between close radial clearance parts

TABLE 1. PUMP DATA

Type Design Suction Pressure Discharge Pressure Pumping Temperature Pumpage Speed Driver Flow

Pacific HMIJ, 8 in. Horizontal, 13 stage, Vertical split between bearings, Barrel type 100 psig 2500 psig 750˚ Hot Oil 3600 rpm 3000 hp electric motor 2000 gpm

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69

at some point in the pump’s life. The hard coating allows a brief contact without removing metal from the rotating element.

INTERFERENCE FIT BUNDLE When added together the manufacturing tolerances on the original bundle could produce loose fitting covers. Assembly and disassembly of loose fitting parts is easier, but maintaining concentricity throughout thirteen stages is virtually impossible. Since weld buildup was already required to correct the axial spacing problems, additional metal was added to the register fit area on each intermediate cover. All fits were turned parallel and concentric to original specifications. The rabbet fit was changed to produce a one to four mil interference, ensuring that all stationary hardware was concentric.

REMACHINE BEARING HOUSING HORIZONTAL SPLIT LINE The pump’s radial bearings are spherically seated babbitt-lined sleeve bearings. Original design requires the bearing crush to be adjusted by varying the gasket thickness on the bearing housing horizontal split line. This is a

70

time-consuming task, and end results are of questionable accuracy. Present day precision bearing manufacturing techniques eliminate the need for an “adjustable” bearing housing split line. Elimination of the split line gasket requires machining of the horizontal split line and reboring of the spherical seat. After machining, the housings are centered and dowled into position on the pump case.

SEALS The frequency of seal failures allowed for the testing of many seal variations. However, the seals’ poor performance was not entirely due to seal design. The casing distortion and pipe strain were the biggest factors in limiting seal life. The seals selected for the upgrade were a metal bellows type with graphfoil gaskets. Axial position was established through Lshaped split rings that were ground at assembly to locate the seals properly. The rotating sleeve was key driven, and the seal was friction driven to eliminate set screws. Seal flush piping was per API Plan 21 for normal pump operation. When this flow was not available, flush was provided from an external

The Pump Handbook Series

source per API Plan 32. Cooling was provided by a bank of natural convection fin tube exchangers and four shell and tube water cooled exchanges. Seal flush flow was maintained at 4.0 to 5.0 gpm per seal, via manually controlled needle valves. Desired temperature range was 200°F to 250°F. The seals were quenched with low temperature steam. Next month we conclude with a detailed look at the field/case modifications implemented to improve pump reliability. Editor’s Note: This article is based on a paper originally presented at the 1994 International Pump Users Symposium and Short Courses sponsored by The Turbomachinery Laboratory and The Texas A&M University System. ■ David A. Johns is a Senior Mechanical Equipment Engineer for Shell Oil in Norco, LA. He is responsible for the Mechanical Equipment and Technical Staff in the Olefins and Utilities section. Mr. Johns received a BSME from Louisiana Tech University.

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R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

Reliability Improvements to a High Pressure Charge Pump BY DAVID A. JOHNS ast month we reviewed OEM and user recommendations to resolve the problems associated with the design and installation of a high pressure, high temperature, thirteen stage charge pump. Reliability of the pump, located in the Hydrocracker Unit at the Norco Manufacturing Complex, Norco, LA, has been a problem for more than 20 years. This second and final installment details the field/case modifications that resulted in significant MTBPF increase.

L

provisions for thermal growth control were inadequate. The existing scheme included: • one centering pin at the bottom of the coupling end (Photo 1) •

an axial key under the thrust end (Photo 2)

•

Both the centering pin and the axial key had the female part on the case and the male section mounted on the pump base.

•

The centering pin and axial key guide block were tack-welded to a piece of 4-inch pipe that was then welded to the pump’s support legs.

•

Full length water-cooled pedestals support four mounting pads on the pump’s horizontal centerline.

•

The coupling-end horizontal support pads were doweled.

•

Sleeves were installed in the outboard pads to allow the feet to move axially and radially. This was accomplished by making the sleeve long enough to provide clearance between the bottom of the hold-down nut and the top of the pad.

FIELD/CASE MODIFICATIONS Everything was going per plan until three weeks prior to the turnaround, when a power failure caused the unit to shutdown. Restart of the pump was accompanied by a seal failure and fire. Additionally, the pump’s drive motor was damaged severely enough to require replacement. At this point, the decision was made to start the unit turnaround three weeks early.

PUMP REMOVAL This was the first time the pump case had been removed since original installation, and upon disassembly and further inspection, the following damage was found: • The hold-down bolts and dowels were bent. •

The coupling-end (suction end) centering pin was stuck to the case, and attachment welds had failed.

Thermal growth of the pump was calculated using the following equation to verify the observations and determine the magnitude of the problem. ∆L = ∆T x ∝ x L

•

strengthened supports for existing bottom centerline pin and key

•

slotting of outboard end feet

•

sleeving of all four hold-down bolts to allow vertical clearance

•

removal of dowels in support feet

RADIAL KEYS The radial keys (Photo 3) are one-inch square and centered with the existing bottom pin. Machining of the case-half keyway was relatively simple, since the case was already in the machine shop. The case showed the centering pin and the hold-down bolt holes to be on the same plane. The base-half keyways were machined after the pump was returned to the field. The pump was

Supports for the centering key and pin were bent.

∆L = thermal growth (in.)

•

Suction and discharge flanges were out of square with centerline.

∆T = T(product) – T(ambient)

Based on the observed damage, it was clear that the pump’s existing

The calculations identified two areas of concern. The axial thermal growth was more than the clearance in the outboard bolt holes; this is the reason the bolts were bent. Since the coupling end feet are doweled and not allowed to grow, the radial thermal growth forced the pump to distort in an oblong manner. Clearly, the dowels were maintaining alignment, but were also causing unhealthy distortion. The question of how to maintain alignment while allowing the case to grow thermally without distortion was answered by borrowing technology from the steam turbine industry. Thermal growth cannot be prevented, but it can be guided or controlled. The following modifications were made to the pump to make it control thermal growth more as a steam turbine does: • radial keys at the coupling end horizontal supports

where

•

EXISTING THERMAL GROWTH CONTROL

Radial growth from the center = 680°F x 7.23 x 10-6 in./in./°F x 14 in. = 0.069 in. (Figure 2)

L = total cold length (in.) ∝ = coefficient of thermal expansion (7.23 x 10-6 in./in.°F) For this case: Axial growth = 680°F x 7.23 x10-6 in./in./°F x 85 in. = 0.418 in. (Figure 1)

The Pump Handbook Series

71

1. Typical coupling-end centering pin

2. Typical discharge-end centering key

3. Radial keys installed in pump pedestal

4. New centering pin supports

positioned on the base, and the key location was transferred to the base. The pump was removed and base keyways machined. The keys are made of 17-4 stainless steel material and installed with line-to-line contact in the base, with approximately 4.0 mils of clearance in the pump.

ing and aligning the pump, sleeves were fabricated to allow 5.0 mils vertical clearance between the holddown washer and the pump feet (Figure 3) to allow unrestricted movement of the feet.

SUPPORT FEET

Design requirements for the new supports included: • quick installation because startup was fast approaching

The outboard support feet were slotted axially to provide enough freedom of movement. All four foot-bolt holes were enlarged to allow clearance for the sleeve and radial growth. This work was performed while the pump was in the machine shop. Spot facing for the washer was also a machine shop task. After level-

72

AXIAL KEY AND CENTERING PIN SUPPORT

•

support strength because the previous beam had bent

•

use of available material

The Pump Handbook Series

•

keeping it level to avoid binding

•

removable keys for future maintenance

The support for the axial bottom key and centering pin was changed from a single piece of structural channel to a stiffer design consisting of 11/2 in. x 4 in. steel plate (Figure 4). The horizontal plate was carefully fitted and leveled with a machinist’s level, and its position was continuously monitored throughout the welding process. Support gussets were also carefully fitted from the center of the new beam to the pump support rails.

FIGURE 1 SUCTION

ANCHOR

DISCHARGE

THERMAL GROWTH 0.481”

Calculated axial thermal growth

•

only two weeks before start-up

•

hot work (welding, torch cutting, etc.) not allowed

•

piping modifications not permitted

•

pump and associated piping located in a congested area

•

Piping loads had to be below API 610 allowable in four different operating modes. The discharge flow had two possible paths, the normal flow to the process or a minimum-flow recycle line. Both lines were long, complex, and poorly designed from a thermal growth standpoint. The four conditions that were evaluated included: • both lines cold

The old centering hardware was tack-welded to the support beam while the new hardware incorporates a flange mounted on the beam. Installation of the bottom centering hardware was a multistep process: • Install pump, level, center on new radial keys. •

Center bottom pin and key, clamp to support beams.

•

Remove pump.

•

Transfer bolt locations to beam, drill and tap.

•

Install pump, checking for center. Tighten bolts on bottom centering hardware.

•

Remove pump, drill, ream and install dowels.

•

Lubricate all sliding surfaces.

•

Re-install pump.

The pump was now free to grow within its guides, both radially and axially.

EXTERNAL PIPE STRESS External piping forces were examined to ensure that they would not limit pump growth or force it in another direction. Computer modeling of the piping revealed loads well over the present API 610 allowable. It

• both lines hot should be noted that pipe stress • bypass line hot, main line analysis programs were not available cold when the plant was designed in the 1960s. • main line hot, bypass line Resolution of the pipe strain cold problem presented several challenges. The first hurdle was convincing the designer doing the pipe analysis that FIGURE 2 the pump was not a fixed anchor. Initial SUC TION analysis showed all piping loads to be within current API 610 limits. However, review of the input data revealed that the discharge piping was modeled as 8.0 in SCH DISC RADIAL 40 while it is actually HAR GE GR OW TH 8.0 in SCH 160. The dif0.069” ference in stiffness changed the piping loads from acceptable to more than five times API 610 allowable (Table 1). If this installation were still in the design stage, reduction of the pipe strain would be a relatively simple task. However, this wasn’t an option, so the team faced the following restrictions on the pipe Calculated radial growth stress reduction work: The Pump Handbook Series

73

Resolution of the vertical forces was straightforward. It required only spring support installation. The existing spring support was replaced by one large enough to support the load. The moments acting on the discharge nozzle were the highest loads and the source of greatest concern. The normal methods of lowering forces and moments by adding, moving, or sutracting guides and restraints produced more workable solutions. Since the moments could not be reduced without cutting pipe, the designer elected to induce equal, but opposite, moments and forces as close to the pump as possible. This was the one solution that met all of the criteria. Since this in effect applied the high loads to the piping instead of the pump nozzles and casing, additional calculations were performed to ensure that the pipe flanges were within allowable stress limits. Inducing equal and opposite moments and forces was accomplished by mounting three spring supports in the horizontal plane. Bolton clamps were fabricated to attach the spring supports to the pipe, and structural steel was added to provide anchor points. The location and spring rating of the horizontal supports are shown in Figure 4. Piping-to-pump alignment was achieved with the spring travel stops installed. Piping was moved around until the discharge flange could be made up without moving the pump. After all flanges were tight, the stops were pulled, and coupling alignment showed very little change. After the system reached operating conditions, the spring cans were adjusted to calculated hot operating loads.

FIGURE 3

.004 -.006”

1/8” MIN.

Vertical clearance of hold-down bolts

Concurrent with the pump work, a new drive motor was also installed. The original drive motor developed a high synchronous vibration after the pump failed. Repair of this motor was not successful, however, and a replacement ”off the shelf” motor that met the user’s requirements was brought in. The only difference was that the new motor was 18 inches shorter in height. The following steps were performed to adapt the new motor to the old base: • Voids in the existing base were pressure grouted. •

An 18-in. high adapter was fabricated using 1.0-in. plate. Internal gussets and cross bracing made

TABLE 1 Load FX FY FZ FR MX MY MZ MR

74

Units lb lb lb lb ft lb ft lb ft lb ft lb

Before Modification 7,433 5,633 -1,650 9,471 -19,089 12,314 22,569 32,022

the adapter very rigid. Inch-thick pads were added to support the motor.

DRIVER REPLACEMENT

API Allowable 1,700 2,200 1,400 3,120 5,200 3,800 2,600 3,500

After Modification -122 -147 55 199 -1,352 956 2,426 2,937

The Pump Handbook Series

•

The adapter was welded to the existing base and filled with epoxy grout.

•

Motor mounting pads were field machined level with 0.125-in. allowance for shims.

•

The motor was positioned on the base and mounting holes located.

•

Hold-down bolt holes were drilled and tapped.

•

Final alignment was performed.

This pump has operated without incident for the past three years. Shaft vibration has re-mained at 0.2 to 0.3 mils since the modifications were made. One unexpected benefit is a dramatic decrease in the operating noise level. Prior to upgrade, the pump had a distinctive roar. Now its only sound is from the motor cooling fans. Performance has not changed, even though the unit has been through at least 10 start cycles. Operators used to approach start-ups with much apprehension. Fire fighting equipment was strategically placed around the pump, and the emergency response team was put on alert. Now, closely following established operating procedures, the

pump is started with no more concern than with any other pump. All modifications made the pump more like a piece of turbomachinery rather than a common industrial pump. The metal-to-metal head fit, stepped shaft, integral wear rings, and the metal bellows seal were much-needed product improvements. However, these modifications would have been destined for failure had not the pipe strain and thermal growth problems been addressed. ■

FIGURE 4

65 95

70 100

60 75 615

105

85

38

90 620

110

#

0 300

55 50

112 610

#

30

40

17

00

# 115

ACKNOWLEDGMENTS

5

The author wishes to acknowledge John Bertucci and Andy Gallien of Shell Oil Company; and Terry McGuire, John Holland, and Ralph Richard of Ingersoll-Dresser Pump. David A. Johns is a Senior Mechanical Equipment Engineer for Shell Oil. He is responsible for the the Mechanical Equipment and Technical Staff in the Olefins and Utilities sections of Shell’s Norco, LA manufacturing complex. Mr. Johns is also a member of Pumps and Systems’ User Advisory Team.

45 40

Y

120

40 10

Z

X NEW SPRING SUPPORTS

Location of horizontal spring supports

SUMMARY

MODIFICATIONS pipe stress reduction through the addition of horizontal and vertical spring supports

BENEFITS reduction of nozzle loads which were causing casing distortion

2.

thermal growth control through the addition of radial and axial casing guide keys

elimination of casing distortion caused by thermal growth

3.

interface fit covers

4.

stepped split-ring shaft

ease and accuracy of assembly

5.

remachine journal bearing of housing to eliminate split line

accuracy and consistency internal alignment gasket

6.

metal-to-metal head to case fit

7.

hard coated spiral groove wear rings and bushings

less damage if rub occurs

8.

integral impeller wear rings

eliminates wear ring breakage

9.

rotating metal bellows seal

field proven reliability

1.

10. LEG thrust bearing taper lock balance drum

The Pump Handbook Series

latest thrust control technology

75

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R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

Reliability-Driven Pump Maintenance o accomplish profit goals when petrochemical companies and refineries are facing increasing labor and raw materials costs coupled with declining or consistent sales, company managers are continually seeking means to reduce environmental, operational and main- tenance expenditure levels. Short term plans generally consist of reducing costs for waste disposal, equipment maintenance, labor and capital investments. Long term plans should focus on employing more reliable and efficient processes and equipment. Establishing a reliable centrifugal pump repair program, one that leads to improved equipment availability and performance as well as reduced emissions, will also help meet these pressing profit goals. Equipment failures often exceed environmental release allowances, impacting environmental costs for waste disposal or regulation violations, in addition to the more obvious effects on maintenance costs and production rates. Surprisingly, most pump failures are not attributable to normal wear and tear, but are instead the result of premature failure of one or more components in the pump. Parento’s Law, more commonly known as the ”80-20 Rule,” states that in any group about 80% of the total effect will come from only 20% of the participants. Translated, this means the bulk of pump repair maintenance costs will be produced by 20% of the pump population.

T

ROOT CAUSE ANALYSIS Reviewing and analyzing the unit maintenance cost and the failure history of the pumps, drivers and auxiliary support systems (e.g., seal pots, oil mist) provides an effective means to justify a pump reliability program.

If the maintenance cost or failure history is not available, the next best parameter to examine is the work order description. This analysis will likely indicate that the majority of the premature failures on centrifugal pumps are mechanical seals and bearings, as we have found. Often this evaluation will also reveal that non-critical or general purpose equipment, such as pumps, motors and turbines, is more costly to maintain than the critical equipment, such as the major compressors. This circumstance is, in part, due to the high visibility of the critical, non-spared equipment as compared to the spared, non-critical equipment. In general, an operating unit has only a few critical machinery trains, typically 1-5 compressor trains; whereas spared, non-critical pieces of machinery are much higher, typically 100-500 pumps with drivers. A reliable pump maintenance program consists of vibration surveillance, problem detection, diagnosis and corrective action steps. Once a problem condition is discovered, diagnosis and inspection are then required to confirm that the detection was triggered by distress and to identify which component prematurely failed. After the failed component has been located and checked, all other components should also be inspected to ascertain that the failure was not triggered by another problem (e.g., fitup, material dissimilarity, component design, component fabrication). This problem identification process is commonly known as root cause analysis.

CORRECTIVE ACTION Once the root cause of failure has been determined, a corrective action plan can be developed to

BY: TERRY HERNANDEZ 76

The Pump Handbook Series

resolve the problem condition. If production circumstances don’t allow sufficient time to execute the upgrade and/or corrective action during the outage period, document the problem, required action, upgrade design details and schedule for future implementation. Design changes or upgrades should always conform to the OSHA 1910 standard. One of the main ingredients for improving pump reliability is the development of a detailed pump maintenance and repair program. This program should consist of field removal, shop overhaul and re-installation procedures. A reliable pump repair program depends upon the following factors: •

thorough technical pump design knowledge

•

access to pump manufacturer’s repair data

•

practical machining knowledge

•

rigorous field performance testing

•

detailed component inspections

•

quality machining and balance equipment

•

sound repair techniques

•

proven quality control criteria

•

accurate repair fits and tolerances

•

sufficient documentation of pump performance, inspections, repairs and their costs.

The primary goal for improving equipment availability and reliability is to repair the machinery back to the original equipment manufacturer’s (OEM) design. If OEM’s design is flawed, the faulty component design should be upgraded to meet the application requirements. On the other hand, when there is a system prob-

lem, this must be corrected for reliability to improve. Clearly, repairing the pump will not improve its reliability if a system problem exists.

SHOP OVERHAUL Preparation Considerations Since the pump reliability improvement program depends upon inspection and diagnosis, a well thought out pump shop overhaul procedure is essential for success. First, the pump must be properly decontaminated, by removing all hazardous fluids or vapors in the pump wettedend area, prior to bringing it into a confined shop area. Decontaminating the pump involves flushing and draining the casing several times. The decontamination process should be performed according to established operating procedures for removing all trace of process chemicals. If the pump must be shipped to an outside contractor’s repair shop, extra caution should be exercised to ensure that all process fluids and/or vapors have been removed. Strict adherence to the decontamination process is necessary in this case because an outside shop may not possess the capability or understanding to deal with the hazards associated with the pumping fluid. Also, if the shop is inside city limits, this may pose environmental problems with city facilities and the possibility of monetary penalties for environmental violations. Finally, a material safety data sheet on the pumping fluid must accompany the pump whenever it leaves company property. Before the pump is shipped out or brought into the shop, drain the lube oil to prevent the possible release of process fluid from oil that may have been contaminated. In addition, an oil sample, taken during draining, may help identify the cause of the component failure. Once the oil is drained, flush and drain the bearing housing several times with an acceptable solvent. To maximize manpower productivity and minimize pump outage time, the minimum spare parts required to overhaul a pump should be delivered to the pump repair shop one day prior to shut-down. All parts should be stored in a plastic tote bin, or the

equivalent, to protect them from getting damaged or lost. If the pump is shipped to an outside contractor’s pump repair shop, the repair parts should be shipped simultaneously. Table 1 lists the minimal spare parts required for overhaul of conventional and sealless pump designs. Any parts not used during overhaul should be properly protected for storage and credited back into the company’s storehouse for later use. Tables 2, 3 and 4 list safety equipment, general tools and lifting equipment and precision tools essential to quality pump repair procedures. Avoiding Common Measurement Errors During pump repairs, mistakes commonly occur while taking measurements with precision tools. The most frequent mistakes are misreading of dial indicators and micrometers, poor indicator bracket support and failure to calibrate micrometers. With dial indicators incorrect readings are often a result of a failure to properly zero the indicator. Inaccurate measurement also occurs when negative and positive readings are added incorrectly. When the indicator stem moves outward, the reading is negative. When the indicator stem moves inward, the reading is positive. The total indicator reading (TIR) is the total movement. To help prevent mistakes, first locate either the lowest negative reading or the highest positive reading and then zero the dial indicator at that position. When taking a reading, watch the needle movement to ascertain that the final measurement reflects the total indicator movement. To prevent dial indicator support problems, avoid the long overhung rods, which cause indicator sag. Also, gently tap the dial indicator to check whether it returns to zero. If the rod is flexible, it may cause permanent offset, and the dial indicator reading will not re-zero. The most common mistake with micrometers is the failure to calibrate the instrument. Another frequent error is incorrect reading of the device. The error is easily made by misreading the inch size or the 0.025” graduated markings. To preThe Pump Handbook Series

TABLE 1. MINIMAL SPARE PARTS REQUIREMENTS CONVENTIONAL PUMP DESIGN (MECHANICAL SEAL OR PACKING) 1) Radial bearings (antifriction) 2) Thrust bearings (antifriction) 3) Mechanical seal or packing 4) Sleeve 5) Shaft 6) Inducer (if applicable) 7) Bearing housing seals 8) Case gaskets 9) Sleeve gaskets 10) Shims for bearings 11) Thrust bearing lock washer 12) All other locking tabs 13) All other gaskets 14) Coupling bolts & gaskets

SEALLESS PUMP DESIGN (MAGNETIC DRIVE) 1) Radial bearings (antifriction) 2) Thrust bearings (antifriction) 3) Radial bearings (product lubricated) 4) Thrust bearings (product lubricated) 5) Sleeve 6) Shafts (internal & outer rotor) 7) Inducer (if applicable) 8) Bearing housing seals 9) Tolerance or expansion rings for product lubricated bearings 10) Case gaskets 11) Sleeve gaskets 12) Shims for bearings 13) Thrust bearing lock washer 14) All other locking tabs 15) All other gaskets 16) Coupling bolts & gaskets

SEALLESS PUMP DESIGN (CANNED MOTOR) 1) Radial bearings (product lubricated) 2) Thrust bearings (product lubricated) 3) Sleeve 4) Shaft (if replaceable) 5) Inducer (if applicable) 6) Case Gaskets 7) Sleeve gaskets 8) Tolerance or expansion rings for product lubricated bearings 9) Shims for bearings 10) Thrust bearing lock washer 11) All other locking tabs 12) All other gaskets

77

TABLE 2. SAFETY EQUIPMENT

TABLE 4. PRECISION TOOLS

• Safety shower • Fire extinguisher • Goggles–chemical resistant • Face shield • Gloves–leather and chemical resistant • Clean paper towels or rags • Eye wash fountain • Fire suit • Acid or chemical suit • Safety glasses • Fresh breathing air

• 2 dial indicators with attachments (Starrett 196 Series with attachments Catalog no. 196AlZ) • Small hole gauges 1/8”-1/2”(Starrett Catalog No. S829EZ) • Last word indicator (Starrett Model 711S with attachments Catalog No. 71lGCSZ) • Indicator contact set (Starrett Catalog No. 25R) • Flex-o-post magnetic indicator base (Starrett Catalog No. 657T Series) • 2”-12” inside micrometers with case (Starrett 124 Series Catalog No.124BZ) • 0”-12” outside micrometers with case (Starrett Series 436RL) • 0”-6” depth micrometers with case (Starrett Series 445 Catalog No. 445DZ-6RL) • Precision ground straight-edge 36” long (Starrett Series 387 Catalog No. 387-36) • Precision ground straight-edge 48” long (Starrett Series 387 Catalog No. 387-48) • Thread gauge, standard (Starrett Catalog No. 472) • Thread gauge, metric (Starrett Catalog No. 156M) • Telescoping gauges, 5”-16” to 2”-1/8” (Catalog No. S579GZ) • Taper gauges, 1/10” -1/2” (Starrett Catalog No. 269A) • 6” optical flat with 1/10” accuracy (Lapmaster P/n LAP5-0009-001-0621) • 6 X 10 monochromatic light (Lapmaster P/n LAP5-0010-004-0328) • Small chain clamp or fixture for checking seal chamber runouts

TABLE 3. GENERAL TOOLS AND LIFTING EQUIPMENT • 5 ton Drott or Carry Deck picker and lifting accessories–slings and shackles • Work bench with a vise and 3 ton overhead lifting capability • Millwright hand tools (e.g., socket set, open end wrenches, hammers, punches, chisels, screwdrivers, files) • Impact wrenches and sockets • Bearing induction heater with thermocouple • Press (10 ton minimal) • Bench grinder and sander • Hand drill and bits • Taps and dies • Abrasive paper or pads • Solvents or degreasers • Lint free towels for mechanical seals vent this error, always take the measurement twice. If possible, have another person repeat the measurement. To prevent misreading common with inside micrometers, use an outside micrometer to take the measurement from the inside micrometer length. Measurement Terminology Misunderstanding of measurement terminology may also lead to misinterpretation of measured quantities. The following terms are defined for the purpose of clarification:

78

* Starrett and Lapmaster catalog numbers are cited as a reference. The listed parts or their equivalents are acceptable. •

Concentricity–the offset from the center. Refers to circular diameter differences.

•

Perpendicularity–the offset from the measurement plane. Refers to the face squareness as measured with a bore center.

•

Diametrical clearance–the total difference between two diameters. Refers to the clearance between mating parts.

•

Radial clearance–the clearance from one surface to another surface. Radial clearance is also onehalf of the diametrical clearance.

•

Total Indicator Reading (TIR)– the total dial indicator movement on the measurement surface. Refers to runout on the circular measurement surface.

•

Eggshape–the total out of round on a component. Refers to measurements of inside or outside diameters of wear rings, impeller bores and bearing and shaft fits.

•

Taper–the diameter change over the measurement fit. Refers to measurement of inside or outside The Pump Handbook Series

diameters of impeller and bearing fits. Criteria for Repair The pump repair technician should have access to pump maintenance and technical information. Table 5 lists some essentials. Prior to disassembling the pump, all joints and piping should be match marked for easier and proper reassembly. Furthermore, all wearable parts that are removed should be collected and stored for inspection by the machinery engineer, maintenance engineer or pump reliability improvement team. A weather proof tag should be attached to each wearable part. The tag should identify the part and list the pump identification number, work order, repair date, unit, pump repair technician(s), and any other information that is available. This information will be useful for identifying the root cause of the pump problem and developing or evaluating design changes. During the overhaul process the machinery engineer or specialist responsible for monitoring pump

maintenance cost or reliability will need to identify the criteria for inspection and dimensional checks specific to the component undergoing repair. (These criteria points will be presented in detail in subsequent segments of this handbook series.) This person should also review, and either approve or reject, any procedure that does not meet the criteria. An economic evaluation should also be performed for each pump. This evaluation should weigh, foremost, the cost of repair versus replacement and will differ considerably for the common ANSI (B-73.1 & B-73.2) and API (Standard 610) type pumps. Figures 1 and 2 show the ANSI and API horizontal shaft with bearing housings designs, which are the most frequently used designs. The ANSI centrifugal pump design is generally employed for light to medium service. These pumps also offer dimensional interchangability for handling corrosive, hazardous or toxic chemicals. For the smaller and medium ANSI pump frame sizes, it is generally more economical to replace the entire pump rather than two or

TABLE 5. MAINTENANCE AND TECHNICAL INFORMATION • OEM pump maintenance and operation manual • Pump performance curve • Pump material of construction • Pump process data sheet • Pump parts list with OEM part numbers • OEM pump data sheet • Pump cross-sectional drawing • Coupling installation manual • Process schematic drawing • Coupling drawing • Mechanical seal materials of construction • Seal flush schematic • Mechanical seal parts list with OEM part numbers • Auxiliary equipment drawings, data sheets, and replacement parts information

FIGURE 1. ANSI HORIZONTAL OVERHUNG PUMP

4 1

13

19 9

2

8

18 10 3

5

14 16

12 11 17 15 7

6

20

ANSI B73.1 Overhung Impeller, Separately Coupled, Single Stage, Frame Mounted 1. Casing 2. Impeller 3. Shaft, Pump 4. Back-Plate, Stuffing Box 5. Packing or Mechanical Seal 6. Bearing, Radial 7. Gland 8. Bearing, Thrust 9. Housing, Bearing 10. Locknut, Bearing more wetted-end components, (i.e., the casing, impeller or back-plate). Because these pumps are manufactured in large quantities, the parts and complete pumps are available within relatively short delivery periods. The short delivery and the relative low cost of replacement parts make it more cost effective to replace rather than repair the parts. However, this may not be the case with parts employing exotic alloys. On other hand, for non-metallic ANSI pumps, the components cannot be repaired and must be replaced. The Pump Handbook Series

11. Ring, Lantern (Packing only) 12. Cover, Thrust Bearing 13. Deflector, Radial Bearing 14. Key, Coupling 15. Seal, Radial Bearing Cover 16. Seal, Thrust Bearing Cover 17. Shim, Frame Liner 18. Lockwasher 19. Adapter, Frame 20. Gasket Conversely, API centrifugal pumps are more economical to repair than to replace. These pumps are generally employed in rugged duty and hazardous applications in the refineries or petrochemical industries, and they are subsequently very durable and more expensive. As a result, the pump delivery is longer, and the parts are more costly (i.e., casing and back-plate) than for the ANSI pump design. Cast parts are generally the longest lead items on the API pumps.

79

FIGURE 2. API HORIZONTAL OVERHUNG PUMP

ing pressure ratings for the standard ANSI and API pump casing designs are 300 psig at 5 1 300°F and 750 psig at 500°F respectively. Above these ratings the pump is considered a 12 6 22 9 special design. 13 14 The conventional overhung impeller design pumps 10 have a radial split casing. Designed perpendicular to the 4A 16 shaft, this casing is suitable for high pressures and/or high temperatures because of the symmetrical bolting. The seal11 3 ing design of the radial split casing is also more compatible 15B 18 with low specific gravity fluids (i.e., 0.5 or less). Most ANSI and API single 8 17 19 stage, overhung pump casings 4 15A 21 are designed as back-pull-out 20 7 2 pumps for ease of mainte4B nance. To meet this criterion, there is one major difference API 610 Overhung Impeller, Separately Coupled, between the ANSI and API radially split casing design. Single Stage, Centerline Support This difference is in the pump back cover and gasket area as 1. Casing 12. Ring, Lantern (Packing only) illustrated by Figures 1 and 2 2. Impeller 13. Cover, Radial Bearing (Ref. 1). The most common 3. Shaft, Pump 14. Cover, Thrust Bearing ANSI design uses the bearing 4. Wear Ring, Casing 15A. Deflector, Radial Bearing frame adaptor, usually cast 4A. Wear Ring, Impeller Eye 15B. Deflector, Thrust Bearing iron, to hold the pump back cover and gasket against the 4B. Wear Ring, Impeller Hub 16. Key, Coupling casing. The gap, indicated by 5. Back-plate, Stuffing Box 17. Ring, Oil ”F” in figure 3, allows uneven 6. Packing or Mechanical Seal 18. Bushing, Stuffing Box torquing of the bolts, which 7. Sleeve, Shaft 19. Seal, Mechanical, Stationary Element may cause the adaptor to frac8. Bearing, Radial 20. Gasket, Casing ture. The API design employs 9. Gland 21. Seal, Mechanical, Rotating Element a confined, controlled com10. Locknut, Bearing 22. Housing, Bearing pression gasket with metal-tometal fits. Some ANSI pumps, 11. Nut, Impeller especially small frame designs, adopt the API design convention by bolting the back cover directpresent on a sealless pump in addiFor both pump designs, the ly to the casing. The common ANSI tion to an internal rotor, product repair’s effects on safety, reliability, casing design is not recommended for lubricated bearings and a liner or environmental hazards and producapplications in which the fluid can containment shell. To begin the wettion losses are also important factors vaporize or in which the process conted end overview, this month’s disto consider. ditions risk over-pressuring the casing cussion will focus on the casing. WETTED END TECHNICAL OVERVIEW and fracturing the cast iron bearing CASING DESIGN adaptor. To avoid this problem, use All wetted end components are confined gaskets, bolt back covers or The pump casing for ANSI, API in contact with the process fluid. In a cast steel frame adaptor on the and sealless pumps serves two pura conventional pump design, the ANSI design, or apply API pumps in poses, first to contain pressure and wetted end consists of the casing, services where the casing may be second to convert fluid velocity or inducer, impeller, back-plate and over-pressured. kinetic energy into potential energy mechanical seal or packing. With the Pump casing designs employ measured as head generated or presexception of the mechanical seal and volutes or diffusers to convert the sure. The maximum allowable workpacking, these components are also

80

The Pump Handbook Series

fluid velocity imparted by the impeller into pressure. Figure 3 illustrates the various designs. Most ANSI magnetic drive and some API pumps adopt a single volute design because of the relative small size, low flow rates and low specific speeds available with this design. The spiral shaped volute collects the discharge fluid from the impeller. The area of the volute increases at a rate proportional to the rate of discharge liquid from the impeller, producing a constant fluid velocity around the impeller periphery. This fluid velocity is then diffused in the casing nozzle to induce a pressure head. The spiral shape of the volute also produces a radial load perpendicular to the shaft, and the bearings must function to absorb or transfer the radial loads to the bearing housing. The greatest radial load is at shutoff (i.e., highest pump pressure). This load, then, decreases as the flow increases to a minimum load at best efficiency point (BEP) flow. At flows higher than BEP, the load again increases because of an uneven pressure distribution. Some API pumps adopt a double volute or splitter design to reduce radial loads on high flow and high head pumps. Canned motor pumps employ a concentric casing design to reduce radial impeller loads. The compromise with these designs that reduce impeller loads is a loss of efficiency as compared to the single volute casing. Figure 4 illustrates some critical dimensions for reducing hydraulic shock and the subsequent noise and vibration resulting from the conversion of velocity energy into pressure. The limits shown have proven to be effective by some pump operators. The vane tip clearance ”B” controls the strength and amplitude of hydraulic shock created by vane passing, while gap ”A” controls the pressure pulsation behind the impeller, a phenomena that is responsible for higher axial dynamic forces.

CASING INSPECTION AND MEASUREMENT CHECKS Prior to removing the casing, visually inspect the suction nozzle, discharge nozzle, auxiliary connec-

tions, casing feet, sealing joints and general overall condition of the casing. The estimated time required to perform the inspections is 20 minutes. Any findings during the inspections should be documented on pump repair data sheets. • Suction nozzle inspection should note any signs of cracks, flange gasket surface damage or other evident distress. Also, inspect the inside surface for erosion, corrosion or wear. On pumps with nozzles 3” or larger, check the impeller to casing clearance on open impeller designs or wear ring clearance on closed or semi-closed impellers. This clearance check provides a good indication of wear and casing alignment or shaft deflection problems. Also, on pumps with inducers, check the clearance between inducer and nozzle. The clearance check on the inducer provides a good indication of shaft deflection or casing alignment problems. •

Discharge nozzle inspection should note signs of cracks, flange gasket surface damage or other evident distress. Also, inspect the inside surface for erosion, corrosion or wear. If possible, during the nozzle inside surface inspection, check impeller discharge passage alignment with casing volute.

•

Auxiliary nozzle in-spection should note mechanical damage, valve leakage, auxiliary connection leakage and cracks in threaded or welded areas of the connection. Also, check the piping thickness for corrosion problems.

•

Casing feet inspection should note corrosion, cracks, twisting or distortion. The check can be performed using a straight edge and a feeler gauge.

•

Casing to back cover joint inspection should note signs of leakage and distortion as well as the gap distance between the mating parts. The gap distance provides an indication of impropThe Pump Handbook Series

er bolt torquing, gasket problems or casing distortion from pipe strain. •

Overall casing condition inspection should note signs of cracks, corrosion, distortion or other evidence of distress.

Once the casing is re-moved, perform additional inspections of the casing internal impeller running surface area, wear ring area, volute, volute throat area, gasket surface and register fit. The estimated time required to perform the internal visual inspections is 20 minutes. •

For an open impeller pump inspect the contoured casing surface where the impeller vanes run for wear, erosion or corrosion. The casing should be replaced or repaired if the grooves are more than 0.125” deep. Also, evaluate the wear pattern to provide an indication of casing misalignment or shaft deflection.

•

For a closed or semi-closed impeller pump inspect the wear ring area for cracks, wear, erosion or corrosion. Also, inspect the wear ring locking mechanism for cracks or other signs of distress. The wear ring should be replaced if the clearance is greater than 150% of normal clearance. For the normal diametrical clearance, consult the pump manufacturer. If the diametrical clearance data is not readily available, apply the API clearance specifications of 0.010” plus 0.001” per inch of ring diameter for rings up to 12”. For rings over 12” add 0.0005” per inch of ring diameter. Also, an added clearance of 0.005” is required for galling materials and for temperatures over 500°F add 0.010”.

•

For a pump with an inducer inspect the inducer running surface area for cracks, wear, erosion, corrosion or other signs of distress. The inducer housing should be repaired or replaced if the clearance is greater than 125% of normal diametrical clearance. For the normal dia-

81

metrical clearance, consult the pump manufacturer. If the diametrical clearance is not readily available, apply the API diametrical clearance specifications. •

•

FIGURE 3. ANSI CASING DESIGN

Volute/casing

Stuffing-box cover

Inspect the gasket surface for cracks, wear, scratches, fluid cuts, pitting or other signs of surface irregularities or distress. The casing should be re-placed or repaired if the surface damage is more than 0.030” deep. Inspect the casing to back-cover alignment fit area for cracks, wear, corrosion, or any other signs of surface irregularities or distress. If damage is evident, further inspection is required to determine whether problems exist with concentricity, perpendicularity or parallelism with other critical areas (i.e., wear ring and support feet).

Frame adaptor/ intermediate

Packing gland

F Gap F invites rupture and uneven torque on bolts

FIGURE 4. API OR UPGRADED ANSI CASING DESIGN Stuffing-box cover Volute/casing

Frame adaptor/ intermediate

•

Inspect the high velocity area, volute area Wear ring and volute throat area (cut-water entrance to volute) for wear, erosion, corrosion, pitting or other signs of distress. The casing should be replaced or repaired if the surface damage is more than 0.125” deep. If the cut-water is eroded more than 0.125”, the dures to identify any cracks. The casing should also be repaired or NDT dye penetrant testing time is replaced. estimated at 30 minutes. If the other non-destructive methods, such as wet • Inspect the internal surface of magnetic particle or wet fluorescent the suction, discharge and dye particle, are to be employed, all auxiliary connection for wear, testing should be performed at one erosion, corrosion, cracks or time to minimize cost and manpower. other signs of distress. The casing Cracks generally occur in the should be repaired or replaced if nozzle to casing, auxiliary connection the wear or pitting is more than to casing, gasket surface, high veloci0.125” deep. ty, volute throat, wear ring, previous Once the visual inspection is repair or impeller rub areas. On magcomplete, the casing should undergo netic materials the casing can be non-destructive testing (NTD) proceinspected with wet magnetic particle

82

The Pump Handbook Series

testing. On non-magnetic materials the casing can be inspected with wet flourescent particle or dye penetrant methods. The results of the test should be documented on the pump repair data sheets even if no cracks are found. In some cases, ultrasonic or x-ray testing may be necessary to identify the problem. A NDT report should be requested if crack inspection work is performed by a third party contractor. If a casing has a jacketed area, the jacket should be inspected for possible fouling, corrosion or other signs of distress. On a non-metallic pump the casing liner should be spark tested to determine if there are any cracks or holes in the liner. Otherwise, for solid non-metallic materials the recommended inspection method is the same as for a metallic casing.

CASING REPAIRS AND MACHINING Once the casing inspections outlined in last month’s article are complete, the repair process can begin. The following lists some general procedures that should be performed whenever a pump casing requires maintenance. 1.

Chemically clean the jacket area (if applicable). This cleaning is generally done by a third party contractor. After completion, inspect the jacket to ensure relative freedom from debris and corrosion. 2.

Remove the wear ring. First grind the tack welds or drill out the pins and then remove the ring by machining with a lathe or boring mill. Alternatively, mill a groove in one or two locations on the ring and collapse it.

3.

While the casing is mounted in the lathe or boring mill, check the casing register fit and gasket surface for concentricity, paral-

lelism or perpendicularity problems. 4.

Once the wear ring is removed, a light cut (0.0010.002”) should be taken in the wear ring fit to make the casing concentric with the registered fits.

5.

Machine a new wear ring with an OD 0.002-0.003” larger than the casing bore.

6.

Replace the wear ring by pressing it into the housing and then locking it into place. The preferred method to lock the wear ring is by tack welding it in 3 or 4 locations. If the wear ring and impeller can’t be welded, the wear ring can be axially drilled and tapped at the joint. After the

8.

the correct clearance, consult the pump manufacturer’s maintenance manual. If the manufacturer’s clearance is not readily available, apply the API diametrical clearance or use rule of thumb – 0.010” plus 0.001” per inch of ring diameter up to 12” or for rings over 12” add 0.0005” per inch of ring diameter. In addition, for galling materials add 0.005” and for temperatures over 500°F add 0.010”.

9.

On a pump with an inducer, the same procedure described in steps 2-7 above can be applied. The only difference between a wear ring and an inducer running surface is the length. The wear ring is usually 1-2”, whereas the inducer can range from 3-6”. The clearance for the inducer is criti-

10. During the machining of the case wear ring or register fit, also machine the support feet if they were distorted or twisted. The feet need to be machined to where 80% of the foot area is cleaned up. They should be parallel to the register fit centerline within 0.002”, or perpendicular to within 0.001”, whichever applies to the casing/feet relationship.

FIGURE 5. CASING DESIGNS

Single volute

Combined circular volute casing

Circular

Double volute

Casing Diffuser

Impeller Typical diffuser-type pump setscrews are tightened, the end of the ring should be peened to prevent it from backing out. 7.

After the wear ring is secured in place, machine the I.D. of the ring to provide the necessary diametrical clearance for the impeller. To determine

cal because the inducer performance is dependent upon fluid by-pass. An inducer clearance should be per the pump manufacturer’s maintenance manual clearance. If the clearance is not readily available, use the same clearance as the wear ring clearance specified in step 7.

The Pump Handbook Series

Often, the casing register needs to be reclaimed because of excessive clearance or corrosion. The fit is typically reclaimed by welding in three or four tabs and then machining the tabs to obtain the correct clearance. The clearance should be 0.0004” or less between the casing register fit and the back cover (API pump) or frame adapter (ANSI pump).

11. While the case wear ring or register fit is machined, the gasket surface needs to be cleaned up. The gasket surface should have a maximum of 0.003” of runout and be free from any pitting or scratches in the sealing surface. In addition, it should be perpendicular to the resister fit centerline within 0.002”. Before a casing is repaired, the machinery engineer or specialist should evaluate the feasibility of repair by welding or coating relative to the casing’s replacement cost. Frequently, it is more economical to repair the casing rather than replace it, particularly if it is part of an API pump or exotic alloy materials are used. The most effective method to repair the casing is by welding and re-machining as required. If done correctly, weld repairing a casing should be considered a long term repair. For special welding considerations, the material should be identified and an appropriate weld procedure followed, in accordance with ASME Section IX standards for pressure

83

about a critical area (i.e., volute throat area), consult the pump manufacturer. In some situations, the casing b3 should be sent to the authoDiffuser or rized OEM pump repair volute shop. Generally, the casing δ drawing, which is a primary source of maintenance infor“E” mation, is proprietary, and “F” the pump OEM won’t Impeller release it to any repair shop b2 or end-user. Casing flanges should be weld repaired when the D2 serrations are damaged on D3 raised face flanges or when Flow the gasket surface is damaged on flat face flanges. For raised face flanges, the faced Split retaining ring serrations finish should be 20-40 grooves per inch, 0.002-0.005” deep and cut Recommended radial gaps for pumps spirally or concentrically. TYPE Gap “A” Gap “B”* - percentage of For flat face flanges, the surimpeller radius face finish should be 125 Minimum Preferred Maximum root mean square (rms). The Diffuser 50 mils 4% 6% 12% end suction flange must be Volute 50 mils 6% 10% 12% perpendicular to the casing support feet within 0.002”. R = Radius of diffuser/volute 3 100 (R3 - R2) *B = The top suction or discharge inlet R2 R2 = Radius of impeller flange must be parallel to the casing support feet within NOTE: If the number of impeller vanes and/or diffuser/volute 0.002”. vanes are both even, the radial gap must be considerably Occasionally, the casing larger (10% minimum) can be repaired by applying a coating to reclaim the worn areas. Since the coating’s chemical containing components, or American compatibility with the process fluid, Welding Society standards for nonsurface preparation, bonding strength, pressure containing components. The temperature control, pre-heating charwelder should be qualified for that acteristics and porosity are critical to specific weld procedure within the its performance, the coating should be last 12 months. Preheating and postapplied by a specialty coating contracweld heat treatment should be tor to ascertain the requirements for applied to eliminate possible heat disperformance are met. Coatings are tortion or residual stress effects. High employed effectively when erosion residual stress may lead to a higher has occurred. However, if the casing cracking tendency or lower corrosion material has been affected by corroresistance. sion, using a coating should be considIf any major repairs are perered only as a short term repair until formed (i.e., weld depth is 20% of the a replacement casing is obtained. casing wall thickness or greater), the If repairs are not feasible comcasing should be hydrostatically testpared to the cost of casing replaceed to 1.5 times the maximum allowment, the casing should be replaced able working pressure. The with the pump OEM’s, rather than a hydrostatic testing requirements replicator’s casing. Since the casing is should be determined by the machina pressure containment component, it ery engineer or machinery specialist is vital to the pump’s performance. based on the location and degree of The cost savings for using a replicaweld repair. If there is any question gap “A” gap “B”

FIGURE 6. IMPELLER GAPS

84

The Pump Handbook Series

tor’s casing is insignificant if the pump does not meet its performance requirements. Even if obtained from the OEM, the end user should always request a positive material identification and a hydrostatic test check to ensure the replacement casing meets the design specifications. When a casing is inspected or repaired, wear patterns or failure should be examined as a means for identifying other possible performance problems. The list of information below is general. There are no good rules of thumb because the effects are dependent on the pump size and the casing conformance to manufacturer’s design. •

Excessive runout on the impeller running surface of the casing on an open impeller will lead to reduced head and efficiency (i.e., increased horsepower). If the pump size is marginal for the process requirements, the reduced head may cause less flow.

•

On closed or semi-closed impeller casings, excessive eccentricity in the impeller to case wear ring area may cause wear ring rubs or excessive clearance on one side. The rub will cause an increase in horsepower. Excessive clearance will lead to reduced head and efficiency (i.e., increased horsepower).

•

Casing running surface out of parallel with the impeller vane surface will lead to the same problems as with excessive runout. Parallelism should be maintained to 0.002” or less.

•

Volute width is critical since the volute converts velocity into pressure. The head will increase as the volute width decreases. This head increase can overload the driver, increase impeller radial load, and increase impeller axial load. There are small reductions in head with the increase in volute width. Since the head varies with the pump size and design, consult the pump OEM for the effects on a specific pump.

•

The volute throat area is critical because of the throttling effect

with increasing flow rates. At shut-off (”0” flow) the head is the same, while at higher flow rates, it is reduced according to the throttling effect. Since the head varies with the pump size and amount of eccentricity, consult the pump OEM for the effects on a specific pump. •

The volute width and throat problems also apply on dual volute pumps.

•

The volute must be concentric with the shaft centerline to prevent the reduction of head and efficiency. Again, since the head varies with the pump size and amount of eccentricity, consult the pump OEM for the effects on a specific pump.

•

The casing must be parallel to the volute centerline because the cocking effect results in a varying volute width. The varying volute width is described above. The head reduction is dependent upon the location of the throat relative to how the casing plane is cocked. Because the head varies

with the pump size and the amount that the casing is nonplanner, consult the pump OEM for the effects on a specific pump.

CASING RE-ASSEMBLY CHECKS During pump re-assembly the casing critical areas should be checked. These include the casing gasket surface, register fit, wear ring and inducer areas, flange gasket surfaces, and impeller running surface. While making these checks, visually inspect the overall condition of the casing, auxiliary connections and nozzles. The following dimensions should be measured and recorded. (The estimated time to make these inspections and dimensional checks is 20 minutes.) • register fit ID • wear ring ID • inducer area ID • impeller clearance • wear ring clearance (after reassembly) • discharge passage to volute alignment (after re-assembly)

Once the pump is re-assembled, use a feeler gauge to determine the radial clearance between the impeller wear ring and the casing wear ring. This measurement is a good indicator of the assembled alignment between the mating parts. The radial clearance should be equal to or within 0.001” around the diameter. If the clearance exceeds 0.001”, consult with the machinery engineer or specialist. Finally, rotate the pump to check for casing rubs. ■

ABOUT THE AUTHOR: Terry Hernandez is a maintenance engineer at Quantum Chemical Corporation’s LaPorte, TX complex.

REFERENCES 1.

Bloch, H. P. and Johnson, D. A. ”Downtime Prompts Upgrading of Centrifugal Pumps,” Chemical Engineering. November, 1985.

Fluid and Material Basics THE PUMPING FLUID In the petrochemical and refinery industries, pumps are employed to transfer a wide spectrum of chemical fluids at various pressures and temperatures. The specific fluid properties and operating conditions stipulate the pump hydraulic and mechanical design required for that particular process. The conventional (mechanical seal or packing) and sealless design pumps are capable of handling many types of fluids. The pumping fluids can be simply categorized as hydrocarbons, aggres- sive chemicals or mild chemicals. • Hydrocarbons are petroleum based products generally classified as light, intermediate or heavy. At

atmospheric pressure and temperature, light hydrocarbons tend to vaporize, intermediate hydrocarbons are liquid, and heavy hydrocarbons are highly viscous or solid. • Strong acids, alkalines or oxidizing agents are generally referred to as aggressive chemicals. These are destructive to the equipment and the environment and dangerous for people who are exposed to them. • Fluids classified as mild chemicals are generally easy to handle and are not detrimental to the equipment, environment or personnel.

The Pump Handbook Series

Hazards produced by pumping fluid leakage include vapor clouds, fire, and toxic fluid exposure. Vapor clouds are a common danger with pumping fluids that vaporize at atmospheric conditions or are exposed to very high operating temperatures. If a vapor release is exposed to an ignition source, the vapor cloud may transform into an explosion or fire, especially if the fluid has high oxidizing or pyro-phoric properties.

EROSION AND CORROSION PHENOMENA Erosion and/or corrosion may cause premature failure of equipment. Erosion results from the presence of solids in the fluid, high fluid velocities or the formation and col-

85

lapse of vapor bubbles, commonly known as cavitation. [See this month’s article ”Hydraulic Instabilities and Cavitation” for more details regarding cavitation-induced failure.] The surface of an eroded component exhibits pitting with smooth edges as compared to the sharp, irregular edges produced by corrosion. Corrosion is generally the result of the fluid characteristics and operating temperatures. The brief descriptions of corrosion types that follow may be helpful in identifying the root cause of equipment failure. • Uniform corrosion, the most common type of corrosion, is a relatively consistent deterioration over the entire surface of the exposed component. The corrosion results from a chemical reaction between the component material and the pumping fluid. The rate of deterioration can be very slow or very rapid, but it is predictable. So, the life of the component can be estimated from corrosion charts and component thickness or from immersion test results. • Pitting corrosion is a localized effect of surface film deterioration. Once the protective surface film is breached, the metal begins to break down, producing small or large pitting holes accompanied by sharp irregular edges. • Stress corrosion cracking is localized corrosion caused by load or residual stresses in the component and the environment. This is common in austenitic stainless steels, which are subject to free chloride attack. • Intergranular corrosion is degradation of metal around the grain boundaries and is associated with certain stainless steels. It usually occurs in pump castings because of the relative large grain size or around welds because of the heat affected zone. • Fretted corrosion is a mechanical-chemical process occurring

86

when two materials rub together causing the material particles to weld. As the welded particles separate from their base metals, they react with the environment forming corrosion debris in the mating joints. • Galvanic corrosion is an electrochemical reaction between dissimilar metals. When an anode material (least noble) is connected to a cathode material (most noble) and both are immersed in an electrolyte medium, corrosion develops. The corrosion rate of the anode increases while the cathode corrosion rate is decreased. This corrosion occurs via the same mechanism that operates a battery. There are various charts available in mechanical engineering handbooks, metallurgical handbooks or from the Hydraulic Institute Standard for determining the galvanic corrosion between various metals based on their nobility.

•

austenitic stainless steels (300 series, alloy 20, and Nitronic® grades)

•

duplex stainless steel (CD4MCU)

•

exotic alloys such as nickel, hastelloy®, titanium, monel®, tantalum and zirconium

Non-metallic (constructed either from a solid material or by lining carbon steel components with a nonmetallic material) •

polypropylene

•

polyvinylidene fluoride (Kynar®)

•

polyphenylene sulfide (Ryton®)

•

polyethylene

•

fiberglass

•

high silicon Lining materials

•

Teflon®

•

glass

• Crevice corrosion is localized corrosion created by the presence of stagnant chemical solution in small, confined areas (i.e., joints).

•

ceramic

•

rubber compounds such as neoprene, Viton®, ethylene propylene and buna.

PUMP MATERIALS OF CONSTRUCTION

Before repair procedures are performed on any pump component, the material should be identified through non-destructive testing with a nuclear analyzer or other chemical identification test kits. This process is commonly known as positive material identification (PMI).

Pump manufacturers have an assortment of materials for use in pump fabrication. Material selection depends on the operating stress and effects of wear from corrosion, erosion, abrasion, temperature, entrained vapors and solid contents. Common materials employed for centrifugal pumps Metallic • cast iron •

ductile iron

•

bronze

•

carbon and low alloy steels (1018, 4140, 4130)

•

chrome steels (11%, 12% or 13%)

•

martensitic stainless steels (400 series grades)

•

precipitation hardening stainless steels (17-4 PH and 15-5 PH) The Pump Handbook Series

✓

R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

Baseplate Design Affects Reliability

B

TAKING AIM ON ALIGNMENT Industry recognizes that mean time between planned maintenance can be extended by reducing internal stress loads and vibration on process equipment. Rotating equipment, such as an ANSI centrifugal pump, is normally shipped on a baseplate along with a motor and then installed, aligned and coupled at the plant site. Misalignment of the pump shaft with the motor shaft adds stress to the bearings of both rotating units. Final alignment is recommended to be below 0.002 in. parallel and 0.0005 in./in. angular. Alignment must be maintained at these levels to maximize MTBPM. Added stress caused by shaft misalignment reduces bearing, coupling and mechanical seal life. This stress increases maintenance and downtime costs, and it increases power costs resulting from added motor torque. Baseplate flatness is a major factor in aligning the shafts to critical tolerances in a timely, costeffective manner. A slight bend in a baseplate can cost several hours

of effort to achieve proper shaft alignment. It is recommended that baseplates have a surface flatness coplanar within 0.062 in. overall for Group I and Group II sizes and coplanar within 0.015 in./ft. overall for Group III sizes. These tolerances allow for reasonable alignment capabilities. However, some customers specify that baseplates be machined coplanar to 0.002 in./ft. They justify the extra cost for machining primarily by quicker and easier alignment to recommended critical tolerances. Rigid baseplate designs are less likely to bend or twist during shipment and handling or the anchoring/ grouting process. Laboratory twist tests conducted on several types of baseplates demonstrate how the pump and motor shafts can be affected (Figures 1 and 2).

BASEPLATE DESIGN IS CRITICAL The baseplate also must be designed to be rigid enough to avoid diaphragming at the center under pump and motor loads. In combination with the pipe loads, the forces can be very high. Diaphragming can cause separation of the base and grout, increased resonance and vibration, and misalignment of previously aligned shafts. Because diaphragming may not show up until the unit is operating and pump and motor torque loads are applied, make sure that this phenomenon is eliminated from the baseplate design by adding reinforcement in larger Group II and Group III pump sizes. The baseplate understructure should provide a means for the grout to hold the baseplate down when all the live loading occurs. ”Bolt binding” at the motor feet fasteners, the ultimate alignment problem, is often caused by twisted baseplates. Bolt binding can be avoided by using baseplates of proper rigid design and quality, and by following correct installation procedures. Using ”centering nuts” in initial assembly is a good way to locate the motor in a position with optimum adjustment capability (Figure 3). Centering motor fasteners to the motor holes and aligning the pump to the motor before shipment should provide ade-

FIGURE 1. BASEPLATE RIGIDITY TEST TWIST MODE A

.070 .060 Deflection (inches)

aseplate design significantly influences the reliability of the rotating equipment it supports. Equipment reliability in turn determines the total cost including installation, spare parts usage, maintenance, potential process losses, operating expenses and power consumption. Today, customers demand extended mean time between planned maintenance (MTBPM) in equipment, and they want pumps to operate reliably for specified periods of time. Each pump’s downtime can cost between $3,000 and $5,000 and even more if process losses are incurred. This downtime affects the customer’s level of satisfaction with the supplier’s rotating equipment. Suppliers of ANSI pumps recognize the importance of quality and the customer’s perception of the entire pump package. At least one pump supplier has addressed pump/motor reliability issues relative to baseplates and has designed, developed and tested a family of baseplates to meet differing customer needs.

D

.050 .040 .030 .020

E

.010

B C

.000

0

100 Load (lbs)

200

FIGURE 2 Maximum Parallel Shaft Deflection At Applied Force

Type A

.002"

Type B

.004"

Type C

.003"

Type D

.016"

Type E

.005"

quate motor adjustment to avoid bolt binding under normal conditions when aligning the motor to the pump in the field. This, however, depends upon the baseplate staying sufficiently flat in shipment and installation. It is also important to align the pump and motor before connecting it to the piping. Once installed, alignment should be rechecked and corrected if necessary. Baseplate rigidity helps ensure that achieved alignment is maintained while starting, stopping and maintaining the pump. Even when pump shafts and motor shafts are properly aligned, misalignment can develop from excessive and fluctuating pipe loads caused by variable process temperatures. Generally, these types of loads are too high to control by simply depending upon

BY: BOB WALLACE The Pump Handbook Series

87

FIGURE 3. CENTERING NUT — FACTORY PRE-ALIGNED

Durco Centering Nut Motor Foot

Baseplate

rigid baseplates, so other equipment considerations such as centerline mounted casings or C flange adapted motors may be necessary. The mass of the baseplate affects its capability to dampen vibrations caused by the pump, motors and associated piping. Baseplate mass will also help protect the pump and motor from plant vibrations. Cumulative levels of vibration affect component life. The greater the mass the better the distribution and absorption of vibrations, resulting in increased component reliability.

GROUT VERSUS STILT MOUNTED BASEPLATES Preferences on the application of grouted versus stilt mounted baseplate designs are split. In either case, added baseplate rigidity and mass are always recommended. Users of stilt mounted designs generally feel there is an advantage in allowing the pump to be relieved of some of the pipe forces by moving to the point of least resistance. Grouted baseplate installations offer maximum rigidity and mass. Vibration absorption is generally superior. However, pipe loads are locked into the pump, causing distortion and loading without relief. Both grouted and stilt mounted versions can provide excellent pump reliability when rigid reinforced designs are applied. Spring loaded versions of stilt mounted designs make the installed pump package act as the expansion joint. This is particularly effective for fluctuating temperature processes. The cost of installation is always a factor, however, and a stilt mounted unit costs less to install. Pump users today recognize the merits of rigid baseplate designs even when following good grouting practices. Grout is used with the intention of adding to rigidity and improving

88

Shims (For Vert. Alignment) Adjustment Allowance

vibration dampening of the installation. Conventional cement based grout shrinks as it dries and can separate from the baseplate. In time, cement based grout can crack and deteriorate. Low shrink grout should be used to minimize separation from the base. Proper venting will help avoid development of air pockets. Pockets mean reduced rigidity and increased vibrations in the equipment. Resonance levels will vary with the thickness of plate construction, so thicker is better. Even though they are more costly, epoxy style grouts are used increasingly to ensure installation longevity, superior bonding to the baseplate, minimal shrinkage, and corrosion resistance. In addtion, to help ensure that the epoxy grout adheres to the metal baseplate, special procedures can be specified that include roughing the underside surfaces, solvent cleaning, and applying epoxy coatings. Stress and vibration reduction is so important in the quest to reduce total costs and extend MTBPM that these added expenses, along with rigid baseplate designs, make sense to more and more customers. Some customers are going to the extent of providing epoxy foundations. Others install epoxy footings. To spend this kind of capital, they must perceive substantial overall cost savings.

HOW MUCH RIGIDITY AND WHAT COST? Two questions users and suppliers face today are: How much rigidity is enough, and how do I write a specification? The American Petroleum Institute Specification API 610 is excellent for grout installation practices, but it includes many standard features that increase costs significantly. Many users don’t require all of these features. Consequently, users The Pump Handbook Series

interested in the benefits of rigid designs often write their own specifications. This proliferation of ”special” baseplate specifications results in costly delays by the manufacturer, who must design the baseplates for specific projects or customers. In addition, interpretation of specifications can cause excessive costs and delays. Eventually, the manufacturer passes these costs on to the user. One additional problem users face if they don’t specify baseplate quality is that in order to reduce costs some suppliers will provide relatively thin cross sections of metal or inexpensive non metallic ”shell” constructed baseplates. This practice usually creates difficult and more costly installations. It may reduce rigidity and cause difficulty in keeping the shafts aligned, increase stress and vibrations, and increase future user costs due to reduced equipment reliability. Laboratory tests of a series of baseplates have provided a family of pre-engineered baseplates featuring various reinforced grouted and stiltmounted metal and non-metallic baseplates with assorted options. The heart of each design is an engineered minimum plate thickness to match the loads encountered with ANSI pumps and motors (Figure 4). Reinforced designs are supplied with added cross member construction. It is intended that customers select from a variety of pre-engineered rigid designs to find one that meets the process and budget needs. The design is complete, so added costs are avoided and shorter delivery dates are ensured. Five basic designs are offered, each with its unique assortment of options. The standard baseplate is the Type A (Photo 1), which is suitable when properly installed and grouted. The other designs add further reinforcement and rigidity to offer long term economics and performance through improved protection from shipping, installation and operation problems. Type B (Photo 2) base is solid aggregate mixed with a corrosion resistant polymer to offer rigidity against twisting and diaphragming forces, in addition to corrosion resistance, vibration absorption, added features and economics. When combined with epoxy grout, the installation becomes an integrated structure. Type C (Photo 3) is the stilt mounted design, rein-

FIGURE 4 Rigid baseplate design begins with thick plate construction • ANSI base nos. 139 to 258 have 1/2 in. steel plate construction. • ANSI base nos. 264 to 280 have 5/8 in. steel plate construction.

Photo 1

• ANSI base nos. 368 to 398 have 3/4 in. steel plate construction. • Non-metallic bases are constructed of 3" to 4" solid polymer-concrete. • Baseplate Types B, D, D and E are reinforced with added structural support for even greater rigidity.

forced completely with welded 1/2 in. plate and additional gusset support. The types D and E designs provide increased structural support and understructure for grout to hold the baseplate down under live load conditions (Photos 4 and 5). Customers who specify rigid baseplates and perhaps machined flat surfaces also laser align their shafts to critical tolerances. They may use epoxy grouts and special base preparation techniques or stilt mounted designs. These companies often employ pump teams whose job is to determine what is needed to reduce costly downtime. They calculate the actual economics of individual options, and they may justify spending extra capital funds initially in return for improved total life cycle economy. Customers who have not developed their own internal expertise such as this can benefit from the experience of those who have. Rigid baseplates, grouted or free to move, with precision shaft alignment are among the many good pump practices to be considered in maximizing a pump’s reliability and MTBPM. Improved reliability means mechanical seal and bearing life are extended, maintenance costs are reduced, and unscheduled downtime and process losses are reduced. The result: economical pump life cycle costing is achieved. ■ Bob Wallace is Director of Marketing for The Duriron Company, Inc., Durco Pump Division.

Photo 2

Photo 3

Photo 4

Photo 5 The Pump Handbook Series

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R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

Exploring Bearing Lubrication Options Know these important factors when selecting a lubrication system for your centrifugal pumping system.

* dm = d + D 2 , mm

Approximate temperature conversions degrees fahrenheit 50

20000 10000 5000 3000 2000 1000 500 400 300 200 150 100 75 50 40 30

90

2300 1250 900 700 470 350 240 190 140

IS

O IS

VG

68

O VG VG 4 32 00 O 0 VG IS 22 O 0 VG IS 15 O 0 VG 10 IS 0 O VG 68 IS IS O O VG V IS G 46 O 32 VG 22

20 15

IS

10 8 6 5 4 -20 -10

0

10

20

30

40

50 60

70 80

90 100

120

0

100 80 60

150

Temperature, Degree Celsius NOTE Viscosity classification numbers are according to international Standard ISO 3448-1975 for oils having a viscosity index of 95.

ered. The viscosity ratio, Kappa, should be the guideline for evaluation of viscosity. Kappa > 1.5 is preferred. The lubricant viscosity should not be too great as this would cause excessive bearing friction and heat. The frequency of oil changes depends on the operating conditions and the quality of the lubricant. Quality mineral oils with a minimum Viscosity Index (VI) of 95 are recommended. Multi-grade oils, and lubri-

BY: SKF USA, INC. 90

120 140 160 175 190 210

Approximate viscosity conversions Saybolt Universal Seconds (SUS)

T

FIGURE 1

Viscosity centistokes (mm2/sec.)

he lubricant separates the rolling elements from raceway contact surfaces in a bearing, and it lubricates sliding surfaces. It also provides corrosion protection and cooling. The principal parameter for the selection of a bearing lubricant is viscosity, n. Lubricating oils are identified by an ISO Viscosity Grade (VG) Number. The VG Number is the viscosity of the oil at 40° C (104° F). The common oil grades are shown in Figure 1. From this chart the viscosity of an ISO Grade oil can be determined at the bearing operating temperature. Rolling bearing lubricant requirements depend on bearing size dm* and operating speed ν but little on bearing load. The minimum required lubricant viscosity ν1 needed at the bearing operating temperature is obtained from Figure 2. The actual lubricant selected for an application should provide greater viscosity n than the minimum required viscosity ν1 (i.e. Kappa, κ > 1.0). Table 1 provides general lubricant Viscosity Grade recommendations for bearings used in centrifugal pumps. These are valid for operating speeds between 50% and 100% of the bearing Catalog speed rating. At lower speeds, higher Viscosity Grades should be considered, and at higher speeds, lower Viscosity Grades should be consid-

The Pump Handbook Series

cants with detergents and viscosity improvers are not recommended. Mineral oils oxidize and should be replaced at three month intervals if operated continuously at 100° C (212° F). Longer intervals between replacements are possible at lower operating temperatures. Synthetic oils are more resistant to deterioration from exposure to high temperature and may allow less frequent replacement. Lubricants may require more

or slightly above the vertical centerline of the bearing. Spherical roller bearings operating in a vertical oil bath should be completely submerged. For spherical roller thrust bearings, the oil level is set at 0.6 to 0.8 times the bearing housing washer height, C. Shaft sealing in these applications is best provided by a thin cylindrical sleeve inside the bearing’s inner ring support. Horizontal oil-bath lubrication represents the baseline of moderate bearing friction. The friction with other lubrication methods can be compared with that of oil-bath lubrication. Vertical oil-bath lubrication produces high friction if one or more bearings are fully submerged, possibly limiting the operating speed.

TABLE 1. RECOMMENDED ISO VISCOSITY GRADE Ball and cylindrical roller bearings VG 46 VG 68 VG 100

OIL-BATH LUBRICATION In horizontally oriented applications, the oil-bath level is set at the center of the bearing’s lowest rolling element when the pump is idle (Figure 3.). A sight glass or window is needed to set the oil level in the bearing. The oil level observed in the sight glass will vary slightly when the shaft is rotating due to splashing and flinging of the oil in the housing. The housing should allow the oil to flow freely into each side of the bearing. The housing should have a bypass opening beneath the bearings to allow the oil to flow freely. The cross-section area of the opening can be estimated according to the following equation:

A = 0.2 to 1.0 -√n dm EQ. 1 where A = bypass opening cross-section area, mm2 n = rotational speed, r/min. dm = mean diameter of bearing = 0.5 (d + D), mm

The small value from the above equation applies to ball bearings and the large value to spherical roller thrust bearings. Intermediate values can be used for other bearing types. If the bypass opening is not provided or not large enough, the oil may not pass through the bearing. This is particularly true for bearings having steep contact angles (angular contact ball, taper roller and spherical roller thrust bearings) operating at high

VG 68 VG 100 VG 150

speeds, in which case a pumping action caused by the bearing internal design may starve the bearing of oil or flood the shaft sealing. A ”constant level oiler” is an oil reservoir mounted to the bearing housing to replenish oil lost from the bearing housing (Figure 4). A sight glass is recommended along with these devices to allow the correct setting and examination of the lubricant level in the bearing housing. The recommended minimum oil volume V for each bearing in the housing is estimated from:

OIL-RING LUBRICATION An oil-ring is suspended from the horizontal shaft into an oil-bath below the bearings (Figure 5). The rotation of the shaft and ring flings oil from the bath into the bearings and housing. The housing channels the oil to the bearings. The oil-ring is made of brass or steel and sits on the shaft. The inner diameter of the oil-ring is generally 1.6 to 2.0 times the diameter of the

V = 0.02 to 0.1 D B EQ. 2 where V = oil volume per bearing, ml D = bearing outside diameter, mm B = bearing width, mm

For applications with vertical shaft orientation, the oil level is set at

FIGURE 2 ν1

ν1

(cS†) 1000

4600 SUS

500

2300

200

930

100

460

50

230

20

100

10

60

5

40

3 10

20

50

100

200

Pitch diameter (mm)

600

Approximate saybolt universal seconds (SUS)

frequent replacement if contamination is present. The most common methods of pump bearing lubrication are: oilbath, oil-ring, oil-mist and grease. Circulating oil lubrication is also used.

Other roller bearings

Centistokes (mm2/sec.)

Bearing operating temperature °C (°F) 70 (158) 80 (176) 90 (194)

35 1000

dmmm

dm = (bearing bore + bearing O.D.) ÷ 2 ν1 = required lubricant viscosity for adequate lubrication at the operating temperature

The Pump Handbook Series

91

FIGURE 3. OIL-BATH LUBRICATION

shaft and can be grooved for best oiling efficiency. Some sliding may occur between the oil-ring and the shaft. To minimize the resulting wear, the shaft must have a fine finish. Oil-ring lubrication reduces the volume of oil making contact with the bearing and therefore the bearing friction. The large size of the bearing housing needed for the oil-ring improves the heat transfer from the bearings and oil. Higher shaft speeds and lower viscosity lubricants are possible with oil-ring lubrication because of the lower friction and better cooling.

OIL-MIST LUBRICATION A mist of atomized oil droplets is conveyed by compressed air to the housing where it is reclassified (precipitated) into larger oil droplets by a condensing fitting and the bearing (Figure 6). Produced by a mist gener-

FIGURE 4 CONSTANT LEVEL OILER REQUIRED OPERATING LEVEL

OIL-RIM LEVEL

ator, the mist is pressurized slightly above ambient pressure. Oil-mist provides fine droplets of clean, fresh and cool lubricant to the bearings. Contaminants are excluded from the bearings by the oil-mist pressure inside the bearing housing. The mist can also be supplied to the bearings when the pump is idle for maximum bearing protection from contamination and condensation. The oil-mist can be introduced into the bearing housing (indirect mist) or directed at the bearing by a reclassifier fitting. In both cases, the housing must be provided with a small vent [3 mm dia. (0.125 in.)] opposite from the point where the mist enters the housing or bearing. This is to allow free oil-mist flow. Directed oil-mist is recommended if the bearing ndm** value is greater than 300,000 and if the bearing supports high axial load. Synthetic or special de-waxed oils are often used for oil-mist lubrication because paraffins in standard oils may clog the small oil-mist fittings. A specialist in oil-mist lubrication should be consulted for recommendations. Bearings can be ”purge” oil-mist lubricated or ”pure” oil-mist lubricated. Purge oil-mist combines oil-mist lubrication for bearings already lubricated with an oil bath. The ”purge” oilmist removes contamination from the bearings and safeguards against the possible loss of oil-bath lubrication. ”Pure” oil-mist lubrication is without an oilbath. Lubricated only by the clean BEARING mist lubricant, HOUSING the bearings are less likely to be exposed to contamination. Pure oil-mist lubrication has been shown to improve bearing life significantly [Reference 1]. The generation of oil-

** ndm is the bearing speed n in r/min multiplied by the bearing mean diameter dm in mm. dm = d + D 2 92

The Pump Handbook Series

FIGURE 5. OIL-RING LUBRICATION

mist must be adequately safeguarded with alarms to avoid bearing failure in the event of mist failure. It is recommended to prelubricate the bearings with similar oil or connect the bearings to the mist for a long time period before pump start-up to ensure satisfactory initial lubrication. Environmental concerns may limit the use of oil-mist lubrication. The bearing housings can be fitted with magnetic shaft seals and oil-mist collectors to limit the emissions. Oilmist lubrication minimizes the bearing friction.

GREASE LUBRICATION Lubricating greases are semi-liquid to solid dispersions of a soap thickening agent in a mineral or synthetic oil. The thickening agent is a ”sponge” from which small amounts of the oil separate to lubricate the bearing. Greases are selected for their consistency, mechanical stability, water resistance, base oil viscosity and temperature capability. Lithium soap thickened greases are good in all these respects and are recommended for general pump applications. Grease consistency is graded by the National Lubricating Grease Institute (NLGI). Consistency selection is based on the size and type of bearing used. NLGI 3 consistency greases are recommended for smallto-medium ball bearings, pumps operating with vertical shaft orientation, and pumps having considerable vibration.

FIGURE 6. OIL-MIST LUBRICATION

FIGURE 7. VERTICAL GREASE LUBRICATION

NLGI 2 consistency greases are recommended for roller bearings and medium to large ball bearings. NLGI 1 consistency greases are recommended for large bearings operating at low speeds. The grease base oil viscosity is selected in a manner similar to that of lubricating oils. The viscosity ν of the base oil at the bearing operating temperature should be greater than the minimum required lubricant viscosity ν1. Greases of different thickener types and consistencies should not be mixed. Some thickeners are incompatible with other thickeners. Mixing different greases can result in a grease with unacceptable consistency. Polyurea thickened greases are generally incompatible with other

metallically thickened greases, mineral oils, and preservatives. The bearing and the adjacent housing cavity are generally filled 30 to 50% with grease at assembly. Excess grease is purged from the bearing into the housing cavity. The period that the grease can provide satisfactory lubrication (i.e. grease life) is dependent on the quality of the grease, operating conditions, and the effectiveness of the sealing to exclude contamination. The General Catalog provides regreasing interval guidelines and it specifies the quantity of grease to be added at regreasing. The regreasing interval (tf) is based on the use of lithium grease with mineral base oil at 70° C (158° F). The regreasing interval can be increased if the operating temperature is lower or if a premium quality grease is used. The regreasing interval is reduced if the bearing temperature is higher. The regreasing interval is reduced by half if the bearing orientation is vertical. It is best to provide a shelf beneath the bearing to help retain the grease. The shelf should have clearance with the shaft to allow excess grease to purge (Figure 7). Excessive bearing temperatures may result if the bearing and the space around it are completely packed with grease. The bearing housing should be designed to purge excess grease from the bearing at start-up and at regreasing.

BEARING TEMPERATURE

In general, the allowable operating temperature of a bearing is limited by the ability of the selected lubricant to satisfy the bearing’s viscosity requirements (i.e. Kappa). Rolling bearings can achieve their rated life at high temperatures provided the lubrication is satisfactory, and other precautions, such as the correct selection of internal clearance, are taken. In some cases, a bearing operated from start-up cannot achieve satisfactory low steady state temperature conditions due to incorrect bearing fitting (i.e. high initial internal preloading) or due to excessively fast pump start-up rate. These conditions can cause a thermal imbalance in the bearing, resulting in unintended bearing preload. This situation is not unusual when the bearing housing is very cold due to low ambient conditions (cold climate or chemical process). The best solution in these instances is to control the bearing fitting (initial bearing clearance and shaft and housing fits) and slow the start-up rate of the machinery to allow the establishment of thermal equilibrium. Machined brass cages (M or MA suffix) may be needed in these applications. For bearings with polyamide cages to obtain longest service life, the outer ring temperature should not exceed 100° C (212° F) in pump applications. Because of this limitation, some pump manufacturers and users do not allow the use of bearings

TABLE 2 Lubrication factor, f0

oil-mist

oil-ring

oil-bath(1)

grease

Single row deep groove ball bearing

1

1.5

2

0.75 - 2

2.3 3.4

3.3 5

4.3 6.5

2.7 4

3.4

5

6.5

4

Double row angular contact ball bearing Conrad with filling slot Single row angular contact ball bearing pair

(1) - Double value for vertical shaft orientation

The Pump Handbook Series

93

with polyamide cages (TN9 or P suffix). In some instances, the operating temperature is the limiting factor determining the suitability of a bearing for an application. Bearing operating temperature is dependent on the bearing type, size, operating conditions, and rate of heat transfer from the shaft, bearing housing, and foundation. Operating temperature is increased when heat is transferred to the bearings from external sources such as high temperature pump fluids and rubbing contact housing seals. Bearing internal heat generation is the product of the rotational speed times the sum of the load-dependent and load-independent friction moments. The bearing load-dependent and load-independent friction moments can be calculated in accordance with the General Catalog (Reference 2). Recommended values of the lubrication factor f0 for the pump

94

bearing lubrication methods appear in Table 2. Bearing operating temperature and the viscosity of the lubricant can be estimated using SKF computer programs. (The above values of f0 are recommended for use with this program.) Higher bearing operating temperatures can be expected when rubbing contact shaft seals are used. In cases of bearings operating in cold climates or with lubricants having very high viscosity, the radial load may need to be greater than the estimated minimum required. Bearings having machined brass cages may also be necessary in these applications. In no case should a bearing be operated at temperatures less than the lubricant’s pour point temperature. ■

Lubrication Conference, San Francisco, CA, August 18-21, 1980. (2) SKF USA Inc. General Catalog (#4000 US) This article has been excerpted from the publication Bearings in Centrifugal Pumps Application Handbook, published by SKF USA, Inc. For a copy of the Handbook or a copy of the SKF USA General Catalog call or fax SKF USA, Inc., P.O. Box 1507, King of Prussia, PA 19406-0907, phone (610)962-4726, fax (610)265-8047.

(1) Large Scale Application of Pure Oil Mist Lubrication in Petrochemical Plants, H. P. Bloch, Century 2 ASME/ASLE International

PUMPS AND SYSTEMS MAGAZINE

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R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

How to Extend Pump Bearing Life By Larry Anderson entrifugal pumps are among the most widely used of all fluid-handling pumps. They are commonly applied in the hydroprocessing, wastewater, pharmaceutical, and pulp and paper industries. A single petrochemical plant may employ as many as 3000 centrifugals. Despite their prevalence, however, little attention is paid to maximizing their life expectancy. Although they are designed for 15 years’ average service life, the typical pump often fails within 18-36 months due to problems with bearings or related components. Proper bearing solutions, along with effective lubrication and maintenance practices, can prevent many of these problems and double or even triple centrifugal pump life expectancy. A pump rebuild can cost as much as $10,000, and hours of productivity might be lost before a replacement pump can be substituted for a damaged one. Given these realities, measures that improve pump longevity can save a plant millions of dollars annually in maintenance costs.

C

are bearings at two positions: a single-row deep-groove ball bearing near the impeller, and an angular contact ball bearing set in the thrust position at the coupling end. The bearing accommodates radial loads and floats in the housing. This allows thermal expansion of the shaft. The angular contact set is fixed in the housing. It secures the impeller in the proper axial position, handling the thrust load and a portion of the radial load. In some designs a cylindrical roller bearing is used at the impeller location. This allows shaft expansion to be accommodated internally. At high speeds, however, cylindrical roller bearings are susceptible to damage initiated by relatively minor shaft misalignments. ANSI-design pumps are very similar in configuration to API pumps, except in one respect. Instead of an angular contact ball bearing set, ANSI pumps have a double-row ball bearing-a single bearing with two rows of balls in the thrust position.

PREVENTING BALL SKIDDING

In both API- and ANSI-design pumps the bearing employed in the impeller position, whether a singlerow deep-groove ball bearing or a cylindrical roller bearing, normally encounters few problems. The bearing in the thrust position faces a much tougher task. It must handle a complex, dynamic combination of thrust and radial loads and hold the entire pump rotor assembly in place under all conditions. Traditionally, several bearing types have been used as thrust bearings, depending on pump design and application requirements. The duplexed angular contact ball bearing set, commonly employed in APIdesign centrifugal pumps, consists of two bearings positioned back to back. Typically, each bearing features a 40° contact angle, meaning that the balls in each bearing roll on an axis that forms a 40° angle with the pure radial position. Bearings with high contact angles, such as this one, are designed to carry axial loads; they are axially rigid and radi-

CENTRIFUGAL PUMP DESIGN Understanding centrifugal pump dynamics and operating characteristics is an important part of extending pump life. During operation, hydraulic conditions in centrifugal pumps generate major thrust (axial) loads. The type and size of a pump’s impeller determine the load’s magnitude and duration. Steady radial loads are imposed by the weight of parts and components. Fluctuating radial loads result from hydraulic and unbalanced conditions. Most centrifugal pumps are designed to either API or ANSI standards. Heavy-duty process pumps are generally designed to API standards, light-and mediumduty pumps to ANSI standards. In API designs (Figure 1), there

Figure 1. Schematic of a single-stage centrifugal pump, designed to API standards, shows portion of the single-row deep-groove ball bearing and angular contact ball bearing set. The Pump Handbook Series

95

0.1270

15 degree bearing

0.1016 0.0762 0.0508

40 degree bearing

0.0254 1.779

3.558

5.338

7.117

8.896

10.675 12.454

THRUST LOAD (kN) Figure 2. Graph demonstrates the high axial rigidity (low axial deflection) of a 40° angular contact ball bearing compared with a 15° bearing.

A

B

40°

15°

Fa

BEARING A

AXIAL FORCE, F

BEARING B

Fa

δ

δ Fa = FA –FB

1

Pr = PB Fb

Figure 3. Schematic shows the dissimilar contact angles in a 40°/15° bearing combination.

ally compliant. Bearings with low contact angles, on the other hand, are designed for radial loads; they are radially rigid and axially compliant. (Figure 2 compares axial rigidity in 40° and 15° bearings.)

96

One serious problem that can affect duplexed 40° angular contact ball bearings is ball skidding. During operation, thrust loads apply pressure against one bearing, causing it to deflect and the other bearing to The Pump Handbook Series

COURTESY OF MRC BEARING SERVICES

AXIAL DEFLECTION (mm)

0.1524

unload. Centrifugal forces acting on the unloaded bearing may cause its balls to run on a skewed axis and begin to skid. Ball skidding produces a microscopic wear or lapping process that damages the bearing raceway. Also, ball-to-raceway friction generates heat that reduces the bearing lubricant’s operating viscosity. This condition can shorten bearing life substantially. The traditional solution to ball skidding in API-design pumps is to use a preloaded set of 40° angular contact ball bearings. When the active bearing in the set is under load, residual preload in the other bearing prevents it from unloading. Determining the correct preload under actual operating conditions, however, is difficult. Important factors in calculating preload, such as the temperature differential between the inner and outer bearing rings and the thrust load on the impeller, are rarely known with certainty. The optimal preload window is narrow, and errors are easy to make. Excessive preload will cause the bearing set to run hot while insufficient preload will fail to prevent ball skidding in the unloaded bearing. A more forgiving solution is a matched pair of angular contact ball bearings with dissimilar contact angles. One such design—the MRC PumPac© system—consists of a 40°/15° bearing combination (Figure 3). This unidirectional system can be used if the direction of thrust is known. The 40° bearing handles primary thrust load. The 15° bearing resists any reversing thrust load during startup and handles radial load during normal operation. The 15° bearing is axially compliant, and its balls are unlikely to skid. The 40°/15° system improves radial stiffness and enables the pump to run with less vibration, maintaining the integrity of mechanical seals. It also runs cooler over a wider range of applications than the 40°/40° arrangement. These advantages enable the bearing to run longer without relubrication and increase the pump’s life expectancy. Ball skidding is also a recurring problem in the double-row ball bearings employed as thrust bearings in ANSI pumps. The two approaches

provides very consistent lubrication. Because oil is an excellent conductor of heat, it also removes heat from bearing balls and raceways and dissipates it through the pump housing. The air-oil mist system is gaining in popularity. This method suspends oil particles in a flow of compressed air. When the mist reaches the bearing, the oil is precipitated out of the air and into the bearing cavity. The system creates a positive internal pressure within the pump, preventing contaminants from entering the bearing. Whichever method is used, the lubricant should be filtered before it is introduced to the bearing. A fivemicron filter is recommended. Recent research indicates that filters with pore sizes larger than this may allow damaging contaminants to pass through. Bearing life expectanA maintenance worker at a paper mill prepares to overhaul a highspeed centrifugal pump with a matched pair of angular contact ball bearings having dissimilar contact angles.

described above preloading and the 40°/15° bearing combination cannot be used in double-rows. However, research is under way to develop effective solutions to ball skidding in these bearings.

INSTALLATION PRACTICES Selecting the right bearing is critically important in centrifugal pump applications, but it is only an essential first step. It is also necessary to follow proper installation and lubrication procedures to ensure that bearings operate as designed. Small bearings with bores of 45 mm or less, whether in the thrust or impeller position, can be mounted mechanically. When using this method, force should always be transmitted through the bearing ring with the interference fit, in most cases the inner ring. If force is transmitted incorrectly through the bearing’s rolling elements, the bearing’s raceway is likely to be damaged. Small bearings usually can be mounted using a hammer and impact sleeve. The sleeve should be large enough to fit over the shaft but must not make contact with the bearing’s rolling elements, cage or seal. Tool kits available from bearing manufacturers contain impact

sleeves that fit many standard bearing sizes. Medium-size and large bearings can be temperature-mounted. This method uses heat to expand the bearing’s inner ring and make it easier to install. It’s important to achieve a temperature differential of about 150 F between the bearing and the ambient temperature for adequate inner ring expansion. One of the most efficient heatmounting devices is the induction heater, which employs an electromagnetic current. It can heat a 100lb. bearing to the optimal temperature in less than five minutes. Ideally, an induction heater should monitor the heating cycle by either time or temperature and demagnetize bearings after heating.

LUBRICATION METHODS Bearings in centrifugal pumps are usually oil-lubricated. The preferred minimum operating viscosity is 70 SUS. If ball skidding is suspected, a higher-viscosity oil can be used to create a thicker film and provide added protection to bearing components. Typical lubrication methods include oil bath, air-oil mist and recirculating oil. The oil bath method

The Pump Handbook Series

Bearing Nomenclature In the rotating bearing industry there are few broadly accepted standards for bearing nomenclature. Each manufacturer employs its own alphanumeric code, a specialized system that designates whether a particular bearing, for example, comes with a machined brass cage or a labyrinth seal. But certain designations, especially those related to basic bearing design, are accepted industry-wide. Bearings with the 5000 designation, for instance, are double-row ball bearings. Other commonly accepted series designations include: 7000 series: Duplexed angular contact ball bearings 8000 series: MRC PumPac© or PumPac© Diamond bearing systems Bearings in the same series, however, may differ widely in design characteristics. If you

97

cy is effectively reduced if these contaminants are present in the lubrication system. After bearings are installed and running, they should be monitored periodically for signs of vibration and overheating. A healthy bearing in a centrifugal pump normally operates at 140°-150° F. Although not damaging to bearing materials, slightly higher temperatures will

98

begin to change the bearing’s preload and operating characteristics. They also can be a symptom of lubrication or seal problems. Temperatures greater than 200°F may indicate actual pump distress. Careful attention to these warning signals will enable plant personnel to optimize bearing performance and to take corrective action well before the onset of failure. ■

The Pump Handbook Series

Larry Anderson is Manager, Technical Services at MRC Bearing Services in Jamestown, NY. He holds seven ball bearing-related patents and has 37 years of experience in the bearings industry.

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R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

Mounting and Clearances for Casing Wear Rings ear rings on centrifugal pumps are important for several reasons: 1. They provide a running joint between the rotating and stationary parts of the pump. 2. The rings control or minimize the hydraulic thrust of the pump. 3. A reduced-pressure environment is provided for the packing or mechanical seal. On one side of the rings is suction pressure; on the other side is discharge pressure. It is desirable to keep clearances between the rings as low as possible to limit leakage from the discharge side back to the suction side. 4. Wear rings protect the casing and impellers from wear and erosion and at the same time simplify pump maintenance by providing replaceable wear surfaces. On most horizontally split pumps the casing wear ring is locked in place with a tongue and groove joint. The case wear rings of a horizontal split pump move or “float” to a running position and are locked in place by discharge pressure. The critical fits are on the bore and vertical plane. The reasons for this practice are: a. It makes internal alignment easier. b. The case gasket can be crushed different amounts without deforming the case wear rings. Normally the case joint requires a 1⁄32″ thick gasket. Asbestos gasket material in use until recently had a thickness of .030″–.038″ with a compressibility factor of about 30%. Thus, the bore of the case in some pumps can vary by 7 or 8 mils. The currently used non-asbestos materials vary in

W

culating water pump the clearance may be as much as 20 mils on the diameter. In no instance should the case rings be clamped when the cover is tightened.

original thickness, and their compressibility varies up to 50%. The effective bore of the case can change according to the type of non-asbestos material used. If the compressibility is too great the case wear ring can be clamped and not be permitted to float. An even worse situation could see the rings crushed into an elliptical shape. The case cover must be put down “blind,” so fits must be “sloppy.” Note in Figure 1 that the top half of the ring is shaped differently to aid in preventing any binding of the rings as the cover is lowered in place. Also note the anti-rotation lugs. The “float” clearance varies widely. On hotter service process pumps it should be 5–8 mils on the diameter of the rings. In a large cir-

WEAR RING SHAPES Many factors combine to dictate wear ring design in pumps. Abrasion Effects. Abrasion is a real enemy of wear ring clearance. Surface hardened materials, Stellite, overlay and heat-treated materials are effective to varying degrees in combating this wear, and new material compositions, such as composites and thermoplastics, are significantly increasing the life and reliability of friction and wear components, including wear rings. To further assist in reducing

One side of groove milled off of top half Impeller Ring Locking Pins

Sealing surface for case ring

LIQUID SEAL

STEPPED TONGUE AND GROOVE (Section half)

STUFFING BOX (INTEGRAL WITH CASING) IMPELLER

Section A-A

SHAFT SLEEVE

STUFFING BOX GLAND

PACKING LANTERN THROAT RING

A

IMPELLER WEARING RING CASING WEARING RING

“Float” Clearance

FIGURE 1. Typical horizontal split pump The Pump Handbook Series

99

Material Case Wearing Ring

Impeller Wearing Ring

Galling Characteristics

Bronze 85-5-5-5

Aluminum Bronze

Good

B-4 Bearium Bronze

Type 410 300 BR

Good

B-10 Bearium Bronze

Type 410 350 BR

Good

25-20 Chrome Nickel

25-20 Chrome Nickel

Poor

Type 410 300 BR

Type 410 350 BR

Fair

Stellite No. 1

Stellite No. 1

Good

18-8 Stainless

18-8 Stainless

Poor

Type 420 400 BR

Stellite No. 1

Good

Cast Iron

Cast Iron

Good

TABLE 1. Galling effects Recommended Minimum Running Clearances For Wear Rings (API STD 610 1981) Diameter of Rotating Minimum Diametral Diameter of Rotating Minimum Diametral Member at Clearance Clearance Member at Clearance Clearance (inches) (inches) (inches) (inches) <2 2.000 - 2.499 2.500 - 2.999 3.000 - 3.499 3.500 - 3.999 4.000 - 4.499 4.500 - 4.999 5.000 - 5.999 6.000 - 6.999 7.000 - 7.999 8.000 - 8.999 9.000 - 9.999 10.000 - 10.999 11.000 - 11.999

0.010 0.011 0.012 0.014 0.016 0.016 0.016 0.017 0.018 0.019 0.020 0.021 0.022 0.023

12.000 - 12.999 13.000 - 13.999 14.000 - 14.999 15.000 - 15.999 16.000 - 16.999 17.000 - 17.999 18.000 - 18.999 19.000 - 19.999 20.000 - 20.999 21.000 - 21.999 22.000 - 22.999 23.000 - 23.999 24.000 - 24.999 25.000 - 25.999

0.024 0.025 0.026 0.027 0.028 0.029 0.030 0.031 0.032 0.033 0.034 0.035 0.036 0.037

Clearances For Bushings In Vertical Pumps Diameter of Rotating Member 0.75 to 1.500 1.500 to 2.500

Clearance 0.04 to 0.06 0.06 to 0.08

Notes: (1) For materials with bad galling tendencies and/or for operating temperatures above 500°F add 0.005″ to these diametral clearances. (2) There should be a 50 Brinell hardness difference in mating materials. TABLE 2. API wearing ring clearances

wear caused by abrasion, spiral grooves are often used. These are cut, either in the inside diameter of the case wear rings, or on the outside diameter of the impeller wear ring, whichever is made of the softer material. Grooves increase the life of wear rings and help prevent galling. Small

100

particles of abrasive matter rotating with the liquid and trying to enter the wear ring clearance find these grooves and follow them to their outlet. Concentric grooves also tend to reduce leakage loss. It is reasonable to expect that spiral grooves would increase leakage loss slightly; howevThe Pump Handbook Series

er, tests indicate a 35% decrease in leakage can be achieved. Apparently, the chambers created by the grooves generate a turbulent flow pattern that acts as a barrier to leakage. Grooving Design. It is a good practice to cut these grooves 3⁄32″ wide by 1⁄32″ deep with a right-hand spiral, as it has been determined that, regardless of whether the spiral runs counter to or with the flow, its effectiveness is the same. Three grooves per inch are preferred. Wear rings, stage pieces, or throttle bushings 2″ or wider can be tooled with one and one-half grooves per inch with equal effectiveness. Galling Tendency Effects. Some wear ring material combinations in common use today gall more readily than others, thereby requiring greater clearance for successful operation. Material combinations in Table 1 classified as “good” rarely gall; those classified as “fair” do so occasionally. Heat Expansion Effects. In addition to selecting materials that will give longest wear life and that will not gall, heat expansion must be considered. Impeller wear rings in hot service pumps must not expand away from the impeller when the unit is brought up to temperature. Materials with a similar coefficient of expansion must be used in the shaft, impeller, impeller wear rings, case wear rings and case.

RUNNING CLEARANCES When establishing running clearances between wear rings and between other moving parts, except bearings, consideration must be given to pumping temperatures, suction conditions, character of fluids handled, and the expansion and galling characteristics of the ring material. Clearance should be as liberal as possible. The important consideration is dependability of operation and freedom from seizure under refinery operating conditions, even at a possible light sacrifice in initial hydraulic efficiency. For cast iron, bronze, chromium hardened 11-13% , and materials of similar galling tendencies, Table 2 shows running clearances typically used for operating temperatures below 500 °F

1. Single suction - Low flows - High heads

s

=2

00

25

N

20

00

15

Ns

=5

10

Ns =

1000

500

Ns = 1

5

N s = 2000 Ns = 2500

0 0

20

2. Single or double suction - Medium flows - Medium head

3. Single or double suction - Medium to high flows - Medium to low head - Francis-type impeller

40 60 80 100 120 140 160 180 200 Percentage increase in wearing clearance Added power resulting from excessive wear ring clearance for different specific speeds.

FIGURE 2. Clearance vs. power consumption

For vertical pumps the running clearances specified in Table 2 do not apply to the clearance of steady bearings or interstage bushings if materials of low galling tendency are used. Generally speaking, a wear ring should be renewed whenever the initial clearance is doubled.

shown in Figure 2. Note that for most pumps of refinery application specific speeds, doubling the wear ring clearance will increase power requirements about 5-7%. ■

William E. (“Ed”) Nelson is a turbomachinery consultant based in Dickinson, TX. He is the author of more than 50 technical papers and a contributor to several handbooks on pump operation and maintenance. Mr. Nelson is also a member of Pumps and Systems Editorial Advisory Board.

EFFECTS OF RING WEAR The effects of increased clearances on pump performance are

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101

COURTESY PACIFIC PUMPS—INGERSOLL-DRESSER PUMPS

Percentage increment in pump power requirements

30

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R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

Factors in Pump Suction Piping Don’t let improper design rob you of system performance. By Kenneth Clarke atisfactory design of the suction inlet will improve pump performance, lead to increased mean time between failures (MTBF) and reduce maintenance costs. An additional benefit to the plant is higher equipment reliability. The hydraulic design of a pump should ensure that the flow approaching the inlet is uniform and steady with a single stream region. The mechanical design should include proper suction pipe fabrication, installation and pump casing support.

S

Fy My Mz Fz

Mx

Fx

REDUCE PIPE STRAIN

102

Pump Size Fx 11⁄2 x 1 2x1 2 x 11⁄2 2x2 3 x 11⁄2 3x2 3x3 4x3 6x4 8x6 10 x 8 12 x 10

120 145 165 175 220 235 260 325 470 610 670 730

Combined Loading at Center of Pump Fy Fz Mx My 295 365 405 435 545 590 650 815 1170 1520 1670 1820

235 285 325 350 435 470 520 650 940 1215 1335 1460

590 730 815 875 1090 1170 1300 1625 2345 3030 3350 3640

295 365 405 435 545 590 650 815 1170 1520 1670 1820

Mz 295 365 405 435 545 590 650 815 1170 1520 1670 1820

Flange Loading Suction Discharge Fr + (Mr/3) Fr + (Mr/3) 325 435 435 435 650 650 650 865 1300 1730 1870 2080

220 220 325 435 325 435 650 650 865 1300 1730 1820

1. Fx, Fy, Fz, Mx, My, and Mz are forces and moments resolved to the center of the pump as shown above. 2. Fr and Mr are the total resultant force and moment imposed on each flange. 3. All forces above are expressed in pounds. 4. All moments above are expressed in pounds-feet. FIGURE 1. Example of engineering standard for maximum allowable forces and moments on pumps The Pump Handbook Series

COURTESY DURCO PUMPS

Process pipe misalignment with the suction pipe will produce pipe strain, which causes the pump shaft to be displaced from its center line position. Resulting pump problems may include impeller/casing rub, motor and premature seal and bearing damage. The suction pipe should be supported independently near the pump so that when the flange bolts are tightened, no strain is transmitted to the pump casing. Acceptable moments and direct flange loads for various pipe sizes are listed with the pump OEM (Figure 1). Pump/motor coupling misalignment and increased casing stresses result when these moments and loads are exceeded. Pumps operating in hot services experience thermal growth; the amount depends on the application temperatures. When the pump casing is supported at the 6 o’clock position, the thermal growth will move the casing upward and strain the suction pipe. Mounting the casing at its radial center line locations (Figure 2) will alleviate this problem. When it is severe, pump cavita-

tion will cause vibration and other internal damage.1 Suction pipe highflow velocities will also generate vibration. If the suction pipe is not correctly supported, vibration can cause pipe strain and casing stress. High suction line velocities and cavitation also produce turbulence. Therefore, suction line velocities should not exceed 10 ft/sec, and consider 5-6 ft/sec maximum for new systems.

CORRECT PIPE SIZE The suction line should be designed with the right pipe size in order to reduce NPSHA (Net Positive Suction Head Available) restrictions.2 Inadequate NPSHA throughout the operating range will reduce flow and lead to pump damage. A suction pipe pressure gauge should be installed for applications where NPSHA is marginal. Generally, suction piping is one or two sizes larger than the pump nozzle. If the pipe is larger, a reducer should be installed in such a way as to eliminate air pockets and uneven flow regions (Figure 3).

Support recommended for thermal effects FIGURE 2. Support recommended for thermal effects Air Pocket

FIGURE 3. Reducer at pump section

RECOMMENDED

NOT RECOMMENDED

7 Pipe Dia. Minimum

Air Pocket

Suction

Suction

FIGURE 4. Suction pipe design If elbow is necessary it should be of the long radius type.

Suction

Suction piping should be supported close to the pump flange to prevent vibration and strain on pump casing.

Path of Water

FIGURE 5. Suction elbow on double suction pump The Pump Handbook Series

103

Suction pipe should have an absolute minimum of five pipe diameters of straight run before the suction flange. Seven pipe diameters are preferred as a minimum (Figure 4). This provision helps make the flow to the pump impeller eye parallel with uniform velocity, and it reduces turbulence.

THE RIGHT PIPE FITTINGS

The number of turbulence or pressure drop-producing fittings in the pump suction line should be kept to a minimum. Globe valves should not be used in the pump suction line because they produce turbulence and high friction loss. When NPSH or turbulence is a problem, a turbulencereducing device should be used and located at about the absolute minimum distance upstream of the pump suction flange. When close-connected suction elbows are used, they should be the long radius type (Figure 5). A Flow Straightener—CRV2— should be installed just upstream of the elbows to reduce turbulence

104

and prevent fluid separation. Without a Flow Straightener, uncontrolled flow through an elbow separates fluid into two stream regions, creating local high velocities and turbulence. Check valves should not be used in the suction pipe. However, they are occasionally used in series-parallel connections to reduce the number of valves to be operated when changing from series to parallel operation. Foot valves should not be utilized in the suction line when the pump is operating against a high static head, even though they are recommended when suction lift is low and priming help is required. If static head is high and the driver fails, the pumped material could rush backwards, causing a heavy water hammer. Experience has shown that when the factors we’ve discussed are considered for pump suction design, or in solving field problems, a significant contribution to increased mean time between failure of the pump results. Remember, however, that these fac-

THE PUMP HANDBOOK SERIES

tors are only a small but important part of the whole pump design. ■

REFERENCES: 1. Karassik, I. J., Krutzsch, W. C. and Fraser, W. H., Pump Handbook, McGraw-Hill Book Company, New York. 2. Garay, P. N., Pump Application Desk Book, Fairmont Press, Inc.

SUGGESTED READING: Kern, R., “How to Size Pump Suction Piping,” Hydrocarbon Processing, April, 1972. Kenneth J. Clarke is a division engineer responsible for rotating equipment at Dupont Chemical’s Acrylonitrile Business Unit, Beaumont, TX. Clarke holds a BS in mechanical engineering.

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R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

Alignment Movement Methods ost alignment articles cover the measurement and calculation phases of the job. Regardless of how these tasks are done, the machine still must be moved to accomplish the alignment. This article describes various ways of accomplishing that task. Our objectives include: 1. Make necessary vertical and horizontal moves precisely and efficiently. 2. Do not damage the machines or injure people while making these moves. Do not retain or introduce strain or distortion, e.g., “soft foot” or excessive piping forces. 3. Eliminate obstacles to movement and provide for referenced return to aligned condition. Several methods will be described—first for the vertical moves, then for the horizontal ones. Vertical position is controlled by shims, and horizontal moves have no effect on these. By contrast, a vertical move usually upsets a previous horizontal adjustment, which must then be repeated.

M

VERTICAL MOVEMENT METHODS Integral Jackscrews. These are threaded vertically into the feet of the machine to be adjusted, generally an electric motor. They often work well and have the advantages of low cost and simplicity. Also, they do not require vertical jacking clearance beneath the machine. Disadvantages include the possibility of rusting, binding and breaking. There is also the need to add a slot to the shims if the jackscrews are not in line with the hold-down bolts. These can also cause motor damage if accidentally left extended during holddown bolt tightening. Prybar. This simple approach needs little vertical clearance. It requires some horizontal space to accommodate the bar’s length, however, and occasionally something slips—which can cause pinched fingers.

Portable Jacks and Wedges. These are available in mechanical and hydraulic versions. Space requirements and load capacities vary. Photo 1 shows a 60-ton hydraulic wedge having only 3⁄16″ collapsed height (Ref. 1). Cost of this wedge is $1,250. Crane or Hoist. These are often inconvenient to arrange, and lifting the machine in this way allows undesired horizontal movement. Later, however, we will describe how to use a vertical lift to assist in precise horizontal movement. Portable Alignment Positioners. These devices come in mechanical and hydraulic lift versions. They require 21⁄4″ vertical clearance and have capacities from 8–40 tons per set (2–4 units). They are used mainly for moving large electric motors for alignment. They accomplish both the vertical lift for shim changes and precise horizontal movement—transverse and axial (Ref. 2). The cost of positioners runs from $2,000–$8,000 per pair. Shims and Shimming. As mentioned, shims are used for vertical positioning of the adjusted machine. They should be of corrosion resistant material, usually stainless steel. Time-saving pre-cut shims are available in a range of outline and slot sizes. If special shim configurations are needed, a set should be made up beforehand in various thicknesses from 0.002–0.125″. Always use a micrometer to check actual thickness. Sometimes shim markings are inaccurate, and often several thin shims will stick together, doubling or tripling the apparent thickness. Check shims for cleanliness and for burrs and edge curl or distortion. Correct these conditions or reject shims having such defects. Make a “shim sandwich,” putting thin shims between thicker ones at top and bottom, as protection against wrinkling. If you are using sizes A through D pre-cut shims (2–5″ wide) of a single brand, and if The Pump Handbook Series

their pull-tabs have parallel sides, you can get the shims evenly stacked and inserted beneath machine feet by gripping at the sides of the tabs with special shim pliers (Ref. 3). Long support pads may cause “heel and toe effect.” This is a form of soft foot that can be alleviated by using a taper ground shim in each pack or, less desirably, by using stepped shims. If pad temperatures will not exceed 150°F, another approach is to use liquid epoxy shimming. However, this will break and require re-casting at each future machine removal/reinstallation.

HORIZONTAL MOVEMENT METHODS Steel Hammers. Photo 2 shows machine foot damage that can occur if a steel hammer is used directly against the machine foot, without an intermediate wood block. We do not recommend using a steel hammer even with a wood block, due to the bouncing involved and the difficulty of maintaining movement control. There is also the possibility of hammering the hand of the person holding the wood. Soft Face Dead-Blow Hammers. These are available in two types—cast lead head, and the plastic-face, hollow-head, shot-loaded design. The latter is shown in Photo 2. Both types work well, but the lead heads distort quickly and must be re-cast. Neither type will damage a machine, and the dead-blow, or nonbounce feature gives good movement control. Jackscrews, Fixed and Portable. These are common and inexpensive and often work well. Problems can occur if they get corroded and stick. Also, if the lug is too thin, it may bend or spring, causing a “stick-slip” movement, which reduces precision. Jackscrew lugs are often welded in place, which can cause problems if they interfere with shim insertion and removal. For this reason, bolting or clamping them in place may be a better choice. Portable Jacks. These can

105

work well for horizontal movement if there is room for them where they are needed, and if there is something to support their base, i.e., “something to jack against.” Comealong or Horizontal Hoist. This method is fairly easy to apply but lacks precision. Stretching of the cable or chain often causes a stick-slip effect resulting in jerky movement. Portable Eccentric Ring-Lugs. This system is used with motor holddown bolts as the “anchors.” It is advertised as suitable for motors from 2–300 hp and higher. Set prices range from approximately $700–$1200 (Photo 3 and Ref.4). Portable Alignment Positioners. See previous description. Crane Lifting Force. Applying an upward force less than the machine’s weight reduces the effective weight of the machine on its supports by the amount of upward force exerted. This makes the machine easier to move horizontally by whatever method is used. We have used this technique in situations where the weight of the machine caused the supporting structure to move, rather than the machine, when a jackscrew was tightened. Taper-Pin Doweling. This returns a machine to the same position on the supports it had before removal. We often use this approach on pumps that always return to their same pedestals. It is recommended that you avoid it on motors because it prevents final trim-up of the alignment. In addition, it works poorly when shim pack thicknesses are changed. It is a good idea to use antiseize compound on the pins. Matching Punchmarks at Motor Feet/Supports. Applying an automatic centerpunch at three motor feet and their supports allows subsequent positioning closely approximating where the motor was, transversely and axially, before it was removed. This can save several hours of alignment time.

HANDLING INTERFERENCES Piping Fits. Pipe should be fitted so strain is not transferred to the

106

connected machine. For new construction or turnaround maintenance with unlined carbon steel pipe, this is usually done by heating the pipe with a torch and bending it into position. In an operating plant, the use of a torch may be prohibited, necessitating numerous trips to a remote pipe shop to achieve a good fit. A way around this is to use a tapered filler ring, or “dutchman,’ between the cocked flanges (Ref. 5). Machine Must Be Lowered, but No More Shims Remain to Remove. This generally requires that the “stationary” machine be raised. To determine the minimum magnitude of moves to accomplish this—see Ref. 6. Machine Must be Moved Horizontally, but Becomes “BoltBound” before Required Position Is Reached. The first tactic to try here is undercut bolts. These should have undercut shanks machined concentric with threaded ends, with gradual section transitions to avoid stress concentrations. Using SAE Grade 8 Alloy permits undercutting somewhat below the root of the thread. After machining, corrosion protection should be restored. Usually this means zinc plating, followed by baking to avoid hydrogen embrittlement cracking. Centering washers made the same way are also helpful (Ref. 7). If further movement is required, the next thing to try is a two-element optimum move similar to that described in the vertical move section (Ref. 6).

CONCLUSION Alignment movement is as important as the more noticed measurement and calculation phases. It deserves careful attention to ensure that the desired end result actually occurs at the machine, not merely on the alignment calculation sheet or CRT display. ■

REFERENCES 1. Hydraulic Wedge (brochure). Hydra Wedge Corp., El Segundo, CA. 2. Alignment Positioner (brochure). Murray & Garig Tool Works,

The Pump Handbook Series

Baytown, TX. 3. Shim Pliers (brochure). Murray & Garig Tool Works, Baytown, TX. 4. Horizontal Alignment Tool Set (brochure). Posi Lock Puller, Inc., Cooperstown, ND. 5. Murray, M.G. “Flange Fitup Problems? Try a Dutchman.” Hydrocarbon Processing. Vol. 59, No. 1 (January 1980), pp. 105–106. 6. Murray, M. G. “Out of Room? Use Minimum Movement Machinery Alignment.” Hydrocarbon Processing. Vol. 58, No. 1 (January 1979), pp. 112–114. 7. Undercut Capscrews and Centering Washers (brochure). Murray & Garig Tool Works, Baytown, TX.

ABOUT THE AUTHOR: Malcolm Murray, Jr., is owner of Murray and Garig Tool Works in Baytown, TX, a manufacturer of several patented machinery alignment tools. Mr. Murray is the author of many machinery related technical papers and does consulting on alignment problems throughout the world.

Alignment Using the C Frame Adapter Every time a pump is torn down and reassembled, both the motor shaft and power end shaft must be realigned. An often overlooked but effective and convenient method of realignment is to utilize a “C” frame motor adapter. It provides easy and accurate alignment of centrifugal pumps and motors during installation and maintains that alignment during operation, provided personnel follow correct procedures. The motor adapter is a machined component that connects a pump’s power end to a “C” face electric motor through closetolerance fits on each end (Figure 1). The adapter maintains accurate alignment despite pipe strain, cavitation, vibration, high temperature and water hammer, all of which cause movement of the entire assembly.

REDUCING MAINTENANCE The C frame adapter is not a new technology, having been used successfully for many years on machine tools and gear boxes. An important advantage of the product is that it significantly reduces the cycle time for placing a pump back in operation. Once pump maintenance is completed, the motor, C frame and power end assembly can be slid back into the casing as a unit. Once the casing and hold-down bolts are tightened, the pump is ready for operation. And it is usually not necessary to waste valuable time checking or realigning the pump and motor.

INSTALLATION PROCEDURES Installed properly with no paint or particles in the fit areas, the adapter will remain in alignment. Of course, the pump must be properly fastened to the base to ensure that no stresses are introduced into the assembly during installation. The following represents typical C frame adapter installation procedures as well as practical tips to avoid misalignment and costly pump shutdowns. • In attaching the C frame to the

bearing frame, be sure there is no paint or dirt on the face or fit where the adapter is to be placed on the bearing frame. There are usually four bolts that hold the two components together. Carefully tighten these bolts in an alternating and evenly loaded fashion. • Place the power end in a vertical position with the C frame on top to eliminate erroneous readings caused by indicator sag and to prevent tolerance stackup in one direction due gravity. Check the runout by clamping the dial indicator to the coupling end of the shaft and sweeping the face and fit of the C frame. • When readings are greater than desired, follow these steps to improve them: Remove the C frame, check the fit for paint or particles, then reattach. Loosen and retighten the bolts from the C frame to the bearing frame in a different pattern. Loosen the shaft cartridge adjusting lock nuts while sweeping the fits on the C frame with the dial indicator attached to the shaft. Gently retighten the lock nuts when the readings have improved. Use caution to allow the cartridge and shaft to center themselves. Don’t exert great stress on the cartridge because this method doesn’t finalize alignment. • Check the C face electric motor runout by attaching a dial indicator to the shaft and sweeping the end bell fit and face. Published motor tolerances can allow up to .005 in parallel and angular misalignment, although the tolerances usually measure less. • A final assembly check with the motor installed is recommended. Indicators or other devices should be attached to both shafts and the alignment verified. These are not easy tasks due to space restrictions within the C frame itself. It is best to check the motor and frame before final assembly to provide reasonable data for alignment verification. There are several fits that will The Pump Handbook Series

contribute to misalignment in both the pump and motor, but there should be some cancellation of error. • To improve motor and pump shaft alignment, one design offers an adjustable C frame adapter. Alignment is accomplished by enlarging the motor fit and using adjusting bolts against the motor for positioning. This compensates for parallelism but may not help the angular misalignment problem. Once the design proceeds to this level, the maintenance specialist must come back to perform alignment with dial indicators. However, some time can still be saved in attaining the final alignment. Motors up to 125 hp can be attached to C frame adapters, depending on the various pump frame sizes. Some pump manufacturers allow the motor to hang freely without foot support up to certain frame sizes, permitting free growth of the entire pump. However, as the motors become heavier, the decision to support the motor feet becomes critical. Pads sized to the correct thickness and attached to the motor feet will remove the strain of the motor weight on the bearing frame. Attaching a dial indicator to the bearing frame before motor assembly will enable you to measure any deflection due to motor weight and improper shimming. The motor pads have been used successfully without bolting them to the baseplate, i.e., the motor is just sitting on the base and fastened only to the C frame. When properly employed, the C frame motor adapter is a timesaving and convenient alignment tool for pump users. Editor’s Note: This article was prepared with the help of Richard P. Antkowiak of A.W. Chesterton Co. (Stoneham, MA), a manufacturer of pumps, mechanical seals, and C frame adapters.

107

Alignment Approaches Given the sophistication of today’s pumps and motors, a mechanic’s attempt to achieve shaft alignment with a heavy hammer or bumping bar seems as archaic as a dentist approaching a tooth extraction with a pair of pliers. Clearly, better methods are available with more precise and predictable results. Motors require adjustments in the vertical and horizontal planes. For accuracy and efficiency, some form of measurement is needed to determine the direction and amount of adjustment. The importance of this undertaking can be seen in assessing the volume of today’s U.S.–based plants that require alignment. Of an estimated 275,000 such facilities, about 60%, or 165,000, require periodic alignment. The remaining 110,000 plants have motors needing adjustment regularly during a 12-month period. Of the motors needing alignment, 37% are under 10 hp; 26% are 15–25 hp; 18% are 30–60 hp; 11% are 75–150 hp; and 8% are motors at and above 500 hp. The process of alignment involves three separate though related operations: evaluating or measuring the relative positions of the shafts; vertical positioning of the motor, and horizontal positioning of the motor. Many tools with varying degrees of sophistication are available for measuring relative shaft positions. These can range from bracket-mounted dial indicators that can measure shaft positions from one-thousandth of an inch to high technology lasers that project beams across shafts for closer measurement. Prices for these tools

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vary from an average $2,800 for a dial indicator set to $13,000 for laser systems and related training required. Naturally, less sophisticated methods and measurements are available with lower price tags. These can be used in a prealignment step to prepare machines for more precise alignment. Refinements will generally involve more intricate, exact methods and equipment. The accuracy of the measurement system requires the ability to move equipment to coincide with the measuring system, otherwise the investment is not realized. Measurement data obtained during the first operation will be used for the vertical positioning of the motor. In the second step, pre-cut shims are placed under each foot of the motor. The motor is then lifted and lowered onto the shims, and this sets it in the desired position. Horizontal positioning of the motor, achieved during the third step, involves fine adjustments, moving the motor from side to side. Adjustments move the motor from a minus measurement through plus measurement and back before coming to rest at absolute center. Unlike vertical positioning, every move on the horizontal plane is critical. It is at this delicate stage that the feet of the motor all too often receive the blows of heavy hammers or bumping bars. Obviously, this stone age approach is inexact and expensive in terms of wasted time, damage to equipment and the potential of injury to the mechanic. One alternative approach is to fabricate a lug that is threaded to accept a jackscrew and to weld this assembly at the feet of every motor. Tightening each of the jackscrews

The Pump Handbook Series

against the feet will eventually move the motor into alignment. While this procedure is preferable to the hammer’s blows, it does have limitations. Not all motor bases are suitable to fit jackscrews. In addition, the cost is high for fitting and/or replacing the jackscrews while fighting rust and corrosion. An up-to-date option that offers the advantages of the jackscrew method while eliminating the disadvantages has been introduced in the form of patented Posi Lock horizontal alignment tool sets manufactured by Enerpac (Butler, WI). These portable jackbolt sets provide precise control for movements as small as l/1000th″. Three sets are available for servicing electrical motors, gear boxes, compressors, fans and other types of rotating machinery requiring precise alignment. Using the Posi-Lock sets for alignment involves a three-step process: •First, place the eccentric socket on the motor mounting bolt head and slide the alignment tool over the socket. •Second, turn the knurled adjusting nut on the alignment tool until it contacts the base of the motor. •And, third, turn the socket until the motor moves into the desired position indicated by the measuring device. The approach enables adjustments to be made and alignment achieved in a matter of minutes. Given the number of motors in medium-size and large plants, the savings in time and labor can be appreciable.

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R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

Foundation Tips Editor’s Note: From time to time, the Shoptech section of Pumps and Systems features Pump Rules of Thumb. These Rules of Thumb are derived from a blend of engineering principles and experience. They are intended to provide a rapid assessment of conditions and can be used as a supplement to general or detailed instructions furnished by the manufacturer. Machines differ in design and construction, but fundamental principles apply to all of them. This issue’s Rule of Thumb on foundations was developed by Ray Dodd of Chevron. There are several articles on the grouting angle but no conversation on the dimensions of the basic foundations. Both dynamic and static forces must be considered for soil support of foundations. Well designed pump foundations and bases should have the required rigidity to withstand the axial, transverse and torsional loadings of rotating machines. Some conservative rules of thumb (as shown in Figure 1) that apply to good foundation design for API or larger pumps are: • The mass of the concrete foundation should be five times the mass of the suppored equipment. • Imaginary lines extending downward at 30° from either side of a vertical line through the machinery shaft should pass through the bottom of the foundation and not the sides.

Base Plate Epoxy Grout Grade Elevation 30°

30°

Foundation

Foundation mass should be approximately 5 times the mass of supported machinery Imaginary lines extended downward 30° to either side of vertical CL should pass through bottom of foundation

FIGURE 1. Typical installation

• The foundation should be 3” wider than the baseplate all the way around machinery up to 500 hp and 6” wider for larger machines. These guidelines can be used for quick evaluation of the pump foundation as a potential source of vibration. Foundation vibrations are usually at rotational speed. If an inadequate foundation is identified as the source of a vibration problem, adding concrete mass may reduce the tendency of the foundation to respond to and transfer vibrations. For example, if a concrete mass equal to the full weight of the machine is added to the base of the machine, the resulting velocity energy will be approximately half that of the original design. If twice the mass

The Pump Handbook Series

is added, the vibration energy will be cut to about 1/3 of the amount experienced on the original without the added mass. While often excellent for minimizing vibration, adding mass is a costly remedy after the foundation installation is completed. On the other end of the spectrum, many ANSI style chemical process pumps in the range of 5-50hp are frequently installed with no specially constructed foundations, or with foundations that do not meet the above guidelines. ■ V. Ray Dodd is the senior equipment reliability specialist for Chevron U.S.A. Products Company.

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Evaluating Pump Monitoring Options By Bob Liddle ptions for monitoring pump performance are as varied as the types of pumps themselves. They can range from low-cost and simple to expensive and complex. For example, various devices can be used to measure the pressures related to a pump medium, including: an on-off pressure switch, a bourdon tube pressure gauge, a piezoresistive pressure transducer. When is low-cost and simple adequate? When is relatively expensive and complex needed? There are no simple answers. The pressure measurement technology needed in each instance depends on the application, the medium being transmitted and the environment in which the pump is operating.

O

INSTRUMENT SIZING/PLACEMENT A starting point for choosing a gauge or transducer is to determine the instrument range required. For a gauge, the best way to ensure proper operation, while leaving enough room for maximum output — or spikes — at startup, is to double the normal operating pressure when selecting an instrument. For a transducer, sizing should be as close to the normal operating pressure as possible. For example, a 60 psi application should get a 60 psi transducer. This ensures much greater full scale accuracy. A well designed and built transducer has a built-in over-pressure tolerance of three times the rated pressure. For both a gauge and a transducer, the issue of instrument placement can be as important as sizing or rating. Avoid locating the device near critical valves that open or close during the process to control pressure. Valves are more likely to cause instrument damage than pumps.

110

APPLICATION CRITERIA

Next, look at the specific application. A small commercial pump on a swimming pool, for instance, may operate fine with a standard, off-theshelf, utility pressure gauge with an accuracy rate of +/–2% of full scale reading.This type of gauge, with brass internals and soft-soldered joints, is satisfactory for many industrial or process applications as well. There are limits, though, to the range of temperatures and types of media such gauges can handle. You may not want to use this gauge type when monitoring pressures on critical pump or compressor applications at power generating plants, refineries or petrochemical facilities. Instead, consider either a more robust and/or higher accuracy gauge, or a transducer that could be electronically linked to an overall plant control system. These cases probably also require an accuracy of less than one-half of one percent of midrange. A good rule of thumb is to buy and install only as much pressure measurement technology as you need for an application. Generally, the lower-tech, lower-risk pumps and systems can be adequately monitored by a pressure switch or a bourdon tube gauge. Conversely, more complex applications may require a piezoresistive transducer or similar technology. Some high-speed compressor applications are better off with a transducer than with a switch or gauge. In situations such as a skidmounted gas turbine used for electric power generation, a transducer can provide continuous information to a central monitoring and control system. The Pump Handbook Series

MEDIA/ENVIRONMENTAL FACTORS

There are other factors to consider when choosing a pressure measurement instrument including: transmission of a corrosive, excessive vibration, location of the instruments and the pump. In situations where corrosives, acids or other potentially damaging fluids are pumped, there are two possible courses of action: • use a compatible construction material for all surfaces that will come into contact with the medium • isolate the instrument from the process stream with a diaphragm seal When specifying a diaphragm seal, carefully select a mounting configuration and a material compatible with your specific requirement. Diaphragm seals can be manufactured from many different materials and contain a variety of fill fluids (ranging from instrument oil to silicone to vegetable oil). In cases of excessive vibration or pulsation, a dampening or snubbing device can be used. However, when appropriate, a liquid-filled gauge can provide some dampening at a lower cost. Stainless steel and other corrosion resistant construction materials should be used in extreme environmental or hazardous conditions ranging from salt water atmospheres to petrochemical cracking operations.

DATA TRANSMISSION AND USE Finally, no discussion of instrumentation is complete without deciding if, and how, data from these devices will be sent and stored.

In some instances, an analog reading from a gauge — recorded by a technician on a plant walk-around — is sufficient. Other situations may require that electronic signals from a series of instruments be sent to a central programmable logic controller (PLC) or a distributed control system (DCS). While the low-tech walk-around and the high-expense control system used to be the only options for collect-

ing data, there is now a middle ground: event monitoring.An event monitoring network collects and stores data for analysis and trending at a fraction of the cost of a DCS. In addition, systems are flexible and can be purchased in modular units that more closely fit the size of the application. From selecting a switch, gauge or transducer to deciding which type of data acquisition system is needed; the better you know the characteris-

The Pump Handbook Series

tics of the pump, the better able you will be to design a system to monitor its operation. ■ Bob Liddle is director of product marketing for U.S. Gauge Division of Ametek.

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Applying Predictive Maintenance Measures Predictive maintenance can increase efficiency in the face of budget, personnel and equipment cutbacks. By Ivan Brown and Brian Barnes

lobal economic pressures and intense competition are forcing private industry to operate with utmost efficiency. In fact, maximum operating efficiency is now mandatory just to maintain financial competitiveness and corporate stability. Achieving increased efficiency involves cost cutting in many areas, including organizational downsizing and restructuring. Thus, maintenance and operations managers are often faced with the difficulty of achieving higher efficiency goals following budget, personnel and equipment reductions. Managers and operators striving for efficient operations are turning to predictive maintenance (PdM) programs and systems as a means to acquire information about the condition of the plant’s machinery. This information is used to maximize the plant’s production based on known or expected reliability of each system tested, and to allocate maintenance resources on a need basis. PdM instrumentation and software make it possible to monitor a plant all the way from the power pole to the loading dock, enabling managers to implement proactive maintenance practices designed to engineer problems out of the machinery. The ability to plan production based on quantified machinery conditions is a major step towards efficiency. The PdM technologies described here are samples of the nondestructive test procedures and reports available to operation and maintenance

G

112

personnel. Pumps are used in all of the following illustrations. However, the physics involved apply to all types of rotating equipment. Pump driving components including motor, coupling and pump are included here as part of entire systems normally examined in PdM testing since any of the three may cause machinery failure.

JUSTIFYING PdM PROGRAMS The following goals represent those most often incorporated in successful PdM applications. 1) Maintenance Management. Information on plant condition allows management to make decisions about production schedules and maintenance activity. Budget justification can be supported with factual data. Resources are directed based on identified need. 2) Avoidance of catastrophic failure. Substantial savings result if machinery is shut down prior to catastrophic failure. A pump bearing failure can result in rotor, impeller and /or casing damage, for example. Detecting incipient machine faults, scheduling repairs and thereby reducing added repair expense is a primary focus of all PdM programs. 3) Quality assurance in new installations and overhauled equipment. Test machinery before ‘sign off’. Establish acceptance criteria and quality assurance programs for purchased equipment before accepting new machinery or overhauls. 4) Implementation of Just-InTime parts inventories (JIT). PdM programs allow JIT to be implemented while reducing risk. They also The Pump Handbook Series

reduce the costs of maintaining large parts inventories. 5) Avoidance of unnecessary maintenance. Calendar-based maintenance often results in the repair of healthy machinery. Calendar-based programs are costly and in some cases introduce problems into smoothly running machines. Calendar-based repairs are justified only with strong historical evidence supporting machine life cycles—for example, in pumps handling corrosive material. 6) Increase in mean time between failures (MTBF). Root cause failure analysis using PdM techniques is a proven method of achieving higher MTBF. As longer MTBF periods become a reality, maintenance schedules can be realistically adjusted. This is a major part of proactive maintenance. 7) Energy savings. Between 5% and 15% of energy usage is wasted in the average facility. Detection of the cause and remedy of the waste is possible using PdM techniques. Examples include coupling misalignment and phase unbalance. 8) Proactive maintenance. Instead of just repairing the defective machine, analyze the available data to determine the root cause of the problem. Then revise or redesign the machine installation to eliminate the cause of damage. In this way, machine life is extended and repair frequency reduced.

VIBRATION ANALYSIS Lubrication programs are considered the backbone of preventive maintenance. Likewise, vibration analysis is the most informative of

the predictive maintenance technologies. It can monitor all moving parts associated with a rotating assembly, and it will detect problem conditions before they become serious. Ongoing industrial PdM program applications using vibration analysis typically include the following tasks: • establish machinery vibration data baseline or acceptance criteria • monitor machinery on predetermined schedule and identify machines that deviate from existing baseline data or exceed absolute alarm levels • analyze data and inform maintenance staff of the source of the problem and its severity; then recommend repair action

benefits of a PdM program. Following repairs, machinery is then retested to verify the diagnosis and the quality of the repair. Figure 1 is an example of a plain language text report.

• centrifugal • propeller • rotary thread • rotary gear • sliding vane • positive displacement A mature automated program may contain hundreds of diagnostic rules for the pump categories listed above, the rulebase automatically analyzes raw data and produces plain language reports containing machine fault diagnosis, fault severity and specific repair/action items. The automated program minimizes man-hours required and enables non technical personnel to access the POTABLE WATER PUMP #1 ACQUIRED: 06-16-1993 19:51:29 SPEED: 3574 RPM SEVERITY: Mandatory, replace pump bearings DIAGNOSIS: Extreme pump ball bearing wear is indicated by

Data collection techniques vary from simple rms overall vibration measurements to sophisticated detection/diagnostic analysis of individual machinery components; from walk around portable data collectors to continuous on-line monitoring systems. Advanced diagnostic programs require specific information regarding pump configuration, number of pump vanes, and timing gear tooth counts, for example. When using computerized diagnostic rulebases, analysts usually use the following pump categories:

104 (+10) VdB [4A] at 3.12 x Pump 122 (+29) VdB [4A] at 9.72 x Pump 116 (+24) VdB [4T] at 13.60 x Pump Slight pump bearing or other internal looseness is indicated by 107 (+7) VdB [4A] at 3.00 x Pump 108 (+8) VdB [4R] at 4.00 x Pump 105 (+4) VdB [4T] at 2.00 x Pump FIGURE 1. Typical vibration analysis data

Plant: CVN Corporation Plant 7 Machine: Circulating Water Pump #101 ■ (10) Pump Imbalance ● (NF) Motor Thrust Bearing Wear

Area: Building 46 ◆ (13) Angular Misalignment ▲ (NF) Motor Mounting Flexibility

Extreme

◆ ■ Serious

■ ◆ ■

Moderate

■ ◆

▲

▲

Slight 9 Apr

20 Apr

2 May

13 May 25 May

5 Jun

●

16 Jun

28 Jun

9 Jul

21 Jul

1992

The Pump Handbook Series

Wear particle analysis (WPA) is a powerful technique for non-intrusive examination of the oil wetted parts of a machine. The particles contained in the lubricating oil carry detailed and important information about the condition of the machine. This information can be deduced from particle shape, composition, size distribution and concentration. The particle characteristics are sufficiently specific so that the operating wear modes within the machine can be determined, allowing prediction of the imminent behavior of the machine. Often, action may be taken to correct the abnormal wear mode without overhaul, such as when abrasive contamination indicates a change of oil and oil filter. Alternatively, timely overhaul can prevent costly secondary damage. A typical application uses trend information from a direct reading ferrograph of oil samples to assess the amount of wear particles present. Alarm limits based on these readings, as well as the length between samples, are commonly set to determine which sample requires microscopic evaluation. Some extremely clean running types of equipment, such as vacuum pumps, require that every sample be given a microscopic evaluation. Other types of equipment, such as centrifugal pumps, may require microscopic examination only once every three months. Being able to see all particles and having relevant equipment configuration information, the analyst can identify problems and recommend appropriate maintenance actions. The WPA and used lubricant industry have grown recently due to the need to extend lubricant life through additives, combined with the soaring disposal costs of used lubricants. Wear Particle Analysis tests include:

1 Aug 1992

FIGURE 2. Vibration fault trend plot

WEAR PARTICLE ANALYSIS

• direct reading ferrography and/or microscopic analysis

113

PdM Express - DLI Engineering Services - Division of Predict Technologies MACHINERY VIBRATION DIAGNOSTIC REPORT CLIENT: FREY FOODS

AREA: BUILDING 10

MACHINE: WASH DOWN PUMP #4

Test Date: 23-JAN-1994 Motor RPM: 3545 Fault Status: SLIGHT

Pump Model No: YTB2B2 CMMS Unit # 2356-89

MACHINE DRAWING Test Locations

1

CONDITIONS DURING VIBRATION TEST: Pump Discharge: 180 psi Pump Suction: 3 psi Motor Lead: 120 amps

2 3

ANALYST NOTES: A slight pump bearing defect is evident, fault trend analysis shows only minor condition degradation since December 15, 1993. Condition should be monitored regularly. RECENT VIBRATION ANALYSIS RESULTS: 23-JAN-1994: SLIGHT PUMP BEARING WEAR 15-DEC-1993: SLIGHT PUMP BEARING WEAR 14-NOV-1993: NO FAULTS

Fault Trend Plot

Wash Down #4

o--- Pump free end bearing wear

Extreme Serious Moderate Slight 11/14/93

12/15/93

FIGURE 3

114

The Pump Handbook Series

1/23/94

Used Oil Analysis tests include: • viscosity • elemental spectroanalysis • chlorine • Karl Fischer for water • particle count • bacteria • physical properties • total acid number • total base number

MOTOR CURRENT ANALYSIS

Electric motor reliability is a concern to the pump operator for obvious reasons. Current analysis is often thought of as the “packaging of all the electrical tests” into one test procedure. Current analysis test data using high resolution spectrum analysis provides the diagnostic techniques that allow motor reliability to be quantified. Database management of motor current data makes ongoing programs easy to manage. Current analysis systems typically include the following tests: • resistance • voltage rotor bar problems • current balance • power factor • inductance balance • capacitance balance Diagnostic outputs include: • electrical hot spots • leakage resistance to ground • air gap eccentricity • rotor faults

THERMOGRAPHY Thermography or infrared imaging is a preventive as well as predictive maintenance tool. It provides both predictive information and fault detection at the time of the test. For the pump operator, thermography is useful for finding electrical hot spots in the motor

and loose or deteriorated electrical connections, and for monitoring bearing and coupling temperatures.

MANAGEMENT ISSUES

PdM technologies have been successfully implemented for many years, and successes are well documented. Starting a new program or upgrading to current technologies requires a commitment from management if it is going to be successful. For every successful program there is one that failed for lack of proper management. Before cost benefits can be realized, an investment in training is needed, and a commitment must be made to allow maintenance staff adequate time to produce good results. Maintenance staffs that we have met are generally enthusiastic when they understand the benefits and potential of PdM programs. The enthusiasm quickly disappears, however, if a PdM startup is just another extra duty for the staff. Proactive maintenance not only results in efficiency that produces cost savings, but it also improves the working environment of the staff. Financial reporting is a management issue that is often ignored with PdM programs. Tracking financial results will justify investment in your programs, indicate if programs are running properly or require revision for your application. Strengthening team building within maintenance departments and other departments often results when documented successes are shared. Top management will support your PdM efforts when the financial impact on the bottom line is made available.

BENCHMARKING FOR PROCUREMENT Benchmarking technical programs is recommended during procurement. The cost of hardware and software is minuscule compared with employee salaries and operating

The Pump Handbook Series

expenses incurred over the life of an ongoing PdM program. Evaluation of data collection time, database creation, report generation and ease of use are some critical areas that warrant careful consideration. The wide selection of equipment, software and technical applications that are marketed by various vendors is mind boggling. Benchmarking will not only save the purchaser from dependence on a salesperson, but will also help define the program from the beginning and help get a program going smoothly.

CONCLUSION The PdM technologies have improved drastically in recent years due to the rapid development of the PC and digital data collection and diagnostic equipment. It is now possible to perform predictive and diagnostic analysis and produce a report on a complex machine in far less time than it took to collect and chart minimal data years ago. However, equipment and software capabilities have advanced faster than many of the experts in the field were able to incorporate the new technologies into their application philosophy. A manager charged with selecting the technology for his program should carefully analyze not only what is available, but best suited for his needs. The emphasis today is on getting information to management and staff at a reasonable cost. Detailed evaluation of report formats and technical content is crucial to maximize return on investment and have a truly successful program. ■ Ivan Brown has been with DLI Engineering Div., Bainbridge Island, WA for 14 years. Brian Barnes has been an analyst with Predict Technologies WPA Div., Cleveland, OH for five years.

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Vibration Analysis Yields Good Vibes New clues help track and solve vibration problems. By D. Gary Powers, P.E.

W

116

PLOT SPAN 0.6

PK VELOCITY IN IN/SEC

ith a high-speed electronic vibration data collector, a user can gather enough data from a pump in minutes to keep an analyst busy for hours. However, as we increasingly rely on data, we can easily lose sight of the basics and their significance. The decision to remove equipment from service for repair should be based more on economic factors than on arbitrary standards of vibration. The causes of vibration are varied. For example, a pump’s design and construction characteristics must be evaluated and its history of operating reliability considered. In every machine, dynamic forces of mechanical and hydraulic and/or aerodynamic origin are present, and a certain amount of vibration is therefore inevitable. To ensure the safety of the unit, the vibration must be kept within acceptable limits. If high vibration exists, or if it increases or decreases markedly over time, problems arise. Detailed measurements and analysis are needed for a reliable diagnosis. This article attempts to answer the question: How do you approach a pump you’ve never seen before that has a vibration problem? Hopefully, the approach we’re taking will be helpful even if you have been fighting a problem for a week or a month. A typical pumping unit can be divided into some or all of the following systems: • the foundation or pedestal • the driver • mechanical transmission of power through a coupling or gear box

0

0

10

20

30 40 FREQUENCY IN kCPM

50

LIST OF SPECTRAL PEAKS Machine: (MBS) PUMP 2789 Meas. Point: PUMP 2789 - #4V - Pump, Pump End Vertical Date/Time: 04-May-93 10:43:16 Amplitude Units: PEAK FREQUENCY PEAK ORDER PEAK FREQUENCY NO. (kCPM) VALUE VALUE NO. (kCPM) 1 7.439 0.0189 2.10 13 22.465 2 8.936 0.0205 2.52 14 23.625 3 9.647 0.0304 2.72 15 24.141 4 14.717 0.0244 4.15 16 24.985 5 15.440 0.0243 4.35 17 26.019 6 16.084 0.0205 4.53 18 28.731 7 17.174 0.0295 4.84 19 31.815 8 17.903 0.5673 5.04 20 32.101 9 19.662 0.0438 5.54 21 32.767 10 20.519 0.0394 5.78 22 34.720 11 20.937 0.0451 5.90 23 35.820 12 21.441 0.0543 6.04 24 39.350 TOTAL MAG 0.6123

60

IN/SEC PK PEAK ORDER VALUE VALUE 0.0353 6.33 0.0281 6.66 0.0229 6.80 0.0365 7.04 0.0191 7.33 0.0189 8.09 0.0223 8.96 0.0244 9.04 0.0190 9.23 0.0208 9.78 0.1379 10.09 0.0254 11.08

SUBSYNCHRONOUS SYNCHRONOUS NONSYNCHRONOUS 0.0087 / 0% 0.5689 / 86% 0.2262 / 14%

FIGURE 1. Vibration spectrum of a blade passage frequency and its harmonics The Pump Handbook Series

70

• •

the driven unit the piping or duct work. A methodical investigation will help identify vibration sources that may be local, within a section or related (directly or indirectly) to causes of vibration in other sections.

LOOK AT THE PUMP Walk around the pump and look at it closely. Is the machine sitting on a concrete pedestal, or is it in a structure 30 feet in the air sitting on two pieces of channel iron that were the prototype for the first “slinky”? How does the piping enter and leave the pump? Are there any noticeable leaks? Are the seals leaking? How is the pump coupled? Is there a gear type coupling, shim pack, rubber grid or “Mike Calistrat Rope” coupling? Is there a belt drive? A gear box? Is the gear box a speed increaser or speed reducer? Do name plates on the equipment give pertinent data, such as rpm, gear ratios and bearings specs, etc? Let’s assume the pump is sitting on a concrete pedestal. Look closely and see if air bubbles are being pumped into the water over the joint between the pedestal and the concrete floor. While looking at the pedestal, examine the baseplate. Are air bubbles being blown in the oil over the grout joint between the baseplate and the pedestal? If the baseplate is not an integral part of the pedestal, it takes very little exciting force to cause large vibration levels in the pump and driver. One of the primary reasons for a pedestal is “mass dampening.” Pedestals normally have three to ten times the mass of the rotating units mounted on them. (i.e. motor/ pump or turbine/gear box/blower). When the baseplate is loose, there is no mass dampening, and any excitation force will cause it to ring like a bell. Some conservative rules of good foundation design include: 1. The mass of the concrete foundation should be 5-10 times the mass of the supported equipment. 2. Imaginary lines extended downward 30 degrees to either side of a vertical line through the machinery shaft should pass through the bottom of the foundation and not the sides. 3. The foundation should be 3″

wider than the baseplate, all the way around a pump up to about 500 hp, and 6″ wider for larger pumps.

TAKE COMPLETE VIBRATION READINGS How many times have you taken vibration readings on a troublesome pump when it was not operating? You may find that the vibration spectrum is in alarm with the same amplitude peaks and frequencies that existed when the unit was operating. The brinniling that takes place in rolling element bearings that have a high amplitude of vibration when they are not operating can cause failures when the unit is started. The seals also may be ruined by the vibration. Ideally, vibration spectra should be taken in the horizontal, vertical and axial planes on each bearing. Take phase angles at 1-times the running speed with a strobe light or other phase reference, such as a key phase for light emitting diode. The amplitudes may be taken in mils displacement, the velocity in in/sec or acceleration in G’s. Frequency range will vary depending on several factors. On some bearings it is desirable to take spectrums in 2 or even 3 frequency ranges. It may be desirable to take one spectrum from 60-6,000 cpm in displacement (mils); one spectrum from 600-60,000 cpm in velocity (in/sec), and one more from 6,0004,000,000 cpm or more, in acceleration (G-forces). Velocity spectra will probably prove to be the best indicator of most problems on 85-90% of the rotating machinery. In machines with sleeve bearings, one spectrum can be from zero (0) -15 or 20 times the rpm. On gear boxes or turbines a second spectrum may need to go as high as 4 times the gear mesh frequency or the blade pass frequency. On machines with anti-friction bearings, 50-65 times the rpm is a good general spectrum frequency. Whether you are setting up a vibration monitoring program, doing a vibration analysis on a single machine, or taking base line signatures, take as complete a set of readings as possible. All of these readings may not be taken each time periodic readings are taken, but they will be available for comparison when a problem does occur. The Pump Handbook Series

It is also important to take vibration readings on equipment prior to a shutdown or a turnaround. When the pumps are placed back in service, an immediate check can then be made to determine if problems were corrected during the outage or if problems exist after the turn-around that didn’t exist before. Problems that can occur during a turn-around that may become obvious if you have good base line data with which to compare the new readings include: • misalignment • piping strain • rubs (seal or internal) • unbalance In the last several years we have found the unbalance that occurs after a rebuild is more prevalent on electric motors than on mechanical rotating equipment. This seems to occur most in electric motors that have been dipped. Typically, the Glyptol Varnish has sagged or run to one side, causing the rotor to be out of balance.

SYNCHRONOUS VIBRATION Vane Pass Vibration. One of the most common vibration problems encountered with centrifugal pumps is a blade passage frequency. This commonly causes noise, vibration or pressure pulsations at frequencies equal to rotational frequency multiplied by the number of impeller vanes. Multiples, or harmonics, of blade passage frequencies often occur and are sometimes stronger than the blade passage frequency itself. Figure 1 shows a typical vibration spectrum of a blade passage frequency and its harmonics. Multistage Pumps. Various means are used in multistage pumps to cancel or compensate for dynamic forces generated at one impeller by forces generated at another. The best known method is to stagger the impellers – that is, to mount them on the shaft in such a way that the vanes of adjacent impellers are not aligned and will not pass the diffuser vanes simultaneously. There have been some pumps manufactured with the vanes on all of the impellers lined up and all of the cut waters in a nice straight line. On one pump with 5 vanes in this configuration, we got a very high 5 times (5×) running speed peak on

117

118

A.

B.

FIGURE 2. A. Before, pump impellers in a straight line. B. After, pump impellers rotated 1/5 turn each

Inlet Guide Vane

GAP "B" GAP "A" CL

both bearings in line with the cut water, and a low vibration amplitude perpendicular to the cut water. The solution was a new shaft with the keys indexed so that the vanes were staggered in such a way that they came in at 1⁄5 the distance between diffuser cut waters, from impeller to impeller (Figure 2). This reduced the vibration to well within the acceptable limits. High Axial Vibration Due to Recirculation. Most high axial vibration in pumps is due to recirculation and can be cured by properly setting the clearances (Figure 3). The following references are quoted from Shop Repair of Centrifugal Pumps by Douglas Latimer, Jr. “Correct any Gap A clearances that do not have the recommended clearance.” Hydraulic instability can set in at flowrates below BEP. Recirculation sets in behind impeller sideplates and can result in high axial vibrations and pressure surges. This instability behind the impeller sideplates can be controlled by Gap A. Gap A controls the severity of the pressure pulsations that give rise to these high axial dynamic forces behind the impeller and shroud. Axial vibrations cause mechanical seal damage and rapid bearing failures, especially thrust bearing failures. The problems can be eliminated by the proper clearances. The effects of Gap A clearance have been shown to be best, with no variation, between 0.035″ and 0.085″, with all effectiveness being lost above 0.135″. So it is generally recommended that Gap A clearances be set at 0.050″ for single stage rotors and between 0.050″ and 0.075″ for multi-stage rotors. There is no need to jeopardize the rotor with clearances below 0.050″, but clearances to 0.035″ are acceptable. “Correct any Gap B clearances that are not in the recommended range. Gap B represents the clearance between the impeller vane tip and the leading tip of the diffuser cutwater.” A hydraulic shock or impact occurs as the rotating vane passes the stationary vane, resulting in excessive noise levels, vibrations and energy losses. Gap B controls the strength and amplitude of the hydraulic forces and can be reduced from 80–85% by increasing Gap B from 1–6%. There is no loss of over-

FIGURE 3. Recirculation corrected by proper clearances of Gap A and B

all pump efficiency or power from trimming these vanes, as is commonly thought. Corrected Gap B’s can greatly reduce operating noise and vibrations and improve efficiency. Misalignment. Misalignment between a pump and motor usually causes high axial vibration. Good alignment, using either dial indicaThe Pump Handbook Series

tors or laser alignment (reserve indicator method) practices, is necessary when a pump is installed or repaired. Even then, misalignment can occur from flexible baseplate design, loose foundation bolts or piping stress. Piping stress, particularly on new construction, probably accounts for more than 50% of mis-

alignment. Thermal expansion either in the piping and/or between the motor and pump can also cause misalignment. Spacing between the motor and the pump must meet coupling specs so as not to exceed the flexible limits of the coupling. The high axial vibration generated by misalignment can occur at 1times running speed of the unit (with shim pack couplings, the vibration can equal the number of bolts times the running speed). If you suspect piping stress, one of the easiest proofs can be to take vibration readings with the unit running and the suction and discharge piping disconnected. Ball and Roller Bearing Failures. Have you ever had an electric motor come back from a repair shop and soon after experience bearings failure? Min. 4% 6%

Diffusers Volute

We had a large electric motor (250 hp) with an eddy current clutch that repeatedly experienced bearing failures. These failures occurred very shortly after the motor came back from a rebuild and balance of the eddy current drum. On the third or fourth rebuild I was asked by the customer to follow the motor through the rebuild shop and check the balancing of the eddy current clutch. The unit was rebuilt and new bearings were installed. Then the eddy current clutch drum was mounted on the shaft, and the motor positioned in the test stand. The drum was balanced by using duct seal for trial weights. When a final correction was to be made by welding a permanent weight to the eddy current drum, the shop foreman brought in the welding machine and attached the Preferred 6% 10%

Max. 12% 12%

Note: The percentages should be increased approximately 4% if both the impeller and diffuser have an even number of vanes. All final clearances, fits and repair work performed should be documented as a permanent record. TABLE 1. Recommended "Gap B's" (given as a percentage of impeller diameter)

A

0

60-60K CPM BAND 5

10

B 0

60-60K CPM BAND 10 20 30

30

ground lead to the lifting ring on the top of the motor housing. He then prepared to place his weld on the drum, which was mounted on the rotor shaft. The foreman did not realize that every time this was done, he had caused an arc to pass through the rolling element bearings. Correcting this, by not allowing the arc to pass through the ball bearings, prevented the motor from experiencing premature bearing failures. Normally, checking where a welder attaches his leads is a concern in field balancing equipment, but this should also be checked in the shop. Probably the best solution would be to have the welder use the ground lead clamp to hold the balance weight in place. Then when he strikes an arc either on the base metal or on the weight, the return path will be very short. A new construction site is a place where all equipment bearings should be very carefully checked on start-up. On nearly every large project there are probably one or two bearings that have had a welding arc pass through them during construction. Water is another major contributor to 5-10% of bearing failures within the first 90-120 days of start-up.

C

60-60K CPM BAND 0

5

30

30 2.9 AVG.

60

3.7 AVG.

(0.6 - 6.1) MIN MAX

After balancing 3.7 Mil avg both fans

60

11.0 AVG.

10

(2.8 - 3.1) MIN MAX

2.9 Mil avg. Background from west fan

(0.8 - 14) MIN MAX

11 Mil avg. both fans

FIGURE 4. Beat frequency plots The Pump Handbook Series

119

This is particularly evident in equipment that has been in the “lay down” yard for a long time. On occasion, we have found blower bearings that showed high vibration amplitude peaks at two times the ball pass frequency. When the bearings were removed, there was an etched water level mark in them. Obviously, they had been sitting for a long period with water inside them before they were drained and filled with oil. Having the etched water mark on opposite sides of the race caused the high vibration amplitude peaks at two times ball pass frequency. Here again, we have a reason for establishing good base line data on our equipment.

SPECIAL CIRCUMSTANCES: BEAT FREQUENCIES There are many occasions when the amplitude of vibration can increase and decrease or “beat” when plotted against time. This can occur even when the filter is tuned to the running speed of the unit. You can hear this occur in cooling towers, pump alleys and jet airplanes, to name a few. The “beat” results when two motors run close to, but not at, the same rpm. When they come in phase, the amplitude is amplified to a peak, and when they are out of phase, they reach a null. When taking readings on two operating pumps that are mounted on the same manifold, it is advisable to install a pick-up on one pump and tune the strobe light very carefully to the running speed of that pump.

120

Then measure the amplitude of vibration against time. Then, without changing any of the set-up, observe the second pump with the strobe light. The shaft of the second pump will appear to rotate at the same rate as the peaks of vibration amplitude on the first pump. In Figure 4 the vibration plots of amplitude versus time were taken on a two-cell cooling tower. In this case, one cell had to be kept operating at all times. The first plot (a) shows the “A” cell shutdown and the “B” cell operating. This is the background vibration reading, or the amplitude of vibration, from the “B” cell fan gear box. This measurement was taken on the “A” cell fan gear box, tuned to the running speed of the “A” cell gear box, even though it was not operating. The second plot (b) was taken with both cooling tower fans operating. Note that it takes approximately two minutes for the fans to come in and out of phase with each other (both fans are running very close to the same speed).

HELPFUL TIP When field balancing with a strobe light and there is a strong “beat frequency,” it may be necessary to choose either the far clockwise or the far counter-clockwise rotation of the strobe flash as the point that will be used for the phase angle while balancing. It is not uncommon for the phase angle to vary by 30–90 degrees. Therefore, an arbitrary direction of rotation must

The Pump Handbook Series

be chosen in order to plot the vector, with the average of the amplitude used for the vector. The third plot (c) shows the “beat” amplitude after the “A” cell had been balanced. The “A” cell fan now can be put back on line, and “B” cell now can be balanced using the same procedure.

CONCLUSION The importance of good base line signatures, consisting of complete spectrums taken on each bearing in the horizontal, vertical and axial directions, cannot be overemphasized. These can be analyzed for existing problems, or be the spectrums to which all following spectrums will be compared. Analysis of vibration readings will help determine the problems with a pump and lead to successful repairs. The new base line signatures obtained after repairs will yield only good vibrations. ■ Gary Powers, P.E., is a vibration consultant and owner of Power Dynamics Engineering in Houston. Active in vibration analysis and field balancing of rotating equipment since 1968, Gary provides vibration analysis and field balancing consulting services to clients in the petrochemical, marine, power generation and related industries.

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R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

Getting the Full Benefit from High-Efficiency Motors When evaluating a changeover, look at two main factors— cost of electrical power and hours of operation. By Richard Cole and Terry Thome

lerted by the Energy Act of 1992, many users of AC induction motors seem to be eyeing premium-efficiency motors with mixed feelings, and perhaps some confusion. When the October 1997 deadline arrives for compliance with the 1992 standards, the clearest winners will be the environment (due to reduced need for more power generating plants) and the petroleum economy (because of reduced dependency on foreign oil)—but what will it mean for pump users? Most obvious is the potential for smaller electricity bills. Motors account for approximately 64% of the electricity consumed in the U.S., at a yearly cost of $112 billion. Simple arithmetic: every 1% reduction in motor demand cuts 0.64%—or $716,800,000—off the industry-wide bill. Higher efficiency motors also run cooler, which lightens the load on air conditioning systems and further reduces plant demand. Lower total demand can lead to lower rates by helping to minimize peak-demand surcharges. As rates increase over

A

HP

HIGH EFFICIENCY

STANDARD EFFICIENCY

10 25 50 100

93.2 93.0 93.6 95.4

88.5 90.2 90.2 91.0

time, these savings get bigger. Through a combination of premium-efficiency motors and electronic controls, however, certain types of pump applications offer more potential to reduce electrical bills than others. But electricity bills aren’t the only place where savings will show up. Premium-efficiency motors are built better and provide longer service life backed by longer warranties, which means lower maintenance costs. On the other hand, because they’re built better, premium-efficiency motors typically cost up to 25% more than a comparably rated standard motor. Buying higherpriced equipment always needs to be cost-justified, so let’s get that out of the way with the following example.

EVALUATING PAYBACK Standard-efficiency motors typically incur yearly operating costs of 10 – 20 times purchase price, compared to 8 – 12 times purchase price for premium-efficiency motors. A 100 hp standard-efficiency motor with an efficiency of 91.0 and a list

price of $3,785, running a 100 hp load continuously (8,760 h/yr) @ 8¢/kW-h, can be expected to rack up $57,450 in one year’s electricity bills. A 100 hp premium-efficiency motor with an efficiency of 95.4 and a list price of $4,404 costs $54,800 per year to operate under the same conditions—a savings of about 4.6% (Figure 1). In this example, the $619 highefficiency price premium returns an annual $2,650 savings, which pays back the price difference in 2.8 months. The savings then goes on to pay off the full price of the premiumefficiency motor in less than 20 months. Note that this example doesn’t factor in any “hidden” savings in peak-demand rate reduction, lower air conditioning and motor maintenance costs, or the return offered by local utilities through energy rebate programs, which can drop the payback period down closer to one year. Most pump users will see greater savings because most motors in industrial pump service are smaller than our 100 hp example—and the biggest gains in efficiency are found

.04/KWH

.06/KWH

.08/KWH

.10/KWH

.12/KWH

$149 $218 $526 $1,325

$223 $327 $790 $1,987

$298 $436 $1,053 $2,650

$372 $545 $1,316 $3,312

$447 $654 $1,579 $3,975

Figure 1. Projected energy savings based on continuous operation (8,750 hr/yr), motor operating at full load The Pump Handbook Series

121

in the lower horsepower ranges. For several important reasons which this article will discuss, however, maximum savings will not always accrue to every installation. Anyone considering a switch to premium-efficiency motors should be aware of these factors in considering a changeover strategy. First, let’s look at what makes a motor energyefficient and why that costs more. Understanding this is an important part of the essential cost-justification.

WHY EFFICIENCY RAISES PRICE Motors lose energy in several ways. The most significant among them are the “copper” losses that result naturally from electrical current passing through wire windings (Figure 2). Premium-efficiency design employs larger-diameter wire, increasing the volume of copper by 35 – 40%. To accommodate larger wire, the steel laminations that support the windings need larger wire slots. This reduces the amount of active steel in each lamination. To compensate for the loss of steel, more laminations must be added. Consequently, the rotor and stator core must be lengthened and the motor’s shell length increased. More metal adds more cost. Next comes magnetic core loss— technically divided into eddy-current and hysteresis losses. In premiumefficiency designs the longer rotor and core generally help decrease magnetic losses, but the make-up of the laminations is the key factor. Most standard-efficiency motors use low-carbon steel laminations around .025” thick, rated for electrical loss at 3.0 watts per pound. Premiumefficiency motors use high-grade silicon steel laminations around .018” Core Loss (20%)

Friction and Windage (6%) Watts Loss (Windings) (74%) Figure 2. Typical distribution of motor losses at full load

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thick, having an electrical loss of 1.5 W per lb. The chemical composition and thinner gauge of premiumefficiency laminations, plus a coating of inorganic insulation on each piece, combine to reduce eddy current losses greatly. However, better steel costs

more. Hysteresis losses, a result of molecular magnetic alignment properties too complex for this brief discussion, are reduced in premium-efficiency motors by special annealing and plating of rotor and stator components,

Effects of variable speed (inverter) controls on centrifugal pump systems driven by premiumefficiency motors Throttling System:

Totals:

Days

GPM

Head

RPM

HP

Total kW-h

109 109 129 18

3,600 3,200 2,800 2,000

115’ 120’ 127’ 133’

1,800 1,800 1,800 1,800

127 120 118 103

445,351 420,804 489,716 59,646

365

1,415,517

Utility Cost ($0.10819/kW-h) = $153,145 Variable Speed Drive (VSD) System: Days GPM 109 109 129 18 Totals:

3,600 3,200 2,800 2,000

Head

RPM

HP

Total kW-h

89’ 71’ 54’ 28’

1,620 1,440 1,260 900

93 61 40 13

326,123 213,909 166,005 7,528

365

713,565

Utility Cost ($0.10819/kW-h) = $77,201 Savings Using Variable Speed Drive System: $75,944

The above example shows two applications using the same size/model pump, in identical systems designed to deliver 4,000 gpm flow at 110’ head. Both systems are profiled in terms of flow groups, according to the annual total number of days the system operated at each stated flow level. Shown for each flow level are the pump speed (rpm) and motor horsepower needed to drive it. The Total kW-h column indicates the power required by the motor for each speed/hp level (Total kW-h = Days x 24 Hrs x hp ÷ 0.746). In a throttling system, lower flow levels are attained by closing a throttling valve, which causes an artificial increase in the system head or pressure. As a result, system hp requirements will decline slightly. In a variable-speed system, when flow is reduced by lowering the motor rpm, the total system pressure is reduced. Because less fluid is being moved at lower head pressure, there is a significant reduction in the system hp requirements. A dramatic difference is evident between the two systems in terms of motor hp and total kW-h requirements, resulting in lower utility costs.

The Pump Handbook Series

plus use of high-purity cast aluminum rotor bars. Friction losses are reduced by higher-grade bearings, and windage losses in fan-cooled motors are reduced by smaller, more efficient fan designs. And, overall, tighter tolerances and more stringent manufacturing process control are applied to minimize losses from unplanned conducting paths and stray load phenomena, both of which are common among motors. While all of these differences in material and manufacturing discipline combine to increase motor price, they make premium-efficiency motors run cooler than their standard-efficiency counterparts. Aside from cutting down on air conditioning costs, cooler operation lengthens the motor’s service life in two important ways. For every 10˚C reduction in temperature, motor insulation life doubles; premium-efficiency motors tend to operate 10 – 20˚C cooler than their standard-efficiency counterparts. Heat also is the primary cause of grease breakdown, which shortens bearing life; premium-efficiency motors tend to run 10 – 15˚C cooler at the bearings. Thus, premium-efficiency motors can provide up to four times longer winding life and twice the lubricant life of standard motor designs.

BEST CANDIDATES FOR CHANGE When evaluating a possible changeover to premium-efficiency motors, look first at two main factors—cost of electrical power and the hours of operation. The utility cost benefit of premium-efficiency motors begins to diminish when industrial power rates drop below 6¢ per kW-h. Note that the true kW-h cost very often is 1 – 24¢ higher than the actual base rate, due to the addition of peak demand charges and other penalties. True kW-h cost is most easily determined by dividing the facility’s total electric cost by its kW-h usage. Rates can vary widely by geographic area; wherever true kW-h cost moves above 6¢, the economic argument for changeover gains strength. In such areas, hours of operation will indicate which motors in your plant might provide the quickest and best opportunities to save money with higher efficiency.

Motors that run at least 2,000 hrs/yr (8 hrs/day, 5 days/wk) are your best candidates. As operating hours increase, the payback period shrinks (Figure 3). Remember that lower rates and shorter running times should not automatically rule out a change. Many local utilities offer energysaving rebate programs that might shorten the payback period enough to make premium-efficiency motors more cost effective. Also consider that the other benefits of better construction, cooler operation and longer life might still make premium-efficiency motors attractive no matter how much they help cut utility bills. A third consideration is the size and type of motor involved. Although the potential for premium-efficiency motor cost savings is significant throughout ratings from 1–125 hp, the greater gains are made in the lower hp ranges (Figure 4). The ratings that are most readily replaced are T-Frame, NEMA Design A, B or C and which operate at 3,600, 1,800, 1,200 or 900 rpm. In applications where changeover to T-Frame is difficult, premium-efficiency direct replacements for older U-Frame motors are available. These eliminate the need for adaptations on the replacement motor to match the original mounting dimensions. Applications in which seasonal or peak-load requirements must be accommodated by an oversized motor, but non-peak periods will force the motor to run at less than 50% of rated load throughout most of its duty cycle, usually cannot be justified on utility cost basis alone. Perhaps the least likely place to consider switching is in harsh environments where a motor’s life expectancy is so short that neither the energy-saving payback period nor the long-life benefits of better construction have any real relevance.

CALCULATING YOUR SAVINGS The process of deciding whether to switch to premium-efficiency motors typically begins with an attempt to calculate the savings that will result. Standard PC-diskette programs are available for this purpose, usually built around the following formula: The Pump Handbook Series

S = 0.746 x hp x C x N 100 – 100 SE PE where S =savings per year hp=motor horsepower C =energy cost ($/KWH) N =running time (hrs/yr) SE =efficiency (%) of standard-efficiency motor PE=efficiency (%) of premium-efficiency motor Note, however, that using this formula (or the diskette programs) often gets users into trouble because it presumes constant load (i.e. does not account for load variation), and it assumes that hp will be a true horsepower value (i.e. the worst-case horsepower requirements of the driven load). One cannot simply plug in the hp nameplate rating of the existing motor for two reasons: 1. The motor presently operating may be an inadvertent or incorrectly specified replacement for what was originally there. 2. The motor presently operating, even if identical to what was originally specified, may not be operating close to its rated nameplate hp. These are critical cautions when considering a switch to premiumefficiency motors. To deliver their promised economies, premium-efficiency motors must run between 75 – 100% of rated load. Operated at less than 50% load, they might even use more electricity than standard motors in the same service. Determining the correct hp value to plug into the formula requires the following “homework.” First, someone familiar with the pumping system must profile that system to determine when and for how long the motor works hardest. Many industrial processes move through regular cycles of varying difficulty in the course of a normal day. Some applications experience different levels of motor load or different operating conditions depending on season or time of month. Likewise, one should determine when the motor workload is lightest and for how long. Unusual aspects of the candidate motor’s load profile also should be noted. Intermittent or pulsating loads, such as occur with reciprocating pumps, may need a motor specially

123

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water or other liquids. As with the duty cycle, such notations should indicate whether environmental factors are continuous or periodic.

exchange systems using water or some other fluid for heating, cooling or steam generation. Such systems traditionally use a motor that runs constantly at full speed and full load, with fluid delivery controlled by pressure- and flow-regulating valves that allow varying amounts of excess fluid to recirculate through closedloop bypass plumbing. By switching to a premium-efficiency motor with variable-speed inverter control, and adding appropriate sensors and transducers to provide feedback signals to the inverter, fluid delivery can be modulated by varying the

“SMART” SYSTEMS SAVE MOST Virtually all pumping applications, especially those with round the clock operation, will gain 5–8% energy savings. As noted earlier, gains will be strongest for motors of 10 hp or less. Outstanding energy savings—ranging as high as 50%— are possible for centrifugal pumps in closed-loop throttling situations such as those commonly found in heat-

Power Costs in Cents per KWH

8 7 ♦

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3 4 5 6 7 Operation in Thousands of Hours per Year

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Figure 3. High-efficiency motor payback based on cost/kwh vs. hours of operation 100 95 ■

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80 75 ♦

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Efficiency (%)

built for that purpose. Replacement with a premium-efficiency motor not built for that purpose may result in higher current pulsations and an apparent low power factor. In addition, measurements taken during the times of hardest service must be compared to the motor’s nameplate rating to determine how closely that relates to a fully loaded condition. This can be done within reasonable margin of error in two ways: 1. Measure the motor’s rpm and compare it with nameplate full-load rpm. 2. Measure the motor’s amperage and compare that with nameplate amps. For both measurements, motor input voltage must be checked at the same time to make sure the motor is receiving full rated voltage. Motor performance varies with the square of the voltage change, so even a small voltage variation, such as a mild brownout, can make a big difference in your measurements. If voltage is chronically erratic or off nominal at the measuring site, consult the motor manufacturer for assistance in determining performance variations during voltage fluctuations. If rpm and amperage measurements taken under full voltage fall close to their respective nameplate values, the motor is running at or near rated load. If they aren’t close to nameplate values, these measurements can be used with standard amp/watt/horsepower charts to approximate the motor’s hp output at the time of measurement. It’s not a bad idea to plot the motor’s duty cycle graphically to show how much of that cycle keeps the motor above 75% of rated load. The savings that result from premium-efficiency motors will vary according to how much of the motor’s operating time is spent above that value. Load horsepower is most reliably determined when the motor can be temporarily replaced with a calibrated motor—one for which the amp/watt/speed profile has been documented—but the expense of doing this can make it impractical. Measurements also should be accompanied by descriptions of the motor’s operating environment, including ambient temperature, airborne dust or falling particulate,

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70 1

1.5

2

3

5

7.5 10

15 20

25 30

40 50

60 75 100 125 150

Horsepower ■

High-Efficiency Motor

♦ Standard Efficiency Motor

Figure 4. Efficiency comparison: industry average high-efficiency motor vs. industry average standard efficiency motor The Pump Handbook Series

routine-maintenance requirements associated with pressure/flow regulating components. Inverter controls typically are not recommended for positive-displacement pumping systems because this type of pump represents a “constant-torque” load. Here, a decrease in speed does not reduce torque, but instead demands a compensating increase in motor horsepower, therefore increased electrical demand. Nor are inverter controls usually recommended for standard motors. The “stepped wave” characteristic of

motor speed. Major savings are possible because centrifugal pumps are “diminishing torque” loads, for which the horsepower demand on the motor varies by the cube of the speed change. Consequently, a decrease in speed reduces torque, thereby greatly reducing the motor’s power requirement. (See accompanying sidebar for an example illustrating this concept.) This type of “smart” centrifugal-pump system not only greatly reduces energy requirements, but also eliminates the cost and high 100 95

Power Factor (%)

90 85

■

■

■

80 75 70

♦

♦

♦

■

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■

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■

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■

■

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♦ ♦ ♦ ♦ ♦ ♦ ♦ ♦ ♦ ♦ ♦ ♦ ■

65 1 1.5 2 ■

3

5 7.5 10 15 20 25 30 40 50 60 75 100 125 150 Horsepower

♦ Standard-Efficiency Motor

High-Efficiency Motor

Figure 5. Power factor comparison: industry average high-efficiency motor vs. industry average standard-efficiency motor 100

Power Factor, Efficiency

90

♦

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♦ ■

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■

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■

100

125

70 ■

60 50 ■

40 30 20

10 ♦ ■ 0 0

12.5

♦ Efficiency

25

50 75 Percent Load

■

Power Factor

Figure 6. Variation of efficiency and power factor with motor load The Pump Handbook Series

their power output tends to make the motor run hotter, shortening its life. Premium-efficiency motors, as noted earlier, run cooler by design and construction, and they are better able to accommodate the steppedwave power.

OTHER SYSTEM CONSIDERATIONS In any type of pumping application, premium-efficiency motors tend to run a bit faster than their standard-efficiency counterparts. This will aid the energy savings effect with positive-displacement pumps by slightly diminishing the motor’s hp requirements. However, in diminishing-torque circumstances in which pumps run at relatively constant speeds without inverter control (such as irrigation or drainage applications), care must be taken to see that the higher motor rpm does not increase the pump’s rpm. Here the cube relationship can work against you; that small increase in pump speed will demand a large increase in motor output hp, in turn causing the motor to draw more power and lose part or all of its energy-saving benefit. In belt-driven applications of this type, load rpm can be compensated easily by using variable-pitch sheaves to bring the pump back to specified speed. However, make sure that the load rpm does not have to be slowed to the point of causing the motor to run at less than 50% of its rated load, for reasons expressed below. Direct-drive applications involving diminishing-torque loads typically require additional engineering evaluation to assure energy-saving benefits. With the benefit of higher power factors than offered by standard-efficiency motors (Figure 5), premiumefficiency designs require less total current for an equal amount of work. Lower current demand on the line means that less energy is wasted due to line losses in all feeder circuits serving the motor, thereby enhancing the efficiency of the distribution system. This power factor advantage begins to erode as motor operation drops below 75% of rated load, and it declines sharply below 50% of rated load (Figure 6), which accounts for the general recommendation to operate the motor at or above 75% of rat-

125

ed load. Note from Figure 6 that premium-efficiency motors actually retain fairly high efficiency through loading as low as 12.5% of rated values, but at that level efficiency is negated by power factor decay. This suggests that in applications where motors must be oversized to accommodate occasional peak loading but can stay below 50% of rated load for prolonged periods, the operating economy of premium-efficiency motors can be preserved by adding simple, inexpensive power factor correction

126

devices to the line. Premium-efficiency motors also draw higher locked-rotor (standing start) current. Installations using thermally protected motor starters usually need to be checked for properly rated heater elements. Likewise, circuit breakers with higher trip ratings may be needed in the motor service line. The main caution regarding the switch to new energy-efficient electric motors is an old, familiar and simple one: look before you leap. Substantial savings will result if

The Pump Handbook Series

enough time and care are invested in determining the correct load values to use in the formula, so the premium-efficiency motor will be properly specified for the job you expect it to do and the environment in which it will work. ■ Richard Cole has spent 12 years with Dayton Electric and is currently Product Manager/AC Industrial Motors. Terry E. Thome is Senior Product Engineer/AC Industrial Motors & Drives at Dayton Electric.

✓

R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

Coupling Strategies Don’t let price be your prime consideration. By Michael M. Calistrat he selection of “the best” flexible coupling for a given application is not a simple task for three main reasons:

T

• Each application is different; the experience gained with one is not necessarily applicable to another one. • The perfect coupling has not yet been invented; each coupling type has its advantages and disadvantages. • Some coupling catalogs either lack all the necessary data for judicious selection, or the data given may be confusing.

It is unfortunate that coupling selection is often left to the discretion of purchasing agents; inevitably, purchase price becomes the prime consideration. The real cost of a coupling can be divided into:

cussed in this article. What is important to consider is the fact that not all 20 hp pumps are the same, and that a turbine’s load coupling and accessory drive coupling are quite different. Whenever considering a given application, the following items should be determined: • power consumed/power available • normal operating speed/speed range • uniformity of torque that coupling will transmit • maximum torques that might occur (peak torques) • type of maintenance available • importance of machines for production, or for safety • availability of spare parts • space available and shaft data Although the items listed above

De-Rating Factor

Of these costs, the purchase price is the smallest! Too many newly installed couplings fail within a few days. Maintenance departments must often stock loads of spare parts, and sometimes millions of dollars are spent in the repair of machines that are damaged when couplings fail.

It’s obvious that a 20 hp pump is different from a 50,000 hp gas turbine; that’s not what will be dis-

POWER CONSUMED, AND POWER AVAILABLE One example will be used throughout this article: a two cylinder, double acting reciprocating water pump driven by an induction electric motor, through a geared speed reducer. The power consumed by the pump is 1,525 hp at 600 rpm, and the electric motor is rated at 1,800 hp at 1,760 rpm. The reason the motor power is larger than needed is that seldom can an electric motor (which is a highproduction item) have exactly the same power as the particular machine it drives. Therefore, motors are selected based on the next larger increment of power available from the manufacturer’s catalog. Actually, motor data are standardized by the National Electric Motor Association

1.3

• the price • the installation and maintenance expense • the expense to repair the coupling during its service life • the cost of consequential damages – those resulting from possible failure of the coupling

DETERMINATION OF APPLICATION CHARACTERISTICS

might seem obvious, they should be carefully analyzed.

1.2

1.1

1.0 0

4

8

12

16

20

24

Hours per Day Service factor that allows for the use of full rated torque only if machines are never turned on. Source: ATR Sales Inc., Atra-Flex catalog. FIGURE 1 The Pump Handbook Series

127

(NEMA). From an energy consumption point of view, the selection of a larger-than-necessary motor is insignificant; motors draw only the energy they are required to deliver. For a proper coupling selection both machines power ratings are important; the continuous coupling torque rating should satisfy the power consumed (in the example, the one of the pump), while the coupling’s peak-torque rating should satisfy the torque available from the driving machine (in this example, the motor’s).

NORMAL OPERATING SPEED AND SPEED RANGE In the example, the operating speeds are fixed: 1,760 rpm at the coupling between the motor and the gear box, and 600 rpm at the coupling between the gear box and the pump. This is not always the case; if the driving machine had been a steam turbine or a variable-speed motor, the speed range would extend both below and above the normal operating speed. Knowing the speed at which a coupling will operate is very important, and it seems unlikely that such data could be omitted. However, machines imported from Europe are designed for an electrical frequency of 50 Hertz; when used in the U.S., the speed increases by 20%. This speed increase did cause the failure of the supplied coupling in one case I studied.

UNIFORMITY OF THE TRANSMITTED TORQUE The torque required to drive a reciprocating pump varies significantly during one revolution of the pump shaft, from maximum when pistons are about half way through their stroke, to almost zero when pistons are at either end of their stroke. Couplings must be able to transmit the maximum torque during one cycle, which is always higher than the one calculated based on power and speed. In addition, the uniformity of the torque supplied by the driving machine must be known. The torque delivered by an electric motor is uniform, but the torque delivered by other machines, such as a diesel engine, is seldom uniform. Finally, synchronous motor dri-

128

ves impose severe torsional excitation on shafts during startup. Special couplings must be selected in such situations.

PEAK TORQUE Too often, only the power (torque) consumed by the pump is considered; however, couplings must be able to withstand large but occasional peak torques. The obvious peak torque (in this example) to be considered is the one that occurs at startup. For instance, induction motors’ torque is about 3 times larger at startup than at normal speed. It is during startup that a motor must, besides supplying torque for pumping, supply torque for accelerating the pump components, the driving gears, and the column of liquid from standstill to normal speed. The energy required for these accelerations can easily require all the torque that a motor can deliver. Therefore, the selected coupling must be able to withstand (for a short time) the maximum torque of the driver, and not only the one required by the driven machine. If the driver is a steam turbine, the peak torque will be roughly the same as the continuous torque because steam turbines’ acceleration is low. Some applications have peak torques that can be even larger than the ones that the driving machines can deliver. This is the case of reversing drives. For example, the drive that moves an overhead crane on its rails can experience extremely large torques whenever an operator reverses the direction of movement without waiting for the bridge to stop. The drive coupling will be subjected at that moment to the maximum torque of the motor, plus the torque generated by the kinetic energy stored by the moving mass of the bridge. Some manufacturers supply couplings that have double the number of flange bolts for crane applications.

TYPE OF MAINTENANCE AVAILABLE To show why the type of maintenance is important in determining the application requirements, I will use the case of repeated coupling failures in a steel mill. Large plants such as this one often have two (or even more) indeThe Pump Handbook Series

pendent maintenance departments – a mechanical and an electrical department, for instance. The installation of electrical motors to gear boxes in this facility presented serious problems: the half-coupling at the gear box was installed by the mechanical maintenance department, while the other half-coupling was installed on the motor shaft by the electricians. No grease was used in this half-coupling, simply because electricians do not have grease. Consequently, after couplings were bolted up, they operated with insufficient grease and wore in a short time. For plants with such a maintenance procedure, nonlubricated couplings would have been a better selection. The skill level of mechanics also plays an important role in the coupling selection. To exemplify this requirement, let’s examine the case of a rapid transit company. This company’s maintenance department had to service couplings installed at each car drive. To simplify installation and removal of couplings, management selected couplings with keyless hubs. The union contract determined that coupling work has a low-skill requirement, however hydraulic installation and removal is definitely not a low-skill job. The conclusion was that the labor contract had to be rewritten so that highly-skilled mechanics could work on couplings. Many more such examples can be described. The important thing to remember is that the type of maintenance available in a plant is an important parameter in categorizing the application.

IMPORTANCE OF MACHINES In the example (a 1,525 hp pump), a plant might have a number of such pumps operating in parallel. A wise manager will install at least one more pump than is actually required – a pump that will be used as a spare, to replace any of the others whenever they are in need of repair. With the existence of a spare, neither of these pumps can be considered critical for production. On the other hand, a sump pump that is used for water removal in a mining operation is obviously a critical element in the system because without it the mine can

become flooded. Evaluating the importance of machines for production or safety is a key parameter in defining the application. In the case of the pump for the mining application, it is important to use a coupling type (or size) that is unlikely to fail between regular maintenance stops, or a type of coupling that, should it fail, can be replaced in a very short time.

Constant Torque . . . . . . . . . . . . . . . . . . . . . . . . . .1.50 API 671 . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .1.75 Moderate Fluctuations . . . . . . . . . . . . . . . . . . . . . .2.00 Three manufacturers of high-speed disk-pack couplings mandate a minimum application factor of 150%, even for best conditions. Therefore, the published torque ratings cannot be used. Sources: • Rexnord Coupling Operation, THP coupling catalog. • Kop-Flex Power Transmission Products, Bulletin HP 103. • Flexibox Inc., Leaflet F100(US)/1192.

AVAILABILITY OF SPARE PARTS

FIGURE 2. Application factors

Plants located in an industrial area need not stock spare parts. This is because power transmission distributors compete for plant business and make available replacement components for couplings. On the other hand, a plant located on an island in the South Pacific must stock spare parts for its couplings, even though it might use the same machines as the plant located in an industrial area.

standard values. In addition, few catalogs provide information on tolerances for shaft diameters. The lack of this information is striking when compared with the ample amount of similar information provided in most anti-friction bearing catalogs. Before starting a size selection for a particular type of coupling, users must ensure that sufficient data for proper selection has been made available by the manufacturer.

DIMENSIONS Coupling selection can be made only after shaft sizes, shaft separation and space available are known. For instance, in the example, the data required are: • pump shaft diameter • gear box output shaft diameter • shaft separation • space available (most gear boxes have their oil pump at this location) • motor shaft diameter • pinion shaft diameter • shaft separation • space available

AVAILABILITY OF COUPLING DATA Couplings are selected by comparing requirements with catalog ratings and dimensions. Some coupling catalogs contain a wealth of data; others are succinct. For example, while all catalogs contain data on the outside diameter of couplings, none includes the diameter of elastomer elements at speed. Without such information it might be difficult to determine if a particular coupling will fit in the space available. Another example is the data on shaft-to-shaft separation. Very few catalogs say whether a coupling can be used if the actual separation differs only slightly from

CONFIDENCE IN CATALOG DATA Coupling catalogs are supposed to provide reliable data on the ability of a coupling to perform in a given application. Unfortunately, some catalogs contain questionable information. The best example of this is shown in Figure 1, which was reproduced from a catalog of elastomer couplings. As indicated, the published torque ratings can be used only if the machines are never turned on! Some catalogs are titled “High Torque” to differentiate the described couplings from the ones in the standard catalog. Reading the fine print of these publications shows that in reality the high-torque couplings are identical, in all respects, to the standard couplings; only the selection method is different. Other catalogs list coupling sizes that are actually not available. For instance, certain diaphragm coupling catalogs allow the user to select packs of as few as three, or as many as 15 disks. Only when placing an order will the user find out that only one number of diaphragms of each size comes in a pack. I could provide other examples of insufficient or inaccurate catalog information. The point is that before actually selecting a coupling size, users should study the data supplied The Pump Handbook Series

in catalogs and determine if their needs can really be met.

SERVICE FACTORS As previously discussed, couplings must be selected for the maximum torque that occurs during one cycle (which most often represents one revolution), and not for the data taken from a machine’s nameplate. The original equipment manufacturer knows this torque but seldom publishes it. This is why coupling manufacturers provide listings of multipliers for the average (or calculated torque), as obtained from a machine’s nameplate. This multiplier is known as the service factor. Users must be distrustful of service factor listings that do not include unity for even the smoothest possible applications. Typical applications where most manufacturers list a service factor of one (1) are: • centrifugal blowers • centrifugal pumps • belt conveyors • centrifugal fans • generators Two service factors must be used: one for the torque consumed, and one for the torque supplied. The two factors are often listed independently and they are additive. (Do not multiply these two factors.)

APPLICATION FACTORS Theoretically, a coupling size can be selected based on catalog ratings; however, this is seldom the case. Just as electric motors are selected based on the next largest size as compared with actual requirements, couplings are selected so that catalog ratings are larger than the actual torque that will be transmitted. The ratio between the catalog-rated torque and

129

COUPLING TYPE

SERVICE FACTOR

SIZE

TORQUE RATING SERVICE FACTOR

APPLICATION FACTOR

SPEED RATING

APPLICATION FACTOR

MAXIMUM BORE DIA.

OUTSIDE DIAMETER

MOTOR COUPLING Operating Power Operating Speed HP/100 rpm Shaft diameter Space available

1,525 HP 1,760 rpm 87 3.5 inches 30 inches

Gear Disk-pack Rubber-tire Urethane-tire

1.75 2.6 1.5 2

1035 G 550/50 PX 200 E 120

131 HP/100 rpm 109 HP/100 rpm 87 HP/100 rpm 135 HP/100 rpm

148% 123% 100% 152%

3,900 6,300 1,300 1,800

217% 350% unusable 100%

4.1 inch 5.1 inch 7.5 inch 7.5 inch

11 inch 15 inch 24 inch 25 inch

PUMP COUPLING Operating Power Operating Speed HP/100 rpm Shaft diameter Space available

1,525 HP 600 rpm 254 6 inches 19 inches

Gear Disk-pack Rubber-tire Urethane-tire

1.75 2.6 1.5 2

1050 G 750/51 PX 280 E 120 1

371 HP/100 rpm 302 HP/100 rpm 320 HP/100 rpm 135 HP/100 rpm

146% 119% 126% unusable

2,900 4,700 1,080 910

480% 780% 180% 150%

7.0 inch 7.5 inch 9.0 inch 7.5 inch

15 inch 20 2 inch 28 2 inch 25 2 inch

NOTES: 1. Largest size available 2. Will not fit in space available This chart helps in obtaining an overview of the parameters needed for selecting the couplings under consideration.

FIGURE 3. Example of coupling data organized for selection process

actual torque (after conditions made based on service factors) is defined as the application factor. The magnitude of an application factor depends not only on the rating steps in the catalog, but to a great extent on these considerations: • previous experience with a vendor • confidence in catalog data • provisions for future upgrading of the machines for which a coupling is selected. For this reason application factors are also called experience factors. The following issues should be considered: • A large application factor ensures that the buyer gets a coupling that can do the job required, with plenty of reserve, for a long time. • A large application factor, on the other hand, means that the selected coupling is larger, heavier and inherently more expensive than a coupling selected using a smaller factor. In other words, safety and peace of mind come at a price. Considering that couplings represent a small expense, compared with the cost of the equipment on which they are installed, price should seldom be a selection criterion for couplings. Each coupling user should establish his or her own application (experience) factor, which can vary from coupling type to coupling type, and from vendor to vendor. The American Petroleum Insti-

130

tute (API) recommends the use of an experience factor of 175% for special-purpose couplings in refinery applications. This large factor is only a guideline for users that have no experience in a particular case. Some coupling manufacturers mandate a minimum application factor of 150% (Fig. 2). The catalog ratings of these couplings cannot be used under any conditions!

SELECTION OF COUPLING SIZE At this point enough data is available to make a size selection for the couplings in the example chosen, even though a decision on the type of coupling is yet to be resolved. This is because there is not enough data to make a judicious type selection. It should be noted, however, that based on purely subjective reasons (such as previous experience, preferences or instinct) only four coupling types were considered for the particular example: • gear-type • disk-pack • rubber tire, and • urethane tire.

GENERAL CONSIDERATIONS Considering that two couplings are required for the example chosen, and that four types are under consideration, it is best to have the data plotted in chart form (Fig. 3). It should be noted that the smallest coupling that fits the requirements is listed. This smallest size might have an application (experience) factor that is too small for The Pump Handbook Series

some users. In such a case the next larger size must be evaluated, particularly for maximum speed and space available.

PUMP COUPLING SELECTION The urethane-tire coupling cannot be used because even the largest size manufactured is rated for a smaller torque than required. Either one of the other three couplings have acceptable torque and speed ratings. The oil pump of the gear box limits the space around the low-speed coupling; because of this, only the gear-type coupling can fit in the space available. Assuming that a disk-pack coupling can fit in the space available, the user’s decision must be based on the type of maintenance available, as well as on non-technical factors such as delivery, price and availability.

MOTOR COUPLING SELECTION Because the rubber-tire coupling cannot operate at the required speed, only three coupling types remain under consideration: the gear-type, the disk-pack and the urethane tire. The choice in this case is simplified by the application: whenever large torque fluctuations occur, it is preferable to have a torsionally soft coupling in the system. Torsionally soft couplings are best installed at the low-torque end (for weight and cost considerations), which is the motor shaft. Therefore, the urethane-tire coupling is selected at the motor. Note that the size selected has no application margin for speed; the manu-

The following factors are applicable for the selected example: Reciprocating pump Power...........................................................................................1,525 HP Speed ...........................................................................................600 rpm Average torque .............................................................................1525x63025/600 = 190,000 in/lbs Service factor for a gear-type coupling.........................................1.75 Service factor for a disk-pack coupling ........................................2.6 Service factor for a rubber tire coupling.......................................1.5 Service factor for a urethane tire coupling....................................2.0 Electric motor Power...........................................................................................1,800 HP Speed ...........................................................................................1,760 rpm Average torque .............................................................................1525x63025/1760 = 54,600 in•lbs Peak torque ..................................................................................1800x63025x3/1760 = 193,370 in•lbs Service factor ...............................................................................0 TABLE 1

facturer must be consulted for assurance that the selection is acceptable.

VIBRATION STUDIES

No decision on coupling selection is final until vibration studies show that resonance frequencies are sufficiently removed from the operating speed. Torsional vibration studies. Torsional vibration studies are easy to perform, now that a number of computer programs are available. To use these programs, it is necessary to obtain the following data: • rotational inertia of all the components in the train • torsional stiffness of the portions of machine shafts which are under torque

• torsional stiffness of the selected couplings, including reliable data on the hub-toshaft connections. • reduction ratio of the gear box • damping factor of elastomer couplings Lateral vibration studies. Lateral vibrations seldom occur at electric motor shafts, gear box output shafts, or at reciprocating pump shafts. This is because such shafts are designed to accommodate belt drives, which impose large radial forces on shafts. Shafts must be sized for these lateral forces; when couplings are used, such shafts are subjected only to small stresses and are unlikely to suffer from lateral vibration. High-speed shafts (input shafts) of

The Pump Handbook Series

gear boxes are slender and could operate near, or even at, a lateral resonance frequency. Therefore, users can take a reasonable chance in not performing lateral vibration studies for the example chosen, with the exception of the pinion shaft. It is always advisable, of course, to conduct complete lateral vibration studies.

CONCLUSIONS As we’ve seen, the coupling selection process is quite involved and requires knowledge that is seldom available to purchasing departments. Undoubtedly, price and delivery are also important, and purchasing agents can be of help with this aspect of selection. The final decision on vendor selection, however, should always involve the people that made the type and size selection, so they can continue to apply their experience factor. ■ Editor’s Note: This article was excerpted from Michael M. Calistrat’s book Flexible Couplings—Their Design, Selection and Use. The most definitive work on the subject we’ve seen, the 575 page publication may be ordered from: Caroline Publishing, P.O. Box 451611, Houston, TX 77245-1611, Fax (713) 437-4656 Mr. Calistrat is an international expert on flexible couplings and the owner of Michael Calistrat and Associates, a consulting firm specializing in design and failure analysis of rotating machinery. He has obtained 17 patents and was Chairman of the ASME International Conference on Power Transmission and Gearing.

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Adjustable Speed Drive Offers Pump Flexibility By Robert C. Waterbury, Senior Editor

D

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Energy Consumed as a Percentage of Unrestricted Full-Flow Energy Consumption Percentage of Flow

Throttle Valve Energy

100 95 90 85 80 75 70 65 60 55 50 45 40

Bypass Valve Energy

100 96 94 93 89 86 83 81 79 76 74 72 71

100 100 100 100 100 100 100 100 100 100 100 100 100

ASD Replacing Throttle Valve

Bypass Valve

105 90 78 66 57 48 41 35 30 25 21 17 14

105 98 93 87 82 76 71 66 61 56 51 46 41

TABLE 1. Relative energy consumptions of different pump flow control strategies in Table 1. According to EPRI, pumping operations can be controlled by “using the differences between a user-selected pressure setting and the measured value of the actual fluid pressure. The two signals then can be compared and the resulting difference used to generate a control signal that is sent to the ASD such that the motor operates the pump with constant pressure.”

PUMP/ASD CHECKLIST But perhaps the biggest questions that pump users have are “How do I know if I need an adjustable speed drive? And if so, where?” Darryl Van Son, manager of market research at Baldor Electric, a major supplier of drives and motor products, has put together the folThe Pump Handbook Series

lowing checklist to help answer these and other common questions. Why should I consider an adjustable speed drive on my pump system? 1. It may be able to save a substantial amount of electrical energy 2. It can help reduce wear and tear on the mechanical components being driven 3. It can help reduce “water hammer” effects due to sudden pressure changes When should I consider an adjustable speed drive? 1. If you have long hours of use

Source: Electric Power Research Institute (EPRI)

rives are a prime example of integrated power and control. It takes both to do useful work—especially in an industrial environment. Power, distributed through devices such as circuit breakers and transformers, may eventually be controlled by distributed control systems and programmable logic controllers. These two domains of power and control, however, are often linked by a drive positioned between them. A drive provides “powered integration” to the working environment. For this reason, drives are used throughout industry—in pumping and production, water and wastewater treatment, manufacturing and many other applications. Adjustable speed drives in particular can be applied to boiler feedwater pumps in power plants and process plants as well as hot water circulation pumps in commercial buildings. Projects sponsored by the Electric Power Research Institute (EPRI) have demonstrated the effective use of adjustable speed drives (ASDs) for boiler feedwater systems, hot water circulation systems, wastewater treatment plants and municipal water pumping systems. According to EPRI, the ASDs used in these applications range from 5 hp up to several thousand. In reducing or regulating system throughput, a user has the option of maintaining full pump speed and restricting the flow of liquid simply by using valves or similar devices. This is not only wasteful but potentially harmful to the operating system. Adjustable speed drives, on the other hand, can control pump speed and help save energy costs, as shown

What do I need to know? 1. Actual voltage range and amp draw 2. Operating hours at each output set point 3. Ambient temperature range 4. Distance between motor and proposed control 5. Quality of incoming power a. disturbances b. outages c. dips d. sags e. spikes, etc. What are the advantages of an adjustable speed drive? 1. Matches speed to output need 2. Can offer infinite adjustments from process inputs (e.g., psi and flow) 3. Can be operated at preset speeds from external command 4. Has soft start and soft stop for smooth output changes 5. Can preset speed range for minimum and maximum output 6. Can save 25-50% of energy usage What are the potential problems and likely fixes? 1. Problem: Line disturbance nuisance trip–off Fix: Add a line reactor 2. Problem: Premature motor insulation failure from high switching frequency Fix: Program ASD for its lowest frequency without audible noise, and add load reactor 3. Problem: Premature motor insulation failure from resonance in the line between

motor and control Fix: Keep line length less than 50’ or add a load reactor 4. Problem: Motor noise at one particular speed Fix: Program “skip frequency” into ASD so it will not stay at that speed 5. Problem: ASD reflecting harmonics back up the power line, disturbing other sensitive equipment Fix: Add line reactor or harmonic filters Van Son adds that adjustable speed motor drives are changing so rapidly that it is difficult to specify the right system. For that reason he believes users should exercise caution in using “canned” specifications offered by vendors. It can be far better, he contends, if the user develops his own specifications that don’t artificially limit possible solutions. In so doing, the user may discover solutions that he didn’t even know existed. New approaches, devices and control algorithms can provide improved performance and efficiency that result in maximum productivity and reliability. And not every vendor uses the same techniques.

ASD PRODUCT APPROACHES Let’s look at several recent adjustable speed drive product offerings to see how they differ in their market and applications approaches. GV3000. Reliance Electric has introduced what is claimed to be the first integrated vector and generalpurpose drive, designated the GV3000 controller. It combines features of both AC and DC drives in applications ranging from 1–150 hp. In the vector mode, GV3000 offers high dynamic response, maintains full-rated motor torque to zero speed, and controls motor speed precisely in both directions using pulse tachometer feedback. It reportedly provides speed and torque regulation equal to or better than possible with DC drives. In the general-purpose mode, GV3000 is suited for a broad range of adjustable speed applications including pumps, fans and blowers in materials handling, food processing and water treatment applications. Reliance chose to offer this package because it incorporates two solutions in one, enabling users to reduce inventory while gaining added operational flexibility. It also offers a standard RS-232 port for communications to a Windows-based Configuration

PHOTO COURTESY RELIANCE ELECTRIC CO.

2. If your energy costs are high 3. If your system has a throttling valve or recirculation valve that regulates the output of the system 4. If you spend long hours at partial output 5. If you have frequent starts and stops on booster pumps

Photo 1. The GV3000 AC drive is the first combined vector and general-purpose ASD. The Pump Handbook Series

133

Executive. And optional high-performance modules allow connection to networks including the Reliance AutoMax DCS and DeviceNet™. Altivar 66. Square D, on the other hand, recently introduced its Altivar 66 line developed jointly with Telemecanique for global markets. The thrust behind this product is to build a single world-class product— one product, one part number—that can be used anywhere in the world for nearly any application. Square D’s approach called for standardizing on a basic core product meeting common industry standards; using modular design concepts to expand product flexibility; and allowing integration with other products using multiple standard communication protocols. The result is a product that can be customized using specific applications modules such as flux vector control, regenerative control, position control, PID regulation, logic functions and crane control. And to accommodate international variations in current input power, the drives are designed to sense the power

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level and automatically adapt so their operation becomes transparent to the user. It becomes, in effect, a “smart” drive. Square D also notes that the modular concept allows operators to purchase the exact amount of functionality needed, with the flexibility to add on as needs increase or change. Altivar 66 products operate in the following voltage ranges: 208– 240 V, 400–460 V, and 575–600 V, at 50–60 Hz. Power range is from 3–350 hp in the constant torque operating mode, and 3–400 hp in the variable torque mode. Frequency range is 0.1–400 Hz.

ASD IMPLEMENTATION AND USE As Darryl Van Son suggests, it is best if the user can develop his own ASD performance specifications according to how he plans to use the devices. But it is also useful to note some general cultural differences that influence the design and use of ASDs. In designing the Altivar 66, for example, Square D chose to employ functional maskability. This means

The Pump Handbook Series

that for basic applications the user is exposed to only a small fraction of the device’s total functions. Higher level functions for more complex applications can be implemented by exercising optional modules or software packages. Other differences abound. In the U.S. there is a preference for using 3–wire control as opposed to 2–wire control common in Europe. Fuses, wires and short circuit protection may be different as well. American operators frequently prefer to operate the drive directly from a keypad, whereas in Europe drives are often integrated into systems that are controlled by engineers from a central location. Likewise, drives are often installed here by technicians as opposed to engineers. The purchase decision really comes down to answering the question: “Does an ASD offer the performance and savings I need without overpaying for unused functionality?” It’s your call, but hopefully the information provided here will help make the decision somewhat easier. ■

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R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

Solid State Controls Tune Pump Operation Solid state motor controls can offer smoother operation, minimize mechanical shock and lower life cycle costs for centrifugal pumps. By Jeihri Robinson echnologies used in designing and implementing centrifugal pumps have advanced significantly in the past decade. With advances in material, manufacturing capability and computer aided design, we can be assured of continued progress. Along with these changes, however, have come additional concerns in using the pump as part of the overall system. One of these concerns involves starting and controlling the motor connected to the pump. This concern includes several aspects of the system. The first is the capacity of the electrical system and whether there is enough power available to handle the load without disrupting other equipment. The mechanical shock caused by across-the-line starting may be damaging to the pump and cause long term maintenance problems. The use of a backup generator also may be an issue, especially in the case of a retrofit. These generators may have serious capacity problems when used on larger horsepower motors. The key consideration in selecting the starting method is installed cost. This differs from the initial cost, which refers to the actual purchase price and cost of installation. A good example would be an acrossthe-line starter compared with a solid state starter. While the solid state device may cost more initially, it does not cause the same kind of mechanical shock associated with the across-the-line. Mechanical shock can have many unwanted effects including water hammer, check valve slamming, cavitation,

T

seal failure and bearing failure. Any one of these problems can dramatically increase the installed cost of the device. Therefore, the pump system is more expensive to operate in the long run. Or special control valves that take up additional pump house space may be needed.

SYSTEM DYNAMICS Let’s analyze some of the more common system anomalies and how they relate to the starting method. The first thing to consider is the system flow dynamics. Hydraulic surging or water hammer results from a rapid change in flow. This can be caused by the opening and closing of a valve or the starting and stopping of a pump. The intensity of the surge depends on the pressure variance across the valve or pump and how rapidly the flow changes. The noise or hammer is only part of the physical problem. Cavitation also can occur, causing premature pump impeller wear and erosion of the pipe inner lining. In the case of a pump, the flow changes rapidly when full voltage is applied to the motor. If we look at the relationship between the motor torque and speed and a typical pump’s torque requirement over the speed range, we can see why the system experiences a rapid change in flow. Figure 1 shows how much excessive acceleration torque is being applied to the pump. Acceleration torque is the amount of difference between the applied torque and the required torque. If the acceleration torque can be lowered, the start will be smoother. So the key to eliminating the potentially harmful effects of this starting method is to reduce the The Pump Handbook Series

applied torque. This can be accomplished in several ways including voltage control, frequency control and mechanical torque dampening. We are focusing here on only the voltage control. The relationship between the voltage and applied torque is: Torque Applied α Voltage2 This means that a small drop in voltage can cause a significant drop in applied torque. For example, a 10% drop in voltage would result in a 19% drop in applied torque. If the voltage is reduced during acceleration of the motor, the current will also be reduced (Figure 2). The relationship between the voltage and current is as follows: Voltage Applied = Current Drawn Voltage Max Current Max The equation indicates that a drop in voltage will cause a proportional drop in current drawn. This is true only during acceleration and does not apply at full speed. When the motor accelerates to full speed, the pump output changes from zero flow to rated flow. The net result is a rapid change in flow over a very short period of time, as shown in Figure 3. To slow down the acceleration of the motor, less torque should be applied. In doing this we must consider some of the issues mentioned earlier, namely power constraints. The amount of current draw is also a consideration when reducing the voltage.

REDUCED VOLTAGE STARTING METHODS The most common method of reducing voltage at startup is through electromechanical means. There are three very popular methods: the Wye

135

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One method that has gained popularity recently is the solid state reduced voltage starter. This device uses silicon-controlled rectifiers (SCR’s) to control the voltage applied to the motor. Basically, the voltage realized by the motor is reduced by phasing back the firing angle. Because of the lack of mechanical constraints, there is no transition when starting. Smooth, stepless starting lends itself to much smoother control of the motor’s acceleration. We have already looked at the torque and flow characteristics of an across-the-line type of start (Figures 1 and 3). Now let’s look at the results of a reduced voltage start. We will use the solid state ramp as an example since it is the best method discussed so far. Figure 5 shows the effect of starting a pump motor with a solid state reduced voltage starter. The solid state starter provides excellent torque reduction initially, but the torque increases too quickly as voltage increases. This is due to the open loop method of control, and it would be the case also with the other electromechanical means. At about half speed the torque starts to increase quickly; this also is the point where the pump really starts working.The net result is an abrupt change in flow (Figure 6). Another solution also involves a solid state reduced voltage starter — but one with a few more features. As solid state technology advances, the capabilities of the solid state starter increase. A number of units are now available with microprocessor control. When the microprocessor is used, a feedback system can be used to create a closed loop acceleration. This would help eliminate the rise experienced after about 35% speed. Figure 7 depicts the torque curve from such a unit. By using a feedback system, much smoother acceleration can be realized during startup without abrupt change in flow.

TORQUE CONSIDERATIONS Torque vs. Speed

180 100 0

The Pump Handbook Series

100 Percent Full Speed

Figure 1

CURRENT CHARACTERISTICS Current vs. Speed 600

Reduced Voltage Start

100 0

100 Percent Full Speed

Figure 2

FLOW CHARACTERISTICS Flow vs. Time 100 Abrupt change in flow

0

STOPPING A PUMP

This does not address the problems that can occur during the stopping cycle. Figure 8 demonstrates the

Typical Pump Requirement

Across the Line Torque Curve

% Full Load Torque

SOLID STATE STARTERS

flow characteristics when power is removed from the motor. The only energy left is the inertia of the pump and motor, which is quickly dissipated by the back wash of water. Figure 9 shows three different stopping methods, including coast to stop, soft stopping and controlled stopping. This chart demonstrates that a closed loop controlled stop can help

% Full Load Current

the closed transition will not remove power and thus eliminate the exessive spike.

% Flow

Delta starter, the part-winding motor starter and the autotransformer. The first and most popular method in the pumping industry worldwide is the Wye Delta starter. The Wye Delta generally requires a special motor that brings the ends of all the windings out to the motor conduit box. By changing the configuration of the connections, two different voltages can be achieved. The biggest problem is that the first voltage is fixed and will not work if additional torque is required. Another method is the partwinding motor starter. This method uses a specially wound motor with two identical windings in the stator. During the startup one set of windings is used to bring the motor up to a certain point in speed. The second set is then brought in parallel to the first. This also has a fixed starting point, but that is generally sufficient for most centrifugal pumping applications. The autotransformer is a very effective method of achieving reduced voltage starting. It is based on the property of a transformer to shift voltages. By changing taps, the following percentages of full voltage can be applied to the motor: 50, 65 and 80%. This method is ideal for weak power lines. It delivers a lot of torque for a very low current draw. Also, unlike the Wye Delta or partwinding starts, the starting torque can be increased by changing the taps on the transformer. Unfortunately, these starters tend to be large and expensive compared to the other two. One drawback of the electromechanical means of starting is what is known as a transition. Figure 4 shows what this transition looks like. The current drawn by this motor is significantly reduced; in this case it draws approximately 200%. The transition point is generally controlled by a timer. An incorrect setting could producce a transition too early, resulting in a current spike. This diagram depicts an open circuit transition, the transition removes power from the motor and then reapplies full voltage. The open circuit transition can result in a current spike far in excess of the locked rotor current rating of the motor as the magnetic field in the motor collapses and re-energizes. You can get a closed transition device. However,

Figure 3

Time

TYPICAL REDUCED VOLTAGE START

Torque vs. Speed 600 Current Drawn

Open Circuit Transition

100 0

100 Percent Full Speed

% Full Load Torque

% Full Load Current

Wye-Delta Start

SPECIALIZED SOLID STATE CONTROL TORQUE CHARACTERISTICS

Across the Line Torque Curve

Solid State with Closed Loop

DIAGNOSTICS 180

Typical Pump Requirement

100

0

Figure 4

100 Percent Full Speed

SOLID STATE TORQUE CHARACTERISTICS

Figure 7

Torque vs. Speed FLOW DYNAMICS OF SPECIALIZED SOLID STATE CONTROLLERS

Typical Pump Requirement

Flow vs. Time 100

Solid State

100

0

100

% Flow

% Full Load Torque

Across the Line Torque Curve

180

Smooth Transition with Closed Loop

Percent Full Speed 0

Figure 5

Time

Figure 8 SOLID STATE FLOW RESPONSE Flow vs. Time

STOPPING CHARACTERISTICS Flow vs. Time

100 100

Controlled Stop

% Flow

% Flow

Abrupt change in flow

0

Soft Stop

Coast to Rest

Time

Figure 6

eliminate the harsh stopping that normally occurs when the motor is de-energized. It also helps eliminate the problem associated with soft stopping. Soft stopping is the opposite of soft starting. Instead of ramping the voltage up, it ramps it down. Because the relationship between voltage and torque is a square function, the torque drops off too rapidly during a soft stop — which causes the quick stopping. By using feed-

id state offers an excellent return on investment. This is especially true if a unit designed specifically for pumping applications with the closed loop control capability is used. Now we will look at some of the other capabilities of the solid state starter.

Time Figure 9

back from the motor, a more sophisticated solid state starter is able to drive the motor to a complete stop and automatically turn it off . This capability saves wear and tear not only on the pump but also the mechanical system, including check valves, seals and piping. Looking at the installed costs of these devices, it becomes clear that the solThe Pump Handbook Series

A significant advantage is that solid state control gives the ability to increase the availability of diagnostic information. This may significantly reduce the installed cost of the device by preventing small problems from becoming large ones. Some of the offerings in these devices include overload, underload, jam detection, stall detection, phase reversal, phase unbalance protection with automatic rebalancing, elapsed time meter and kilowatt hour meter.

LOOKING AHEAD Solid state starters can provide smooth, stepless control in starting and stopping pump motors. Diagnostic features can save a great deal of panel space and money on the initial installation by eliminating the need for additional relays. Some devices even have the ability to communicate to programmable logic controllers or personal computers. The solid state motor starter has come a long way from the early reduced voltage starters, and it continues to advance in capability. These newer devices can save time, money, space and down time. With smaller packaging and easy set up their popularity in the pumping industry, particularly in the new specialized units, will continue to grow. These devices may some day replace most existing reduced voltage starters. ■ Jeihri Robinson holds a BSEE from Drexel University. He has 9 years electrical industry experience with Bell Labs, Furnas Electric and Allen-Bradley Company, and he has worked with solid state products for the past 5 years. He is currently a Product Marketing Specialist for solid state power controllers and protective devices at Allen-Bradley.

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Pumps for Haz/Rad Wastewater Service The composition of hazardous/radioactive wastewater streams varies widely, making pump selection and implementation a major challenge. By Peter Y. Burke azardous/radioactive wastewater systems are designed to collect, segregate, store and process potentially hazardous and/or radioactive chemicals contained in water medium. The composition of the waste streams varies depending upon the installation producing the waste, making standardization of pumping equipment challenging. Wide variations often exist in chemical composition and radionuclide concentrations. Table 1 lists three stream classifications which most often result from the basic process of reducing the volume of hazardous and radioactive waste contained in the water medium. This is achieved by employing various processes such as centrifuging, filtration, reverse osmosis, UV oxidation, demineralization, degassification and evaporization. Influent waste pump applications handle waste in a diluted condition with near waterlike properties. Pumps in concentrate waste service must be carefully selected based on corrosion/erosion resistance, temperature limitations and fluid pumping properties. Effluent or discharge service is the least difficult application. Haz /rad waste processes often produce water chemistry conforming to strict federal and/or state requirements, and such water is readily handled by a variety of pump types.

H

HYDRAULIC CONDITIONS Similarly, pressures and flow rates vary as the composition of the streams. Most often the process equipment is configured in a series arrangement requiring wastewater

138

Stream Type

Source

Characteristics

Applicable Pumps

Influent waste

Floor drains Wash water Dewatering Storage tank

Low conductivity High suspended solids Low dissolved solids High micron particles

Centrifugal, horiz. Centrifugal, vert. Air op. diaphragm

Concentrated waste

Filter backwash Evaporator bottoms R.O. concentrate

High conductivity High suspended solids High micron particles High dissolved solids

Centrifugal, horiz. Air op. diaphragm Peristaltic

Effluent

Demineralizers Distillate Oil separator Charcoal filters

Low conductivity Low suspended solids Low micron particles Low dissolved solids

Centrifugal, horiz.

TABLE 1. Haz/Rad wastewater stream types flow on a tank-to-equipment-to-tank basis. This typically necessitates that the pump total dynamic head (TDH) requirements be based on pipe and fitting resistances, valve and eductor pressure drops, and elevation differences. Therefore, the flow and TDH combination is frequently well within the capability of single stage centrifugal pumps—with the exception of reverse osmosis system requirements. For the sake of brevity, this topic will not be addressed here.

as 60 years in conjunction with proposed second generation nuclear power plants currently on the drawing board. This extended life requirement is passed down to equipment suppliers, who must guarantee a similar longevity for their equipment with replacement of only normal, prestocked wear parts. Equipment may be evaluated on the basis of total cost to install, operate and maintain over the entire plant life.

DUTY/LIFE

MAINTENANCE

Pumps in haz/rad wastewater service are generally operated continuously or intermittently 24 hours per day and 5–7 days per week. Processes are often shut down several times per year for maintenance, but pumps are usually installed with a 100% capacity standby spare. Current trends are moving toward facility design lives of 30 to as much

In addition to the ease of maintenance requirements typical of most industrial installations, haz/rad wastewater service requires that maintenance workers be exposed to radiation levels “As Low As Reasonably Achievable” or ALARA. In other words, the equipment and the installation should be designed to minimize residual waste in the

The Pump Handbook Series

equipment after isolation and drainage and to minimize the time required to perform the required maintenance tasks.

EFFICIENCY Currently, efficiency is not a salient parameter in pump selection in haz/rad service. However, evaluation of the total cost to install, operate and maintain the pump over the life of the facility is becoming increasingly important. Should this trend continue, the first cost of the pump and driver will be outweighed in favor of energy cost savings evaluated over the project life. Contributing to the reduction in operational costs will be the use of more efficient pumps, variable speed motor drives and high efficiency motors.

tion, which raises the temperature of the barrier fluid in the seal chamber, thereby decreasing its density. The lighter liquid has a tendency to rise to the top of the supply tank and be replaced by cooler buffer fluid from the bottom of the tank. The primary source of cooling is heat dissipation from the pump shaft and convection from the reservoir. If further cooling is required, cooling coils within the buffer tank can be incorporated.

ed and thus conserved. When an external source of pressurized seal water is unavailable, a thermal convection cooling system (Figure 2) can be installed, provided a pressure source is available. This system is referred to as an API 53 or ANSI 7353 plan. A supply pressure typically 15–25 psi above the “stuffing box” pressure is required by means of an external gas source. Heat is generated by seal rota-

Pressure Gauge Pressure Source

Vent Valve

SELECTION GUIDELINES Pressure Switch

Reservoir

Block and Bleed Valve

Level Gauge Level Switch

Slopedown 1/4”π

Cooling Coil

Pumping Ring

Drain Valve High Pressure Buffer Fluid

12 to 30 inches Leakage

Here are guidelines for selecting pumps in haz/rad waste service: 1. When considering pumps equipped with mechanical seals, make design provisions for the seal leakage. Two constraints make seal selection challenging in influent or concentrate service. First, external, outboard leakage from the pump seals must be minimized and contained. Mechanical seals should be selected with the objective of limiting leakage to occurrences during startup, transient and shutdown conditions. Pump bases should include provisions to collect and transfer leakage to a central equipment drain tank for reprocessing. Secondly, seal flush flow should be minimized to avoid the addition of secondary waste into the process flow, which necessitates the enlarging of process equipment to treat more waste—a counterproductive approach. A preferred seal configuration is back-to-back double seals with an external, barrier fluid seal system (Figure 1). The advantages of this arrangement are: • Barrier fluid leakage into the process is minimized. • The process seal is protected from abrasive particles by the barrier fluid. • The overboard leakage is limited to the barrier fluid, not process fluid. • Barrier fluid can be recirculat-

Pump Motor Pumpage

Drain

FIGURE 1. Back-to-back double seal configuration with an external barrier fluid seal system The Pump Handbook Series

139

Purge Port (Optional) Mechanical Seal Vent Nut

Mating Ring

O-Ring Spacer

Pin

FIGURE 2. Canned motor pump with recirculation loop separated from the process fluid by a mechanical seal The pump/seal manufacturer should be consulted with regard to quality and quantity of seal water, recommended instrumentation and cooling requirements. 2. Carefully evaluate fluid properties when considering sealless pumps. At first, sealless pumps—that is, canned motor and magnetic drive types—would seem like the logical way to eliminate the need for seal water. However, the user must realize that if the waste is either influent or concentrate, neither pump type may provide satisfactory service. Consider the effects of the following fluid characteristics: Operating temperature Abrasive particles Particle size Ferritic particles Where process temperatures are known to exceed pump bearing limits or if crystallization within the pump is likely to occur, both canned motor and magnetic drive pumps may be configured with “recirculation cooling circuits.”

140

In applications where waste contains abrasive particles with a known history of causing accelerated internal pump wear, abrasion resistant materials such as silicon carbide are offered by pump manufacturers. Again, an internal recirculation loop can be provided in which pumpage compatible fluid acts as a buffer to prevent the pump’s critical internal components from exposure to the process fluid. Some canned motor pump manufacturers offer a recirculation loop separated from the process fluid by a mechanical seal (Figure 2). This results in an acceptable amount of process dilution as only a minute amount of recirculation fluid leaks from the motor bearing area into the process. Particles in excess of 100 micron (.004”) diameter can lead to premature bearing failures in sealless pumps. Mechanical filtration should be provided upstream of the pump at a level consistent with the pump manufacturer’s recommendation. The recirculation loop described is THE PUMP HANDBOOK SERIES

also an effective configuration to prolong bearing life. Magnetic drive sealless pumps must be avoided in waste water service where ferritic particles are present. These particles, if allowed to enter the magnet portion of the drive, will affix themselves to the magnets and cause damage to the primary containment. Be certain the ferritic content is understood. Slurries encountered in concentrate wastewater service may vary from 1–10% suspended solids by weight. This is often achieved by utilizing a “preconcentration” mode to raise the percent solids from the existing influent level to a concentrate level appropriate for an evaporation/solidification subsystem. The maximum solids percentage by weight must be communicated to the pump manufacturers in order to understand what special provisions will be included in their offering to accommodate this fluid property— i.e., open impellers, recirculation loops or high current alarms. This data should be included on the data sheet contained in the pump specification. 3. Consider the advantages of an air-operated diaphragm and peristaltic pumps in abrasive, slurry concentrate service. While centrifugals may be the pump of choice in haz/rad wastewater service, handling slurries with large particles is often best accomplished with flexible member, positive displacement types: air-operated diaphragm or peristaltic (a.k.a. hose) pumps. Often they are referred to as “sealless” pumps, which tends to be a misnomer. Both types of construction use a flexible member functioning as a piston that also seals the pumpage from atmosphere. Therefore, unlike the canned or magnetic drive pump, they rely on a flexible seal for containment. The flexible seal has a predictable, finite life, and provisions must be made accordingly. The advantages of using these pumps are their ability to handle: Abrasives High solids content Large particles Dry running Aggressive, corrosive fluids Operation at reduced capacity

The peristaltic pump is particularly advantageous in concentrate service where accurate, variable flow control is required for evaporation/ solidification service. Variable frequency or speed control systems provide a reliable, inexpensive means of controlling the waste processing subsystem, thus avoiding a potential problem of plugging a control valve. 4. Install leak detection monitoring with alarms, shutdowns or switchovers for flexible/diaphragm type pumps in continuous duty applications. As previously stated, provisions for flexible member pump failures should be incorporated in haz/rad wastewater service where pumps operate in continuous-duty service. In addition to providing a low discharge pressure alarm or shutdown, a variety of leak detection systems are available. For example, one system offered for air-operated diaphragm pumps includes a double diaphragm with a graphite sensor inner layer. In the event of a process-side diaphragm leak, conductivity is established within the graphite pad, and an amplified signal is used to indicate a failure condition. Leak detectors for peristaltic pumps typically consist of a float switch to sense the presence of pumpage within the hose lubrication chamber. In addition to leak detection provisions, regularly scheduled maintenance will minimize the possibility of unexpected pumping element failures. However, as an additional precaution with air-operated diaphragm pumps, the exhaust air should be vented away from the equipment to prevent waste fluid from being sprayed on workers in the event of a diaphragm failure. This is preferable to an off-gas ventilation system 5. Use oil-free pumps to eliminate additional secondary waste. The addition of treatment chemicals, oil or even water in a waste treatment process is often counterproductive. Consequently, air-operated diaphragm pumps that require oil lubrication of the motive air should be avoided. Most air-operated diaphragm pump manufacturers currently offer nonstalling, nonlubricated air valves that can tolerate normal plant service air without the added requirement of an oil lubricator. Similarly, rotating pumps should

include sealed-for-life bearings or slinger rings and oil collection devices that will contain oil leakage rather than permit it to leak to the waste stream, equipment base or drainage system. 6. Evaluate the effects of pumping action on downstream process equipment. Often overlooked is the effect of pumping action on downstream process equipment. For example, centrifugal pumps produce a high frequency, low amplitude pressure pulsation that is preferred in most processes. However, because the pumping action has an homogenizing characteristic, it would be undesirable to install a centrifugal pump upstream of an oily water separator or coalescing filter, yet this pump type would be highly desirable in concentrate service to provide fluid (eductor) tank mixing prior to evaporator feed. Reciprocating piston or diaphragm pumps produce low frequency, high amplitude pressure pulsations that should be avoided upstream of critical membrane services. Pulsation dampeners may be beneficial depending on the distance to the downstream equipment. However, they should be evaluated for potential plugging and ease of maintenance. 7. Evaluate proposed haz/rad pumps from a maintenance perspective. Maintenance features should support the ALARA principle. Pump casings should include high point casing vents and low point casing drains where practical to facilitate complete pump case drainage, thereby minimizing worker exposure. Spacer couplings and back pull out centrifugal pump designs eliminate broken pipe connections and prevent pump/motor alignment from being disturbed. Pump baseplates should include provisions for collecting seal overboard leakage, which can be routed to the drainage system. Flanged pump connections are preferred over threaded joints for ease of pump removal and replacement and for reduced particulate buildup. Most pump manufacturers provide flanged pump cases as standard although often small case pumps (below 1”) may require nonproduction fabrication. Cavities in the pump casing, where crud can accumulate, should be avoided. THE PUMP HANDBOOK SERIES

Stainless steel internal parts may reduce the tendency for such buildup. Totally enclosed fan cooled (TEFC) pump motor enclosures facilitate wash-down and thereby reduce operating personnel exposure during maintenance handling. 8. For vertical sump applications, select pumps with seals and bearings elevated away from the liquid level. Vertical pumps are frequently specified for influent service where wide variations in fluid properties are anticipated. Suction screens may be required to prevent debris from damaging the pump. Consult with the pump manufacturer for maximum permissible solid size. Special recessed impeller vertical pumps are available for this difficult application. Some feature one piece, nonclogging casings, open impellers, replaceable wear rings and—importantly—bearings and seals that are elevated above the pumpage. This is achieved by designing the pump shaft to operate below the first critical speed. These “stiff shaft” designs allow the bearings to be located above the sump deck. A generous clearance bushing located immediately behind the impeller prevents excess backflow of the pumpage upward toward the bearings. This pump configuration is typically limited to sump depths of approximately 10 ft. For deeper sumps, cutless rubber bearings are utilized for shaft support and are often supplied with clean flush water to ensure reliability. 9. Insist on hydrostatic and performance testing in accordance with a nationally recognized pump code or standard. There are no industry standards or specifications written exclusively for pumps required in haz/rad wastewater service per se. Standards such as ANSI 55.1 and 55.6 “Solid Radioactive Waste Processing System for Light-Water-Cooled Reactor Plants” address only a few of the key requirements. Therefore, the pump manufacturer’s standards coupled with the purchaser’s requirements are typically the basis for procurement. However, very importantly, hydrostatic testing, performance testing and inspections should be conducted in accordance with the latest revision of the Hydraulic Institute Standards or their equivalent.

141

Test results should be presented specifically for each pump, dated and certified according to pump serial number. The results should be reviewed by the purchaser, approved and filed with quality assurance records for the project. Witnessing is recommended to ensure that tests are conducted prior to shipment and also serve as a final inspection for dimensions, overall appearance and tagging information. The haz/rad waste treatment

142

industry today is a hybrid, consisting of technology adapted primarily from municipal waste treatment and the chemical and power industries. In the absence of unique standards, it is clear that the purchaser and the pump manufacturer must be in tune with the application requirements in order to avoid potential installation and operation problems. The purchaser must communicate the requirements accurately to the supplier. Conversely, the supplier is responsible to provide the product

The Pump Handbook Series

and documentation to satisfy the purchaser’s requirements. ■ Peter Y. Burke is a registered professional engineer with ADTECHS (JGC) Corporation (Herndon, VA), a company specializing in hazardous and radioactive waste treatment process design and construction. Mr. Burke has written numerous articles on pump and compressor system applications, design and maintenance.

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R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

Alignment Tolerances for Spacer Couplings How good is good enough? By Jack Essinger ood alignment of rotating equipment is essential to safe and trouble-free operation. Virtually everyone agrees with that statement. The question, however, is how good is good enough? Alignment tolerances that are too restrictive are hard to achieve and likely to cost more than will ever be realized from improved machine performance. Tolerances that are too loose will result in premature failure of the equipment. Fortunately, good guidelines have been developed and published in recent years. Adherence to these time-tested guidelines will virtually assure satisfactory results. It is noted, however, that applying proper tolerances is but a small part of the total alignment process. Readers unfamiliar with this process will find the material referenced at the end of this article informative.

G

by the offset of the flexible coupling element at Machine A. In this instance, the coupling angularity at Machine B is 20 mils divided by 20 inches, or 1.0 mil per inch. Figure 1(b) represents misalignment of the equipment in the horizontal plane. The misalignment at Machine A is 5 mils divided by 20 inches, or 0.25 mils per inch. The misalignment at B is 16 mils divided by 20 inches, or 0.8 mils per inch. For users who choose to treat vertical and horizontal misalignment as separate entities, the four calculated angularities can be compared with a tolerance specification to determine suitability of the alignment. In this example the largest misalignment is in the vertical plane at Machine B, 1.0 mils per inch. If

the applicable alignment specification allows 1.0 mil per inch or greater, then the alignment would be deemed satisfactory. If the specification requires less than 1.0 mil per inch, alignment improvements would be required. For greater accuracy and more uniform results, vertical and horizontal components can be combined into a single value representing the true misalignment at each end of the coupling. The true offset of the spacer with respect to Machine A is:

√10 mils2 + 5 mils2 = 11.2 mils, or 0.56 mils per inch.

ANGULAR MISALIGNMENT OF SPACER COUPLINGS For machines using spacer-type couplings, misalignment must be judged on the basis of the angularity of the coupling spacer with respect to each of the coupled machines. Figure 1(a) represents misalignment of Machines A and B in the vertical plane. The angularity of the coupling spacer with respect to Machine A is determined by the offset of the flexible coupling element at Machine B. This angularity is most easily expressed in terms of the offset per unit coupling spacer length – i.e., the offset divided by the coupling spacer length (spacer length is the axial length between flexible coupling elements). In this example the offset is 10 mils (1 mil = 0.001 inch). The angularity is 10 mils (offset) divided by 20 inches (spacer length), or 0.5 mil per inch. Similarly, the angularity of the coupling spacer with respect to Machine B is determined

The true offset of the spacer

10 Mils

Machine A

ine

ch Ma

20 Mils

B

20 inches

16 Mils

5 Mils

Mac

hine

Machine A

B

20 inches

Figure 1(b) Horizontal Plane

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143

00 0 0 ,0 00 00 00 10 9, 8,0 7,

30

0

00

6,

0

00

25

5,

20 18 16

4,

0

00

00

6 3,

0

00

3,

14

0 00 0 0 ,8

2,

10 9 8 7

1

6

1,

5

0 1,

0 20 00

4

Speed - RPM

Coupling Span in Inches

12

3

2

1 1

2

3

4

5

6

7

8

9 10

15

20

25

Tolerance (Offset in Mils) Figure 2. Alignment tolerance chart

with respect to Machine B is:

√20 mils2 + 16 mils2 = 25.6 mils, or 1.28 mils per inch. If desired, angularity in mils per inch can be changed to degrees merely by using a factor of 1 mil per inch = 0.0573 degrees. The misalignment at Machine A in this example is (0.56 mils per inch) x (0.0573 degrees per mil per inch) = 0.032 degrees. At Machine B, the angularity is (1.28) x (0.0573) = 0.073 degrees.

ALIGNMENT TOLERANCES For industrial equipment the amount of misalignment that can be tolerated is a function of many variables including rpm, horsepower, coupling type, spacer length, design of the coupled equipment and expectations of the user with regard to service life. Since it is not practical to consider all of these variables in a reasonable and useful alignment specification, considerable simplification is required. Recommended alignment tolerances are found in a number of publications including coupling manu-

144

facturers' catalogs and bulletins, technical publications by societies such as the American Petroleum Institute, and technical articles and publications such as those listed at the end of this article. For general-purpose couplings, tolerances specified by manufacturers usually have more to do with the capability of the coupling itself than with the requirements of the coupled equipment, and they tend to be more liberal than other specifications. For high-performance, engineered couplings, allowable tolerances are usually shown on the coupling drawing. In the 1970s alignment tolerances based on rpm and coupling spacer length were published. Figure 2 is an Alignment Tolerance Chart published by Dynamic Measurement and Control Company. Tolerances shown in it are typical of those published in other sources of that era, and these tolerances have been widely used for many years with generally excellent results. Many of them were based primarily on experience with lubricated, gear-type couplings. Experience has shown, however, that they are equally applicable to the vast majority of non-lubricated couThe Pump Handbook Series

plings. They can, in fact, be used with virtually all couplings that employ flexible elements. They should not be used for solid couplings. As an example of the usage of this chart, consider a 3,600 rpm pump with a 5" spacer. The allowable offset would be 8 mils, and the allowable misalignment would be 8 mils divided by 5", or 1.6 mils per inch (0.092 degrees). For a compressor with a 12" spacer operating at 7,000 rpm, the allowable offset would be 10 mils. The allowable misalignment would be 10 mils divided by 12", or 0.83 mils per inch (0.048 degrees). Although not specified in the chart, it is generally accepted that the misalignment tolerances are to be applied to total coupling offset, not to vertical and horizontal offsets individually. For the sake of simplicity, there is a trend toward using a single allowable offset for all equipment regardless of operating speed. The most often-quoted figure is 0.5 mils per inch of coupling length. The new API RP-686 / PIP REIE-686 Recommended Practices for Machinery Installation uses this tolerance. Most specifications, unfortunately, are unclear as to whether the 0.5 mils per inch applies individually to the vertical and horizontal components of misalignment, or to the resultant of the two components. If it is applied individually, then 0.5 mils per inch in each plane translates to a maximum allowable offset of 0.71 mils per inch. In either case, the tendency these days is to use tolerances that are more restrictive than those in use for the past several years, especially for equipment operating at 3,600 rpm or less. While there is nothing wrong with aligning equipment to more stringent tolerances, there is a great deal of experience to suggest that the less restrictive tolerances established in the 1970s are entirely satisfactory for a very high percentage of industrial equipment.

PRECAUTIONS IN APPLYING TOLERANCES While any of the tolerances specified above appear to be fairly straightforward, the user should be mindful of a number of pitfalls that

Figure 3(a) 5" Indicator Separation Machine A

Machine B

5"

Figure 3(b) 10" Indicator Separation Machine A

Machine B

10"

Figure 3. Typical reverse-indicator arrangements

may result in alignment that is not as good as the tolerances imply: Failure to Account for Thermal Movement Failure to account for thermal movement of equipment under operating conditions, or the use of inaccurate thermal growth data, are common causes of alignment problems. A conscious effort should be made to account for thermal growth of equipment operating at other than ambient temperatures. Even moderate temperature changes can result in significant changes in alignment. If the equipment is large, high speed or vital to plant operation, or if it operates at temperatures far from ambient, then the thermal growth should be measured and accounted for in the alignment process. Ambiguous Specifications Regardless of the specification being

applied, it is essential that it be clearly defined. If an allowable tolerance of 0.5 mils per inch of coupling length is used, for example, it must be clear as to whether the tolerance applies per plane, or to total coupling offset. Specifications That Do Not Consider Spacer Length The length of the coupling spacer is fundamental to meaningful alignment specifications. Specifications that do not recognize spacer length will result in inconsistencies and should be reviewed carefully by the user to assure that they yield acceptable results. When using dial indicators, the currently-preferred method for obtaining alignment data is the reverse-indicator arrangement (Figure 3). In this approach, coupling offset can be approximated directly (1/2 of the largest total indicator reading) if the dial indicators read directly on the couThe Pump Handbook Series

pling flanges as shown in Figure 3(a). The maximum coupling offset is 6 mils, and the misalignment is 6 mils divided by 5", or 1.2 mils per inch. Note that there is a slight error in this calculation because the flexible elements are not directly at the dial indicators, but the error is too small, to be of concern. If the indicator separation is greater than the distance between flanges, however, as is often the case, then this simple calculation is no longer valid. Referring to Figure 3(b), shaft alignment remains unchanged, but the indicators are separated by 10" instead of 5". If the specification allows an alignment calculation based upon dial indicator reading divided by indicator separation – as does API RP-686 / PIP REIE-686, for example – then the apparent misalignment would be 8 mils divided by 10", or 0.8 mils per inch, which is incorrect. A proper calculation must recognize spacer length as well as dial indicator separation. Similarly, some currently-used specifications are based on a maximum offset of the coupled shafts at some location (typically at the midpoint of the spacer), plus some maximum angularity of the coupled shafts. Since coupling length is not considered, such specifications normally result in progressively larger misalignment as spacer length decreases. Again, the user should review such specifications carefully to assure that they will accomplish the desired results. Laser Alignment Systems Laser alignment systems are currently popular and may offer significant advantages in some applications. The user, however, should have a good understanding of the specific system being used. Interpretation of laser alignment data is normally done automatically by the computer that is an integral part of the alignment system. Interpretation varies from one manufacturer to another, however, as do alignment results. Shortcomings as mentioned above are inherent in some laser systems. It is incumbent upon the user to understand how the laser system interprets data, and to assure that the interpretation meets user requirements.

145

AXIAL SPACING

plings vary greatly depending upon design and manufacture. It is difficult to apply generic alignment tolerances to such couplings. ■

Maintenance of proper axial spacing is important to all couplings, but tolerances are quite dependent upon the specific coupling and cannot be treated in general terms. Gear-type couplings, for example, can normally accommodate fairly large axial tolerances. Diaphragm-type couplings require much closer tolerances. Manufacturers' recommendations should be followed.

Jack Essinger, P.E. is an engineering consultant with Acculign, Inc., Willis, TX. He holds a B.S. degree in Mechanical Engineering from Arizona State University, and worked as a machinery engineer for Shell Oil Company for 29 years.

NON SPACER-TYPE COUPLINGS

REFERENCES

As with axial spacing, alignment tolerances for non spacer-type cou-

146

Dodd, V.R., Total Alignment, Tulsa, Oklahoma: Petroleum Publishing (1980).

The Pump Handbook Series

Essinger, Jack, "Tutorial on Shaft Alignment, "Proceedings of the 12th International Pump Users Symposium," Turbomachinery Laboratory, Texas A&M University, College Station, TX (1995). Jackson, Charles, "Reverse Indicator Method of Alignment," Proceedings of the Second International Pump Symposium, Turbomachinery Laboratory, Texas A&M University, College Station, TX (1985). Piotrowski, John, Shaft Alignment Handbook, New York, New York: Marcel Dekker, Inc. (1986).

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R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

Coupling Alignment— Let's Talk Tolerances By Jack Bolam ou finally got the interview you have been waiting so patiently to receive. It's quite unusual that the interviewer has brought his entire family along, but that's OK. The session is going just fine, and you're feeling very good about it. Just as the interviewer turns to say something to an acquaintance, his little girl walks up to you and kicks you in the shins. You wince with the pain and hold back the urge to get angry. That is tolerance. When I wish to determine the operating temperature of a bearing, I cautiously place my finger on the bearing cap. If I have to pull it away immediately, the bearing is operating above 120° F. If I can continue to touch it for a few seconds before the discomfort sets in, it is at 120°, the highest level of heat that I can tolerate. Any higher than that is immediately painful. High tolerances on components let you put them together easier, but then the final product wears out faster. On the other hand, if the parts have tolerances that are too tight, they are difficult, if not impossible, to assemble. The right tolerances are designed into the various components not only to facilitate assembly but also to insure the reliability of the assembly. As you can see by the examples, tolerance can sometimes be a negative word. I believe that it is such a word when it comes to precision coupling alignment. How tight is too tight? As most of you know, there are two types of misalignment, parallel and angular. Usually a combination of both types is found. Parallel mis-

Y

alignment is usually expressed in thousandths of an inch of "offset." Angular is stated in thousandths of an inch per inch "angularity." Alignment tolerances are normally stated in these same terms. Why are there coupling tolerances? How should they be applied? I try to adhere to the following rule when it comes to using the tolerance charts. A coupling should not be misaligned when the coupled units are operating at design conditions. The only time the coupling tolerance is used is to misalign the units deliberately to allow for normal machine changes that occur when the train is started up. Thermal growth or shrink, the shaft finding its running position in an oil film bearing, gear shaft movement under load – these are all compensated for by deliberate misalignment. So many times we hear that the final alignment is within the coupling tolerances, only to find out after a short time that a bearing fail-

ure has occurred, or that the seals are leaking again. The unit has to be shut down for more repairs. Was the alignment within the stated tolerances? Yes! Was the coupling damaged? Possibly! Remember that the coupling is designed to operate misaligned for short periods of time; seals and bearings are not. Was it a precision final alignment? No! Let's look at an example of a coupling manufacturer's published coupling tolerances and see how close we must be to be within tolerance. The manufacturer of a commonly used coupling lists operating alignment limits in inches. Yes, it does say "operating" alignment limits. For a coupling with a 2 1/4" bore, the offset or parallel limit is listed as .010". The angular limit is also listed as .010". What would this look like on a graph? In this illustration (Figure 1) we are not compensating for any machinery movement. The pump centerline and the motor centerline

Figure 1. Example of published coupling tolerances. The Pump Handbook Series

147

will be in perfect alignment if the indicators show us that the actual motor centerline location is on or "near" the one on the graph. This discussion is all about "how near should we be?" The parallel tolerance allows us to be 10 mils (.010") high or low. Since there are no +'s or –'s on the chart, I assume that the tolerance applies to above or below or to the left or the right of the ideal centerline. With the precision alignment methods being used today – i.e., the reverse indicator method or the laser alignment method – this can easily be corrected to within 1 or 2 mils of the desired centerline. Now let’s look at the angular tolerance. It is also stated as 10 mils (+/–). Since 10 mils is not an angu-

lar measurement, I assume that it means 10 mils per inch. Notice that the 10 mil per inch lines are almost perpendicular to the machine centerlines. To correct a 10 mil per inch pure angular misalignment in the illustration, we would have to add or subtract 350 mils (a little more than 11/32") at the outboard motor feet and 150 mils (a little less than 5/32") at the inboard feet. Carpenters use closer tolerances than this to build steps. As I said earlier, in the real world a pure offset or a pure angular misalignment does not usually happen. It is normally a combination of both. If deliberate misalignment, for those changes that naturally occur, is included with a reasonable tolerance

The Pump Handbook Series

(I do not consider any of the tolerances I have seen so far to fall in this category), I still find it difficult, without graphing, to know if I am "within tolerance." Couplings may be built to operate under the conditions published by the coupling manufacturers, but bearings and seals are not. Early seal failures, bearing housing and bearing wear, coupling wear, and coupling failure are all results of being within tolerance. ■ Jack Bolam is President of VibraCon Predictive Maintenance Consultants, Loveland, Co. He has 36 Years’ experience in rotating machinery and has worked for Dresser-Clark, Eastman Kodak, and Fluor Daniel.

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R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

Machinery Alignment vs. Coupling Misalignment Do we align machines, or do we align couplings? By Michael M. Calistrat lignment occurs when two lines are superimposed and form a single line. Misalignment is an error in alignment. It is a measure of how much two lines are away from forming a single line. Evidently, two lines can be at an angle to each other (as they are in angular misalignment), or can be at a distance from each other (as they are in offset misalignment). When we are talking about machinery, and specifically about two machines being brought into alignment, the lines we refer to are the centerlines of the machine shafts. We can move one or both machines in the attempt to bring the centerlines "in line," and this process is called machinery alignment. While machinery alignment is easy to understand, an explanation of coupling misalignment requires a simple sketch (Figure 1). Note that in this figure the couplings are not shown; rather, only the two shafts (which are offset from one another), and the coupling’s spacer are shown. Coupling misalignment is the angle that the spacer’s centerline forms with the centerline of either shaft. Therefore, each "half-coupling" can (and usually does) have a different misalignment. What is important to understand is that coupling misalignment is measured in degrees, not in inches! Because of this, we should never talk about "coupling alignment." That is a misnomer. Couplings were invented and couplings are still used for accommodating the misalignment between two shafts. What many people do not realize is that all couplings resist being misaligned, and the restoring

A

forces and moments involved in this resistance can damage bearings, seals, and even shafts. Most (but not all) coupling manufacturers include in their catalogs data about how much misalignment their couplings can stand – in other words, the couplings can be misaligned and still have a satisfactory life. "Satisfactory" is a relative term, of course. I have talked to users who feel that they got their money's worth if a $30,000 coupling lasted nine months! Maybe because of this, the misalignment capability of a coupling becomes a marketing ploy. You may have talked to a coupling salesman who says: "Buy my coupling because it can accommodate 50% more misalignment than the one you are using now." Why shouldn’t we utilize a coupling's ability to withstand misalignment? The answer is that the restoring forces and moments are proportional to misalignment: the larger the misalignment, the larger the forces on bearings and other moving parts. It results that if you want a long life from your machines, you should align them perfectly. True? Well, it depends.

There are only two types of couplings in this world: those that accommodate misalignment through the sliding of an element over another (such as gear-type couplings), and those that accommodate misalignment through the flexing of one or more components (such as the diskpack coupling). Evidently, whenever we have relative sliding, we better lubricate the surfaces! A coupling that accommodates misalignment through flexing works better and lasts longer than a sliding one at small misalignments. It will impose no restoring forces and last forever at zero misalignment. Therefore, your efforts to reach "perfect" alignment will be amply rewarded. A coupling that accommodates misalignment through sliding of one surface over another needs some sliding to keep the surfaces lubricated. Without any motion the lubricant is expelled from spaces between the surfaces, and the coupling wears out quickly because of "brinelling" and fretting. Although it may be an anathema to many, lubricated couplings last longer if they are subjected to some misalignment. Let’s summarize. 1. NEVER rely on catalog mis-

Coupling misalignment

Shaft Offset

Figure 1. The Pump Handbook Series

149

0

150

00

25

5,

20 18 16

0 00 4, 0 0 6 3, 0 00 3,

14 12

0

00

2,

10 9 8 7

0 80

0

20

1,

0

00

1,

5 4

Speed - RPM

1,

6

3

HOW WELL SHOULD WE ALIGN MACHINES? How much misalignment is acceptable? Individuals and organizations have issued guides, which are often called "alignment tolerances." One such guideline, in graph form, is shown in Figure 2. It was issued by Dynamic Measurement and Control Company, and reproduced in an article published recently by Pumps and Systems. This graph is to be commended because it includes speed, which is too often omitted in determining alignment. However, the graph is misleading because it lacks "coupling diameter." The disk pack of a misaligned coupling is shown in Figure 3. When the coupling rotates, the disks flex back and forth, once per revolution. The fatigue induced in the disks is a function of two variables: the amount of flexing and the number of cycles accumulated. The amount of flexing, in turn, is a function of the angular misalignment at the half coupling (there are two disk packs per coupling), and of the coupling diameter! The larger the coupling, the larger the flexing. Therefore, the larger the coupling, the smaller the angular misalignment should be; otherwise, the larger the coupling, the shorter its useful life. The reader might expect now that I will give a new graph, which will be perfect. Sorry to disappoint you; I believe that all "universal" guidelines can be misleading. The acceptable error (tolerance) in aligning machines

0 00 0 0 0 ,0 ,00 ,00 ,00 6,00 10 9 8 7

30

Coupling Span in Inches

alignment data – neither for aligning your machines, nor for help in deciding on couplings to buy. However, if you know that the machines will be misaligned beyond what a certain coupling can accommodate, forget that coupling. 2. Align the shafts of your machines as well as possible, but trying to have zero misalignment is seldom cost effective, and in the case of lubricated couplings it is actually counterproductive! 3. In bringing machine shafts into alignment, we line up the machines; let’s stop using the term "coupling alignment." Actually, most people align their machines before the coupling is even installed!

2

1 1

2

3

4

5

6

7

8 9 10

15

20

25

Tolerance (Offset in Mils) Figure 2. Alignment tolerance chart

depends on so many factors that a rule of thumb cannot be given. But most importantly, machines must be intentionally misaligned, because the intent is to have good alignment when the machines are at operating temperatures!. If one uses the graph of Figure 2, or any other guideline, to align his machines, when they are cold, he is in for an unpleasant surprise. Here is my conclusion: Cold alignment tolerances should be indi-

disk pack Figure 3. The Pump Handbook Series

vidually determined by engineering or maintenance management for each machine; universal tolerances are for the birds! ■ Michael M. Calistrat is an international expert on flexible couplings and the owner of Michael Calistrat and Associates in Houston, Texas. His firm specializes in design and failure analysis of rotating machinery. He has obtained 17 patents and is author of Flexible Couplings Their Design, Selection and Use. For more information, write Caroline Publishing, P.O. Box 451611, Houston, TX 77245-1611, fax (713)437-4656.

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R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

Build a Better Foundation to Reduce Costly Downtime By Richard D. Myers entrifugal pumps usually lead the list of failureprone equipment in refineries and major chemical/petro-chemical plants, where downtime expenses average $250,000 per hour. Pumps are also a major contributor to emergency shutdowns and production delays in pulp and paper mills, in which downtime expenses average $100,000 per hour. Pumps commonly have precisely machined mounting surfaces on their bedplates. At installation, a pump bedplate is connected to a concrete foundation with an epoxy or cementitious grout. Ideally, vibrations that occur when a pump motor kicks in or shuts down, or when pumping conditions vary from shutoff to full flow, are transmitted through the bedplate to the foundation and down through the subsoil. Properly installed, a structural epoxy grout will bond to a concrete foundation and to a pump bedplate with tensile and shear strengths that are much higher than concrete, effectively transforming a bonded baseplate into a monolithic structure that includes the concrete foundation. The single-block monolith dampens vibrations produced by the pump. Effective damping (frequency attenuation) reduces pump-shaft vibrations caused by resonance. It also helps to absorb shock loads. The end results are a significant increase in the mean time between failures (MTBF), reduced maintenance expenses and longer life for mechanical seals and bearings.

C

amplify vibrations developed at resonant frequencies When a cementitious grout is used, voids frequently occur because that grout cannot provide a perfect nonshrink bond to a steel baseplate. Cement mix water will then bleed into the voids, accelerating rust and metal deterioration and thus increasing the amplification of the wiggles, wobbles and shakes that cause damage and downtime. Voids can also occur if a structural epoxy grout was installed without sufficient vent holes in the baseplate, allowing air to be trapped and thus producing "bubbles" in the grout. Other conditions that cause voids include improper baseplate preparation (e.g., dirty, greasy, not sandblasted),use of a primer that will not bond to the metal baseplate and/or to an epoxy grout, insufficient static head pressure when a liquid epoxy grout is poured, and any baseplate movement during grouting or during the 24-hr epoxy cure time. Since cementitious grouts require a 20-day cure, the cost of downtime can run into hundreds of thousands of dollars in small plants and into many millions of dollars in major facilities if epoxy grouts aren't used to install new pumps or to retrofit or repair existing installations.

FILL VOIDS TO KEEP PUMPS ON LINE

WHEN IN DOUBT, DON'T DRILL OR FILL

While the best prevention is

If you are not certain there is a

Vertical Plane before grout after

"SOFT FOOT" AMPLIFIES DESTRUCTIVE VIBRATIONS Grouting a pump baseplate to a foundation requires careful attention to detail. Voids between a baseplate and its foundation will cause a soft-foot condition, which will

careful, professional epoxy grouting, voids in a foundation system can be filled after the fact by pressure injecting a liquid epoxy grout into each void. By filling all voids, you can provide intimate contact between a pump baseplate and its foundation, turning a poor grouting job into a successful retrofit that should prevent problems until the next scheduled shutdown. The first step in filling voids is to locate them. It's a simple process. All you need is a small ball-peen hammer and a marking pen. A ballpeen hammer will produce a better ring than a large hammer, large bolt or other striking tool. Tap the hammer on the baseplate and listen. If you strike above a void, you will hear a ringing sound instead of a dull thud. By tapping out from any spot where you hear a ring, you can easily determine the extent of the void. Mark that area. (This sounding procedure should not be conducted until an epoxy grout has cured for at least 24 hours.) The next step is to drill injection and vent holes through the baseplate. A small void may require only one injection hole and a single vent hole. Large voids will require several vent holes around the perimeter of each void and an injection hole in the center. Injection holes should not be more than 12-14" apart. You may need to sketch several layouts to pinpoint one that will work best.

Horizontal Plane before grout after

Vibration Velocity, mm/second Pump #1 Pump #2

35.610 4.983

1.939 0.443

5.417 5.647

2.264 0.477

Table 1. Effect of grouting on pump vibration The Pump Handbook Series

151

void, do not drill that area. An indefinite ringing sound, something between a dull thud and a clear ring, may result from cross-bracing under the bedplate or from vertical gusset reinforcements near motor or driven mounts. Drilling and pumping epoxy into such areas may, very likely, disbond the baseplate and cause additional voids. If access is available from above the baseplate, injection and vent holes should be drilled vertically. Otherwise, the holes can be drilled at any angle, even horizontally if the configuration of the pump baseplate and drive machinery permits. After drilling the holes, determine the depth of each void by inserting a stiff wire into the hole and measuring the depth of penetration. If the voids are less than 1/4" deep, you can use a two-part (resin and hardener) epoxy grout. To inject the mixed liquid epoxy, you will need to drill and tap every injection hole for a either a 1/8" or 1/4" pipe fitting. (Zerk fittings should be substituted if grease guns are used to pour liquid epoxy into the voids.) Tap each vent hole, too, because it may be used as an injection hole in the final stages of epoxy placement. Deeper voids require a threepart epoxy (resin, hardener and filler/aggregate). Voids deeper than 1" also require an epoxy grout formulated for high-flow and they necessitate larger injection holes.

INJECT EPOXY TO FILL VOIDS The liquid epoxy is injected from a gun into fittings screwed into the injection holes. If the job requires fewer than 25 injection holes and less than two gallons of epoxy, you can use a hand-held grease gun. The advantage of a grease gun is it's cheap. If the injection process is delayed and the epoxy hardens in the gun, simply throw it away. For larger jobs, use a pressure pot and hoses. If compressed air is available, up to two hoses can be used on a pressure pot. If you use a pressure pot, be sure the pressure under the baseplate does not exceed 25-30 psi. Greater pressures may deform or delaminate a securely bonded monolith. For even larger

152

ESCOWELD 7550E

ESCOWELD 7560E

Compressive Strength (ASTM C-109)

12,500 psi

12,500 psi

Compressive Modulus of Elasticity (ASTM C-109)

680,000 psi

610,000 psi

Flexural Strength (ASTM D-790)

7,500 psi

5,600 psi

Tensile Strength (ASTM D-638)

3,100 psi

3,100 psi

Adhesive Strength to Concrete

better than the strength of concrete

better than the strength of concrete

Linear Shrinkage (ASTM D-2566)

less than 0.0002 in/in

less than 0.0002 in/in

Cure Time @ 75°F

18 hours

24 hours

Pot Life @ 75°F

20-25 minutes

40-45 minutes

Density

1.65

1.65

Viscosity, Centipoise @ 72°F

5,000

6,348

Typical Depth of Pour*

1/16" to 1/2"

1/2" to 2"

*Depth of pour is the most important selection criterion. Other high-performance ESCOWELD epoxy grouts are formulated for single-layer pours up to 18-20". With expert supervision, crack-free single pours up to 4 ft have been executed. All high-performance epoxy grouts are highly resistant to most oils, lubricants and many chemicals, and they will not support combustion.

repair jobs, you may need an airoperated drum pump, similar to

Because the epoxy grout should be injected into large voids through

Table 2. Low viscosity epoxy grouts for thin pours

those found in service-station grease racks. If you use such a pump, install a pressure regulator on the air side to prevent over-pressurizing the grout and distorting the baseplate. The pump should have no greater than a 20:1 ratio, and it must fit into a five-gallon bucket. If you use anything other than a throw-away grease gun, you will need to clean the equipment frequently to keep the epoxy from setting up inside the equipment and ruining it. Therefore, be sure your equipment is compatible with cleanup solvents.

A LABOR-INTENSIVE OPERATION

Be sure you have enough people and equipment to do the job in one continuous process instead of stretching it out over a long time.

The Pump Handbook Series

several holes simultaneously, multiple guns will be needed. Use an accurate dial indicator to measure any movement at pump couplings. There will be very little movement. A dial indicator should also be mounted within 6" of the point of injection to make certain pressure is not developing under the baseplate. If the dial indicator moves a maximum of 3 mils, stop injection of the liquid epoxy and allow the pressure to dissipate. Remember, this procedure is not intended to develop pressure under the baseplate. The only reason to use a grease gun or other "pumping system" is to place the liquid grout in each void. Measure the resin and hardener accurately. Improper proportions will prevent proper curing. Mix the material in small batches (e.g., only

as much as you will need for 30 minutes, and never more than one gallon). Structural epoxies set up quickly, so don't mix too large a batch. Remember, too, that hand-held grease guns don't hold much epoxy. Before using a grease gun, remove and discard the end cap and spring plunger. During the application process, hold the grease gun vertically while a helper pours epoxy in through the open top. Your helper must be sure to maintain a continuous liquid flow into the gun so that no air is entrapped with the epoxy and injected below the baseplate. During the injection process there should be no external pressure on the baseplate. Do not stand or lean on the baseplate.

LITTLE PRESSURE, LOTS OF PATIENCE This is a time for minimum pressure and maximum patience. Begin injecting the epoxy at one of the central injection holes and continue until material comes out of the open vent hole. If the pump base slopes, as with an API pump base that tilts

down from the driven end, begin at the low end and work upward. If you are pumping the liquid epoxy and you feel pressure, as you will if you're filling a void only a few thousandths of an inch deep, stop pumping periodically to let the pressure within the void subside. Never force epoxy into a void. Remember, high pressure can distort a baseplate. Instead of forcing the epoxy in, use a start/stop action until the void is full and epoxy begins to flow from vent holes. For large voids be sure to move from hole to hole smoothly. The goal is to fill each void sufficiently from each peripheral point, while monitoring overflow from the vent holes. Once a void is filled, leave the fittings undisturbed and fill all open holes. If a slow leak develops into an adjacent void or into an underlying foundation crack, you can pump in additional epoxy. If leakage continues, let the grout set up; then attempt a second injection either through the original hole or through a new one. This procedure is not as compli-

The Pump Handbook Series

cated as it sounds. By injecting liquid epoxy, many users have repaired loose bases of pumps and other equipment satisfactorily. Pressure injection, however, is not a substitute for proper grouting. ■ Richard D. Myers is the Escoweld Marketing Manager for ITW Philadelphia Resins. He has more than 32 years of experience in the industrial coatings and grouts industry.

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R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

Flow Sensing For Pump Protection Precise flow monitoring is a primary line of defense against pump and process malfunction. By Greg Ochs

P

Pressure. Measuring the intake and discharge pressure with pressure switches is a method based on normal pumping operation. A low pressure alarm on the intake can indicate a loss of fluid while a high pressure alarm on the discharge may indicate a blockage. Pressure detection can be difficult to implement and unreliable if process conditions or the media are not consistent. More important, pressure measurements do not necessarily indicate the presence of flow. Current. Measuring the pump motor current determines the flow rate based on the normal pump operating parameters. The pump with no fluid will have a low current load with no force on the impeller. Proper pump operation with flowing media will result in a normal current load. A blocked pump will exhibit a high current load. This system can be difficult to implement, however, if the media viscosity changes. Pressure and current are not directly related to actual flow conditions during the pump's entire life. Measurements of motor current are based on conditions at the time the protection scheme is implemented. Over the life of the pump and in changing service conditions, wear and tear will change pump characteristics, and this can result in false alarms. Protecting costly pumps with this method may not ensure PHOTO COURTESY OF FLUID COMPONENTS INTL.

rocess sensitive applications and costly equipment require pump protection to ensure product quality and to minimize maintenance or re-placement costs. There are various pump protection techniques, and it is important to weigh performance requirements against the cost of individual protection schemes. Balancing the appropriate level of pump protection versus capital investment depends upon the type of pump used and how critical the application is. Low cost pumps or noncritical applications may or may not require a protection scheme. More costly pumps and critical installations demand reliable protection. Implementing an appropriate protection plan at the time of pump installation avoids serious problems including cavitation, water hammer, vibration, leakage, emissions and more. In designing pump protection, one must first examine the type of pump and the characteristics of the media or substances that will be pumped. Is the pump a positive displacement pulsing type or a progressive cavity type? Is the media a refined low viscosity liquid or a coarse, dense raw material? Answers to these questions will help in selecting the best pump protection scheme from the techniques discussed here.

TECHNIQUES There are several ways to determine pump flow. Three of the most common technologies used are pressure, current and flow. Each has advantages and disadvantages and can be an appropriate choice depending on the type of pump, the media involved and the nature of the application.

154

FLT93-S high temperature flow switch The Pump Handbook Series

reliable pump shutdown in a timely, orderly manner. Flow. Pumps are designed to move substances. Reliable flow detection is best obtained by monitoring actual flow in the operating environment. Monitoring the actual flow represents a direct relationship to the performance of the pump. There are both intrusive and non-intrusive devices for monitoring flows. Non-intrusive technologies offer simple installations that clamp onto a pipe rather than penetrate it. They are advantageous if the media is lethal or corrosive. However, some substances are not conductive or change conductivity. This causes unreliable flow detection. Intrusive technologies offer direct contact with the media. This produces responses relating directly to flow conditions. These technologies can be separated into two types: mechanical and solid state. Mechanical types include paddle, target, impeller and others. Solid state types include thermal, ultrasonic and vortex. Both mechanical and solid state options offer relatively simple installations due to insertion configurations with retractable glands. Mechanical technologies offer a passive approach to flow monitoring. With no power requirement, a contact closure is provided to indicate a change of state from flow to no-flow. The drawback with mechanical techniques is that they are mechanical. The mechanical movement is subject to wear and breakage. The life expectancy of pivots and springs depends on switch cycles. The overall assembly can be a concern. The seal around the switch movement can offer a leakage path from the process to the environment. And the joints and seals are often affected by corrosion or leakage in high pressure applica-

stands out is thermal dispersion sensing. Thermal dispersion flow switches, flow meters and liquid level /interface controllers offer the greatest advantage in solid state designs. These devices can be referred to as a fit-and-forget technology. The benefits they offer include: • Ultra low flow sensitivity • High turndown Figure 1. The sensor portion that is exposed to the process consists of two all-welded stainless steel thermowells housing two temperature sensing devices.

tions. Mechanical measurement is also viscosity dependent. High viscosity and debris in the media can cause fouling or breakage of the mechanical apparatus. In a high temperature process, the mechanics can lock. Breakage and fouling can lead operators to think there is no flow loss since there is no reliable indication or alarm for a failed device. Mechanical devices also have a defined life cycle for the movement. In pulsing or dosing pumps, the switch indicates a flow, then a no-flow and so on, at each stroke. This pulsing flow can lead to short installed life for the switch due to the limited life cycle of such mechanical devices. This repetitive cycle provides a pulsing contact that should be interpreted by a secondary device or distributed control system (DCS) to determine a fault condition. Solid state technology offers an active method of flow monitoring. The fact that solid state devices require power can be a drawback, but their benefits far outweigh this issue. Solid state, no-moving parts technology offers the highest level of reliability for flow monitoring. With no joints, pivots or orifices there is nothing to foul, clog or jam. This inherently provides for minimum maintenance and long operating life. Solid state design also allows the use of rugged, all-metal seamless sensor head construction. The solid state technology that

• Wide process temperature operation • All metal construction (Hastelloy, Inconel, etc.) • Suitability for hazardous locations, T4 temperature rating These benefits are independent of viscosity and temperature changes within the process. The thermal dispersion design is a reliable hybrid technology incorporating basic thermodynamics and electronic principles. The sensor portion that is exposed to the process consists of two all-welded stainless steel thermowells housing two temperature sensing devices (Figure 1). A low power constant wattage heater is associated with one of the temperature sensing devices. This element is called the active element; the other element is the reference element. This configuration creates a temperature differential between the two temperature sensing devices. The differential is greatest in a no-flow condition and decreases as the flowing media passes over the probe and cools the heated element. To determine an alarm condition, the differential is converted to an electronic signal and compared to an adjustable setpoint. This technique operates independently of process temperature. The reference sensor is always at the process temperature; the active sensor is at the process temperature plus the added heat of the heater. This two-sensor design ensures a distinct temperature differential on the assembly indepenThe Pump Handbook Series

dent of the process. The temperature differential is proportional to the flow rate. The sensitivity of the devices allows for a setpoint rangeability from 0.01-3.0 fps. Flow rates outside of range will not damage or affect flow performance. The design of the sensor assembly itself ensures high reliability. With only three components in the all-welded metal sensor head, failures due to vibration, temperature, pressure or high viscosity are extremely unlikely. In conditions other than loss of flow, cavitation also can be detected by thermal dispersion. When a pump begins to cavitate, the ability to move media decreases. This decrease produces a change of flow that is detected by the thermal flow switch. Detection is due to the rangeability and sensitivity of the signal generated by the thermal differential. Thermal dispersion offers high reliability in most process applications. In pulsing or dosing pumps, the pulsing flow is seen by the thermal switch as an average flow rate. The pulsing flow cools the heated element at a certain rate. Use of the adjustable setpoint allows the switch to ignore the pulsing, but detect the loss of flow. The no-moving parts design offers high service life because there is nothing to wear or break. The wet/dry condition of the switch offers added benefits for pump protection. The temperature differential in dry conditions is higher than in wet conditions, and the temperature differential for a set no-flow signal is higher than a wet flowing signal. The use of dual switch points allows the pump operator to determine if the pump is primed with liquid before it is started. The operator can then detect if the pump is operating. Wet/dry detection helps identify changes of liquid to gas and steam void conditions in condensate systems. The electronics associated with thermal devices can be located at the sensor or in a remote location and safe area. Electronics independent of the sensor make field calibration possible

155

trol circuitry and dual relays that can be configured to alarm as independent SPDT relays to sense both flow rate and temperature. FCI's FLT93 switch is especially effective in pump protection applications that involve monitoring product flow, in addition to pump seal and/or lubricant leak detection.

SAMPLE APPLICATION Figure 2. Electronics independent of the sensor make field calibration possible while the sensor remains installed within the process.

while the sensor remains installed within the process (Figure 2). Suppliers of thermal flow instruments have begun integrating simultaneous flow, liquid/level interface and temperature monitoring/switching functions in pump protection applications. This integration of three functions in a single state-of-the-art flow switch provides precision accuracy, simplified installation/maintenance and low life cycle costs. In pump protection applications the flow accuracy of FCI's multifunctional FLT switch, for example, is within ±2% of the setpoint velocity of a ±50ºF (±27ºC) operating temperature range. Repeatability is ±0.5% of reading. Level accuracy can be specified at a resolution as low as ±0.25 inch (±0.64 cm), with repeatability of ±0.125inch (±0.32cm). Standard temperature accuracy is ±2.0ºF (±1.1ºC) with a ±1.0ºF (±0.56ºC) repeatability. In terms of pump protection, the FLT switch features intelligent con-

156

In producing methane from coal beds, a process known as coal bed degasification, water removal is important. The pumps used to remove water from the bottom of the wells are typically progressive cavity pumps set at 1,000-5,000 ft deep. In the early production phase water is pumped in hundreds of barrels per day. As water production drops below 10 barrels per day, control of the pump is critical. The typical control method is to slow the pump for continuous duty. These pumps have a metal-on-rubber construction. When a pump runs dry, it heats up and causes damage to the rubber lining. Running in continuous duty mode at low rates is not the most effective way to operate the pumps. Using an FLT switch allows personnel to operate the pump on an efficient duty cycle for best economy and extended pump life. To do this, the switch is installed at the pump discharge. The pump is set to pump at high rates. When the pump starts, the flow of water is indicated by the switch. When the flow stops, the switch detects this and shuts off the pump. A time delay relay is used in conjunction with the switch to

The Pump Handbook Series

restart the pump. The cost savings are in power usage and equipment cycle life. The old operation was low rate on a continuous duty cycle. With the switch and relay incorporated into the control scheme, the pump operates just 5-10 minutes every hour. This is one sixth of the previous operating time, a difference that reduces pump wear and extends pump service life.

PUMP INSURANCE Pump protection is like insurance. One never knows its true value until something goes terribly wrong. With proper planning, a protection scheme will provide years of trouble-free detection and pump operation. Thermal dispersion instruments offer a universal approach to pump protection. While the cost of solid state technology may be higher initially than other techniques, its greater reliability for critical applications over the life of pumps makes this technology well worth the investment. ■ Greg Ochs is a senior engineer at Fluid Components International. He has been with the company for more than 10 years and has prior experienceof six years in instrumentation and process control. He holds a B.S. in computer science and a degree in electronics technology.

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R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

Reliability Profile #1: Champion International Corporation By Robert Matthews

Facility: Location: Number of Employees:

Champion International Newsprint Plant Sheldon, Texas 825+

Primary Product(s) Manufactured:

newsprint, sheeted pulp and specialty grades of paper

Estimated Number of Pumps per Application:

1250 100 400 200 100 300 30 30 90

pumps millwide water stock hydraulics and lubrication steam and condensate chemical lift sump miscellaneous

The Pump Handbook Series

Editor's Note: Plant Profiles is a new series that will appear periodically in Pumps and Systems. It will show how specific plants have tackled pump maintenance and reliability problems, and it will explain the strategies and actions taken to achieve success. Our first profile has been put together by Robert Matthews, Pump Shop Technician and Maintenance Training Coordinator at Champion International's Sheldon, Texas plant, just outside Houston. Bob details the steps he and his associates took that led to a significant increase in pumping system life and a reduction in maintenance costs. We hope you find these "plant snapshots" useful and informa-

157

tive.

PUMP RELIABILITY STRATEGY

Our goal is to explain to you how Champion International's Sheldon, Texas plant increased the efficiency of the pumps that we use in our mill. For many of you, the solutions we employed may be nothing new, but for others they may just be the answers you are looking for.

TIGHTER SPECS About five years ago, our mill hired its own maintenance work force, an action which resulted, among other things, in changes and improvement in our pump shop. Before putting the new maintenance workers on-line, we were averaging approximately 30 centrifugal pumps in the shop per month. In 1991 this number dropped to 12-15 per month, and it went to even less the following year. A key strategy was to change to tighter specifications in the rebuilding of our pumps. Some of the major areas in which the specs were changed included bearing fits, shaft TIR and alignments. In the last half of 1994, we also started measuring housing fits during pump disassembly. Since 1992, these changes have enabled us to repair pumps that we were previously sending out to other shops. They have also provided us with an opportunity to work on a wide range of pumps.

BEARING ISOLATORS Another change that increased pump and shaft life came in the form of bearing isolators. With the lip seals gone, we don’t trash as many shafts that have been cut too deep or repaired too many times. Bearing isolators have reduced oil consumption and contamination. Using the isolators and sight glass combinations has also enabled us to reduce the number of oil bowls that we were using.

STAINLESS STEEL WET ENDS Why stainless steel in a paper mill? I get this question all the time, and I have many answers for it. The two best answers are consistent grade and additive changes. Stainless steel costs more than cast iron or bronze, but when we can stock 1/3 fewer parts and pay less in inventory tax, it is clearly a cost savings. This strategy has also led to reduced paper handling, less stocking in shipping and receiving areas and less mixing of parts. Another consideration is that paper grade changes and changes in paper manufacturing methods call for different chemicals, and stainless steel works well in these applications.

MECHANICAL SEALS Our mill was designed in the 1960s when mechanical seals were few and far between. The packing

method of sealing worked well, but manufacturing processes changed based on environmental concerns. We started using seals where we had to. Today, we are constantly adding more seals for cost savings, such as reductions in water, fiber and energy losses. Cartridge seals are the most common in our plant, along with some single and double seals. Again, we try to use the same size, type and material in order to reduce stock, confusion and the number of suppliers.

COATING INSIDE CASTINGS I'm sure many of you have experienced the following: After cleaning the inside of a bearing housing that has had a complete bearing failure, metal particles reappear. You can clean the housing until your blue in the face, and after a close inspection you will still see grit inside. To control this situation, we paint or coat the inside of our porous metal castings with Glyptal. The coating traps particles so they don’t escape and contaminate the oil, and it makes it easier to clean the casings when the pump is repaired. Recently, we have started using electroless nickel as the coating on cast iron and cast steel pump parts. After all, we are using SST wet ends that don't need painting so why not use a pump that doesn't need paint. Our plant has many areas that require continued repainting because of chemical attack, for example. The cost of one paint job after another adds up quickly, and a small study has revealed how much the constant repainting has cost us. An added plus is that the looks of the electroless nickel coated parts is very nice.

FINE NICKEL ANTI-SEIZE In our pump shop reliable means reusable. Many new pump assemblies arrive at the plant with carbon steel bolts threaded in or boss fit, and without anti-seize. Later some corrosion damage or galling occurs. When using stainless steel on stainless shafts, fine nickel anti-seize has reduced galling of these parts. In addition, split case pump gaskets with anti-seize make it easier to part or separate.

PACKING CHANGES New packings are common,

High pressure water pumps.

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suction pumps. In fact for more than four years these pumps have run almost continuously, stopping only for a day every 2 months for short shutdowns. Prior to the shafting changes, these pumps experienced bearing failures every 2-3 months.

DYNAMIC SEALS Thickstock pump.

but there some good new fibers such as GFO, a Gortex product. These new fibers can take a 1-14 pH range, which meets most of our mill pump needs. This reduces the types and quantities of packings in inventory, and replacements are needed less often. Waterless packings have an advantage, but make sure you never over tighten them.

EXPANSION JOINTS By installing expansion joints in our TMP (Thermo Mechanical Pulp) unit on several pump discharge piping arrangements, we have reduced pipe strain to within acceptable alignment standards.

HEAVIER SHAFTING AND BEARINGS We have improved our pump life by moving to stronger shafting, particularly in the high speed end

We have reduced stuffing box pressure with dynamic seals on our de-ink (recycle paper) plant pumps and on some of the paper machine pumps. The stuffing boxes are shorter and require less packing, and waterless packing can be used.

VIBRATION ANALYSIS In 1989 we started a vibration analysis program at our mill, with a running history on 5,142 points monitored. Original alarm level was set at 0.3 in/sec, and on the initial route 122 points were in alarm. Tolerance is based on weight of rotor and normal running speed according to the API standard. The formula used to calculate it is:

These numbers dropped because we held to specifications up to six times greater than OEM standards.

CONCLUSION What has really made the difference at our plant has been quality feedback from what is happening in the field. By listening to the individuals involved with the machinery, we have significantly reduced our pump rebuild problems. ■ Robert L. Matthews is Pump Shop Maintenance Technician and Maintenance Training Coordinator at Champion International Corporation's Sheldon, TX newsprint plant. He is a frequent contributor to Pumps and Systems and a member of our User Advisory Team.

4x Weight (in grams) rpm We are still monitoring the 5,142 points and finding only 10 in alarm.

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R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

Tips and Trends in Sanitary Pumping Sanitary pump manufacturers target pharmaceutical applications, share selection tips. By Robert C. Waterbury, Senior Editor

A

SELECTION CRITERIA Clean-in-place/steam-in-place. For sanitary and especially sterile applications it is important that the pump be designed for clean-in-place (CIP) or steam-in-place (SIP) capability. These abilities offer the user major labor savings in cleanup and maintenance. They also minimize damage to pump parts due to frequent handling and disassembly/assembly. As APV Crepaco points out, this is important

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in continuous processes where piping systems and equipment are cleaned daily. Construction materials. The 3A standards require that metal product contact surfaces be of AISI 300 series stainless steel construction, or offer equivalent corrosion resistance metallurgy that renders the metal inert, nontoxic and nonabsorbent. Typical alloys meeting these specifications include stainless steels 316, 316L, 174PH and Nitronic 60. Similar additional 3A standards are applicable for elastomer and plastic construction

PHOTO COURTESY OF APV CREPACO

ccording to APV Crepaco Inc, sanitary pumps handle materials ranging from alcohol to glycerine and from shaving cream to peanut butter. They must accommodate corrosives and abrasives as well as soft drinks and fruit juices. Thus, sanitary pumps require a wide variety of options in order to deal with different fluid applications. It is perhaps not surprising, then, that both positive displacement and centrifugal pumps play major roles in sanitary pumping applications. To make an informed selection of pump type and options, one first must be intimately familiar with the particular fluid and its pumping and handling characteristics. Once these are considered, there are some general selection criteria that apply to both centrifugal and positive displacement sanitary pumps. Sanitary pumps must meet the requirements specified in the 3A Standard for Centrifugal and Positive Rotary Pumps. Depending upon the application, they also may have to comply with other FDA regulations, current good manufacturing practices and ANSI/ASME standards for food, drug and beverage equipment.

Photo 1. Model 8V2 aseptic centrifugal pump from APV Crepaco

materials. Surface finish. To ensure continued sanitary or sterile conditions, the surface finish or polish of the wetted areas of the pump should not allow microbes or bacterial growth to become trapped. Thus, surfaces should be finely polished and free of cracks, crevices and occlusions that could harbor contamination. Mechanical polishing to a fine micro-inch finish is the commonly accepted method. According to Greg Sturicz of The Pump Handbook Series

Tri-Clover, pump manufacturers generally offer a 150 grit or 35 micro-inch finish. Other options include 180, 240 or 320 grit with 25 micro-inch, 15 micro-inch or 10 micro-inch finish, respectively. Mechanical seals. The seals perform a dual function by providing a sterile barrier to the atmosphere and protecting the fluid quality from outside contamination. They must be easily cleaned and not harbor bacteria. They can be of single or double design and internally or externally mounted. However, they must meet 3A requirements calling for sanitary, packingless design with all product contact parts demountable and accessible for cleaning or inspection. Seal materials commonly used include tungsten, silicon carbide, ceramics and carbongraphite. Associated mounting and supporting elements – including cups, o-rings, set screws and so forth – must also exhibit compatible fluid and performance characteristics. Shaft and impeller design. Whether it is centrifugal or rotary displacement, the sanitary pump should have a shaft and impeller design that offers efficient, low-shear fluid acceleration. Impellers may offer several available modifications including a full or semi shroud placed at the back. This creates a high positive pressure on the inboard mechanical seal, preventing contamination from entering the pump through the seal. Construction material and surface finishing of the impeller or rotor may match that of the pump if it is compatible with the pumped fluid. Rotary clearances must be held constant to prevent galling, and the retention mechanism must remain sanitary.

PHARMACEUTICAL APPLICATIONS Many pumps commonly used in food, beverage and chemical applications have been adapted for use in demanding pharmaceutical production environments. Companies such as APV Crepaco have developed models specially for the pharmaceutical industry. Crepaco's V2 series pumps are an example. Pharmaceutical needs are classified on the basis of the products and their manufacturing requirements. Ethical pharmaceuticals, for instance, consist of products sold to dental, medical or veterinary professionals. They might include such items as plasmas, serums, antibiotics, vaccines, narcotics and hormone products. They are available only through a doctor's prescription, and distribution is limited. Companies producing ethical pharmaceuticals include HoffmanLaroche, Merck, Glaxo Wellcome, Eli Lilly and Baxter International. Proprietary pharmaceuticals are non-prescription products sold over the counter to the public. These include cold remedies, analgesics, ointments, laxatives and cough suppressants. Companies in this market include Miles Laboratories, Plough Inc., Sterling Drugs and McNeill Labs. The U.S. government also classifies pharmaceutical products based on SIC (standard industrial classification) codes, reflecting whether the preparations are biological chemicals, and botanical products in nature. Accordingly, APV has divided its centrifugal pharmaceutical pump products into three categories: standard 3A, pharmaceutical pump and water for injection (WFI). The following descriptions illustrate how the duty levels and pump products differ. Standard 3A pump products offer a finish on all wetted parts (casing, impeller shaft, etc.) equivalent to 150 grit properly applied to stainless steel sheet. This level is intended for use by proprietary pharmaceutical manufacturers for feed stocks and service lines. Certain options are available as noted in Table 1. The pharmaceutical pump offers a level of surface finish on all wetted parts equivalent to 180 grit polish plus electropolish. Also, a special casing drain connection and the Type 5 seal are provided. The casing drain

PUMP SPECIFICATIONS - PHARMACEUTICAL DUTIES FEATURE

STANDARD 3A

PHARMACEUTICAL

WFI

Construction Material

CF-8M investment cast stainless steel (AISI 316 wrought equiv.)

CF-8M investment cast stainless steel (AISI 316 wrought equiv.)

CF-8M investment cast stainless steel (AISI 316 wrought equiv.)

Surface Finish-Product Contact Areas

150 grit polish

180 grit polish + ectropolish

320 grit polish + electropolish

Casing Drain

optional

front outlet, sanitary

front outlet, sanitary

Mounting

pedestal or close coupled

pedestal

pedestal

Pedestal Base

mild steel painted

stainless steel

stainless steel

Seal

Type 1 thru 5

Type 4 or 5

Special Type 4

Flush Media

water, solvent

water, distilled water

WIF at 176°F

Cleaning/Sanitizing

CIP, not sterilized

CIP, hot water sterilize

CIP, steam sterilize

Typical Applications

soft water supply UF supply solvents fermentation broth process water emulsions RO effluent

deionized water fermentation products pyrogen-free water saline solutions plasma

water-for-injection parenterals final rinse water sterile water post sterlization cooling water sterile products

Table 1. Pump specifications for pharmaceutical duties

is located at the bottom for complete drainage and includes a front facing 90º ell and sanitary ferrule. This design is used mainly by ethical pharmaceutical manufacturers. The water for injection (WFI) pump is a demanding service unit whose shaft sealing involves use of hot, high purity water as a buffer or flush liquid. A special double mechanical seal is required to counteract the high (176ºF) temperature and poor lubricity of WFI. In many cases a 320 grit (12 micro-inch) finish plus electropolish is specified for all metal product contact surfaces. An M series rotary displacement aseptic pump, offering guaranteed product purity for both pharmaceutical and food processing applications is also available from APV. Special features include: ■

■ ■ ■ ■

Double mechanical shaft seals, with front cover and inlet/outlet connections fitted for steam or sterilizing chemical flush Product contact surfaces polished to 12 micro-inches (320 grit) and passivated 316 stainless steel rotors suitable for operation at up to 300ºF Differential pressure capability up to 290 psi in selected models Viton or EDPM high temperature o-rings used throughout In the case of the rotary displaceThe Pump Handbook Series

ment M series aseptic pumps, both static and dynamic seal designs are available. The standard dynamic seals are Durametallic RA mechanical. This is an externally mounted, balanced design with positive drive. Seal materials include carbon, hardened stainless steel, tungsten carbide and silicon carbide. Secondary o-ring seals are EPDM, Nitrile, Viton and Kalrez. The double mechanical seals with sterile media exclude contaminants. Shaft seals on this series have faces that combat abrasion and corrosion. Static shaft seals include single oring and packed gland designs. The o-ring design meets sanitary requirements. The packed gland is for nonsanitary uses. The M series pumps are furnished with replaceable stainless steel shaft sleeves. This is because the sleeve is easier and less expensive to replace than a shaft. For more aggressive applications the sleeves are hardsurfaced to extend service life. The hardened sleeves are recommended for use with static seals. M series pumps feature a modular design. The bodies, shafts, shaft sleeves and seal chambers can be changed out along with the seal types and materials for maximum flexibility in building a unit suited to a particular application.

SEAL SELECTION IS KEY Greg Sturicz of Tri-Clover points

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■

■

■

■

■

■

■

The Type D external balanced seal is a multi-purpose seal designed for extended service in applications ranging from dairy products to vegetable oils and beverages. It is also applicable to acid cleaning solutions and detergents and is well suited for CIP service. The Type DG clamped-in seat assembly provides extended service life for seals and backplates in corrosive or abrasive environments. Users can select seat and seal materials to suit their application. Type F and FG seals are identical to types D and DG, with cascading water to assist in cooling or cleansing seal faces. The Type E water cooled balanced double seal is suited for slurries and heavy-duty vacuum applications to 28" Hg as well as tacky materials and other products at temperatures up to and exceeding 212ºF. The Type H John Crane double type 9 seal is used in the pharmaceutical industry. It is designed for WFI applications and requires a sealing barrier of WFI discharge water. The Type Y John Crane double type 8 seal is an upgrade of the John Crane double type 9. The only difference is that the double 8 uses o-rings as a secondary seal, whereas the double type 9 uses PTFE wedges as the secondary seal. The Type R Chesterton 241 double seal with flush features two seals in one housing – essentially a double tandem combination seal. Also designed for WFI applications, the seal requires the

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PHOTO COURTESY OF TRI-CLOVER

out that proper seal selection and application is the key to successful sanitary pump operation. A good pump and seal match can pay off in terms of improved operating efficiency and lower maintenance. A poor match, however, can adversely affect productivity, product quality and operating and maintenance expense. Thus, it is necessary to have a wide variety of seal options available. Following is a typical list of seal options available for implementation on various Tri-Clover sanitary pumps.

Photo 2. Tri-Flo CL series centrifugal sanitary pump

sealing barrier of WFI discharge water. In addition, of course, there are optional non-sanitary seals including types A and B packing glands. And a wide range of other commercial seals commonly used in food, beverage and chemical processing applications is available as well. Sturicz provides several rules of thumb for evaluating seal options. 1. Mechanical seals usually provide longer seal life than packing gland seals. They also have a much lower leakage rate. 2. If several different mechanical seals will do the job, remember that the type requiring fluid flushing and hardened seal faces will normally last longer. 3. Maintenance can be simplified with minimal spare parts if you consider all applications and product characteristics and then select compatible types of seals. And don't forget that the pump manufacturer can provide helpful information in selecting the proper seal and other options for your sanitary pumping application.

SANITARY CENTRIFUGAL PUMP PRODUCTS Although numerous options are available to help customize sanitary pump solutions, it is not simply a matter of selecting the right options. Manufacturers have in most cases sought to design sanitary pumps to perform certain types of jobs. The Pump Handbook Series

For instance, Fristam Pumps offers several series of sanitary centrifugal pumps. The FP series is designed for heavy duty and high efficiency, the FZ series for heavy duty self-priming, and the FM series for multiple stage high pressure applications. The FP series, for example, features a unique impeller shape and close clearance between impeller and cover so less net positive suction head (NPSH) is required. It is also designed to offer excellent shear characteristics for applications involving movement of cream, whole human blood and other biogenetic solutions. Its double mechanical seal makes it virtually impossible for air to get into the product and contaminate it. At the same time, it easily handles viscous solutions up to 1500 centipoise (about the same consistency as honey) and has even been used for a thixotropic pear slurry, which has an apparent viscosity in the 4000-8000 centipoise range. When the interior of the FZ heavy duty self-priming pump is wet enough to form a liquid ring it will self prime, lifting water up to 23 ft. It pumps products with large amounts of entrained air or even air in the line and still maintains its prime. With its external seal option, the FZ can operate in either direction, allowing a single pump to empty or fill tanks and silos. The FM series sanitary pumps are designed for multiple stage high pressure applications. The internal seal design and heavy duty construction

■ Sanitary finishes. A 316 stainless steel construction with standard 150 grit interior product finish, and optional mirror polishes and electropolishing on product contact surfaces ■ Double seals. A double mechanical seal at the impeller shaft and a double seal connection between the pump casing and backplate ■ Seal materials. Standard materials are carbon versus silicon carbide seal faces with EPDM oring secondary seals; options include silicon carbide versus silicon carbide seal faces or Viton o-ring secondary seals. ■ Choice of sterile media barriers. Low pressure steam, hot conden-

sate or other external sealing liquids, or product from the pump discharge, can also be used as the flush media as in water for injection. And this is just a sampling of the many centrifugal pumps and options offered for use in sanitary applications.

ROTARY DISPLACEMENT SANITARY PUMPS A new line of positive displacement sanitary pumps for high pressure, high performance applications was recently introduced by Tri-Clover. The new pump design provides flexibility with a wide selection of sizes, capacities, tri-lobe or bi-lobe rotors, and other options to accommodate a broad range of applications. The new line includes the TSK and TSR models, all of which have CIP capability and comply with USDA and 3A requirements, as well as international hygiene standards. They feature 316 stainless steel construction for corrosion resistance and are designed with a sealed spline area with rotor nut o-rings at the back of the rotor for easy cleaning and rotor disassembly. The five TSK models are termed "ultraclean" due to a design that separates processed product from potential sources of contamination. The pump's spline is totally enclosed in the rotor, eliminating exposed rotor nuts, splines and front cover recesses—areas where product can be retained and contaminated by bacterial growth. APV Crepaco also offers what it terms a MicrobeClean stainless steel lobe pump. The MC series features a one-piece rotor/shaft sleeve design with no joints, seals, threads or splines in the product zone. It is a self-draining, self-venting unit based on vertical, top entry design and internal body profile. The standard finish is 78 micro-inch (300 grit) on product contact areas with passivation. An optional finish of 63 micro-inch (320 grit)

The Pump Handbook Series

PHOTO COURTESY OF FRISTAM PUMPS

allow them to withstand up to 1000 psi at the pump inlet. This construction enables a series of such pumps to build up to 1250 psi – a level that cannot be reached by other pumps of lighter construction. Tri-Clover recently introduced its Tri-Flo CL series of centrifugal sanitary pumps featuring 316L stainless steel construction and full CIP capability in casing diameters ranging from 8.5-15.75". High efficiencies are a result of tight tolerances between the impeller and front cover. And the pumps can be equipped with a variety of seals including external balanced seals, single and double shaft water cooled balanced seals and a broad range of commercial seals. APV Crepaco offers its W series high efficiency centrifugal pumps for use in food, beverage, pharmaceutical and chemical industries. Within this series is its Wi UNIversal inducer pump for use with viscous or aerated products, the WHP high pressure pump, the W-140/50 multi-stage pump and the Wa series aseptic pump. The Wa design is for use in totally sterile processing environments. Its features and options include:

Photo 3. FZ-Series, heavy duty selfpriming sanitary centrifugal pumps

with electropolishing is available. The standard single balanced mechanical seals are externally mounted; interchangeable double mechanical seals are available for duties where a sterile fluid barrier is necessary. The MC pump is available in fully aseptic construction with double seals to the front cover joint and port connections and mechanical seals for saturated steam or sterile circulation.

ADAPTING TO APPLICATIONS Considering the many pump designs and options mentioned in this brief article, it is obvious that centrifugal and positive displacement pumps can meet nearly any of today's sanitary pumping needs. It does require careful evaluation, however, to match the pump to the particular production need. Obviously, the requirements of a line producing quinine are much different from those of a line producing toothpaste. And yet both involve careful consideration of sanitation. ■

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R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

Computer-Based Pump Reliability By Steven J. Hrivnak oday’s competitive business environment requires data to help make business decisions. In years past, many business decisions were made without the data to support them. Now, with the use of new computer programs, maintenance engineers can use this information to identify which equipment generates the most work orders.

T

Recent analysis shows that pumps are at the top of the list for rotating equipment work orders. A pump improvement program initiated at Tennessee Eastman Division to reduce pump work orders resulted in new specifications for alignment, baseplates, piping and grouting. Proper installation and application of ASME B73.1M pumps have significantly reduced work orders at Tennessee Eastmans’s division over four years on 91 pumps that were installed or reinstalled to new standards. A computerized maintenance system, reliability program implementation and standards development were the keys to improving the reliability of the 91 ASME B73.1M pumps.

BACKGROUND Tennessee Eastman Division of Eastman Chemical Company has more than 30,000 pumps. Approximately 14,000 are centrifugal type pumps, of which 8,000 are of ASME B73.1M design. (Note: ASME B73.1M design pumps are also known in industry as ANSI B73.1M.) In the late 1980s computerized statistical work order analysis showed that pumps created the second largest number of work orders (Table 1). The pump population had migrated to free-standing baseplates because of lower cap-

164

From 7/1/93 to 12/31/ 93 ■ Building: ALL ■ Area: ALL ■ Equipment: (Type / Number / Problem Part):

Pareto Chart of Equipment Types - By Frequency Equipment Type Pipe Pumps Tanks Dryers Conveyors

Frequency 9532 7862 3266 2173 1620

Percent % 13.91 11.47 4.77 3.17 2.36

Table 1. Plantwide work order Pareto chart

ital costs. Valves were placed on the pump flanges to reduce product loss during maintenance. To complicate matters, straight edge alignments were used routinely due to the perception that proper alignments seldom improve results. Pump reliability became such a problem that most pumps were spared. If a spare pump was put into service, the primary pump repair work order became top priority with approved overtime. Maintenance crews were sometimes forced to work shifts to accommodate top priority pump work orders.

COMPUTERIZED MAINTENANCE SYSTEM In 1984, Tennessee Eastman implemented a computerized maintenance work order system. It was called the Maintenance Management Information System (MMIS). The VAX based system allowed operations, engineering and maintenance to enter work requests from any terminal location in the plant. The computer program contained such features as crew work management, backlog age, work order planning, priority sorting, time charging, equipment nameplate data information, problem codes, action codes, equipment history cost and work order histories. For the first time in the plant’s 70-

The Pump Handbook Series

year life, all work request histories were kept in one database. In 1991, a Work order Analysis and RePorting (WARP) system came on line. It gave direct statistical access to work order histories. Employees could get Pareto charts on equipment failures in an operating area, division or the entire plant. High maintenance equipment could easily be identified. Number of work orders, hours spent on a piece of equipment, actions taken and dollars spent could be accessed and used to justify changes or modifications (Tables 2 and 3). The WARP system provides maintenance, operations and engineering with the data needed to prioritize where maintenance improvement dollars should be spent. It also supplies the data to prove that existing installation and maintenance practices produce equipment reliability problems. Poorly designed freestanding baseplates cause high vibration and maintenance headaches. Similarly, air voids under grouted baseplates are poor vibration attenuators and require high maintenance. Cost information applied to this data helped convince management to increase funds for upgrades and improvements.

RELIABILITY PROGRAM INFORMATION Coupling Alignment. The alarmingly high number of pump work orders prompted Eastman’s Plant Maintenance Division to target pump alignment as the solution to improved reliability. Management purchased computer alignment tools and taught computer alignment to maintenance mechanics in hopes of solving pump reliability problems. The standards adopted for alignments were 0.003 or less for vertical and horizontal offset and 0.0005 angular for 1800 rpm. At 3600 rpm, the standard

WARP DATASET SUBSETTING OPTIONS

Table 2. MMIS request choice screen PARETO CHARTS

tor was made from bent plate, hollow tubing, C-channel or solid blocks. Little thought was given to stiffness to maintain coupling alignment. There were no standards for the designs. Inspection of a group of 30 free-standing pumps typically found about five pumps supported on only three stilts; the fourth vibrated loosely in the air. Analysis of the plant’s ASME pump population showed that, on average, free-standing baseplates with poor piping had one work order every 6.3 months (Photos 1 and 2). Some pumps had as many as 35 work orders in one year! The analysis also showed grouted pumps with reasonable piping had one work order every 35 months (Photos 3 and 4).

Table 3. MMIS Pareto chart selection screening

tightens to 0.001 or less for vertical and horizontal offset, and 0.0002 angular. A mechanic with alignment instruction experience began aligning the 200 pumps in one department. As a result, he encountered many obstacles in trying to align pumps associated with equipment installation (Table 4). Through this experience and others the Plant Maintenance Division discovered that alignments did solve some problems but it was not the only answer; only a piece of the puzzle. Some free-standing baseplate designs could not hold alignments over time while others could. The question was then raised, Conditions found on first 27 pumps aligned: 60%

Bolt Bound

11%

The Motor Base Was Too High

89%

Were Free-Standing Design

100%

Had Either Pipe or Conduit Strain

100%

Motors Without Alignment Jack Bolts

Table 4. Alignment conditions found

"What is the best way to install a pump?" This initiated an information search for the best pump installation practices. Engineering standards did not provide answers or Photo 1. Free-standing installation with were years out of date. A few poor piping dedicated individuals in maintenance and engineering began a journey to pump reliability, but the obstacles were numerous. Other Obstacles. Many existing baseplates were either supplied by the pump vendors or designed inhouse by mechanics or engineers. The design was a C shaped or Omega shaped 1/2 stainless bent plate. The superstructure that supported the mo- Photo 2. Free-standing installation with poor piping The Pump Handbook Series

165

Photo 3. Grouted installation with good piping

Three situations around the plant paved the way for changes to existing practices: • It was discovered that the coupling alignment would move 0.010" with existing free-standing baseplate designs if a stilt was loosened. This information convinced management that a solution was needed to maintain alignments. • A civil engineer raised concerns about the continuous repair of corroded floors underneath freestanding baseplates. • Rotating equipment engineers noticed high vibration data on many ASME pumps. Management subsequently funded a test to install several pumps correctly. This management buy-in in one department allowed a trial pump installation of a select group of pumps to see if correct installation would improve reliability. To do this, several areas besides alignment had to be addressed. These included baseplate design, piping design and grout placement Baseplate Design. Eastman’s information search on the correct way to install pumps yielded some best practices on baseplate design for improved reliability. A team was organized to standardize baseplate designs and apply the best practices. Certain fundamental rules were found to be overlooked when exam-

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ining existing baseplate designs. The foundation mass ratio, the weight of the foundation in comparison to the weight of the machine, driver, baseplate and the liquid or material in the machine, must be a certain value in order to minimize vibration. For rotary equipment such as pumps, this ratio should be three to one. For reciprocating equipment, this mass ratio should be five to one. Free-standing pump designs cannot meet this requirement. Therefore efforts were focused only on grouted baseplate designs. A set of five drawings was eventually developed showing all pump and motor combinations for ASME B73.1M type pumps. The drawings contained the following design points: • The baseplate shall be fitted with one 4" grout fill hole uniformly distributed for every 10 square feet of baseplate surface and/or per subdivided section or raised welded cavity. • Vent holes 1/2" in diameter shall be provided for each bulkhead compartment at all corners, high points and perimeter edges of the bulkhead. Perimeter vent holes in the baseplate shall be on 18" centers maximum spacing. Any angle or "C" channel added for stiffness will require vent holes on both sides. • Vertical jacking screws 1/2" diameter minimum shall be proThe Pump Handbook Series

vided around the baseplate perimeter 3" from each anchor bolt location to facilitate alignment of the baseplate in the vertical direction. • Machined mounting surfaces for the equipment and driver shall have horizontal jack screw positioning bolts 3/8" diameter minimum. • All welding on the baseplate shall be completed prior to machining equipment and driver mounting surfaces. • Machined mounting surface shall be coplanar to 0.002". • Machined mounting surfaces shall extend 2" beyond pump and driver feet on all sides with a 125 micro inch Ra finish. When machining a bent plate as a mounting surface, maximum material removal must not exceed 1/16". • Provide 1/8" minimum shim adjustment under driver feet for alignment. • If a spacer coupling is used to couple the equipment to its driver, add an additional 1/4" axial clearance to the spacer length in the baseplate design. • Anchor bolt holes shall be 1/4" larger on the diameter than the anchor bolts. • Tapped equipment bolt holes in the baseplate should be of sufficient depth to allow for plenty of thread engagement. On 1/2" thick or smaller baseplates, weld 1/2" or 3/4" thick square plates under or on top of the baseplate to increase the thickness before drilling, tapping and machining. Tapped depth should be one and one half fastener diameter. • All bulkhead cross bracing on the underside of the baseplate shall

Photo 4. Grouted installation with good piping

have a 2" x 6" opening to allow for grout flow from bulkhead to bulkhead. • All corners of baseplate flanges shall be radiused a minimum of 1". All surfaces which will be in contact with the grout shall be rounded to eliminate stress risers. • Angle, channel or stud anchors shall be welded to the bottom surface to act as a shear key in the grout. • Gussets shall be welded inside the driver superstructure to increase stiffness and vibration damping. Most of the above baseplate design points can be found in Process Industry Practice RESP002. Pipe Standards. Proper piping costs no more than improper piping, and provides far better results. During the time the baseplate standard was developed an ASME B73.1M pump piping standard was also created. Before the new standards were written, existing practices put valves on the suction and discharge flanges or elbows directly on the

pump. Many times the pump casing was the only pipe support within 20 feet of the pump. On one new installation where the pipe was out of alignment by about 1/2" on both the suction and discharge flanges of a 6x 4-10 pump, the pipes could still be easily bolted to the pump. The pipe was disconnected and the pump was aligned to class one tolerances. Then the pipe was reconnected and the alignment was rechecked. The horizontal offset moved from 0.003 to 0.025 ! The pump was being twisted. The frightening realization was that even though the pump was twisted, the motor could be moved enough to get class one alignment. The piping was reworked the correct way and the pump operated for two years without a failure. Simple rules that existed for years were typically ignored. Proper piping minimizes shaft, casing and baseplate deflections as well as cavitation. This increases seal, bearing and coupling life. A pump pip-

PIPE HANGER OR SPRING SUPPORT, DIRECTLY ABOVE PUMP

VENT VALVE IF REQUIRED

CHECK VALVE, IF SPECIFIED

SPOOL PIECE, FIELD FIT 8 PIPE DIA. STRAIGHT RUN OR 18", WHICH EVER IS GREATER DISCHARGE PRESS, GAUGE, 5 PIPE DIA. FROM PUMP DISCHARGE

SPOOL PIECE, FIELD FIT 8 PIPE DIA. STRAIGHT RUN OR 18", WHICH EVER IS GREATER SUCTION PRESS, GAUGE, 5 PIPE DIA. FROM SUCTION VALVE ECCENTRIC REDUCER

PIPE SUPPORT, 1 1/2 TO 3 FEET FROM PUMP NOZZLE

SUCTION VALVE, 1 PIPE SIZE LARGER THAN SUCTION FLANGE OR FULL PORT

Figure 1. Pipe installation diagram The Pump Handbook Series

ing standard was written to incorporate best practices. The purpose of the standard was to minimize piping and nozzle loading and provide uniform flow into the impeller eye. Piping Design. The new piping standard combined available information on best practices and contained the following (Figure 1): • Discharge and suction lines shall be straight for a minimum of eight (8) pipe diameters or 18", whichever is greater, in the runs adjacent to the pump flanges. (This minimizes the possibility of cavitation and smoothes out the flow entering the impeller.) The pump suction line shall be constructed to provide a spool section for strainer removal. • Discharge valves shall be located directly after the required straight run. Suction valves shall be located directly before the straight run. Suction valve port area shall be equal to or greater than the suction line area. Suction piping shall be sized for 5 ft/s (bulk velocity) or less. Hot condensate service and service for other liquids within 10ºF of the boiling point at suction pressure shall be sized for 2 ft/s or less. • Reducers in the horizontal suction lines shall be eccentric and located directly after the suction valve. If the liquid is within 20ºF of boiling, place the eccentric reducer at the suction flange of the pump. • Lateral suction piping shall be supported and guided no closer than 1.5 ft to a maximum of 3 ft from flanges. Vertical discharge piping shall be supported from above using spring hangers. Hangers shall be sized to exert minimal vertical loads on the pump nozzle in the cold condition. Piping shall be guided to help prevent transfer of piping moments to pump nozzles. All pump loads shall be below those determined by the pump manufacturer and verified by calculations for both hot and cold service. Exceptions: Support design may differ from above description if acceptable loading is verified by calculations and approved by the appropriate engineer. Special care should be exercised in support design in lines with fast acting and shock inducing check valves. Installation. Installation require-

167

CL + / - 1/8”

+ / - 1/32” INDEPENDENT OF FLANGE DIAMETER Figure 2. Piping flange tolerance

ments include: • Install piping to within the eight pipe diameters of the pump. Grout the pump baseplate and align the motor using the reverse indicator method to within 0.002" TIR. Bolt the suction and discharge stub ends and flanges with gaskets to the pump using four bolts. The spool pieces connected to the pump shall be field fabricated; field-fit and tack weld the piping to the stub ends, and then match-mark the flanges. Finally, remove the spool piece and final weld the piping to the stub ends. For reference purposes, the following installation guidelines should apply: • If the flanges cannot be aligned by hand (not with hand tools) for insertion of the bolts, the spool piece will be reworked.

x + / - 1/8”

• If a set of flange faces is not parallel to each other to within 1/32" across the length of the faces, the spool piece(s) will be reworked (Figure 2). • If the spool piece, which should be eight pipe diameters or greater in length, is more than +/- 1/8" of true length, the spool piece will be reworked (Figure 3). • The pump shaft deflection shall be checked using the reverse indicator method during flange bolt tightening. Coupling alignment must remain within 0.002 TIR. Miscellaneous. Other miscellaneous considerations include: • Baseplates shall be grouted and not free-standing. • For new pump installations, a strainer shall be installed to trap foreign material. The strainer should be removed and cleaned every eight hours until it no longer traps material. It should then be removed from the piping. A spool piece should be installed for easy removal of the strainer. • Pressure gauges or taps for pressure gauges shall be installed. (Pressure gauges can be used to troubleshoot pump problems). Grout Standards. Some excellent onsite grout training in which mechanical engineers, civil engineers, mechanics, installers and engineering contractors were all invited to learn at the expense of the Rotating Equipment Team produced a team of representatives that combined to write a new grout standard. The specification provides requirements for cementitious, epoxy, and corrosion resistant grouting of rotating and reciprocating equipment baseplates to concrete slabs or raised foundations. The purpose of following these procedures was to minimize equipment vibration and alignment problems and increase bearing, seal and coupling life. The specification contains the following:

Slab or Foundation Preparation and Protection

• After concrete has been properly cured and attains its design strength, typically 28 days, scarify existing slab or foundation surface that will come in contact with the grout using a small chipping hammer to expose aggregate (1" minimum Figure 3. Piping length tolerance

168

The Pump Handbook Series

depth). High pressure air clean chipped area to remove all dust and grit. • The cleaned, scarified surface of the foundation should be protected from water, oil, dust, contamination, etc. The preferred method is to place the grout directly against the clean foundation. (For epoxy grouts, if the time interval between cleaning and grout placement is sufficient for the contamination of the prepared area by water and oil, the scarified area may be coated with the grout manufacturer's recommended primer approximately 3- 5 mils thick. Under no circumstances will a generic epoxy paint be used. If the prepared area becomes contaminated, clean before grout placement. • The prepared area for epoxy or corrosion resistant grouts shall be kept dry. Epoxy grout will not bond properly to damp or uncured concrete. • For cementitious grouts, the prepared area should be wetted prior to grout placement per manufacturer's recommendation. • The prepared area shall be kept within the manufacturer's temperature limits. The area is to be shaded from direct summer sunlight 24 hours before pour and 48 hours after pour.

Anchor Bolt Sleeves

• Clean out any foreign material in the anchor bolt sleeve if used, and fill with a non-bonding urethane foam.

Grout Thickness and Flow Lengths

• Lacking information from the grout manufacturer, a minimum thickness of grout should be 1" with a maximum grout flow length of 1 ft. For each additional foot of grout flow length, the foundation to baseplate clearance should be increased by 1/2". The maximum allowable grout flow length shall be 5 ft. • The maximum grout thickness shall be per the manufacturer's written instructions and recommendations.

Reinforcing Steel

• In cementitious and polymer grout pours 4" and greater, concrete reinforcing steel shall be provided in accordance with ACI 318-89 Section 7.12 requirements for temper-

ature and shrinkage. Reinforcing steel shall conform to ASTM A615, Grade 60 and shall be fabricated and placed in accordance with ACI 301 and Specification 03001, unless otherwise specified on the construction drawings. Consult the manufacturer for reinforcing in epoxy grouts. BaseplatePreparation. This section contains information that should be completed before grout mixing begins. • Verify that the baseplate has a grout fill hole of adequate size and location to permit proper placement of the grout. If the baseplate does not already have a grout hole, cut a 4" diameter grout hole in the center of each cavity. • Verify that the baseplate has 1/2" diameter grout vents located in the corners of each cavity where air may be trapped. If each baseplate cavity does not already have 1/2" diameter grout vents, drill 1/2" diameter holes every 18". Place vent holes on both sides of any angle or "C" channel added for stiffening. This will allow air to escape while the grout is poured and avoid air pockets. • Remove all sharp edges of the baseplate (chamfering from 1/4" to 1/2") that will be in contact with the grout. The radius of all corners shall be 1" minimum. This will prevent crack propagation in the grout. • Verify that the baseplate has jack screws for leveling adjacent to each anchor bolt. If the baseplate does not have holes for jack screws, drill and tap a 1/2"-13 UNC or larger hole 3" away from each anchor bolt hole. • Clean and dry the underside of the baseplate to remove dirt, oil and other contamination. • When using epoxy grouts, the underside of the baseplate shall be sandblasted to white metal and a 3 mil anchor profile minimum (Soda ash and bead blasting are unacceptable) to remove rust and scale. The preferred method is to have the bare metal of the baseplate in contact with the grout. However, in the interest of corrosion prevention, if the time interval between cleaning and grout placement is sufficient for the formation of rust, the bare met-

JACK SCREW

ANCHOR BOLT

MOUNTING PLATE

SANDBLAST EDGES AND BOT. OF MOUNTING PLATE TO OBTAIN A 3 MIL. PROFILE

2"0 ROUND x 3/8 THK

GROUT (REBAR NOT SHOWN FOR CLARITY)

WRAP AS MUCH EXPOSER THREAD AS POSSIBLE WITH DUXSEAL TO PREVENT BONDING OF THE GROUT

ENSURE ANCHOR BOLT SLEEVE, WHEN SPECIFIED, HAS BEEN FILLED W/NONBONDING URETHANE FOAM BEFORE POURING GROUT

CONC. BASE DETAIL

FORM SEAL JOINTS WITH RTV

1/2"0 BLEED HOLE AT 8" O/C 3/4" CHAMFER STRIP FORM

TYPICAL GROUTING FORM

Figure 4. Typical grouting details

al may be coated with the grout manufacturer's recommended primer. Under no circumstances should a generic epoxy paint be applied to metal surfaces to be set in epoxy grout. A rust-free, bare metal surface is the preferred alternative that is simple, safe, and yields structurally superior results. Baseplate Installation. Following items pertain to installing

the baseplate (Figure 4): • Place 2" diameter by 3/8" thick stainless steel pieces with sharp corners removed on the scarified concrete under the jack screw locations. Level the baseplate using the jack screws against the 3/8" thick pieces. The preferred method is to level the baseplate without the equipment. If the equipment must be mounted while leveling,

Table 5. Machinery leveling tolerance The Pump Handbook Series

169

Anchor Bolt Tightening

• After the grout has cured, tighten the anchor bolts to the applicable torque values (Table 6).

Void Testing

Table 6. Anchor bolt torque values

pregrout the 3/8" thick pieces to stop movement under the increased load. Level the machined area of the baseplate to within manufacturers' specifications. If manufacturer's specifications are unavailable, use the following parameters (Table 5). Hold the baseplate in position against the jack screws with the anchor bolts. • Prior to pouring grout, wrap the baseplate leveling jack screws and anchor bolts with putty to prevent the grout from adhering. Coat equipment fastener bolts and coupling guard bolts with antiseize and install all the way into the baseplate to keep grout from clogging the threaded bolt holes and adhering to the bolts. (If grout adheres to anchor bolts, proper torquing may break them. If grout adheres to leveling jack screws, they cannot be removed after grout cures. Grout is designed to support equipment weight and absorb vibration; leveling screws are not.)

Wood Forms

• Coat the inside of the wooden form's vertical face with one coat of form oil for cementitious grouts or three coats of paste wax for epoxy grout (to prevent sticking). Forms should be waxed before installation to prevent accidental application of wax to surfaces where the grout is to bond. • Seal all form joints with silicone rubber sealant caulk to prevent grout from leaking out.

Grout Placement

• Grout should be mixed and placed per manufacturer’s instructions. An 18 to 24" long by 4" head

170

pipe and funnel is recommended to facilitate grout placement. Put a flange on one end and physically hold the flange down over the grout hole or build a wood bracket to hold it in place. Pour the grout through the funnel into the head pipe. (The head pressure will help the grout flow and push the air out from under the baseplate.) Before removing the head pipe, plug all grout vent holes with rubber or wood stoppers. After removing the head pipe, plug grout fill holes with plywood. (This holds the grout in place and eliminates air pockets beneath the plate.) Caution: do not use a vibrator on polymer concrete or epoxy grouts as it will separate the aggregate from the resin. • When using cementitious grout, use a high frequency vibrator to remove entrapped air bubbles.

Curing

• The grout should be cured in accordance with the manufacturer's specifications. (Typically, cementitious grouts are moist cured and epoxy or corrosion resistant grouts are air cured.)

Leveling Jack Screw Removal

• Remove jack screws after the grout cures, degrease jack screw holes and, in non-corrosive areas, fill holes with silicone caulk. In corrosive areas, fill holes with a mix of silica powder and the resin used in the grout. If jack screws cannot be fully removed, retract jack screws 1/2" and saw off flush with baseplate. (This allows the baseplate to rest on the grout rather than the jack screws.) The Pump Handbook Series

• Tap the baseplate carefully to "sound out" voids with a small ball peen hammer. A small hammer will give a better ring than large hammers, bolts or other metal objects. The sounding procedure should be done after the grout has cured per manufacturer's cure time. • When tapping the baseplate with the ball peen hammer, a definite ring is heard if a void is present. The areas where the baseplate is in full contact with the grout will have a solid sound like a "lead nickel". Sometimes the sound will not be so definite. Do not assume there is a void at these locations. Occasionally the cross bracing under the baseplate will give a different sound than in the center of the steel plate. Vertical gusset reinforcement near motor or driven mounts will sometimes ring as you tap the base. Do not confuse these sounds as voids. If a void is discovered, contact the design engineer for proper repair. Injection repair of voids can distort baseplate if not vented and done correctly. Part 2 of this article will examine the successful implementation of the progam and its results. N ew computer programs enable maintenance engineers to focus effectively on repair needs. Part 1 described computerized maintenance, program reliability implementation and standards development as the keys to improving the reliability of 91 ASME B73.1M pumps. Part 2 concludes with a discussion of actual program implementation, expectations and operating and financial results.

PART II ROAD TO SUCCESS As part of the program, several problem pumps in two test areas were reinstalled with satisfactory to exceptional results. Engineering management learned of the Rotating Equipment Team’s work to shift from using free-standing designs and allow for longer straight-run

piping, an approach that conflicted with the Capital Team’s goal of low cost. The Rotating Equipment Team was asked to produce data which supported moving away from lowcost free-standing baseplate designs and to use proper but more expensive piping design (Table 7). The data were collected and presented. By reinstalling the test pumps, a failure rate of 9.38 per year per pump was reduced to 0.88. Management not only supported these efforts but commissioned the Rotating Equipment Team to persuade people in operations and capital projects that change was in the best interest of the company. However, it took several years and much more data to accomplish this.

PUMP RELIABILITY PROGRAM Some of the first pump rebuilds were conducted in Division B. Promising results prompted requests to rebuild the 1200 pumps in that division to the new standards. The area maintenance department formed a team from maintenance and engineering to create a process to control the requests. They developed the "Green Sheet," which is simply a green piece of paper that attempts to put boundaries on the rebuild process (Figure 5). Maintenance workers identified problem pumps using WARP (Work order Analysis and RePorting) and collected as much information as possible, recording it on the Green Sheet. The sheet was then sent to a production engineer so operational data such as flow, head, fluid, vapor pressure, hours of operation and instances of deadheading could be added to the sheet. The sheet was then forwarded to Plant Engineering people who determined whether existing pumps were the correct size. Not all pumps were rebuilt based on number of work orders and cost history. Many were redone because the existing pump and baseplate were non-repairable and had to be replaced.

PUMP SIZING AND EDUCATION While the Green Sheet process was being used, training in pump reliability was also being addressed. Pump sizing, installation and reliability were important, but training

Table 7. Department A pump reliability data

was definitely needed. This became a turning point in the company’s pump knowledge. Over the last couple of years, the Rotating Equipment Team has sponsored in-house training on baseplate design, up-to-date grout techniques, piping, and pump sizing and operation. The Plant Maintenance Division also formed a team with a mission to train maintenance mechanics in the concepts of pump and seal operation to solve pump reliability problems using data, a pump troubleshooting matrix and a seal selection guide. They also developed training specifically for operators to change operating practices that cause pump reliability problems. The team of six mechanics and

The Pump Handbook Series

three engineers created a diagnostic class especially for maintenance mechanics. Without the buy-in and involvement of mechanics, pump reliability on a large scale would not work. The class teaches what happens to flow, head and horsepower when an oversized impeller is installed. It also indicates what happens if motor speed is increased; if a pump is deadheaded; how line sizes are chosen; what goes into sizing a pump; how to install a pump; and how to troubleshoot a pump. The course is also excellent for engineers inexperienced with pumps. To date, more than 800 mechanics and 150 engineers have been shown the proper techniques of pump installation and operation. Such instruction was

171

SEAL SPECIFICATIONS:

Figure 5. Pump data collection sheet

essential in changing the organizational culture.

RESULTS The results of class one alignments, good baseplate and piping design, up-to-date grout techniques and pump theory classes have been excellent. Based on successes with the first several installations, more than 600 pumps have now been installed on capital projects or reinstalled since 1991. To verify the standards work, studies were conducted over the last four years on 91 pumps in one operating area that were installed to the new standards. Each pump had to be visited to grade it on installation and verify nameplate information. Therefore, not all pumps were included because of the amount of work

172

involved. The pumps included in the study (Table 8) are listed by the MMIS (Maintenance Management Information System) identification number. Also listed is the pump size, horsepower, rpm, rebuild date or installed date, a pipe and baseplate rating and the number of seal, bearing, coupling or gasket failures in 1991, 1992, 1993 and 1994. The base and piping were rated on a scale from 1 to 5, with 1 being the worst and 5 the best. If the base was free-standing, it was scored a 1, and if it was grouted by the standard, it was given a 5. But if the grout application was poor, the rating would be 3. Similarly, if the piping had elbows, valves or obvious strain, the rating would be 1, and if the piping was installed to the stan-

The Pump Handbook Series

dards, it received a 5. Not all the pumps in the study were installed or reinstalled in 1991. Some were done in 1992, others in 1993. Each work order was read to determine if it was related to the pump's reliability. All work orders not related to the pump were eliminated. Work orders to insulate, steam trace and replace valves were not counted. Only work orders on seals, head gaskets, bearings and couplings were counted. Some installation dates could not be absolutely identified and therefore were left blank. Exact records were not kept on hydraulic fit of the pumps. But more than half of them were checked for hydraulic fit. Most that were checked required smaller impellers, and in a few cases a smaller pump. Although precise records were not maintained, it is believed that adjusting the hydraulic fit helped increase Mean Time Between Failures (MTBF). In 1991 only 79 of the 91 pumps existed. They had 150 work orders, or 1.89 work orders per pump per year, or one work order every 6.32 months. With the application of the new pump installation standards in 1994, 91 pumps including the original 79 had only 31 work orders. This corresponds to 0.34 work orders per pump per year, or one work order per pump every 35.2 months. This is an 82% increase in pump life. Pump vibration as a result of this study was also reduced greatly. For instance, the 81-WASH-G-DF15 pump vibration changed significantly, as seen in Table 9, before and after. And a smoother running pump before the re-installation, 81-WASH-GDF18, ran even smoother. On the average, pumps installed to the new standards run less than 0.1 in/sec. One tank farm with 21 pumps, most of which have free-standing baseplates and poor piping, averages 0.178 in/sec overall vibration, where a group of 40 grouted pumps with good piping averages 0.071 in/sec (Table 10). Pump work orders in the division dropped from 3206 in 1991 to 2770 in 1994. Even though more pumps were added to the population, work orders decreased 14% (Table 11). Approximately 9% of the pumps in the division were installed to the

REBUILT PUMPS

REBUILT PUMPS MMIS ID#

SIZE

HP

DATE BASE RPM REBUILT RATING

The Pump Handbook Series

173

Table 8. Pump failure spread sheets

PIPE FAILURES PER YEAR RATING 1991 1992 1993 1994

MMIS ID#

SIZE

HP

DATE BASE RPM REBUILT RATING

PIPE FAILURES PER YEAR RATING 1991 1992 1993 1994

new standards, with a 14% reduction in overall work orders. Managers consider this a major accomplishment.

COST ANALYSIS When management was first approached in 1991 regarding the proposed baseplate, grout and piping standards, data showed that a typical single seal failure for 316 stainless steel construction cost about $1,000. Assuming that a twoyear seal life instead of six months was possible, and using the $1,000 cost per seal failure, $3,000 would be saved in two years or $7,500 in 5 years. To realize a 15% return on capital investment with a five year life, engineering could justify spending $4,400 to reduce the seal failures. An estimate in 1991 to install a pump with the proper piping and the new style baseplate with grout on a capital job was about $3,000 more than previous installation costs. The analysis helped convince management to adopt the standards. A more recent analysis of WARP data (Table 12) shows an average single seal failure today

Table 9. Pump G-DF15 vibration data

Table 11. Overall work orders in Division B

Table 12. Average seal repair cost

Table 10. Vibration in poor installation vs. good installation

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The Pump Handbook Series

Table 13: Average pump reinstallation cost

costs $1,970. Assuming a three year seal life instead of six months, and using $1,970 cost per seal failure, $9,850 will be saved in three years and $16,416 would be saved in five years. For a 15% return on a noncapital investment with a five year life, maintenance can justify spending $11,600 to reduce seal failures on existing pumps with seal life of 6 months. A WARP analysis of the average cost to retrofit an existing pump to the proposed standards is $11,576 (Table 13). Managers typically approve capital investment or maintenance expense projects if there is a 15% or greater return on investment. Collecting data that meets or exceeds a 15% return on investment over five years was a key factor in acquiring the money and adopting the standards. Implementing the new standards on the studied pumps enabled the company to avoid at least 119 work orders in 1994. Applying the average seal failure cost data (Table 12) to these situations a 1994 savings of $234,430 was calculated. This maintenance money is now used for other reliability projects in the department.

CONCLUSIONS A computerized maintenance management software program was needed to determine that pumps generated the highest number of work orders. This created an atmosphere for change. The Rotating Equipment Team then worked with maintenance staff to pilot test a few pumps to identify best practices learned through training. Astounding results were documented and presented to management several times over the past few years. Persistence and thoroughness in data collection and documentation created systemic change and decreased seal, bearing, coupling and gasket work orders 79.3% on 91 pumps. Today the increased cost to install a pump the correct way on a capital project is between $4,000 and $5,000. Reinstallation, although justifiable, is more than twice that amount. This becomes another rea-

Photos 1 and 2. Two examples of good piping and base construction

son to spend the money during the capital phase of the project. In 1991 two-year seal life was a goal. Today three-year seal life is a reality. As more run time is experienced on these properly installed pumps, four- and five-year seal life may become a reality. ■ Steven J. Hrivnak is a senior mechanical engineer at Eastman Chemical Company's Tennessee facility with responsibilities in plant and maintenance engineering focusing on

The Pump Handbook Series

rotating equipment reliability. He is a Registered Professional Engineer and has a B.S. degree in mechanical engineering from Virginia Polytechnic Institute and State University. Editor's Note: Reproduced with permission of the Turbomachinery Laboratory, from Proceedings of the 13th International Pump Users Symposium, Turbomachinery Laboratory, Texas A&M University, College Station, TX, pp. 115-124, Copyright 1996.

175

✓

R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

Proper Repairs Avert Failures By William E. (Ed) Nelson entrifugal pumps often fail prematurely because of poor installation design or because repair activities do not correct the causes of failure. Pump users need repair guidance because maintenance is the most expensive aspect of turbomachinery utilization. Furthermore, the hazardous nature of hydrocarbon processing has prompted the government to require that repair facilities maintain detailed maintenance instructions based on type of pump and manufacturer. This series of articles is designed to help users develop their own much-needed maintenance guidelines. Pumps encountered during repair activities are made by diverse manufacturers and represent standards of different vintages. Two of the most common standards used by the hydrocarbon processing industries were developed by the American Petroleum Institute (API) and the American National Standards Institute (ANSI). American Petroleum Institute (API) pumps are designed expressly to handle hydrocarbon liquids. They normally have carbon steel pressure containing parts, closed impellers keyed to the shaft and confined casing gaskets. While guidelines and practices vary slightly among owner companies, API pumps are normally found in services with pumping temperatures above 350ºF and/or working pressures above 350 psi. Most of our discussion focuses on API type pumps. American National Standards Institute or ANSI pumps are designed to handle chemical process liquids. Built to ANSI standard B73.1, 2 or 3, they normally have ductile iron or stainless steel pressure containing parts, compressed casing gaskets and open impellers screwed to the shaft. Generally, they are used with operating

C

176

service temperatures below 350ºF and working pressures below about 300 psi.

DESIGN DIFFERENCES Problems encountered in pump repair often involve design differences. For instance, a radially split pump as shown in Figure 1 departs from the design practices of the axially split pump. API standard 610 requires radially split pump casings under the following conditions: • when pumping temperatures are above 400ºF (a lower limit should be used if thermal shock is likely) • when pumpage is flammable or toxic with specific gravities less than 0.07 • when pumpage is flammable or toxic at rated discharge pressures greater than 1000 psig Radially split pumps use metallic spiral-wound confined gaskets between casing sections, Figure 2, whereas the gasket on an axially split joint is not confined. Thus, the probability of a proper and continuous seal is greater with a radially split casing than with an axially

split design. Also, conversion of the filler in the spiral-wound gasket to nonasbestos materials has had less impact on the performance of the gasket than the compressed sheet gasket material.

RADIALLY SPLIT PUMP PROBLEMS Problems inherent in radially split cases can be troublesome if not recognized promptly, Figure 1. Bearing alignment. In radially split designs the bearing brackets are bolted to the casing or heads. Both the bearing bracket and the head or heads are removed during disassembly. Thus all internal alignment must be reestablished each time maintenance work is performed. This is a time-consuming and exacting procedure that is not spelled out well in most maintenance manuals. Inner casing alignment. The liquid path of a radially split pump consists of inner case guides contained within the pressure walls of the outer casing. Alignment of the inner casing can be poor. In some single stage designs the guide is cast into the head. In other designs it is a separate piece.

HEAD

GASKET

GASKET

Figure 1. Radially split API design pump The Pump Handbook Series

BEARING BRACKET

Internal leakage. The high pressure (discharge) and low pressure (suction) compartments of both single and multistage pumps are separated by a single invisible gasket. If the inlet guides or the bundle do not fit tightly in the case bore, the gasket may not function properly, permitting internal leakage. Also, the gasket may be damaged as it passes across the suction nozzle opening in the case — a condition that is hard to avoid and equally hard to detect. Furthermore, it is difficult to compress both the external and internal gaskets the proper amount by tightening the head. Thermal distortion of case. Heat distortion imposes a severe strain on a pump shaft and bearings and usually results in permanent damage. Uneven heat expansion of the pump case is the most serious cause of mechanical failures in hotservice pumps. Case distortion due to pipe strain. Because the pump casing carries the bearing housings, any distortion of the outer casing due to excessive pipe strain is reflected in Metallic spiral

Nonmetallic filler

Figure 2. Cross section of spiral-wound pump head gaskets

the location of the rotating element centerline as established by the bearings. Misuse of dowel pins. Internal alignment problems can arise after a severe pump bearing failure. Frequently, replacement shafts and the bearing bracket bore clearances are built up with weld metal and machined. Proper running clearances are established for the shaft, Conditions found on first 27 pumps aligned: 60%

Bolt Bound

11%

The Motor Base Was Too High

89%

Were Free-Standing Design

100%

Had Either Pipe or Conduit Strain

100%

Motors Without Alignment Jack Bolts

Table 4. Alignment conditions found

bearings and deflector in the shop. If the old dowels were used for final positioning of the bearing housing, misalignment of the bearing housing with the shaft can result. The can lead to a loss of internal clearances and to bearing failures. Vibration. An unbalanced pump rotor caused by a shift in the center of mass of loose impellers produces vibration. Loosely fitting nonrotating diffuser sections and rotating impellers can generate vibrational waves that travel freely through a pump, its piping and foundation.

PROCESS PUMP GASKETS Radially split casings. Spiralwound pump head gaskets consist of V-shaped pre-formed plies of metal wound in a spiral with a soft separation of nonmetallic fiber, Figure 2. The V shape offers spring-like characteristics as the metal and soft filler plies flow into gasket surface finish irregularities to provide sealing. The inner and outer metal-tometal plies must be under equal compression. The compressibility of a spiral-wound gasket is controlled for a specific bolt loading of 30,000 or 45,000 psi by varying the number of metal-fiber wraps. The standard gasket is good for temperatures up to 750ºF and should not be reused. Most pump head gaskets are 0.175" thick and should be compressed to 0.130 +/-0.005". Because the ID of the spiralwound gasket decreases during compression, a clearance of 1/16" on the ID is needed to permit equal compression of all plies. The OD can fit in the head recess more snugly, but a tight fit is not allowed. Axially split casings. The ban on using asbestos as a gasket material has caused serious problems because no substitute for asbestos has demonstrated clear superiority to date. The major deficiency in nonasbestos gaskets is long-term high temperature stress-relaxation. This tendency necessitates frequent bolt tightening, and there is risk of gasket blowout. A wide variation exists in the quality and ruggedness of available axially split casing joints. The normal practice with these pumps is to machine the bore with a 0.030" The Pump Handbook Series

shim between the halves of the casing. The shim is discarded after boring. A 1/32" compressed sheet gasket is used between the flanges. This material has a thickness of 0.030-0.038" as manufactured and a compressed thickness of 0.0250.032". With some nonasbestos substitute gasket material, the compressed thickness can be less than the desired amount. The gasket is unconfined. Therefore, the proper bolting procedure and sequence must be followed carefully to ensure proper sealing, particularly on multistage pumps and in high pressure applications. The tightness of a horizontally split casing joint can be lost if the pump is thermally shocked (sudden introduction of a hot fluid into a cold pump), and as a result the unconfined gasket can blow out and cause a major leak. The tendency of asbestos substitutes to relax under stress increases this problem. The outer portion of the upper casing flange of some horizontal split pumps appears to be distorted because the flange is not flat by several thousandths of an inch. The appearance of the casing is not distorted. The mating surface on the top half is machined on a taper as shown in Figure 3 so that as the casing bolts are tightened the gasket becomes fully compressed in the critical areas. The flange is given a crown so that the outer edge on one side (either side) is tapered 0.0005" per inch of distance from the edge of stuffing box bore to the outside edge of the flange, with a minimum of 0.007" on small pumps. This method is generally used on pumps designed for high pressure and high temperature, such as those in boiler feedwater service. The casing should not be remachined to remove the taper.

INSPECTION PRIOR TO PUMP REMOVAL If the pump needing repair is an old one, compare it to designs currently being specified. An inherent hydraulic problem may be causing the failures. Thermal growth misconceptions can result in improper alignment targets, causing vibration

177

"X"

Stuffingbox Bore

.0005 Per Inch Taper x Length "X"

Figure 3. Crowned faces on parting flanges – multistage axially split pump

applying pumps. The Seventh Edition also recognized Gap A and Gap B impeller-to-casing clearances and provides limits for impeller applications. The suction piping and valves. This is unfavorable dynamic behavior of piping due to loads from dynamic, static or thermal causes including resonance excitation. Instrumentation for control of pump flow. Control system/pump interaction during startups or other periods of low flow can cause pressure pulsations. High pressure pulsations can occur due to hydraulic instability of the entire pumping system. Alignment anchoring devices. Once alignment is established, dowels or other devices hold the pump alignment. The Seventh Edi-

problems. Before removing the pump for repairs, evaluate the following components of the pumping system. The foundation. Many API pump baseplates were poorly designed prior to the 1981 Sixth Edition of API 610. The current edition states that "the pump and its baseplate shall be constructed with sufficient structural stiffness to limit displacement of the pump shaft at the drive end of the shaft or at the register fit of the coupling hub to the values shown in Table 8 of the Specification." Inspect the baseplate of the failed pump for signs of poor installation. The driver. Excitations from the vibrations of the driver (motor, steam turbine, gearing) can be transmitted to other components. Two-pole (3600 rpm) motors in particular are prone to produce mechanically and elec- Figure 4. Typical head flow characteristics of a centrifugal pump indicating several flow rates trically induced vibrations. Mechanical power transmission. Excitations may come from tion of the standard states that the coupling area, especially due to "Cylindrical dowels or rabbeted fits misalignment of the driver or eccenshall be employed to align casing trically bored coupling hubs. This is components, or the casing and covincorrect positioning of the driver er, and to facilitate disassembly and and pump such that the distance reassembly." The thermal growth between shaft ends (DBSE) exceeds guiding system of the pump must be the axial flexing limits of the coueffective. pling. These areas are aspects of pump The driven pump. Design of installation that are external to the the pump greatly influences the rotating element that is normally hydraulic interaction between the removed for repairs. When not proprotor and the casing and thus their erly evaluated, they can result in problems. The Sixth Edition of API repeat failures. 610 and earlier specifications did not OPEN PUMP INSPECTION address the possible need for an NPSH margin over the NPSHR. Nor Before opening, check the area was the concept of suction specific of the pump curve to determine speed and its effect on the range of where the pump is actually being stable flow a factor in selecting and operated. Liquid flow in the impeller

178

The Pump Handbook Series

internal channels is a very complex phenomenon, especially at off-design operating conditions. The internal flow is unstable and unsteady, with violent changes sometimes occurring from channel to channel. The curve shown in Figure 4 tabulates the turbulent flow problems encountered in various operating ranges of a pump. All such problems can be detrimental to the mechanical aspects of the pumping system and could cause pump failure. The problems accumulate to destructive proportions at low flows. Examine the liquid channels of the impeller(s), the casing and adjacent piping for signs of the failure cause. Flow disturbances can produce: Impeller erosion. Turbulent flows can cause severe erosion of the impeller vanes on both the leading and trailing edges in a very short time (less than nine months). Impeller failure. Low frequency hydraulic pulsations may cause fracture-type failure of the shrouds or covers of the impeller. These pressure pulsations can be in the magnitude of up to 20% of the total head at a frequency of about 1.0 to 10 Hz, with less than 6.0 Hz being most common. An impeller failure may occur after only a few hours of operation at low flows. High failure rate of mechanical seals. Hydraulic pressure pulsations in impeller channels are very destructive to mechanical seals and may cause opening of the primary sealing faces. High bearing failure rate. The pump rotor is moved axially by the impeller-induced hydraulic pulsations and can cause impact failure of bearings, particularly ball thrust bearings. In double suction pumps the pulsations are out of phase on each side of the impeller and can be at varying frequencies, causing the rotor to shuttle back and forth.

IMPELLER ATTACHMENT METHODS In single-stage single-suction API pumps the impeller is keyed to the shaft to prevent rotation. Both the impeller and a "hook" type sleeve are held in place axially by a fastener threaded in a direction that

is counter to pump rotation. This reduces chances of the fastener loosening. The design incorporates a square key to drive the impeller, a Woodruff key to drive the shaft sleeve and a socket head cap screw fastener to secure the assembly. Flow disturbances can also loosen fasteners – a situation that leads to unsecured impellers, a poor impeller hub bore-to-shaft fit and leakage under the shaft sleeve. The shaft and/or impeller hub bore will then fret. In this instance, the gasket fit faces of the shaft and hook type shaft sleeve may have to be trued up. The actual fastening method is not clear in most pump drawings, but standard items purchased in bulk by the pump manufacturer are generally used. The material can be too hard ( >RC 25) to be compatible with many pumped products such as chlorides. The fastener may then break during a run. This is especially common with socket head fasteners that are countersunk in contoured cones. Standard socket head cap screws are in the RC 40 range. Lock washers used under the head of hex headed cap screws are also hardened and susceptible to failure in some working environments. Many ANSI impellers are screwed onto or into the end of the shaft without a key. The screwed impeller is particularly vulnerable to backward rotation of the driver or a backflow of product.

SINGLE-STAGE DOUBLE-SUCTION BRONZE IMPELLERS When removing a hollow castbronze impeller, remove the plug in the passageway leading to the hollow area before heating to allow water/steam to escape without causing a rupture or explosion. If the impeller has been in hydrocarbon service, remove the hydrocarbon before heating, Figure 5. If it is necessary to straighten bent impeller shrouds or vanes, do so when they are cold. Hot bronze is apt to shatter if hammered or shocked. Likewise, do not try to remove a heated bronze impeller by hammering, as it may crack or shatter. Use a steady force such as a press or a strong-back.

cutback on the impeller diameter because:

Figure 5. Venting of bronze impeller

EFFECTS OF GAP A AND GAP B Careful machining of the volute or diffuser tips to increase Gap B (vane to volute or diffuser tip) while maintaining Gap A (shroud to casing) has been used for some time to reduce vane-passing frequency vibration. The pulsating hydraulic forces acting on the impeller can be reduced 80-85% by increasing the radial gap from 1% to 6%. There is no overall loss of pump efficiency when the diffuser or volute inlet tips are recessed. Some slight efficiency improvement results from reducing various energy-consuming phenomena: the high noise level, shock and vibrations caused by vanepassing frequency, and the stall generated at the diffuser inlet. Recommended dimensions for radial gaps of the impeller to casing are given in Table 1. Pump impeller trimming. The head and flow developed by a centrifugal pump can be adjusted by trimming the impeller diameter. The resulting performance of the Type Pump Design

Gap A

Diffuser Volute

50 mils 50 mils

Gap B - Percentage of Impeller Radius Minimum

*B=100 (R3 - R2) R2

4% 6%

Preferred

6% 10%

Maximum

12% 12%

R3 = Radius of diffuser or volute inlet R2 = Radius of impeller

NOTE: If the number of impeller vanes and the number of diffuser/volute vanes are both even, the radial gap must be larger by about 4%. Source: Dr. Elemer Makay

Table 1. Recommended radial gaps for pumps

pump is the subject of much confusion, however. Other publications offer guidelines on the effects of impeller trimming on head, NPSH, flow, efficiency and vibration. Calculations based on "affinity laws" generally dictate too great a The Pump Handbook Series

• The "affinity laws" assume that the impeller shrouds are parallel; actually, the shrouds are parallel only in lower specific speed pumps. • The liquid exit angle is altered as the impeller is trimmed, so the head curve steepens. • There is increased turbulence at the vane tips as the impeller is trimmed due to increased shroud-to-casing clearance or Gap A. • Impeller diameter reductions for radial designs should be limited to about 70% of the maximum diameter; for pumps of higher specific speed values (2500-4000), trimming should be limited to about 90% of the maximum impeller diameter. What to trim. No hard and fast guidelines exist for the mechanical aspects of impeller trimming. For volute type pumps it is best to cut the impeller vanes obliquely as shown in Figure 6, leaving the shrouds unchanged, or cut the vanes only, as in Figure 7. Either method tends to even out the exit flow pattern and reduce recirculation tendencies in the exit area. Gap A should be about 0.050" (radial) for minimum axial vibration. In most diffuser type pumps it is best to trim only the vanes in order to control tip recirculation and the ill effects of an increased Gap A. This cut yields a more stable head curve because flow is more evenly distributed in the exit area. Structural strength of the shrouds is a factor in the decision. There may be too much unsupported shroud left after a major reduction in diameter. The double oblique cut of Figure 6 leaves the shrouds unchanged and solves the structural strength problem as it improves the exit flow pattern. Correcting vane shapes. Cutting impeller vanes results in blunt vane tips that cause disturbances in the volute and hydraulic "hammer" even when the impeller OD is the correct distance from the cut water (Gap B). Corrections can be made by tapering the vanes through "over-

179

Gap "B"

D’ D

Gap "B"

Gap "A"

D’ D

Gap "A"

Figure 6. Trimming impellers – oblique cuts of vanes

Gap "A"

D’ D

Gap "B"

Figure 7. Trimming impellers – cutting vanes only

Gap "A"

D’ D

Gap "B"

Figure 8. Sharpening impeller vanes by overfiling Normal Sharpening

Original thickness

Original outlet width New outlet width Mill or grind away

Max. sharpening Leave at least 1/8"

Figure 9. Sharpening impeller vanes by underfiling

180

filing" or removing metal on the leading face of the vane as shown in Figure 8. By sharpening the underside or "underfiling" edge of the vane as shown in Figure 9, the outlet area of the liquid channel can be enlarged. This will result in about 5% more head near the best efficiency point, depending on the outlet vane angle. At least 1/8" of vane tip thickness must be left. Sharpening the vanes also improves the efficiency slightly.

Figure 10. Installing wear rings by radial pinning

IMPELLER WEAR RING MOUNTING Impeller wear rings are shrunk on and locked in place to prevent movement due to rubbing or a partial seizure. Interference fit between impeller wear rings and impeller is about 0.001-0.0015" for ring diameters ranging from 2-1/2" to 6"; 0.002-0.0025" is used for rings 6-1/2" to 12" in diameter. Two methods are used to lock the rings. One is to drill and tap either radially or axially for screws at the joint between the impeller and the impeller wear ring. This is illustrated in Figures 10 and 11. After the screw is tightly bottomed, it is sawed off, peened and dressed down with a fine file. API 610 specifications limit the diameter of the pin to 1/3 the width of the ring. This method has some problems. Threads must be tight, which causes the wear ring to bulge at the pin with soft materials and may crack stellite and colmonoy hard surfacing, as in shown in Figure 12. The second method substitutes tack welding for the screws. Thermal stresses and metallurgical changes in the impeller metal must be avoided. API specifications disThe Pump Handbook Series

Figure 11. Installing wear rings by axial pinning

Figure 12. Installing wear rings – axial pinning problems

courage this method, but properly done it is a good method. After the rings are locked in place by whatever method, they should be checked for radial runout. ■

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R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

Proper Repairs Avert Failures, Part 2 By William E. (Ed) Nelson inely tuned techniques for disassembling and assembling single stage horizontal pumps can greatly minimize shop returns. And attention to details, along with the use of simple tools such as a dial indicator, can make all the difference in the effectiveness of planned maintenance programs for pumps. First, the pump shaft must run "true" to achieve proper mechanical seal performance. The shaft must be coaxial with the seal gland bore and perpendicular to any seal faces held in the gland. If the shaft is not "true" the rotating seal face will oscillate axially as it attempts to maintain face contact with the stationary member. In addition, this oscillation will cause "fretting" or chafing of the shaft sleeve under the static seal of the rotating element of the mechanical seal. Thus, checks are needed to confirm that the shaft is running true. Shaft runout. Shaft runout, deflection or lift should be checked by mounting dial indicators at each end of the bearing housing and locating the stems (Figure 1). Rotating the shaft by hand will show radial runout at the seal end. Observing both indicators will reveal whether the shaft is bent. Lightly lifting the shaft may show a greater reading than shaft runout. This indicates bearing wear or a poor bearing fit in the housing. End play. With the dial indicator mounted on the pump bearing housing and the stem located against the shoulder of the shaft or sleeve, attempt to move the shaft from end to end. End play normally will be between 0.001-0.004" (Figure 2). Different pump manufacturers allow various tolerances for thrust bearing thermal growth. Excessive end play is detrimental to the mechanical seal. Normally, hydraulic forces keep the shaft thrusting in the same direction.

F

A

Figure 2. Checking shaft end play

B

Figure 1. Checking shaft runout or deflection

Start/stop routines, however, can damage the contact faces of the seal. Sleeve concentricity. If the shaft runout is satisfactory, the shaft sleeve should be mounted in position. Allowing for tolerances on both the bore and the OD of the sleeve, the total runout should not exceed 0.004" (TIR) (Figure 3). If runout exceeds this amount, it should be corrected. Excessive runout can cause oscillations of the seal faces and variation of the fluid film thickness (a major problem in the pumping of light hydrocarbons).Vibrations can also occur, possibly causing bearing failures. Stuffing box bore concentricity. This check is frequently overlooked on between-bearing, double suction, radially split pumps. Because the bearing brackets must be removed during repairs, their position can shift. If the stuffing box bore is not concentric, the bearing brackets can be moved slightly and redowelled to center the shaft in the stuffing box. Again, TIR should not exceed 0.004" (Figure 4). Excessive runout can also cause tracking problems with the seal face, resulting in uneven wear and eventual failure. Stuffing box face squareness. A final dial indicator check is The Pump Handbook Series

Figure 3. Checking sleeve concentricity

made with the pump completely assembled but without the seal(s) installed. Mount the indicator on the shaft end and the stem located on the stuffing box face(s) (Figure 5). When the dial indicator and shaft are rotated, TIR should not exceed 0.003". If the stuffing box face is not normal to the centerline of the shaft, the stationary seal face also will not be square. This situation will cause wobbling of the rotary faces. Excessive runout can cause a loss of contact and produce fretting wear between the seal faces.

BALL BEARINGS A ball bearing is a piece of precision equipment manufactured to extremely close tolerances. To obtain maximum service from a bearing, the shaft and housing must be machined to the same exacting tolerances that are used to make the bearing. One of the bearings is fixed axially while the other is free to slide. The outboard bearing (the one closest to the coupling in a pull-out design) is fixed axially. The inboard bearing slides freely within the housing bore to accommodate thermal expansion and contraction of the shaft. Because the outboard bearing is fixed in the housing, it

181

Rules of Thumb: Bearing Housing Fits 1. Bearing OD to housing clearance. About 0.00075 inch loose with 0.0015 inch maximum. 2. Bearing housing out of round tolerance is 0.001 inch maximum. 3. Bearing housing shoulder tolerance for a thrust bearing is 0 to 0.005 inch per inch of diameter off square up to a maximum of 0.002 inch.

Table 1. Figure 4. Checking concentricity of stuffing box bore

Figure 5. Checking squareness of stuffing box face

carries both the axial and radial thrust. The numerical code used in bearing identification is fairly standard among the various manufacturers. However, the alphabetical prefixes and suffixes are not. Make sure that all code numbers and letters are clearly defined and understood when making substitutions. Ball bearing fits. Always check the bearing housing and bearing clearances during pump assembly. Unfortunately, many pump manufacturers do not indicate the proper bearing fits for either the shaft or the housings to guide shop repairs. The original dimensions of both the housing and the shaft will change from time to time due to oxidation, fretting, damage from locked bearings and other factors. Every bearing handbook has tables that aid in selecting fits. The vibrational effect of looseness on the bearing fits is different for the housing and the shaft. Housing fits. Ball bearing fits in the bearing housing are neces-

182

sarily loose for assembly. If this looseness becomes excessive, however, vibration at rotational speed and multiple frequencies will result. Do not install bearings with OD outside of the given tolerance band because this may result in either excessive or inadequate outer race looseness. Table 1 offers good guidelines for looseness. Shaft fit. A loose shaft-to-bearing bore fit will give the effect of an eccentric shaft. The objective of the shaft fit is to obtain a slight interference with the antifriction bearing inner ring when mounted on the shaft. The bearing bore should be measured to verify inner race bore dimensions. Do not install bearings with an ID outside of the given tolerance band, for this can result in either excessive or inadequate shaft tightness. Table 2 offers guidelines for bearing shaft fits.

PUMP GASKET COMPRESSION Many single stage process pumps incorporate the back pull-out feature so that the pumping element can be removed without breaking the piping connections. This construction also permits centerline mounting of the casing Rules of Thumb: Bearing Shaft Fits 1.

Fit of bearing inner race bore to shaft is 0.0005 inch tight for small sizes: 0.00075 inch tight for large sizes.

2.

Shaft shoulder tolerance for a thrust bearing is 0 to 0.0005 inch per inch of diameter off square up to a maximum of 0.001 inch.

Table 2.

The Pump Handbook Series

to minimize thermal expansion problems and allow operation at higher pressures and temperatures for hydrocarbons. One important fact cannot be overlooked in working on this type of pump: the head must be made up square. This is important for two reasons: • to obtain correct gasket compression for pressure containment • to maintain correct internal alignment and positioning of the rotating element with respect to the stationary casing components Starting in 1981, API Standard 610 specifications required that the heads be machined so that when they are pulled "metal-to-metal," the gasket is compressed the correct amount. Pumps manufactured be-fore 1981 may have a separation when the gasket is made up properly. The head can become cocked if this separation is not maintained uniformly using feeler gauges. ■ William E. (Ed) Nelson is a pump and turbomachinery consultant who retired from Amoco Oil Co. with more than 36 years of experience. He is a founding member with more than 20 years of service on the Turbomachinery Symposium Advisory Committee, has authored more than 40 technical papers and is the author of Centrifugal Pump Sourcebook (McGraw-Hill, 1992). He received a B.S. in mechanical engineering from Texas A&M University and is a Registered Professional Engineer in Texas. Editors' Note: This series of articles on proper pump repair procedures (including last month's gasket and impeller coverage) is based on a tutorial by the author entitled "Monitoring Repairs to Your Pumps, presented at the 1995 International Pump Users Symposium. The figures are reproduced with permission of the Turbomachinery Laboratory, Texas A&M University.

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R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

Proper Repairs Avert Failures, Part 3 By William E. (Ed) Nelson e now move from single stage to multistage double casing type pump repair.

W

Tool requirements. Special tools are needed to assemble a multistage double casing pump such as the one shown in Figure 1 properly. These include: • micrometers for measuring bores of individual diffuser cover bushings and wear rings • a mandrel for checking bushing alignment in all cover or diffuser sections after they are stacked • the plates and rods that compose the inner casing stacking jig • rails used to insert or extract the inner casing from the outer casing • several mandrels used to balance impellers and pressure reducing sleeve (balancing drum) • pressure reducing sleeve (balancing drum) lock nut tool • air ring used to cool the shrink fits on the impellers

SHAFT RUNOUT CHECK The grooves cut for split impeller locator rings tend to produce shaft runout. Checking this runout is important, and the procedure should include the following steps: • Check the shaft for proper

diametral fits. • If the shaft has been thermal spray or chrome coated on the journals, an area adjacent to the journals should be used to support the shaft as long as there is no runout. • Despite its symmetrical nature, the shaft must be balanced. • Prepare halfkeys for the keyways of the bare shaft. These should be carefully taped in position using high-strength fiberimpregnated tape; several turns are usually required. Note that the tape sometimes fails while spinning in the balancing machine. It is thus important that adequate tape be used to protect personnel against the hazard of flying halfkeys. • Mount the bare shaft, with halfkeys in place, in the balancing machine with the supports at the journal locations. Spin the shaft at 300-400 rpm for approximately 10 minutes. Shut down and check the radial runout (TIR) at mid-span using a 1/10 mil diameter indicator; record the angular position of the high spot runout value. Spin the bare shaft at a speed of 200-300 rpm for an additional 5 minutes. Shut down and again check the radial runout (TIR) at mid-span; record the angular position of the high spot runout value. Compare the results obtained after the 10-minute and 5minute runs; if they are the same, Discharge

Suction

the bare shaft is ready for further checking and balancing. If the results are not repetitive, additional spinning is required. This should be continued until two consecutive 5-minute runs produce essentially identical results. • Check the radial runout (TIR) of the bare shaft in at least three impeller locations approximately equidistant along the bearing span and near the shaft ends. Record the angular position of the high spots and runout values at each location. The shaft is satisfactory if both of the following conditions are satisfied: (a) the radial runout (TIR) at the section of the shaft between journals does not exceed 0.001", and (b) the radial runout (TIR) outboard of the journals does not exceed 0.0005". • With the balancing machine operating at about 800 rpm, make the required dynamic corrections to the bare shaft using wax. When balance is satisfactory, remove material at the face of the step at each end of the center cylindrical section of the shaft. Under no circumstance should material be removed from the section of the shaft outboard of the journal bearings. • If a balancing machine is not available for shaft runout checking, use the following method: (a) mount the shaft on "V" blocks at bearing areas to confirm mechanical runout; maximum runout should be 0.001", and (b) with the shaft still mounted in the blocks, check electrical runout at the probe burnished points; maximum runout 0.00025".

Thrust Bearing

PRELIMINARY ROTOR STACKING (IMPELLERS ONLY) The pump rotor must be assembled for balancing and runout checks without the diffuser sections in place. This involves the following considerations: Figure 1. A typical multistage double casing pump The Pump Handbook Series

183

• The rotor assembly should be stacked vertically with the shaft held in a special fixture. • All impellers, the thrust collar and pressure reducing sleeve (balancing drum) should be heated to 250ºF for mounting; under no circumstances should the temperature exceed 450ºF. • Measure the bore of the impellers, thrust collar and balancing drum to assure uniform expansion. • Install the match-marked split rings and key for the first stage; it may be necessary to use a dab of grease to hold these pieces in place. • Install the first stage impeller using these procedures. 1. Heat the impeller to about 250ºF. 2. Place the air ring on the OD of the impeller vane tips with the air turned off; cooling air flows through the impeller vanes to suction eye. 3. When moving the impeller into position, it will be easier to "hit" the key in the shaft if the keyway position of the impeller is marked with a felt tip pen; the keyway in the impeller isn't visible while installing it on the shaft. 4. When the impeller counter bore makes contact with the split rings, maintain constant downward pressure on the impeller and turn on the cooling air; after about 30 seconds the downward pressure can be released (note that the air must remain on until the impeller is at ambient temperature). 5. Carefully remove the rotor to "V" blocks for runout indication; with the journal areas mounted on the blocks maximum runout should be about 0.001". 6. If the rotor shows no appreciable runout changes after mounting the first impeller, remaining impellers may be mounted without individual checks. 7. Install the remaining impellers on the shaft using these same procedures. 8. After the rotor is stacked, check it for runout; if its position is satisfactory, then tighten the impeller nuts and check runout again. 9. If all checks are satisfacto-

184

ry, final balance the rotor. 10. If excessive runout is present, check the faces of the shaft nuts; reclaim the faces as necessary (all runouts must be brought into tolerance before balancing the rotor).

ROTOR ASSEMBLY BALANCING After the rotor is assembled, it is ready for final dynamic balancing. Observe the response of the balancing machine at several speeds and select the best speed. For most rotors of the lengths and diameters encountered in this style pump, 500-800 rpm is probably the best balancing speed. The following steps can help ensure that rotor assembly balancing is accomplished properly. 1. Spin the rotor and record the first run numbers. If the first run is below 25.0 gr/in of imbalance on either plane, proceed to step 2. If the imbalance exceeds this limit, follow the procedures below: • Reduce the balancing machine speed to 200 rpm. • Heat each impeller backside (discharge side) fit with a #3 Rosebud torch tip placed 1.0" from the hub bore for 30 seconds. Heat all impellers this way in succession. The rotor must be rolling at 200 rpm throughout the entire heating and cooling process. Do not stop turning the rotor or allow it to turn any faster than 200 rpm until the entire element has cooled to ambient temperature to prevent any rotor bow. • When the entire rotor is at ambient temperature ,make a balancing run at about 1100 rpm. If readings are still above 25.0 gr/in, unstack rotor to determine the cause of the excessive unbalance. If readings are below 25.0 gr/in, proceed to step 2. 2. Install suction nut hand tight and make another balance run. Readings should go down when mass is added. If the readings go up, the nut is probably out of balance (step 10). Make additional runs using putty to determine correction location and weight. Complete step 3 before drilling correction hole. Drill correction holes only in line with the existing spanThe Pump Handbook Series

ner wrench holes on the suction locknut. 3. Tighten suction nut with spanner wrench. If readings go up, fits are off, so try another nut or correct the existing nut. 4. Install pressure-reducing sleeve or balancing drum only after preheating to 250ºF. Tighten the locknut snugly to hold the pressure reducing sleeve in position until it is cool. After the element has come to ambient temperature, remove nut and make another run. Readings again should go down. Proceed to step 5. If readings go up, check to see if the pressure reducing sleeve is out of balance or if fits are off. 5. Install the pressure reducing sleeve locknut snugly and make another run. Readings should go down. If they do, proceed to step 6. If they go up, either the locknut is out of balance or the face fit to the pressure reducing sleeve is off. In this case tighten the nut and make another balance run. If the readings go up again, the fits are still off. Try another locknut or correct the face fits on this one. If readings go down but are still higher than before the nut was added to the element, make a correction in the nut. Use the same procedure as for the suction nut to determine correction location and weight. Drill correction holes only on the small-diameter fit of the pressure reducing sleeve locknut (see step 10). 6. Tighten pressure reducing sleeve locknut. Make another balancing run. If the readings go up, locknut fits are off. Try another locknut or correct fits. If the balancing run readings stay the same or go down, proceed to step 7. 7. Preheat thrust collar to 250ºF. Install it along with the locating spacer and snug the locknut firmly. Make another balancing machine run. If the balance readings go down, proceed to step 8. If they go up, try tightening the locknut some more. If imbalance readings go up further, this means the locknut faces are off. Repair the thrust collar, spacer sleeve or locknut fits. If readings go down proceed, to step 8. 8. Do a six-point residual imbalance test. Record results on

residual imbalance form and run sheet. 9. Scribe all components with an air pencil for precise reinstallation at final stacking. 10. After the rotor is dynamically balanced, reconfirm the runouts.

RULE OF THUMB DIM “Z”=DISCHARGE IMPELLER-DIFFUSER PASSAGE OVERLAP. NOTED: REMAINING PASSAGE OVERLAP POSITIONS WILL BE “Z” PLUS THERMAL GROWTH FACTOR.

“X”

Note that the suction locknut and pressure reducing sleeve locknuts are the only components on the rotor that are dynamically balanced. All other parts are precisionbalanced, and no corrections should be made on them after assembly on the rotor. If locknuts require balancing, remove only enough material to reach last balance amplitudes and phase readings (before nut was installed).

FINAL ROTOR AND DIFFUSER STACKING Use the following guidelines for final assembly of the rotor and diffusers: • Carefully unstack the rotor. ARRANGEMENT OF FOUR GASKETS AND THREE SPACERS

243

169

1233

“Z” INACTIVE THRUST BEARING

INNERHEAD EXPANSION GASKET SPACER INNERHEAD GASKET INNERHEAD GASKET RETAINER

Figure 3. Expansion gasket and spacer arrangements

COVER WEAR RING

“Y”

ACTIVE THRUST BEARING SPLIT RING PRESSURE REDUCING BUSHING PRESSURE REDUCING SLEEVE

IMP SPACER SLEEVE

LOCATOR RING

DISCH END

SUCTION END

Figure 2. Impeller-diffuser passage overlap in a multistage pump

• Vertically stack the rotor and diffuser elements using the same procedure as the impeller-only stacking elements described above. • Vertically lift check the shaft after each cover is installed to check the bumping or float distances; jack can be placed under coupling end of shaft; design lift should be maintained throughout the stacking procedure. • Because diffuser covers should have 0.005" interference fit, some heat may be needed to install them; if the cover doesn't drop into fit, heat around entire cover circumference for about one minute with a small rose bud; then lightly tap with a hammer; verify alignment of cover to outer casing key as each succeeding cover is installed while the lower cover is still warm. • After the pressure-reducing sleeve is installed and cooled, tighten nut to match-marks established during balance procedure; leave nut snug until pressure reducing sleeve or balancing drum is cool; take bump or float dimensions so that dimension "Z" (Figure 2) can be set readily by the thrust collar position. • Install inner casing stacking jig after element has cooled to ambient temperature.

PUMP ASSEMBLY

169 243 1233

GAP "A"

GAP "B"

THRUST COLLAR

• Mount the rotor on the "V" blocks at the journal areas and reconfirm the mechanical runouts; maximum runout should be about 0.001" again; check the thrust collar face runout at this time – maximum 0.0005". • Reconfirm the electrical runout at the probe burnished areas; maximum runout should be 0.00025".

GAP "A" = IMPELLER OD & DIFF/VOLUTE TIP ID GAP "B" = RADIAL GAP BETWEEN IMPELLER & DIFF/VOLUTE

RATIO X/Y - 1.15 - 1.20

At last we come to the final pump assembly elements. These include the following items. • Install the expansion gaskets and spacer between the inner casing and the outer head (Figure 3); note that various combinations of The Pump Handbook Series

gaskets and spacers are used (four gaskets and three spacers; three gaskets and two spacers). CAUTION: Ensure that the inner head gasket spaces that were supplied with the pump or exact replicas are used in reassembly or serious problems could result. Improper makeup of the intermediate diffuser covers, the suction spacer and the discharge diffuser spacer could cause liquid recirculation within the pump, and this in turn will decrease efficiency and cause erosion, particularly at the high pressure end. The thickness of each spacer should be measured, along with the gasket thicknesses, to achieve approximately 0.015-0.020" compression on each gasket (about half the compression when used as a head gasket). • Install head and tighten; most heads will go to a metal-to-metal fit internally so the gap used at the head will be approximately 0.0480.058". • Install bearing housings with the lower half journal bearing in place; move the coupling end housing until suction impeller is centered in wear rings within 0.001", using full-length feeler gauges; move thrust end housing up or down until pressure reducing sleeve is centered in the mating bushing within 0.001", again using feeler gauges; record feeler gauge clearances; take dial indicator readings on seal chamber bores and faces; record all readings. • Ream bearing housing dowel holes.

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• Install thrust collar with the locating spacer and the inboard thrust pads to check rotor position (the locating spacer should be machined to control final position). • Check the journal bearing shell "crush" or "clamp" by adding a 0.005" shim to the bearing cap gasket (normally 1/64" thick). • Remove the bearing housings; install mechanical seal assemblies and leave them loose on the shaft until bearings and thrust are finally set. • Machine the bearing cover or add gasket thickness to control thrust bearing clearance or float to

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about 0.010-0.012". Having covered both single and multistage horizontal pump repair, we will discuss special problems presented by vertical pump repair in Part 4. ■ William E. (Ed) Nelson is a pump and turbomachinery consultant who retired from Amoco Oil Co. with more than 36 years of experience. He is a founding member with more than 20 years of service on the Turbomachinery Symposium Advisory Committee, has authored more than 40 technical papers and is the author of Centrifugal Pump Sourcebook (McGraw-Hill,

The Pump Handbook Series

1992). He received a B.S. in mechanical engineering from Texas A&M University and is a Registered Professional Engineer in Texas. Editors' Note: This series of articles on proper pump repair procedures (including last month's gasket and impeller coverage) is based on a tutorial by the author entitled "Monitoring Repairs to Your Pumps, presented at the 1995 International Pump Users Symposium. The figures are reproduced with permission of the Turbomachinery Laboratory, Texas A&M University.

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R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

Proper Repairs Avert Failures, Part 4 By William E. (Ed) Nelson ertical pumps present repair and maintenance problems that are different from those of horizontal pumps, which were discussed in Parts 1, 2 and 3. Many more parts are required to rebuild a vertical turbine pump because, typically, four stages of this mixed flow design are required to produce the head of one stage of the radial flow horizontal pump. With more wetted parts, a vertical process pump composed of many alloys is also more expensive than an equivalent horizontal process pump. Vertical pumps usually consist of three major components as shown in Figure 1. Head and driver assembly. This consists of an electric motor and a cast or fabricated base from which the column and bowl assembly are suspended. Column and shaft assembly. The column pipe that suspends the pump bowl assembly from the head assembly serves as the conductor for the liquid from the pump bowl to the discharge. Within the column pipe is the line shaft that transmits the power from the driver to the pump impellers. The line shaft bearings in the columns are generally lubricated by the liquid being pumped. Pump bowl assembly, the pump proper. Each bowl unit or stage contains an impeller.

V

VERTICAL PUMP PROBLEMS The vertical pump's rotor is not gravity stabilized. When problems arise, the gyroscopic effect of rotation can cause significant damage to the rotor and the casing. Large cost savings are possible if repairs offer improved reliability and extended runs. Critical speeds in vertical pumps. Critical speeds are not usually a factor in centrifugal pumps. The stiffening and dampening effects that result from a pressure

drop across the narrow spaces between wear rings and interstage bushings are significant and raise the critical speed far above what is known as the "dry" value. The bowl assembly may have a high "wet" critical, but the column assembly can have a very low critical (Figure 1). The only liquid force influencing this column shaft is a small amount of "hydrodynamic wedge" in the bushings. Vertical pumps are manufactured with large spaces between the shaft guide bushings. This reduces the wet criticals of the shafting to below operating speeds. The resulting "whip action" drastically increases the chance of seal failures and other mechanical problems due to shaft runout. Problem pumps can be helped by shortening the bearing span. Column spacing between shaft guide bushings should not exceed values indicated in paragraph 2.9.2.1 of API-610, (Figure 2). Adherence to these guidelines should keep the first critical of the drive shafting at least 130-150% higher than its operating speed and thus allow for the effects of wear on the bushings. Startup considerations — upthrust. When a vertical pump is started, it has a momentary upthrust hydraulic action. The equipment including the motor must be designed to take the upthrust. When a vertical pump is operated at very high capacities continuously, it will likely have a continuous upthrust. Resulting damage can include buckled line shafts, seal leakage, impeller rubs and motor bearing failures. Impeller locking devices. Five types of locking devices are used to position impellers on the shaft. These generally determine how the shaft is disassembled. • split collet – normally limited to smaller pump impeller diameThe Pump Handbook Series

ters (less than 6-8") • locknuts – mounted on each end of shaft and spacers between each impeller • "Gib" key – an ell-shaped key fitted in a stopped keyway • spring ring – a snap ring located in grooves in the shaft • threaded pin – a pin that goes through the impeller hub into the shaft • split rings – stops located in a wider groove and retained by a shrunk-on or bolted-on keeper ring Bearing troubles. The pumpage-lubricated lower bearing of a vertical turbine pump is especially vulnerable because there is almost no pressure differential across the bushing to provide lubrication. A flushing line from the discharge as shown in Figure 3 will extend the pump life considerably. If the fluid being pumped contains abrasive material, a clean fluid from another source can be used if the process fluid will not be contaminated by the flushing fluid. A cyclone separator such as that used on mechanical seals can be used to clean the pump discharge so that it can be used for flush. External piping must be added.

BUSHING AND BEARING MATERIALS Before replacing the bushings, evaluate the original material selection and its suitability for your current application. • Bronze – probably the most common material used, it should not be employed in any corrosive service, above 200ºF or below -20ºF, or in sandy dirty water sump service. •Carbon – normally the second choice after bronze, it should be used for corrosive products and when product lubricity is low and temperature is below -20ºF; do not use for dirty pumpage

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and remember that its coefficient of thermal expansion is very low. • Cast iron – use in mildly corrosive service (sour crude), when temperatures are above 200ºF, when particulate matter is present in the product and when lubricity is fair to good. • Rubber – applicable only to water pumps; use in conjunction with open impellers in sandy water or dirty sump service; do not use in hot service or where products detrimental to neoprene are present. • Teflon, glass filled or other – use only as a last resort and with engineering guidance; carbon is the best alternative in services where Teflon is specified. • Metallic filled carbon – a graphite that has been filled with a metal – typically copper, nickel, babbitt or powdered iron – to alter its characteristics are the alloying metals; commonly referred to by the trade name "graphalloy."

Electric Motor or Other Approved or Listed Driver

Discharge Assembly

CRITICAL CAN BE IN RANGE OF RUNNING SPEED

Top Shaft

Column Pipe

Bearing Retainer

BEARING SPAN Shaft Coupling Impeller Shaft

VERY HIGH CRITICAL

FLUSHING FLUID UNDER PRESSURE

Bowl Assembly

COLLETS BEARINGS

Suction Manifold

MAXIMUM SPACING (inches)

Figure 1. Rotor dynamics of a typical vertical pump 100 90 80 70

0 rp 120 0 rp 180

60

m

m

50

m 0 rp 360

45

Figure 3. Vertical pump bowl designed for flushing fluid

35 30

PUMP DISASSEMBLY

25 20 1.0

1.5

2.0

2.5

3.0

3.5

SHAFT DIAMETER (inches)

Figure 2. Maximum spacing between shaft guide bushings for vertical pumps

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The Pump Handbook Series

4.0

The basic construction principles of a vertical pump determine the appropriate repair methods. Some of the differences actually indicate from which end disassem-

bly should begin. Pump bowls may be : 1) individually screwed together, 2) bolted together with a series of bolts around the bowl flange or 3) bolted together with tie bolts that reach from the bottom of the pump to the top through the bowls (sometimes called channels) or case. Method 2, using a bolt circle in a flange, is encountered most often.

DISASSEMBLY PROCESS Before beginning disassembly, prepare a data sheet for the pump. Record the name plate information, noting any instructions regarding "lift" and other items pertaining to the particular pump. Then complete the following steps. • Set the pump on the floor with the discharge flange facing down. Grind flats at all mating joints in the discharge column and the pump bowls, 180º from the discharge. Starting at the bottom or suction, match-mark all joints at the ground flats and record stencil configuration as to letter height, etc. Don't confuse previous markings with current ones. • With the pump thrusting toward the suction, extend the upper shaft length to the motor adapter face and adapter face configuration. • Measure and record total pump shaft float. • When disassembling a pump that has a split collet bushing to lock the impeller on the shaft, measure its distance from the end of the shaft to the impeller and record this for reassembly. • The bowl and impeller matchmarks and numbers should be carefully recorded during pump disassembly for proper reassembly. Measure and record the float at each stage. • Match-mark the impellers and the collets (or other mounting devices such as keys, locking pins, split keepers and keeper plates) before removing the impellers from the collets and shaft in order to maintain their relative positions. This will help in machining new wear rings and in positioning the impellers for balancing. If this is not done, past

experience shows that the impeller skirt will run out. When reinstalling the impeller on the collet, be sure that the matchmarks are lined up. • Match-mark line shaft couplings to their mating shaft on each end. Important: Line shaft couplings should be removed using pipe or chain tongs on the mating shafts only. Wrenches applied to the coupling, can collapse the coupling causing it to seize on the shaft. If normal efforts (including heating) to remove the line shaft coupling fail, split the coupling with a cutoff wheel in a grinder. This enables you to salvage the more expensive pump shaft sections. • Check the rigid coupling halves from the motor and the pump to be sure they run true. The spacer spool should also be checked for squareness. All couplings should be steel. Cast iron is subject to frequent failures. • Check the shaft for straightness after complete disassembly of the pump.

AFTER DISASSEMBLY • Take all clearance dimensions (i.e., impeller wear ring to bowl or channel, spider bushing to shaft, throat bushing to shaft). • Properly size all wear ring and bushing clearances using the clearances specified by API 610, correcting as needed for product temperature. • Check all bearing spider bores for parallelism with the shaft column. • Lock in all bushings in the spider using a method consistent with the material and design of the bushing. Most vertical pumps have brass throat bushings and brass spider bushings. • Because of the tightness of their fit the shaft, all bushings should have a spiral groove on the bore to provide for lubrication. • Most vertical pumps with bottom suction (or normal flow pattern) should have a sand collar and a plug in the bottom bowl. • All impellers should be individually balanced to compensate for uneven wear of the casting, pieces broken out of the vanes, etc. The Pump Handbook Series

• Where possible, a one-piece shaft is preferred. If couplings are needed, a radial weep hole should be drilled in the center of the coupling. Screw the shaft together before assembling the impellers to check the shafts for runout. This can be done on knife edge rollers. • A sleeve shrunk on the shaft in the area of the spider bushings or an eccentric spider bushing can be used instead of shaft replacement where shaft wear is a problem. • Wear rings should be grooved in the direction of rotation, starting on the outboard side of the wear ring and threaded toward the impeller. A coarse pitch with two or more leads will reduce leakage across wear rings by as much as 35%.

SHAFT INSPECTION The pump section and the drive shaft should be supported in a horizontal position on knife edge rollers for checking. Sag. This is the vertical movement or deflection of the shaft on the rollers resulting from gravity, a flexing under its own weight. The condition is not permanent, however, due to the elasticity of the material. If the shaft is rotated, the sag will move to a new spot. Sag is neither runout nor shaft bow. Shaft bow. This is the amount of permanent deflection of a shaft beyond the elastic limit of a material. The shaft has permanently yielded to an external force. Unlike sag, this bow will not relax or return to a new position as the shaft is rotated. The deflection must be removed by either thermal or mechanical means. Shaft runout. This is the measure of shaft roundness combined with any permanent bow or bend. The effect of runout is greatest on a mechanical seal or packing due to the orbital movement or path of the shaft. A shaft with 0.003" runout will move the seal faces 0.003" per side or a total of 0.006". Permissible runouts should be limited to about 0.002" with some small leeway for lower pressures and speeds of less than 3600. If the shaft must be replaced due to excessive wear in the bushing areas and/or bows, the shaft quality should be selected carefully. Following is a sug-

189

gested specification for purchasing stock:

MACHINING OF SHAFT SECTIONS Use care in handling shafts during the machining process to avoid damage or bending. When the shaft is "chucked" in a lathe, use soft copper pads under the jaws and support the shaft to avoid bending it as it is rotated. Pump Shaft Quality (PSQ) Type 416 Stainless Steel Process: Hot rolled bar, turned, ground and polished Specifications: ASTM - A581, A582 and AMS 5610 Diameter Tolerance: 0.750 thru 1.500 +0.000/-0.002" 1.625 thru 2.437 +0.000/-0.003" 2.687 thru 2.937 +0.000/-0.004" 3.000 thru 4.000 +0.000/-0.004" Length: 20-24' random lengths Straightness Tolerance: 0.0015" per foot – guaranteed FOB point of shipment Out of Round Tolerance: one half of the diameter tolerance Surface Finish: 16 RMS Chemical Analysis: C MN P S SI CR MO Min. 15 12.00 Max. 15 1.25 .06 1.00 14.00 .60 Mechanical Properties: Tensile Strength 100,000 psi minimum Yield Strength 85,000 psi minimum Brinell Hardness 262 maximum

REASSEMBLY PROCESS

The pump section should be assembled in a vertical position. The bowls and shaft of a multistage vertical pump tend to develop a bow if assembled horizontally. With only about 9.0 mils bowl bushing clearance, the shaft comes in contact with the bushing after two or three bowls are installed because many bowls have a clearance of 0.002-0.003" in spigot fits. Vertical assembly randomly distributes this clearance. Vertical pumps have been shown to develop more than a 1/4" curvature from the top to bottom of a 16-stage section after assembly. The jack bolt in the suction piece must be used to position the end of the shaft in order to maintain accurate spacing between the impellers and the casing . Any load applied on previously installed impellers can bump them off the collet while the next impeller is installed. Impellers have come loose in this manner and destroyed themselves as well as the bowls when the pump is run.

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Impellers also have been cracked in the hub area due to overtightening of the collet. Collet-held impellers require more care and attention in mounting than those using split rings, ell keys or snap rings as locking arrangements. All bushing and wear ring fits should be coated with nickel-based anti-seize compound to help avoid damage during startup. After the pump is reassembled, it should be checked to determine if it has the proper lift. Rotate the rotor assembly by hand to check for possible rubs. After the pump bowls are assembled, it may be necessary to assemble the column sections horizontally because of their excessive length. The assembled portions should be rotated 180º as each additional component is installed. This will stagger alignment clearances and fits through the length of the columns and cause less total deviation of the center lines. Upon installing the discharge head, compare the shaft extension with the measurement taken before disassembly. Deviations of 1/16 - 1/8" are acceptable. Minor corrections can be made by installing or removing gaskets of differing thicknesses from mating faces. As an extreme, the upper shaft section may have to be machined. The shaft extension should be supported and the shaft locked for shipment. If it is necessary to leave the jack bolt in the suction bell in place, a warning tag should be attached indicating that the bolt must be removed prior to installation. If the bolt is left in place at installation, it will wreck the pump on startup.

VERTICAL PUMP ALIGNMENT AT INSTALLATION Because vertical pumps have relatively loose fitting bearings and the shafts are flexible and fairly small in diameter, alignment of the motor with the pump is a problem. The pump shaft should really be an extension of the motor shaft. The "coupling" is rigid and does not flex or in any way allow for misalignment. To ensure that the pump assembly is correctly positioned, a special bushing must be used when reinstalling the pump (Figure 4). The split bushing is preferred. The Pump Handbook Series

Figure 4. Alignment bearings for vertical pumps

FINAL OBSERVATIONS Achieving quality, consistent pump repair is simply a matter of paying attention to details. This involves the conscientious measurement of various parts to ensure a proper mating relationship. Above all, those involved in pump repair must be aware of and concerned about following proper procedures at all times. ■ William E. (Ed) Nelson is a pump and turbomachinery consultant who retired from Amoco Oil Co. with more than 36 years of experience. He is a founding member with more than 20 years of service on the Turbomachinery Symposium Advisory Committee, has authored more than 40 technical papers and is the author of Centrifugal Pump Sourcebook (McGraw-Hill, 1992). He received a B.S. in mechanical engineering from Texas A&M University and is a Registered Professional Engineer in Texas. Editor's Note: This series of articles on proper pump repair procedures is based on a tutorial by the author entitled "Monitoring Repairs to You Pumps," presented at the 1995 International Pump Users Symposium. The figures are reproduced with permission of the Turbomachinery Laboratory, Texas A&M University.

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R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

Evaluating Motor and Drive Solutions Pump motor and drive solutions span a wide range of options.

perating cost and downtime are major concerns in any industrial operation. Selecting the correct motors for pumping applications influences both the cost of energy used for operation and the life of the motor in the application. Often the least expensive motor is not necessarily the best solution as it may not provide the lowest overall operating cost. Because this article focuses on motor and drive solutions, let's first look at the motor solution options.

the use of these motors. With the Energy Policy Act of 1992, standard efficiency motors as we know them are on their way out. The new federal law stipulates that motor manuacturers and importers of motors and motor-driven equipment cannot make standard efficiency motors after October 1997. Nearly every continuous duty, 3 phase ac motor from 1 through 200 hp will be required to meet or exceed energy efficiency per Photo 1. High efficiency motor and inverter NEMA table 12.10 for Enerdesigns are still popular for many ENERGY EFFICIENCY gy Efficient Motors. applications, in anything but clean Energy efficient motors are areas, water, dirt and contaminants SPECIAL MOTOR ENCLOSURES popular on most continuous duty are free to enter the motor. Totally AND OPTIONS applications. Utility companies have enclosed fan-cooled motors resolve begun offering rebates to encourage Although open drip-proof this by providing a solid frame and endplates. Most energy efficient ENERGY POLICY ACT OF 1992 TEFC motors actually have rugged Minimum Nominal Efficiencies of Electric Motors cast iron frames and endplates. NOMINAL FULL-LOAD EFFICIENCY In the milling and chemical OPEN-MOTORS ENCLOSED MOTORS industries, machines and equipNO. OF POLES ment are exposed to harsh chemi6 4 2 6 4 2 MOTOR H.P. cals and vapors. Special mill and 1 80.0 82.5 — 80.0 82.5 75.5 chemical designs offer cast iron 1.5 84.0 84.0 82.5 85.5 84.0 82.5 motors treated with special epoxy 2 85.5 84.0 84.0 86.5 84.0 84.0 3 86.5 86.5 84.0 87.5 87.5 85.5 finishes inside and out to prevent 5 87.5 87.5 85.5 87.5 87.5 87.5 internal and external corrosion. 7.5 88.5 88.5 87.5 89.5 89.5 88.5 These motors use plated hardware, 10 90.2 89.5 88.5 89.5 89.5 89.5 plastic cooling fans, shaft slingers, 15 90.2 91.0 89.5 90.2 91.0 90.2 20 91.0 91.0 90.2 90.2 91.0 90.2 stainless steel nameplates and 25 91.7 91.7 91.0 91.7 92.4 91.0 winding thermostats to maximize 30 92.4 92.4 91.0 91.7 92.4 91.0 performance and operating life. 40 93.0 93.0 91.7 93.0 93.0 91.7 Higher temperature insulation is 50 93.0 93.0 92.4 93.0 93.0 92.4 60 93.6 93.6 93.0 93.6 93.6 93.0 provided not because the motors 75 93.6 94.1 93.0 93.6 94.1 93.0 run hotter (they actually run cool100 94.1 94.1 93.0 94.1 94.5 93.6 er), but to provide greater insula125 94.1 94.5 93.6 94.1 94.5 94.5 tion safety margin. Insulation life 150 94.5 95.0 93.6 95.0 95.0 94.5 200 94.5 95.0 94.5 95.0 95.0 95.0 may double for every 10ºC cooler a motor runs. These features help Table 1. NEMA Table (12.10) Minimum Nominal Efficiencies of Electric Motors extend mill and chemical duty

O

The Pump Handbook Series

PHOTO COURTESY OF BALDOR ELECTRIC COMPANY

By John Malinowski

191

facturer should be contacted to check on the horsepower required to operate the pump at the slightly higher speed. On belt driven pumps, a pulley change, or on direct drive pumps an impeller change, may be required if motor operation is desired at its rated horsepower.

PHOTO COURTESY OF BALDOR ELECTRIC COMPANY

ADJUSTABLE SPEED DRIVES

Photo 2. High efficiency motor and in-line pump

motor life. In the textile industry special cooling fans provide lint-free operation. Many different pumps are used in the food processing industry. A special motor has been designed to handle equipment washdown with caustic solutions and high pressure water. This type of motor is called a washdown duty motor. Stainless steel shafts with lip seals and slingers prevent water or chemicals from entering the motor. The inside and outside surfaces of washdown motors are painted with an FDA-approved epoxy paint. Frame joints are sealed with a special compound. All hardware is made of stainless steel. Weep holes are provided as an exit for any condensation that might collect inside the motor. For more severe applications, a paint-free version of the washdown motor is available. This is sealed in the same way as the regular washdown motor except that the motor housing is stainless steel and the endplates are constructed of aluminum treated to prevent chemical attack. Poultry processors prefer this type of motor because a scratch through the epoxy paint of a regular washdown motor can create problems as a result of exposure to

192

severe chemicals and very high water pressures during hose down. The ultimate sanitary motor is the all-stainless steel motor. The motor housing and endplates are made of stainless steel with the nameplate information etched into the frame. Stainless motors are commonly used in pharmaceutical plants. All of the motor types described are stock items. In addition, special options are available to tailor motors more closely to the specific application. Oversize bearings and special shafts and mountings are among the popular options.

MOTOR SPEED In comparison to standard efficiency motors, high efficiency motors have less slip. This increases the base motor speed slightly. Typically, 1725 rpm motors operate at 1780 rpm in a high efficiency design. Likewise, 3450 rpm de-signs will operate at approximately 3525 rpm. This higher speed must be considered in centrifugal pumps, in which the horsepower requirement increases by the cube of the speed change. This may result in motor operation inside the service factor of the motor. The pump manuThe Pump Handbook Series

Centrifugal pumps are the ideal application for adjustable speed drives because they offer the ultimate in energy savings. As motor speed is throttled down, the horsepower requirements are reduced by the cube of the speed change. An ac inverter is commonly applied to pump applications because it can be added to most existing ac induction motors. In the past, control valves were used to adjust the flow from a pump, with the motor and pump continuing to run at rated speed. As the valve was closed, little energy saving was possible from the motor. With the inverter, the adjustment mechanism can be a simple potentiometer knob or a programmable logic controller (plc), which automatically adjusts the pump as required by the application. For example, a 100 hp motor operating a centrifugal pump is required to produce only about 30 hp when operating at half speed. If the motor is operated 24 hours a day, 365 days a year, it uses about $56,000 worth of electricity a year. Such a motor throttled back to half speed saves about $39,000 per year at $0.08/kilowatt. Payback of the inverter does not take long, and the savings continue to accrue thereafter. But caution is advised. Inverters can not only reduce the speed of a motor by adjusting the voltage and frequency; the motor can also be made to operate above its base speed. And because the horsepower requirements increase by the cube of the speed change, it doesn't take much extra speed to overload the motor and risk premature failure. Also, an inverter causes extra heating due to harmonics. The signal of the inverter to the motor is an artificial sine wave. It is

PHOTO COURTESY OF BALDOR ELECTRIC COMPANY

Photo 3. Precision metering pump with vector drive motor

wise to purchase a motor and control that have been tested together from one manufacturer. Another consideration when using inverters is to compensate for dV/dt spikes, which may be several hundred volts and will degrade the motor insulation system. A new spike-resistant insulation can be specified on the magnet wire provding more than 100 times more resistance to this condition than standard wire. Long wiring runs between the motor and control aggravate this condition with voltage ring-up. In this case, line reactors should be placed between the motor and control. Special inverter-duty motors are available. These incorporate a constant speed fan to provide cooling at low motor speeds, whereas the shaft mounted fan of a TEFC motor does not have much air velocity. Another product recently introduced is an energy efficient inverter-duty motor with an inte-

The Pump Handbook Series

gral inverter control. Built in one package, this ensures compatibility between the motor and control because all parameters are adjusted for that particular motor. With the control on the motor, only ac power must be brought into the package. This type of motor will become increasingly popular in pumping applications.

CONCLUSION The pump user should be aware of the speed at which his pump operates. Simple motor replacement may cause motor overloading problems resulting from even a small in-crease in motor speed. On the other hand, if an inverter is used to reduce the pump speed, dramatic energy savings can be achieved. ■ John Malinowski has 20 years of experience as a drives specialist with Baldor Electric Company. He graduated from Rock Valley College in Rockford, IL.

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R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

Slurry Pump Wear Factors High velocity, large solids size and high concentrations increase wear. By Graeme Addie, Krishnan V. Pagalthivarthi and Robert Visintainer ipelines are used to transport solids by forming a solids-water mixture called a slurry. Slurry pipelines can extend to 10 miles or more in length. Except where the solids are very fine and the distances more than 10 miles,centrifugal slurry pumps are the preferred driving force. Above approximately 100 microns in size, solids settle out. Thus, if a certain minimum average velocity is not achieved, the pipeline plugs, and no transport occurs. The velocity required for transport increases with the size of the solids, which may reach 6" or occasionally even be larger. Because the goal is to transport solids, excess water transport is dead load. Therefore, solids concentrations of 30-50% by weight are common. The combination of a relatively high velocity, large solids size and high solids concentration increases wear in the driving pump or pumps. We will examine the wear experienced by pumps of this type, and we will describe design technologies that attempt to minimize it.

P

dredging operations in which slurry pumps and pipelines transport solids from the bottom of harbors and rivers to separate spoil areas.

SLURRY PIPELINE BASICS Pipeline transport of crushed rock or conventional settling slurry as a mixture of water and solids requires meeting or exceeding a minimum mean mixture velocity called the deposit velocity, Vsm. The deposit velocity varies with the pipe size, particle size, solids specific gravity, particle shape and concentration. A real slurry is composed of various sizes and shapes of particles, so the deposit velocity in practice is not one number. Rather, a bed of deposits forms over a velocity range. The head loss characteristic for most settling slurries at different delivered concentrations is assumed to be U-shaped (Figure 1), with min-

imum head loss value as indicated. For applications with constant speed centrifugal pumps, operation is usually recommended at a velocity slightly higher than the minimum head loss velocity shown in Figure 2 to avoid possible instabilities. To avoid stability and deposit velocity problems, pipelines must be designed to operate at a velocity at which no deposit occurs and there is a stable intersection of the system and pump characteristics. Varying transport rates and conditions require that safety margins be incorporated. These, however, must be minimized to optimize energy use and reduce wear.

SLURRY PUMP DESIGN A centrifugal pump has two main components. The first is the rotating element comprising the shaft and impeller, including the

INTRODUCTION Slurry pipelines handling relatively coarse particles (greater than 100 microns) normally use centrifugal slurry pumps as the driving force for distances up to about 10 miles. Slurry pipelines are found in the coal, copper, iron ore, aggregate, tar sands, phosphate and other mining industries. Applications include the transport of coal through a sizing plant, transport of raw phosphate matrix from the dragline to the wash plant and the disposal of various waste tailings such as in the tar sand industry. They also include

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Figure 1. Definition sketch for limit of stationary deposit zone The Pump Handbook Series

Settling Slurry at Constant Concentration

P r e s s u r e

Stable Operating Point

Water

Unstable Operating Point Constant Speed Pump Curve

VS Vsm Solids Start To Settle

Minimum Head Loss Velocity Deposit Velocity For Different Size Particles

Mixture Velocity Vm Figure 2. System and pump characteristics showing conditions for stable operation

vanes that act on the fluid. The second is the stationary element comprising the casing or shell that encloses the impeller together with the associated stuffing boxes and bearings. The hydraulic design of a centrifugal pump involves dimensioning the impeller and the shell to provide required performance characteristics. Pump designers normally offer several combinations of component dimensions that can be used to achieve a specified performance. The combination selected depends on the intended application and on any hydraulic or mechanical limitations. A number of limitations are imposed in slurry pumps. These include the need to pass large solids, the requirement for a robust rotating assembly because the slurry density exceeds that of water and the desirability of thicker sections to minimize the effects of wear. Centrifugal pumps come in a wide variety of arrangements to suit different applications, and they may comprise a number of stages of impeller and collector. Slurry pumps offer less variety. They are normally the single-stage end-suction type and usually of radial or mixed-flow configuration. They commonly have volute-type collectors, but these are often modified to

concentric or semi-concentric forms to reduce the effect of wear on the shell. Sections of a representative end-suction single-stage volute-casing pump are shown in Figure 3. The pump volute or casing converts the kinetic energy of the fluid leaving the impeller into pressure energy. In an idealized pump, it is assumed that there are no losses in either the casing or the impeller. In practice, however, hydraulic losses occur in all wetted passages of the pump. The head-capacity curve of an actual pump is determined by subtracting losses from the idealized pump characteristic. This is fig-

ured on the basis that there is only a single discharge for which the shock loss at the impeller inlet is zero. Stepanoff, in his Centrifugal and Axial Flow Pumps, 2nd Edition, provides empirical relations and coefficients for calculating impeller diameter, outlet blade angle β2, outlet width, shell throat areas and other key dimensions for water pumps of different relative proportions. Figure 4 is a typical performance chart for a centrifugal slurry pump. Note that as β 2 increases, so does the head produced. Values of β2 usually range from 15º to 35º. Slurry pumps commonly have β 2 values around 30º. Slurry pumps require thick sections and flow passages capable of passing large spheres. As a result, they have dimensionless head coefficient values that differ from those of simple water pumps. A slurry pump is thus likely to require larger impeller outlet widths than a water pump. The combined action of the impeller and casing determine the location of the pump's best efficiency point (BEP). The largerthan-normal impeller outlet widths in slurry pumps tend to shift the discharge at BEP to larger values (and to flatten H-Q curves). But this shift can be compensated for by changing the casing geometry. In the idealized case, an impeller with a large number of infinitely-thin frictionless vanes would produce the highest effi-

Radial Outlet Discharge Exit

Suction Inlet Impeller Eye

Volute Casing Figure 3. Sections of a representative slurry pump The Pump Handbook Series

195

90–

600 RPM

70% 80% 83%

80–

84%

83%

70– 80 W 0K

500 RPM 60–

40 0K W 80%

50–

40–

400 RPM

0

20 KW

30–

300 20– RPM –12 10–

AT NPSH PM 600 R

NPSH AT 300 RPM

0– 0

200

400

+

600

800

1000

1200

–8 –4

NPSH (METRES)

TOTAL DYNAMIC HEAD (METRES)

84%

–0

1400

LITRES PER SECOND

Figure 4. Pump performance chart

ciency in a pump. In practice, for a water pump this number ranges from five to nine. In a slurry or sewage pump it may be reduced to three or four in order to pass coarse solids and accommodate extra vane thickness. Fewer vanes result in a steeper H-Q (head quantity) curve and reduced efficiency. The reduced efficiency can be held to 1 or 2% in most cases. The number of vanes used is determined after considering the size of solids to be passed, the vane thickness, the location of the inlet and the overall vane shape design. To minimize efficiency loss, the vanes should have the correct inlet angle for shock-free entry of the fluid at the design point. The outlet angle should be set to give the desired performance, and the shape of the chamber between the inlet and outlet should minimize the rate of change of velocity. As the impeller is distorted to increase passage width and wear

196

life, so is a slurry pump collector. Slurry pump casings vary from the true volute water pump TH type shown in Figure 5 through the semivolute CH type to theessentially annular A type. Each combination of the types illustrated in Figure 5 has its own hydraulic performance and wear

SHELL TYPES T - TIGHT CUTWATER Volute Type

C - CONVENTIONAL Semi-Volute Type

A - ANNULAR Annular Type

Figure 5. Types of shells and impellers The Pump Handbook Series

characteristics. The HE/T combination generally has the highest performance but is not necessarily the most forgiving for wear. The ME/C combination is capable of respectable efficiency and at the same time offers more dependable wear performance. The hydraulic performance of three different types of casings (shells) with the same impeller is shown in Figure 6, and the performance of different shells and impellers appears in Figure 7. The maximum efficiency obtainable from properly-designed water pumps is referenced in the literature. Slurry pumps rarely achieve these values because of their larger mechanical losses (due to larger shafts and bearings) and the hydraulic compromises necessary to improve wear life. Using the latest design technology, including numerical methods of hydraulic design, it is now possible to get to within a few percent of the maximum obtainable values and still have a slurry pump that provides good wear performance.

PUMP SELECTION AND APPLICATION As indicated, pumps must be selected by matching their head-discharge performance to the requirements of the piping system. For settling slurries in particular, selection of appropriate operating conditions raises special considerations. In the case of long lines, the total system head is more than can be handled by a single pump. It is then necessary to use several pumps in series. Once the operating conditions are established, pump selection in its simplest sense amounts to deter-

IMPELLER TYPES HE - HIGH EFFICIENCY Twisted vanes in rounded and narrowed meridinal section. ME - MEDIUM EFFICIENCY Twisted vanes in rounded meridinal section. RV - RADIAL VANES Radial vanes in rectangular meridinal section.

mining the specific performance of each available pump for the head and flow required, and selecting the one best suited to the duty. Generally, the smaller the pump the cheaper it will be. However, reducing pump size implies higher speed for a given discharge. Thus, in a water pump application, NPSH and mechanical considerations limit speed and set the overall pump size. In slurry pump applications, NPSH determines limits only in special cases. And the shaft speed is limited by wear life, which often decreases as speed increases. Therefore, slurry pumps are typically larger with lower rotational speed than

water pumps with equivalent head and discharge. Actual wear data are seldom available for a specific pump type, shaft speed, and slurry duty. However, we shall discuss new wear modeling techniques that can be used to transfer wear information from one configuration to another to select the most even or best wearing pump, and in certain cases to calculate expected wear lives. In general practice, pump selection begins with some sort of semiempirical selection guide. (See Figures 8 and 9 and Tables 1 and 2.) The initial classification as to Service Class (1, 2, 3 or 4) is based on Figure 8 with, if necessary, an

adjustment for the abrasiveness of the slurry (Table 1). From all of the available pumps, a set of acceptable selections is then established using the discharge branch velocity, suction branch diameter (Figure 9) and other limits given in the tables. Particular attention should be paid to the impeller's peripheral speed and percentage of BEP operation for the various shell types. When determining service class using Figure 8, multiply both the d50 of the slurry solids and the specific gravity of slurry by the correction factor. Selection is made based on the efficiencies, power requirements SLURRY SPECIFIC GRAVITY

2.0

Slurry Pump Range Service Class Chart

1.9 1.8 1.7 1.6

For use as a first guide only assume 2.65 SG silica-based solids. Adjust rating to account for solids of different hardness or density, or for other unusual operating conditions using Table 8.2

1.5 1.4

4

1.3 1.2

3 2

1.1

1

1.0

50 100 1000 5000 AVERAGE D50 PARTICLE SIZE (microns) 200 100 48 35 28 14 TYLER SCREEN SCALE

8

Figure 8. Service class chart for slurry pumps ADDITIONAL CLASS 3&4 PUMP SPEED LIMIT 2400

Figure 6. Hydraulic performance for various shells

2000 1500 1200 CLASS 3

800 400 CLASS 4

0 0.1 4

0.3

0.5

0.7

0.9 (m)

8 12 16 20 24 28 32 36 (in)

Figure 9. Pump speed limit chart for classes 3 and 4 Service Type

Figure 7. Hydraulic performance for various shell-impeller combinations The Pump Handbook Series

Correction Factor

Normal silica slurries such as dredged river material, taconite, tailings, etc.

1.0

Dredged coral, bottom ash, copper mill circuit slurries and slurries known to be abrasive

1.2

Coal slurries, slurries in which the solids easily breakup and slurries containing slime

0.8

Table 1. Slurry abrasiveness adjustment

197

SERVICE CLASS Shell type Maximum Discharge Velocity Maximum Throat Velocity

1

2

3

4

m/s

12

10

8

6

ft/s

40

32

27

20

m/s

15

12

9

6

ft/s

50

40

30

20

20-120% 30-130% 50-140%

30-110% 40-120% 60-130%

40-100% 50-110% 70-120%

50-90% 60-100% 80-110%

43 8500

38 7500

33 6500

28 5500

28 5500

25 5000

23 4500

20 4000

Recommended Range: Percent of BEP flow rate

Annular (A) Semi-volute (C) Near volute (T)

(a) All metal Pump: m/s ft/min (b) Rubber Lined Pump: m/s ft/min

Table 2. Recommended operating limits for slurry pumps

and available driver speed limitation considerations. This is a tedious process if several hundred possibilities must be considered. Initial branch checks may be used to limit the available choices and attendant evaluation. A more common approach now, however, is to use a computer program to scan the possible choices and identify acceptable ones in order of efficiency. The highest efficiency selection may or may not be the pump with the best wear life. In the past, experience and the above guidelines were used to screen possibilities further and arrive at the best overall selection for both hydraulic performance and wear life.

PUMP WEAR TECHNOLOGY Improved impeller efficiency usually results in improved wear. The same does not necessarily hold true, however, for the shell collector. Whether for this reason and/or its greater cost, most research work so far has concentrated on the casing. In the early 1980s, Roco and Addie developed computer models aimed at predicting the wear inside the casing of a centrifugal slurry pump. The original model computed the mixture velocity distribution within the casing, assuming the mixture to be slippery as opposed to sticky. The mixture velocity was obtained by solving the equation of

198

motion written in the stream function formulation, and by simultaneously satisfying the pump flow rate and the slip factor of the pump impeller. Numerical predictions were provided by applying the Galerkin Finite Element Method using triangular elements. The relative velocity of solid particles in the carrier liquid resulted from the dynamic equilibrium of particles under the influence of inertial and drag forces. The convective velocity of the solid phase determined the non-uniformity of concentration in the pump volute. The concentration was calculated by solving the diffusion-convection equation for the averaged properties on the volute width. An upwind finite element approach was employed for numerical analysis. The erosion wear rate on the interior was obtained using an energetic approach. The stream function formulation necessitated the a priorital specification of the stagnation point (i.e. cutwater). Some parameters in the calculation model were derived from experimental investigations. The slip factor of the impeller was obtained from measurements of the hydraulic performances of the pumps at the GIW Industries Hydraulic Laboratory. An erosion test loop system was used to establish the experimental ratio of the erosion rate to the energy dissipated by particle impingement onto a solid wall. The Pump Handbook Series

This early model also assumed uniform outflow from the impeller, and the impact wear mechanism allowed some flow normal to the casing wall. The model used a stream function formulation, sought as a superposition of two solutions. The procedure required the explicit specification of the stagnation point in addition to test results of head-quantity characteristic at the specified speed. The results were in the form of wear rate plots along the casing wall. A later model using higher accuracy nine-noded isoparametric quadrilateral elements eliminated the flux of slurry through the wall. Furthermore, this model used the potential function formulation, which eliminated the need to specify the stagnation point since it is determined as part of the overall solution. The nodal velocity of solid particles is calculated by solving a force balance between the inertial (F i ), pressure (Fp) and drag forces (F d), neglecting gravity, acting on the particle: Fi + Fp + Fd = 0, (1) where the drag force depends on the local solids concentration. The local concentration follows the convective diffusive conservation equation, Vs . ∇C – ∇.D ∇C = 0 ~ ~ ~ ~

(2)

where C is the local concentration, D the overall turbulent diffusivity, VS is the solid velocity and ∇ is the ~ ~ gradient operator in two dimensions. Because the drag forces depend on the particle velocity and concentration, an iterative solution is necessary alternating between equations 1 and 2. Initially, a uniform concentration equal to C vd (delivered concentration) is assumed across the entire plan view section. The iterative procedure yields the converged solution for solid velocity and concentration. Wear is calculated by correlating the particle velocity, angle of impact and concentration with the wear coefficients for the relevant combination of slurry and shell material.

Photo 1. GIW pump in matrix service.

The total wear rate in microns/hour, consisting of a summation of sliding and impact wear rates, is plotted as a function of the length along the shell. As wear progresses, the shell shape changes, and so does the local wear rate. The wearout mode of the program determines a period of time over which the wear rate may be assumed fairly constant. It adjusts the shape of the shell at the end of this period, recalculates the element mesh, and recomputes the modified particle velocity, concentration and wear rate, repeating this process until a specified maximum amount of wear occurs. This solution provides an estimated time to wear through a given thickness and identifies where shells wear out, taking into account the effect of the changing shell geometry. Briefly, the wear rate due to directional impact at an angle, α, is approximately given as: S ~ φα ( ρsCvVα )Vαm-1 (3) where ( ρsCvVα ) is the mass flux at angle α, m ≅ 3, and φα is the wear coefficient measured in (m-s2/kg), ρs is the solids density, C v is the concentration by volume and V α the velocity in the direction of the angle. An impact probe having a specified angle at its tip is exposed to a stream of fully turbulent homogeneously suspended slurry with

known C v and V α . By measuring the wear rate sufficiently close to the tip and applying a correction factor, φα is readily determined. Earlier, the simplifying assumption that, φ α = φ ο Cos m α +φ π/2 Sin m α (4) was made, where φο and φπ/2 correspond to α = 0 and α=π /2, respectively. While there is no mechanistic basis for this ad hoc hypothesis, wear experiments were necessary for only two angles, α = 0º and α = 90º. To correlate the sliding component of wear, a new sliding wear machine was built at the GIW hydraulic laboratory. Four rotating samples are exposed to an open channel flow of slurry, sliding past the wear surface while the coriolis force causes the slurry to press against the sample. The specific energy, Esp, is given as: Esp = ( ρs ρL )ω 2 R Cv (Q/h)(t/s), (5) where ρL is the density of liquid, ω is the angular velocity (rad/s), R is the average radius of rotation of the sample, Q is the flow rate per sample, (h.s) is the total wear area and t is the duration of wear. The specific energy is used to correlate the flow parameters to casing wear. This method of calculating the slidThe Pump Handbook Series

ing wear coefficient has proven more reliable, more repeatable and more representative of the wear in pump casings than its earlier counterpart. Moreover, it yields measurable wear in less than an hour. The analytical numerical models discussed so far deal essentially with a 2-D cross section normal to the pump axis. Two phase flow prediction also was conducted to study secondary flows in radial cross sections of the pump casing via a quasi-3D turbulent flow model. This methodology has been used to predict wear rate along the casing periphery in a radial cross section. Wear prediction is intrinsically related to the attendant two phase flow that causes the wear. In recent years, many research workers have focused on the overall problem of two phase flow in general. However, it does not directly pertain to slurry pump wear. Part 2 of this article will explain practical factors affecting slurry pump wear.

PART II W ear of pump components in slurry pipeline systems is a major issue. As might be ex-pected, wear var-ies from pump to pump and from application to application. Phosphate matrix pumps in Florida with branches of 18" and impellers of 44" diameter typically wear out one casing, two impellers and three liners in a year. Pumps operating in alumina red mud service, on the other hand, may last several years, large dredge pump impellers may last years or weeks depending on the service. Nevertheless, several typical or recurrent wear patterns are observed in slurry pump casings and impellers and thus merit attention. Specifically, slurry-pump casings often experience maximum wear in the outer radius or belly, a zone in which sliding beds of solids form as a result of centrifugal forces. A specific area on the circumference of the casing generally experiences maximum wear, but the location of this area can shift with operating conditions and casing geometry. For increasing flow rates and casings of more annular geometry, the location

199

this process. As discussed earlier, noting the wear pattern experiences of maximum wear tends to shift construction materials can also be and applying the known operational toward the pump discharge and varied to improve wear resistance. and geometric considerations that away from the tongue. These trends Typically, some balance between can alter those patterns. Then, are shown in Figure 10. strength, toughness, serviceability through experimentation, the casing Another common wear pattern and cost must be struck with no one size, geometry and operating condiin slurry-pump casings is gouging or material holding the advantage in all tions can be modified in a way that extreme localized wear in the casrespects. will increase casing life. ing side wall just downstream from Impeller wear in a slurry pump the tongue (Photo 2). Gouging is iniFIELD EXPERIENCE VS. is often closely related to its tiated by the three-dimensional AVAILABLE TECHNOLOGY hydraulic efficiency. This seems eddy generated where the tongue reasonable since improved parts the fluid. The severity of the While quasi 3-D models do hydraulic efficiency generally coingouging depends on the pump exist, the current technology does cides with reduced velocities in the design but is typically determined not attempt to model the turbulence recirculating eddies that cause localby the ratio of operating flow rate to generated at the casing tongue at off ized wear. However, a wellbest efficiency flow rate. Specifiduty flows, or the effect of varying cally, pumps operating well pressures at the impeller outShell Shape below the best-efficiency flow let tips and other detrimental rate recirculate large volumes effects. 12 5 T % of flow past the tongue, With only 3 vanes, in 11 increasing the velocity and many cases pressure fluctua0% 70 size of the gouging eddy. tions are significant. It is thus % C 95 of Gouging eddies may also uncommon that a pump oper% Q BE occur at other locations in the ates exactly at its design point P 80 60 % % slurry-pump casing or the where turbulence and recirA impeller. In some instances, culation flows are minimal. 50 % such eddies can be identified Given this, it is interesting to 4 as originating in geometric see how well the available 0% O discontinuities and can be technology works. eliminated with design modA 30" (0.762m) diameter ifications. In other situations, branch, 80" (2.03m) diameter they can be traced to operaimpeller dredge pump reporttional considerations, such as ed as operating at 360 rpm Wear shifts Danger of Wear shifts Danger of Best toward Chance for tongue tip toward sidewall the pump being run well and passing a 45,000 U.S. gpm tongue Even Wear discharge wear gouging away from BEP flow. In all (2840 l/sec) flow of slurry of Wear Trends cases it is vital to have a sound 400 micron d50 size particles understanding of the large- Figure 10. Trends in location of casing belly wear. at a concentration of 1.1 SG designed slurry-pump impeller often scale flow patterns within the pump. was numerically modeled to detercan sustain considerable wear Such an understanding depends on mine the calculated wear rate. before its pumping capacity is careful observation and analysis Having measured actual wear reduced to an unacceptable level, using both basic fluid mechanics and on the pump, these figures were and it may appear to be worn out numerical modeling. A healthy sense then compared with the original long before it needs to be replaced of three-dimensional visualization is analysis. The results are shown plotfrom an operational standpoint useful in suggesting design or operted on a relative scale in Figure 11. (Photo 4). The two worst wear areas ational modifications that can The analysis predicted both the area on an impeller are usually near the improve overall wear performance. of maximum wear and the trend of inlet and the outlet. At the inlet, Gouging eddies in the casing decreasing wear around the volute eddies are induced by changes in also may affect flow patterns in the towards the discharge. flow curvature or by intrusion of the space between the impeller and the The actual wear pattern does, recirculating flow from the gap suction-side liner, producing sechowever, show large fluctuations between the impeller and the sucondary gouging. Such localized wear not predicted by the analysis. Recall tion liner. High velocities at the can be severe (Photo 3). In this that the analysis represents a preimpeller outlet are the major cause instance a practical method of diction of wear based on the origiof problems there. extending the liner life is to rotate it nal design hydraulics. Under field The key to improving overall 180º when one-half to three-quarconditions wear itself affects the wear performance of any slurryters of its estimated life has elapsed, hydraulics and can be expected to pump component is to identify the thus moving the wear zone to a new alter wear rates. Turbulence caused causes of localized wear and either portion of the surface. by gouging wear can lead to the eliminate them or spread them over Since the slurry pump casing is uneven wear seen in this case. Also, a wider area. Design geometry and a consumable part, its wear perforthe effects of recirculatory "secoperating conditions play a role in mance often can be improved by ondary flow" (in a cross section per-

200

The Pump Handbook Series

pendicular to present analysis) may be important. In another case, an 18" (0.46m) branch, 39" (.99m) diameter 18x18 WSO44 type 3ME/A16 slurry pump operating at IMC in Florida and pumping phosphate matrix was returned to the lab for measurement of wear after being in service for 300 days. The subject 16" (.406m) wide shell was measured for wear at 10º intervals along the casing centerline, and the results were plotted as indicated in Figure 12. A "wear out" analysis solution was also run for the pump at the reported conditions of 590 rpm, 12,000 U.S. gpm (760 l/sec) flow and 1.3 SG. A properly sized foundry sand was found to be the best simulation of the abrasiveness of the phosphate matrix. This solution is plotted in Figure 13. Considering various uncertainties in this analysis (such as casting irregularities, variations in duty conditions and the approximate nature of the finite element solution), the actual and calculated results are encouragingly similar. Note that the program not only predicted the general shape of the wear curve, but also the small peaks in wear just past the tongue and before the discharge (points B and F in the figures).

Photo 2. Extreme gouging of side wall.

APPLYING WEAR TECHNOLOGY Despite its limitations, existing casing wear calculation technology appears in many cases to be capable of predicting the form of wear and to a lesser degree its magnitude. As seen from numerical modeling studies and field experience with a given fixed slurry size, shape, abrasivity and concentration inside a given shell, the wear varies dramatically as to both its rate and location inside the shell with different flows (Figure 14). Available technology can be used, therefore, in the first instance as part of the pump selection process. It subjects existing designs to the operating conditions in order to minimize wear around the shell and in some cases direct it to where it will have least effect. Altering the shape of the col-

Photo 3. A gouged suction liner.

Photo 4. Worn impeller may still be operational. The Pump Handbook Series

201

D

F ACTUAL WEAR

RELATIVE WEAR RATE

CALCULATED WEAR RATE

LENGTH ALONG SHELL

Figure 11. Actual vs. calculated wear in dredge pump 80.

lector or applying it directly to create alternative or new designs for a specific duty are other possibilities. For example, the calculated effect of a change in pump geometry in the form of a tongue cutback for a 6" discharge, 8" suction, diameter impeller GIW LCC 150-500 (6" discharge, 20" impeller) type CH/3ME pump is shown in two different plots (Figures 15 and 16). The first plot, Figure 15, shows wear on the vertical axis at various concentrations and radial shell locations for the original design. Figure 16, using the same scale and set of axes, represents a modified shell shape. It shows that the wear is reduced overall and that the maximum wear in the tongue area is reduced by more than 100%. Identical operating conditions were maintained in this comparison. Different shapes of collectors, while perhaps improving wear, produce different efficiencies that must be compared with the different wear costs in an overall analysis. With respect to overall size or simply the relative proportions of the impeller and shell, modeling the effect of impeller diameter versus speed to produce the same head can also be done to optimize the geometry of the design. On the transport side of the equation, the size and shape, concentration and possibly even the abrasivity of the slurry can be modeled into the total transport cost

202

ed wear and capital costs.

equation to come up with the best overall compromise. Operationally, whether to use fixed or variable speed pumps can also be considered with respect to the type of pump, efficiency, expect-

FINAL THOUGHTS Slurry pipelines are widely used. To be efficient transporters, however, they must operate at high

WEAR DEPTH (Inches)

C

LENGTH ALONG SHELL (Inches) Figure 12. Measured wear of 39 inch annular pump. 3.20–

2.80–

F 2.40–

B WEAR DEPTH (Inches)

TONGUE TIP

AB

HS ONT 10 M NTHS 9 MO

2.00–

1.60–

1.20–

0.80–

0.40–

0.00– 0.00

'

'

40.00

'

80.00

'

120.00

'

160.00

'

200.00

LENGTH ALONG SHELL (Inches) Figure 13. Calculated wear of 39 inch annular pump. The Pump Handbook Series

'

240.00

'

280.00

concentrations – a condition that causes considerable pump wear. Slurry pump design is similar to that of water pumps but must be modified to resist the high wear associated with pumping high concentrations of coarse solids. As with all centrifugal pumps, correct application is critical with slurry pumps because of its effect on wear. Technology exists to calculate wear inside the largest, most expensive pump component: the casing. This can be used to model application, design and slurry particulars. Experience so far with the method shows that in spite of limitations and uncertainties, it provides useful answers in most cases. More work is needed, however, to model impeller wear and for more flow visualization work to check and refine the modeling methods described here. ■

Figure 14. Variation of wear rate along the casing for various flow rate values.

D

Krishnan V. Pagalthivarthi, on leave from the Indian Institute of Technology in Kanpur, India, is currently working at GIW Industries. Editor's Note: This article is based on a paper presented during the 1996 ASME Fluids Engineering Division summer meeting, and is reprinted with permission of the American Society of Mechanical Engineers.

Y

C

E

X

O

A B

G

LEN

GTH

ALO

GIW LCC 150-500 6000 SFPM - 100% BEPQ FLOW 270 MICRON SAND SLURRY STANDARD WEAR RATE INDICES

NG

F

SHE

LL

BY TION

UME

VOL

TRA

CEN

ON %C

Figure 15. Wear profiles for different concentrations. RELATIVE WEAR RATE

Robert J. Visintainer is a graduate of the Georgia Institute of Technology School of Mechanical Engineering and of Miami University, where he majored in physics. He has been employed at GIW Industries for 12 years in slurry pump mechanical and hydraulic research and design. For the last six years he has been the chief engineer responsible for design and drafting.

RELATIVE WEAR RATE

Graeme R. Addie is Vice President of Engineering and Research and Development at GIW Industries, formerly Georgia Iron Works. He holds a post graduate degree (MIE) in mechanical engineering from the Royal Melbourne Institute of Technology in Australia and is the author of numerous papers on slurry pipeline and pump performance.

TONGUE OUTBACK AREA

LEN

GTH

OF S

HEL

L GIW LCC 150-500 6000 SFPM - 100% BEPQ 270 MICRON SAND SLURRY MODIFIED SHELL GEOMETRY, CUTBACK TONGUE

UME

VOL

RA

ENT

ONC

%C

BY TION

Figure 16. Wear profiles for a slight cutback at tongue. The Pump Handbook Series

203

✓

R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

Tools That Do Our Work for Us - Or Do They? By Jack Bolam ost presentations I make these days have something to do with precision shaft alignment. I know that 8090% of the problems I find while doing vibration analysis are, in some fashion, misalignment-related. How can that be? The modern computers and lasers we are using cannot be wrong, can they? We must remember that shaft alignment is far from an exact science. There are basics that must be considered, and every alignment method has its faults to overcome. Indicator brackets sag. Shafts move axially in their bearings. Dust, heat and steam will bend a laser beam. So we must know the limitations of the equipment we are using. Most of us who do alignments learned our method from someone else, usually a more experienced mechanic. Where did that mechanic learn his method? It was probably passed down from another more experienced mechanic. Some of the basics get lost in this word of mouth process; others were never passed along. Machinery speeds are also increasing. Thus, making precision alignment more critical, and formal training of the forgotten basics more of a necessity than it ever was. There is a definite need to change the way we look at machinery setups. In the past, pumps were mounted on cast iron bases and bolted and grouted to the top of isolated concrete pedestals. Today, I come across bases that are made of bent sheet metal, machines with no bases at all – just all-thread with a nut above and below the support feet. I have also encountered a large compressor bolted to a one inch plate that was

M

204

“glued” (at one time) to the floor. Obviously, these examples are not good supports for rotating equipment because they make it almost impossible to align the machinery. Whatever happened to precision machining the bases and the feet? In my presentations as well as in my classes, I point out three major basics that are usually skipped over – either to save time or because they are not known to begin with. I talk about soft foot, bracket sag (not in the laser) and thermal growth or shrink. The one that is most elusive and causes the most problems is soft foot – which is not just an alignment problem but also a machine installation problem. What is soft foot? What causes it? How do you recognize and eliminate it? Let’s pretend that we have just been seated at a table in a fancy restaurant. We order our drinks and are well into the conversation when they are served. A point has to be made so, for emphasis, you lean on the table. Rather than being solidly supported, it moves under your weight and spills the drinks. The table has soft foot. The leveler on one of the feet is out of parallel with the other three feet, or the floor beneath the table is uneven at that point. My favorite “fix” for this kind of soft foot is to fold a napkin and stuff it between the foot and the floor. The same situation is taking place every day in new and old machine installations. Cast iron is being replaced by bent thin plates or all-thread. Machines are being built lighter, and machining, grouting and gusseting are kept to a minimum. The Pump Handbook Series

The savings that are supposed to be derived from modern efficiency machinery are being lost to high vibrations and severe misalignment. Soft foot is the primary cause. Soft foot comes about with the mating of two non-parallel planes. On the machinery this could be one or more bent or improperly machined feet. A weakly constructed base changes shape when the hold down bolts are tightened. Pipe strain and severe coupling alignment create soft foot. All of the above can contribute to this phenomenon, and it only takes one to deform a machine casing. When is a good time to check for soft foot? If given the opportunity, I do it at the time of installation. I will set the machine on the base and line up the hold down bolt locations. Using feeler gauges, I will then check for any clearance beneath each corner of each foot. If I find any clearance, I shim to fill the gap. If the machines are not yet piped, you should insert a 1/8 inch precut stainless steel shim beneath each foot of each machine. This will lessen the chances for adjustment problems during the alignment process. The next step is to find out if bolting the unit down causes any problems. Tighten all the hold down bolts. Mount a dial indicator to measure the vertical movement on one foot. Carefully loosen the bolt on that foot. Log the TIR (Total Indicator Reading). Retighten that bolt and move to the next one. Follow this same procedure for each hold down bolt. I like to see less than .002 inch TIR during this process. If a foot is found reading higher than this, you should loosen that bolt again, check

the corners with feeler gauges and shim accordingly. After an adjustment is made, the entire procedure has to be completed again until the .002 inch maximum is attained. It isn’t difficult to see the amount of work and time required to check for and remove soft foot from a new installation. And it can be even more time-consuming on older pump installations. It is usually necessary to remove the machinery and clean the bases before any checking is started. Two shafts cannot be aligned if there is soft foot present, regardless of the alignment method you are using. What about the unit where a pump is mounted directly on the motor shaft? The same procedure detailed above should be followed.

The soft foot is still distorting the motor casing, causing internal wear and reducing bearing life, as well as wasting energy. Today, we are looking for reliability from our rotating equipment. In its purest form reliability means zero unexpected or unplanned shutdowns. The removal of soft foot is one giant step in the right direction.■ Jack Bolam is President of VibraCon Predictive Maintenance Consultants, Loveland, CO. He has more than 36 years of experience in the rotating machinery industry and has worked for Dresser-Clark, Eastman Kodak and Fluor Daniel. He is a frequent contributor to Pumps and Systems and can be reached at (970) 663-2118.

OBTAIN VIBRATION RESULTS

CLEAN THE BASE AND FEET (FLAT FILE, SCRAPER, CLEANER)

High Axial? 2X Operating Frequency?

Dirt Rust Paint Burrs

GATHER OPERATING TEMPERATURE DATA For Thermal Growth Calculations MAKE NOTE OF ANYTHING ABNORMAL Shims or Lack of Shims Has Unit Ever Been Aligned? CONDITION AND TYPE OF SHIMS? Recommended Pre-Cut Stainless Steel

INSPECT THE HARDWARE Bolts Stretched Special Washers Cupped Enlarged Bolt Holes? Coupling Condition? Keyways and Keys

CONDITION OF SUPPORT

Edges Rolled Over Fretting Ridges Key Length Grids, Spacers, Teeth, Etc.

Dirt, Rust or Paint? Looseness? Missing Areas AMBIENT CONDITIONS

Ridges Cracks or Breaks

Water Dripping? Steam? Heat? Is Protective Covering Needed?

ROUGH ALIGNMENT Set Proper Coupling Spacing Straight Edge and Feeler Gauges

Table 1. Pre-alignment checklist - operating

Table 2: Pre-alignment checklist – shutdown

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R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

Canned Motor Pump Application Profiles Plant operators and a sales engineering professional share some of their insights from years of hands-on experience. Editor’s Note: Although the concept of the canned motor pump has been around for decades, it has only been since the early 1980s that this sealless centrifugal design has had a major impact on process pump operation. Since then, significant advances in design and monitoring options, coupled with years of practical experience, have helped users improve performance and reliability. In an effort to share some of this field knowledge, Pumps and Systems recently spoke to a few plant professionals about their canned motor pump applications experience. Participants included: Steve Hrivnak, a Senior Mechanical Engineer at Eastman Chemical Company’s Tennessee facility (and newest member of Pumps and Systems’ User Advisory Team); Carlton Burrell, a Plant Engineer at Bayer’s Baytown, TX unit; Robert (Bob) Martelli, an Engineering Specialist in Facilities Engineering at Dow Corning in Midland, MI, and, from the vendor side, Jaike Williams, President of Fluid Solutions, a distributor in Chapel Hill, NC, and former sales engineer with more than 15 years of canned motor pump experience.

canned motor or mag drive pumps specifically to meet emissions standards. Where we are using sealless pumps, it has been for safety of personnel or for the environment, not so much the cost. The cost difference between a conventional sealed pump and that of a canned motor design is not as significant as it used to be.

Pumps and Systems: What factors related to costs and operation were evaluated before pump selection?

Burrell: Price and availability of spare parts are important to us. We prefer a domestic manufacturer, with pump efficiency a key factor.

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Hrivnak: Most of our liquids have viscosities above .7 cP, meaning that single mechanical seals typically work well. So even though we have 30,000 pumps at this site, we do not have a lot of

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Martelli: Most of the pumps we look to purchase today are below 50 hp and develop less than 400 ft of head. We have had trouble over the years with poor mechanical seal lubrication of most silicon based fluids we need to pump. This resulted in many seal failures – with seal life of a few days to less than 6 months. While some consideration to cost was given, reliability was and continues to be a large factor in selecting canned motor pumps. We started replacing pumps below 20 hp with canned motor pumps about 20 years ago. And they did last longer than sealed pumps. We typically got a year’s life out of them between repairs. Some units have run 2-5 years on a set of carbon bearings.

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Williams: It is my impression that can pumps were a bit oversold on the sales side of the equation when they first became perceived as a sealless environmental solution, and the users sometimes had overblown expectations of product capabilities – a bad combination

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The Pump Handbook Series

that caused some improper applications.

Q:

P&S: In what type of service are the canned motor pumps at your facility being used?

Hrivnak: Our canned motor pumps are used for hot oil service, typically above 500ºF.

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Martelli: We have canned motor pumps in most siliconbased fluids service whose viscosity is suitable for centrifugal pumps. We use non-metallic lined magnetic drive pumps for services where 316 SS and Hastelloy alloys are unsuitable, such as aqueous acids. Burrell: Our can pumps move toxic chemicals, mild slurries and clean solvents.

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P&S: What have you learned about the application of canned motor pumps?

Hrivnak: Once we work the bugs out in a new system, we see service lives for our canned motor pumps at about 5 years between bearing replacements. Hot oil systems can have problems with collecting noncondensibles. We have learned that certain designs keep the gases from being trapped in the bearings. As you know, you have to keep the bearing lubricated and cool. Also, in hot oil service, you must keep the liquid in the bearings below about 350ºF to get good lubricity. We

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usually do this by recirculating some of our pumpage from the bearings to a cooler and then taking it back to the bearings. We have to place a temperature switch in this line to determine if cooling is taking place. We have had several instances where the cooling water side of the exchanger has plugged; so a cleanable cooler is also a must for our type of service. Also, hot oil systems do create coke, and a canned motor pump has to be designed to minimize the chance of dirt getting into the bearings. Martelli: A little over 10 years ago, we started performing vibration monitoring on canned motor pumps. It took 2-3 years of work to convince operations and shop folks that this technique could predict an impending failure. This has had the largest impact on preventing catastrophic failures. However, we have found that this technique does not work on magnetic drive sleeve bearings. I believe that the rotor mass is too small in most magnetic drive pumps for bearing wear to impact overall vibrations. Because of this, we do not have many magnetic drive pumps in service with carbon bearings. On some installations, especially hot reflux pumps, the systems was designed for a pump with a lower NPSHr than what canned pumps typically provide. Running a pump at reduced speed with VFD, and possibly operating the motor with a subcooled flush can overcome the NPSH limitation. A sub-cooled flush arrangement takes the process fluid normally being pumped and runs it through a heat exchanger to drop the temperature 10-20ºC prior to flushing the motor end of the pump in the normal fashion. Doing this can prevent the fluid from flashing in the motor/bearing area of the pump.

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Burrell: Every different application must be evaluated separately with the pump vendor. Process fluid or external flush must have some lubrication properties. In the case of a clean external flush liquid, the characteristics of the liquid – boiling point, vapor pressure, etc. – must also be considered.

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Williams: As far as my applications experience with users, I have always felt that carbon graphite as a bearing material is suitable for a great majority of applications and wonder what is truly gained by using silicon carbide material. I have yet to see a convincing study that differentiates the two in canned motor pumps.

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P&S: What are the best attributes of your canned motor pumps?

Hrivnak: For us, the lack of leakage to the outside environment and long running times are the two most important.

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Martelli: Our experience has been that the most successful canned motor pumps include oversized bearings with high flush rates. Using this design has helped us achieve more successful start-ups that have led to longer service life. Of course, we cannot afford any leakage of the fluids we deal with, so eliminating routine leaks to the atmosphere is also a big plus.

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Burrell: We like the ease of repair – there are fewer parts than in mag drives. Also the pumps are compact, and not much space is needed to install or remove them. But parts can be expensive, and normally the delivery times are long in the case of a major overhaul.

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Williams: Besides the inherent safety, the single best feature of a canned motor pump – properly applied – is the drastic reduction in maintenance dollars. Period. This is a driving force in Europe and Asia, but until recently it has been secondary in North America. I believe sophisticated users of these pumps have now learned this and are putting this knowledge to good use.

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Q:

P&S: Besides the usual concerns of pump operators, are there additional issues relating to canned motor pumps that have required your attention? The Pump Handbook Series

Hrivnak: Worn bearings are easy to replace, but in several instances, before we had all the needed instrumentation in place, we broke bearings because of loss of cooling or vapors collecting in the bearings. These failures necessitated the re-canning of the rotor and stator. They’re expensive to repair and require 6-10 weeks of downtime. Catastrophic failures are hard to repair!

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Martelli: Motor temperatures for heat transfer pumps are very important. Heat transfer fluids at high temperature (400ºF and above) are as difficult to pump as light hydrocarbons and can result in reduced bearing life. At Dow Corning, we have one system circulating Syltherm XLT (R), a lower temperature range fluid, with a pump motor temperature of 180ºC (356ºF). This is a pump with ceramic insulation on the motor windings for high temperature service. After extensive dimension checking on the pump and balancing to ISO quality grade 2.5 on the assembly, and pumping 100% of the time near BEP, the bearing life is still only 14 months. 1x vibrations are 0.03 IPS at running speed with no unusual harmonics or sub synchronous peaks right after a bearing change. Vibrations start rising at 1x running speed after about 10 months of operation. This is in complete contrast to another pump on the same system that runs with a temperature of 100ºF in the motor fluid area. It has operated for more than 5 years on the original set of bearings, including all the start-up flushes! The difference is the temperature in the bearing area of the pump. The viscosity difference is 1.4 cP at 100ºF vs. 0.26 cP at 380ºF.

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Burrell: All systems related to the canned motor pump must be considered and evaluated prior to purchase. One relevant question focuses on the properties of external flush. Is a cooling jacket needed? The flow rates of auxiliary systems, instrumentation for auxiliary systems, predictive maintenance requirements, and what type of bearing wear indicator is used are additional considerations.

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Williams: Industrial users learned of canned motor pumps’ sensitivity to solids the hard way, especially at start-up. The nature of the bearing clearances makes them susceptible to failure in a plant initiation phase. Pipe line trash, weld slag, etc., is always ingested into the pumps. With conventional pumps the seal would usually be affected – an easily replaced maintenance item. However, if startup filters were not utilized in a canned pump’s initial run, the bearings and liners would often fail prematurely. Canned motor pumps will always be somewhat expensive to repair in the event of a rupture of either the rotor or stator liner. This expense, plus the decontamination cost, will often negate the repair, and it becomes more economical to replace the entire unit.■

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R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

Winning Maintenance Strategies at Thorn Creek Cutting wastewater pumping expenses requires a grasp of details as well as the big picture By George Kracke

fforts to improve wastewater pump operation have reduced downtime and improved effluent quality for the Thorn Creek Basin Sanitary District (TCBSD) near Chicago. TCBSD serves about 100,000 people and was organized in 1921 under state statutes to provide wastewater treatment services to incorporated areas in six communities. The district’s wastewater is treated at a computerized central facility in Chicago Heights. TCBSD’s fees are among the lowest in Illinois. This is a result of its large scale operation, use of modern technology, sound planning and efficient operation. Charges are based on actual water usage. This allows customers to control their wastewater treatment costs. Effluent quality entering Thorn Creek from the treatment plant consistently exceeds environmental quality standards and has enabled TCBSD to win Illinois EPA and various technical society awards. The current wastewater load is 17.1 mgd at a rated capacity of 40.25 mgd for the District’s treatment facilities.

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Operation Four off-site lift stations, each with three lift pumps, discharge the effluent into large diameter mains leading back to the district plant. Gravity mains also feed directly into the plant. The plant processes effluent through screens to eliminate large objects that could cause pump problems. The system has

about 120 pumps, most of them of the centrifugal vertical lift type. These range from 3/4 hp to 100 hp in size and include Aurora, Chicago, ITT Marlow, Fairbanks Morse, Allis Chalmers and other makes. The variety of pumps affects both service and parts stocking. Ten of the district’s 40 employees work in maintenance. Because of the variety of pumps, it is difficult to have expertise in any one type. However, they do possess knowledge and training that comes with experience, and regularly attend schools and seminars to increase that knowledge. In our experience, the training offered by pump suppliers and manufacturers has been particularly valuable in assisting us through normal operation and maintenance situations. Standard procedures include preventive maintenance inspections at regular intervals. These include both visible and audible inspections. Flow is also monitored continuously by computer to check both current flow and pump output. The Pump Handbook Series

This information is used to determine pump service requirements. When a pump fails, the rebuild is made. If visible or audible flow problems are detected, the pump is taken down for repair. We do have the luxury of “triple redundancy,” in which one pump in a station of three is able to carry the normal load. This is an advantage that most companies do not have. Most problems are found to be seal related. In pumping waste, much of the debris is captured in the area around the seal and impairs seal flushing. Another cause of failure is the impeller itself. This is due to age or debris that causes unbalance and negatively affects the seals and bearings. Most obvious problems such as bearing failure and impeller damage are detected through preventive maintenance inspections.

Types of Pumps Nearly every type of pump imaginable is represented in our inventory of 120 units ranging from fractional to 100 hp. This includes vertical low lift, centrifugal — both vertical and horizontal, diaphragm, progressive cavity and plunger types. Nearly as many manufacturers are represented as there are types of pumps. Reliance on low bidder contracts as well as 60 years of operating experience have contributed to a diverse inventory that tests operation and maintenance skills. Most pump repairs and rebuilds

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are done on the premises. Over many years plant personnel have acquired the skills needed to make these repairs. The latest technologies are used as appropriate in pump rebuilds including new seal types, bearings, lubricants and pump materials. Many of the original pumps installed at our plant are still in service. This testifies to the original quality of manufacture as well as to the maintenance procedures. Pumps range from 1” to 12” in size, with most in the 4” to 8” range.

Spares and Parts Inventory Dealing with so many different pumps has a negative effect on spare parts. We have been able, however, to standardize on a seal supplier and have converted many pumps from packings to mechanical type seals. A nearby supplier also provides our bearings in one or two days. Some important applications require that we keep a spare shaft and other internal parts on hand as well. Only about five pumps require this level of support. We have an established network of suppliers for bearings and seals as well as machine shops, pump rebuild shops and others that we call on as needed. Again, this emphasizes the importance of developing a good vendor base and establishing good working rapport.

Pump Repair Methods Standard lock out and tag out procedures are used to remove a pump when repair is deemed necessary. The pump is then brought back to the maintenance shop and cleaned up for disassembly. Care is exercised when removing parts to avoid additional damage. All parts are then cleaned for identification and inspection. Procedures are instituted to help determine the cause of component failure. Previous operating history and preventive maintenance records are used as a starting point. The shaft, impeller and volute are examined for wear and repaired as necessary. Repairs may include rebuilds or replacements as economically feasible. When repairs of rotating parts are made, the entire assembly is sent out for balancing. This pro-

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cedure has proved to be both prudent and cost effective. The pump is re-assembled and care is taken to install the bearings, seals and other parts correctly according to the service manuals supplied by the pump manufacturer. Because the motor has a long life span and seldom fails under normal conditions, it is frequently not considered a part of the system. But it should be. The electric motor that drives the pump is also examined during the repair and is evaluated for preventive maintenance work. Records are consulted to determine the last time it was serviced. If the service record indicates more than 7 years for dry locations or 5 years for wet, the motor is sent to the local motor repair facility. The motor is disassembled, cleaned, dipped and baked. New insulation and bearings are added as well. Additional procedures specify that we do not repair motors that are 5 hp or less unless they are special duty or a unique design. Otherwise, all motors in this category are replaced. For most rebuild situations the bearings, seals and gaskets are replaced. Even though only a seal may have failed, the bearings and other parts are also replaced. This precaution is normally taken if the system has been in service one year or more since last repair. This is economically feasible as the bearings cost only $150, not including labor. The measure is well worth the extra effort. Most seal failures occur in less than a year due to debris that causes the flushing water to cease flowing properly. Many procedures have been instituted to improve the life and reduce the downtime of our pumps. These changes include: • Improved seal water supply source. Seal water should be clean and free of obstructions, pressure should be adequate to flow through the seal, and use of visible water wheels is recommended to indicate flow. Seal water is routinely inspected each day. • Change impeller size to ensure proper operation. This is a The Pump Handbook Series

one-time repair. • Adjust discharge valves to maintain proper head so the pump will operate within its curve. • Change inlet and outlet piping where necessary to provide laminar flow. This reduces cavitation. • Use electrical measurements to make sure the system is balanced. Plant gasses can corrode starters and contactors, and electrical problems can lead to motor failures. Also take into account that copper wiring can be attacked by hydrogen sulfide and cause a current unbalance because of resistance. • Inspect and run all pumps weekly; listen for changes in noise or vibration; and heed the advice and observations of the operators because they live with the systems every day.

The Big Picture Common sense, good troubleshooting skills, attention to details and a strong work ethic are important. Setting up a good control system using reliable software such as Datastream is also very helpful. This enables the user to go back and check operating history to help make critical decisions. Accurate records of repairs and relevant maintenance information are likewise essential. And the ability to form a team of reliable workers and suppliers is vital because pump malfunctions are an inescapable part of life.■ George Kracke has been involved with plant maintenance for more than 20 years and is currently the Superintendent of Maintenance for Thorn Creek Basin Sanitary District (Chicago Heights).

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R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

Innovations in Multiphase Hydrocarbon Operations A comprehensive review of practical field issues related to the planning, installation and operation of helico-axial multiphase pumps. By Charles de Marolles and Jacques de Salis

his article reviews practical field issues related to the planning, installation and operation of the helico-axial multiphase pumps. Interest in multiphase production, which leads to simpler and smaller in-field installations, is primarily dictated by the need for more cost effective production systems. Multiphase pumping is essentially a means of adding energy to unprocessed effluent. This makes it possible to transport a liquid/gas mixture over long distances without the prior separation. Under normal operating conditions, the Poseidon helico-axial pumps, are largely unaffected by process fluctuations (changes in pressure, liquid or gas flow rate) at the pump inlet. These pumps have demonstrated stable behavior. They can adapt on their own to instantaneous changes in pumping conditions. Indeed, a multiphase pump set is designed to operate under changing/fluctuating process conditions. An important issue related to pump operability and flexibility has to do with the driver selection. In some cases a fixed speed drive provides sufficient operational flexibility. In other cases variable speed should be chosen. Pump operation and control

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strategies are presented and discussed, and recommendations and guidelines for the field system designer are given.

Introduction In recent years a number of oil companies have shown considerable interest in multiphase pumping. The ability to boost hydrocarbon fluids directly from the well before the gas, oil and water have been separated gives rise to a number of advantages: • Development costs can be dramatically reduced with the elimination of platforms or processing facilities, especially in marginal, hostile environments. • The pump can be installed as part of a conventional process to boost the pressure upstream of a high pressure separation • The production of reserves from marginal fields can become economically viable. • Reducing well head flowing pressure can increase ultimate well yield. • Gas currently being flared can now be recovered. • Permits pumping from single wells or manifolds to (close-by) existing production facilities. When the production of a marginal field or a group of remote wells is considered with an existing cenThe Pump Handbook Series

tral gathering system, the traditional field development options are: • natural flow • artificial lift • in-field separation with crude oil transfer pumps, gas to flare,or gas compression systems. With the recent field deployment of numerous multiphase pumps (MPP), effective new approaches to field development and production have been demonstrated. Typically, helico-axial MPP have accumulated more than 40,000 operating hours with an average availability of approximately 90%. Considering multiphase flow transient behavior, the operability of MPP is highlighted in the case of declining offshore fields. Different levels of pump control are defined depending on the operational requirements.

Pump Specification Multiphase pump selection cannot be based solely on one defined operating point (main duty point) as is commonly done for process pumps or compressors for refinery duty. Actual operating conditions change during the field life and may be different from predictions. Therefore, multiphase pumps should be de-signed for different operating parameters. They should have a

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Pump Design General

The helico-axial pump (Figure 1) is a multistage pump capable of handling raw effluent – that is, a mixture of oil gas, formation water possibly containing H2S, CO2 and a certain amount of solids. The Sulzer Pumps standardized helico-axial multiphase pump range consists of eight frame sizes covering total volumetric flowrates (oil, water and gas) at suction conditions from 22,000 bbl/day (146 m3/h) up to 220,000 bbl/day (1,458 m3/h). It is based on the latest second generation helico-axial hydraulics (Poseidon license) developed by the Poseidon Group (Institut Francais du Petrole, Total and Sta-

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(Photo Courtesy Sulzer Pumps)

wide working envelope, in other words. Flow rates and pressure predictions are bound with uncertainty. The design of production facilities is generally based on extrapolations from the results of exploration well tests, delineation wells and reservoir characterization. As a result, when the production wells are drilled and completed, actual production data may be different from predictions. Furthermore, in the field, life production data (oil flow rate, WC, gas fraction, pressure, temperature, etc.) will change. This long term evolution is induced by the natural reservoir depletion. Medium term fluctuations may also be encountered during transient operations (startup, shut-down, well-testing, pigging operations, etc.) where large gas pockets or liquid slugs can form. Well instabilities can also contribute to changing flow and pressure conditions at pump suction. Some wells at end of life, with or without bottom hole activation, demonstrate an unstable cycling behavior characterized by an active period (producing) followed by an inactive period. Therefore a multiphase pump should be designed for variable inlet conditions. It should also offer the possibility of a large operating domain (in terms of flowrates and GVF variation). Operational flexibility is preferable to efficiency optimization for a narrow operating band only.

Figure 1. Section of MPP multiphase pump – low speed version

toil). Two mechanical designs are available: high speed (up to 6800 rpm) and low speed (below 4000 rpm). The following description concentrates on the main characteristics of the low speed range, which is considered for the case study included in this article. The main advantages of the helico-axial hydraulic pump are : • the ability to pump any GVF from 0 (100% liquid) to 1.0 (100% gas) on a continuous basis. • mechanical simplicity and reliability (one single shaft, rotodynamic principle). • compactness • self-adaptation to flow changes • tolerance to solid particles (open type axial impeller, without tight clearances).

Compression Stages Each compression stage consists of an impeller mounted on a single rotating shaft, followed by a fixed diffuser (Figure 2). The impeller blades have a typical helical shape. The profile of the open type impeller and diffuser blading arrangement is specifically designed to prevent the separation of the gas-liquid mixture during the compression process. The hydraulic passages accommodate solid particles in suspension, and special care has been taken in the design to prevent their accumulation in the pump casing. Due to the compression process the gas volumetric flow rate decreases as it goes through the pump. Therefore, the pump is generally equipped with different series of hydraulics – different series of geometries – whereas one series would have identical compression stages. The changing geometry from one series to the next provides The Pump Handbook Series

Figure 2. The helico-axial hydraulics (orange: stator, yellow: rotor)

adjustment for the decreasing volumetric flowrate. Important features of this design are the hydraulic flexibility and the wide range of duties that can be met by a single pump. Also, the multiphase pump can be easily retrofitted to take account of changing reservoir characteristics during the production life of the field; in particular, the oil flowrate can be maintained longer.

Mechanical Design Based on Sulzer’s long established experience with injection pumps and centrifugal compressors, a modular design for the multiphase pump range has been developed and successfully introduced to the market. The design is in general accordance with API 610. It is a multistage barrel in-line pump with axially split inner casing. The in-line arrangement is required to facilitate transport of the solid particles contained in the pumped fluid. The impeller tip clearance is in accordance with API 610 requirements. The diffusers are in two parts and located in an axially split inner case. This design allows easy inspection of the rotating parts. Further, the rotor does not have to be dismantled during the pump assembly. Therefore, a low residual rotor unbalance can be achieved, guaranteeing a low vibration level. For ease of maintenance, the hydraulic parts (complete rotor with

Instrumentation for Pump Monitoring Similar to other pump units used in oil production, multiphase pumps are typically equipped with following conventional instrumentation suitable for hazardous area and outdoor operation : • pressure and temperature gauges at suction/discharge branches • bearing bracket velocity sensors • speed measurement (if variable speed drive) • bearings temperature monitor • temperature and pressure sensors for the lubrication/sealing circuit The driver is also monitored. For unmanned operation, automatic monitoring of the complete pump/ driver unit can be done with a PLC.

phase pump package is generally mounted on a horizontal skid (Photo 1). The following items are located on the skid: • Multiphase pump • Buffer tank upstream of pump (if required) • Lubrication/sealing fluid system with coolers • Driver • Pumpset-mounted instrumentation Various drivers can be used:

Pump Control & Operation Strategies General

• fixed speed electrical motor • variable speed electrical motor • gas or diesel engine • hydraulic turbine • gas turbine

Depending on the production cases and operational requirements, various control strategies can be proposed. Standard practice calls for the appropriate level of mechanical protection for the pump driver unit. Typically, trigger values are used for alarm to prevent undue shut-downs and for trip to avoid damages to the equipment.

A variable speed drive provides optimum control and very high operational flexibility since the pump output can be adjusted at any time to accommodate short term changes in production requirements and/or long term field decline. Compared to the traditional method of separation, in which all the wells have to work with same inlet separation pressure, a variable speed multiphase pump would allow a controlled reduction of well head pressure. Induced benefits may include reduced loss of production due to water breakthrough, postponed stimulation and maintenance of gravel-packed wells, and longer production periods for low pressure wells. However, a variable speed is not a requirement for helico-axial pumps, which are self-adapting as discussed.

Buffer Tank The buffer tank is basically a static mixer of small dimensions that can be installed upstream of the pump suction nozzle in case of slug flow. Its main functions are to break the energy of liquid slug fronts, smooth out any fluctuations in the incoming flow and act as a sand trap. Potential rapid repetitive torque changes (pulsations) produced by sudden changes in mixture density, with its associated pressure changes as produced by a strong slugging behavior, would be eliminated. Obviously, when instantaneous production flows do not vary much from the average values of liquid and gas flowrates, there is no need to install a buffer tank.

Driver Sizing The helico-axial MPP can run

Package Design Like any pump unit, a multi-

from a mechanical point of view in 100% gas and also in 100% liquid. The power required by the pump for those two conditions is very different, as the mixture density would change by a factor of 100. It is not economical and for all practical purposes not required to size the driver for full power for the maximum output in 100% liquid. If one would do so, the driver would be oversized and work most of the time at part load.

(Photo Courtesy Sulzer Pumps)

impellers, diffusers, inner case and suction casing) are designed as a pull-out assembly. The shaft sealing is achieved with mechanical seals. To avoid gas leakage or solid particles contaminating the bearings, and also to allow dry running conditions, a pressurized external fluid is injected. This maintains an overpressure of 5 bar between the seal fluid circuit and the suction/discharge nozzle. A seal fluid reservoir is located on the skid. One combined seal and lube oil system is used for lubrication/cooling. Operating conditions sometimes dictate the need for materials with excellent corrosion and erosion resistance. Effluent may contain hydrogen sulphides with water and chlorides, sometimes combined with high temperatures. This is a corrosive situation. Sand may be entrained by the effluent, resulting in erosion of pump component surfaces. To compensate for aggressive effluents of this nature, all wetted parts would usually be made of duplex stainless steel. Exchangeable wear rings can be mounted in front of the impellers and diffusers.

Photo 1. P302 package system at SulzerWorks in France The Pump Handbook Series

Self-Adaptive Capability The self-adaptive capability (stable behavior) of helico-axial multiphase pumps has been observed both on multiphase flow loops and in the field. One particular example is the succession of liquid slugs and gas pockets with a frequency of less than a minute while the pump runs

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Process Control

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75

QG(m3/n) qG x 1.25

QL(m3/n) qL x 2

Under normal conditions a constant output (oil 25 50 production) over BUFFER TANK 0 DAMPING EFFECT time is generally P 40 required – that is, 40 the pump operates 20 P at a constant speed. Q Q The well output at Q T Q Buffer Pump Motor a given time de0 1 2 Tank Time (min) Po pends essentially Figure 3. Buffer tank damping effect on the wellhead flowing pressure 4 (WHFP). To inN crease the well outT put, one would 2 have to lower the WHFP. This can be done by increasing 16 P the pump speed P (Figure 4). A speed P ∆P 14 6 4 reduction (rise), P 2 12 decreases (increasQ P 0 10 es) the pump dif15 30 45 60 N Time (min) ferential pressure. T As the pump disFigure 4. Speed variation pressure rise charge pressure stays in first at constant speed (Figure 3). The approximation constant, the suction buffer tank smoothes the GVF flucpressure will therefore rise (be lowtuations. The gas and liquid flow ered). rates at buffer tank inlet (QG, QL) Fixed speed installations are posvary widely while at pump inlet they sible as long as one can tolerate a cer(qG,qL) are dampened. tain amount of back pressure When more gas (gas pocket) fluctuation on the wells. For more enters the buffer tank and the liquid critical applications variable speed level in the tank tends to drop, the drive allows full process control. instantaneous GVF of the mixture Manual process control, by adjusting entering the pump rises, and therethe pump speed set value, enables fore the differential pressure generoperators to produce the required ated by the pump diminishes. The output. This would also allow, with outlet pressure (p0) stays constant in time, adjustments for natural field a first approximation – due to pipe depletion. inertia and pressure regulation of A manual control system would production separator downstream of still offer the required mechanical the pump. Pump suction pressure protections, but it allows only a man(p1) increases. As the pump suction ual speed setting. The system will pressure increases (this may or may not be able by itself to adjust the not be combined with more liquid speed automatically. The pump will entering the buffer tank), the efflueither run at the predetermined speent (mixture) density will rise, and ed or be shut-down. Obviously, the reverse mechanism will take speed setting can be changed manuplace. ally by the operator at any time. The pump discharge differential A more elaborate process conpressure (dependent on mixture dentrol can be achieved through autosity) adapts itself. This is the selfmatic governing of pump speed. This adapting nature of the helico-axial allows a person to set and maintain a multiphase pump. requi-red production output by 150

50

P (bar)

T (dNm)

O

I

L

L

G

T(100.Nm) N(1000 RPM)

G

O

∆P(bars)

P (bars)

O

I

214

The Pump Handbook Series

avoiding interferences from upstream or downstream elements (changes in setting of manifold valves, chokes, well control) and by preventing unnecessary shut-downs.

Conclusions Helico-axial rotodynamic pumps have the flexibility to handle wide ranges of flowrates, both steady-state and transient. They also cope well with slugging. In some cases fixed speed operation provides enough operational flexibility. In other cases, where back pressure control on the wells is more critical or adaptability to evolution of operational parameters is essential, a variable speed system is preferable. Pump control and monitoring can be more or less elaborate depending on the application, expected well behavior, pump redundancy, importance of maintaining a continuous production flow, manned or unmanned facilities.■ Charles de Marolles is Technical Coordinator for Sulzer Pumps Europe and, since 1992, has been responsible for the company’s multiphase pump mechanical development (onshore, offshore and subsea). Jacques de Salis has been with Sulzer Pumps for more than 12 years and is currently Program Manager for the multiphase pump product in France. Editor’s Note: This article is based on the authors’ presentation on the subject at the 13th International Pump Users Symposium and an article (SPE 36591) Multiphase Pumping – Operation and Control, 1996 Annual Technical Conference, Denver, CO. Portions have been reproduced with permission of the Turbomachinery Laboratory, Texas A&M University, College Station, TX, pp. 164-169, Copyright 1996.

Field Parameters and Pump Selection To size a MPP for a particular application, the following data is required as a minimum: • Oil flowrate • Standard Gas Oil Ratio (GOR at standard conditions) • Water cut (WC) • Pump suction pressure • Required pump discharge pressure The pump size is determined by the total volumetric flowrate at suction condition. To estimate this total flowrate, one needs to determine the Gas Volume Fraction (GVF) or the Gas liquid Ratio (GLR) at suction conditions. Terms in use are summarized here: GOR: measure of gas quantity in a well stream expressed in standard cubic feet of gas per barrel of oil, scf/bbl (Sm3/m3). The standard conditions are generally a temperature of 60ºF (15ºC) and a pressure of 14.7 psia (1 bar abs). GVF: ratio of volume of gas to that of the total volume of the effluent (oil, water and gas) at pump suction temperature and pressure, expressed in volume/volume (bbl/bbl or m3/m3). GLR: ratio of the volume of gas to the volume of total liquid (oil and water) only. WC: ratio of the volume of water to the volume of total liquid. The following interrelationships can be used: • rough approximation: GLR = GOR x (1-WC) x [(460 + temperature in ºF) / 520] x (14.7 / absolute pressure in psia)/ (5.615 cft/bbl) • exact: GVF = GLR / (1+GLR) Example based on the case study at the end of this article: GOR = 560 scf/bbl WC = 0.50 Suction temperature = 100ºF Suction pressure = 147 psia The approximation becomes: GLR = 5.37 vol/vol, hence: GVF =84.30 %. Note that the approximation does not take into account the phase changes between the gas and the oil and therefore is inaccurate. Nevertheless, it is a first evaluation when no other data (effluent composition) is available.

The Pump Handbook Series

CASE STUDY Sulzer P302 Pump for the Pecorade Field P302 is a nominal 40,000 bbl/d (total volumetric flow rate at suction conditions) helico-axial multiphase pump based on the Poseidon hy-draulics. It is driven by a 600 kW variable high speed electric motor. The P302 pump is de-signed to suit the design point and working domain shown in Table 1. These specifications are based on the requirements for the Pecorade field (located onshore, France). The P302 features significant advances in terms of helico-axial multiphase pumping. It was designed for relatively low suction pressure (60 psig) and high compression ratios (up to 6). The pump was delivered to Institut Francais du Petrole (“IFP”) at the end of 1993. It was then installed in the IFP multiphase loop at Solaize (France) for bench testing. Pump tests were carried out under steady state and transient conditions. The pump was tested both under single phase conditions (liquid or gas) and with a multiphase mixture (fuel-oil and nitrogen) at various GVF and suction pressures. Theoretical predictions were compared to measured data. Experimental measurements agreed very well with the theoretical prediction, and the achieved pressure rises exceeded the design specifications. The pump was commissioned in June 1994 on the Pecorade field, which is operated by Elf Aquitaine Production. It has been running in a fully operational field environment, logging more than 13,000 hours as of December 1996.

Pumpset Package The P302 multiphase pump with its electric drive and lubrication unit is mounted horizontally on a common baseplate. It is 6.3

215

m long and 2.3 m wide overall. Total weight is 9.2 metric tons, baseplate included. The control panel and electrical utilities are installed at some distance in a non-hazardous area. The pump rotor, diffusers, bearings, shaft seals and end cover are designed as a cartridge assembly that is introduced into the casing to reduce maintenance downtime. Materials were selected for hot sour fluids (high H2S content: 9% mole) that may contain solid particles. All wetted components are made of duplex stainless steel. A chromium oxide coating has been applied on the wear parts. Bearings and the mechanical seal are designed to be lubricated by formation water supplied from an external source. Sea water could be used in platform applications. The lubrication system is pressurized above the pump internal pressure to prevent any ingress of effluent. The pump is driven by a 600 kW variable high speed electric motor. It is an asynchronous two pole 660 V motor. The speed can vary from 2500 to 6000 rpm. It is adjusted by changing the supply frequency between 0 and 113 Hz. The electric motor is air-pressurized for use in hazardous area, with protection according to class IP 55 EExp II T3. It has forced lubricated cylindrical radial bearings. Frequency adjustments are obtained by a thyristor controlled inverter. The 750 kVA inverter is inserted between the constant voltage (660 V) and frequency supplied by the network and the asynchronous motor. The frequency inverter, transformer and control panel are not designed to operate in an explosive atmosphere, so they are located in a remote safe control room.

Field Experience The Pecorade field is producing about 90,000 bbl/d (600 m3/d) of fluid (oil and water) and 3,533,000 scf/d (100,000 Sm3/d) of gas from four wells. Three wells are gas-lifted while one is pumped. The effluent temperature is about 140ºF (60ºC). The oil has a gravity of 35ºAPI. Oil and gas are transported separately to the

Units

Design Point

Suction pressure

psig

Discharge pressure

Lacq plant through two 6” pipes over a distance of 37 km. A specially built manifold on Pecorade facilitates operation of the pump upstream and downstream of the test separator. The piping was designed for both hydraulic performance testing and field operating. Hydraulic performance tests were conducted first. The test separator was then connected upstream of the pump. This configuration brought a controlled and steady state multiphase flow to the pump. This first phase, which consisted of several stop/starts per day, was completed at the end of November 1994. It showed excellent hydraulic performance exceeding the specification (Table 1). For given suction conditions (pressure, GVF, flow rate) the calculated and measured pump speeds were compared for the same pressure rise achieved during the tests. Measured speeds are generally between 5 and 10 % lower than the calculated ones for GVF up to about 0.95. For a given rotation speed, the measured multiphase pressure rises are above the calculated ones. In summary, the P302 pump features excellent pressure rise capabilities in multiphase flow, even at low suction pressure and at high gas volume fractions. A second phase of endurance testing was then started. Problems were limited to the partial removal of a protective coating (without damage of the underlying duplex stainless steel) after 2350 hours operation and a slight displacement of an impeller on the rotor shaft during a long operating in surge/severe slugging conditions after 4000 working hours. The pump has been dismantled and inspected twice, showing no signs of wear or erosion or corrosion. By December of 1996 it had accumulated about 13,000 hours, with an average availability per month of more than 90% allowing for production related downtime. At present the pump is used for normal production operations. It is connected to the well’s manifold and run continuously on a 24 hr basis. The P302 is operated by the field staff. The concept has been validated and valuable experience gained.■

Working Domain

Operating Domain

109

72 - 218

43 - 246

psig

362

100 - 493

261 - 653

%

86

66-91

50 - 100

Total flow at suction

bbl/d

26,000 - 40,000

15,000 - 55,000

6,000 - 53,000

Speed

rpm

4600 - 5500

3000 - 6000

2500 - 6000

Hydraulic power

kw

230 - 300

100 - 500

100 - 500

GVF at suction

Table 1. P302 pump design and operating specifications

216

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R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

Key Indicators - The Measure of Reliability Chart your reliability progress and pinpoint problem areas by relying on key indicators. By Eddie Mechelay ncreased reliability – the promised land for today’s reliability professional. Much effort is given by magazines (including Pumps and Systems), trade journals, trade shows and salesman on what is the best technology available. These efforts concentrate on new equipment, equipment upgrades and repair techniques. However, is this all that the reliability engineer must concern himself with? Let’s imagine the following conversation. Billy the bright young reliability engineer is talking with his boss, Harry, the maintenance manager. “So Billy, how is the reliability department doing these days?” “Well, boss, things are going great.” “Is that right? How do you know?” “In the last three months, Harry, we’ve done much better utilizing our predictive maintenance programs.” “Really? How do you know?” “Well, we haven’t had the same number of bearing failures.” “Really? How do you know?” Get the hint? Too often we concern ourselves with solving problems and finding solutions that we forget to measure how well we are performing. How do we determine if we are at maximum performance if we don’t measure items that indicate how well we are doing our job? One way to pinpoint problem areas is to rely on these key indicators: • Mean Time Between Failure • Mean Time Between Repair • Horsepower Availability • Cost per Connected Horse power

I

• Items not Available in Warehouse • Repeat Work Orders on Specific Equipment • Mechanical Seal Failures and Cost • Motor Repair Cost • Number of Bearing Failures • Number of Coupling Failures • Overall Plant Vibration

Mean Time Between Failure Mean Time Between Failure, or more commonly MTBF, is quoted and discussed almost as much as quality. What is it? What does it mean? And what is the industry standard for it? These questions are commonly asked. How does MTBF relate to Mean Time Between Repair? My answer to all of these questions is that people at each location must determine for themselves the meaning and interpretation of these indicators. Definitions of “failure” range from complete wreckage of equip-

ment to saying that if the repair is scheduled within three days, it is not a failure. A very basic definition for a failure can be the most enlightening. If a piece of the equipment failed due to mechanical failure or if the equipment must be removed from service over any one shift, it is classified as having failed. A repair is any activity that is outside of running maintenance such as changing oil. This could include replacing coupling elements that have exceeded the recommended running life or pulling and cleaning check valves on positive displacement pumps. All of these “failures” correspond to either the pump or the motor failing. If you wished, you could double your MTBF just by tracking the pump and motor as separate items. This is not recommended since problems with the motor can lead to pump failures and vice versa. The goal for measuring MTBF can vary greatly between individu-

MTBF and HP Availability Log Sheet Equipment I.D.

Day Out

01-P-101 02-P-900 03-P-434

Day In

Total Days Out of Service

HP

HP Days Failure Lost or Repair

1/5/97 1/6/97

1

25

25

F

1/8/97 1/10/97 1/9/97 1/9/97

2 0

75

150 0

F R

Table 1. The Pump Handbook Series

217

als and corporations. Many companies use it as a bragging tool, others utilize it to compare different locations within a larger corporation, and still others use it for what it should be used for – helping the reliability engineer at a specific location monitor his or her performance. These goals usually determine the method for measuring MTBF. One large oil refinery measures MTBF by dividing the number of failures in the past year into the total number of pieces of equipment. This gives a rolling average. Although it would give an overall indicator of a site’s performance, it would not assist the reliability engineer in setting priorities on specific equipment within his or her location. With this thought in mind, another method may prove to be a more effective indicator. That method is to measure the MTBF of each piece of equipment and then average them for a plantwide indicator. This method has the distinct advantage of indicating what equipment most needs attention by reliability engineers. Computing this number can be cumbersome, however, and if the data is unfiltered it can lead to inaccurate numbers. Unfiltered data would be that which is generated using a computer based maintenance work order system that generates a report of failures. If a coding system is not utilized while closing a work order, failures could be attributed to the wrong equipment. As an example, many plants identify their rotating equipment with a number scheme such as 01-P100. The 01 designates the area of the plant. The P says the equipment is a pump. And the 100 is the equipment number for the pump. However, many plants do not have a designation such as 01-PIPE – which would include pipes, valves and fittings. What would happen if the suction valve on the 01-P-100 pump were to fail? If the work order system did not have the 01-PIPE designation, then a valve failure would be written to the pump. If this information is not reviewed prior to an MTBF report being generated, then

218

the report could be misleading. Many might consider reviewing this data as time consuming or tedious. But is it really – as opposed to spending effort trying to solve a pump problem that does not exist? For locations without sophisticated maintenance software, the failures could be logged on a data sheet in the repair shop and then logged at the end of each month into a spreadsheet (Table 1). Using this technique, the MTBF of the equipment would be the number of failures divided into the number of years the equipment has been in service. If one wanted to build a record of MTBF, it would require reviewing all equipment histories and then logging the data. This is not a fun task, but it is well worth the effort in order to determine the plant’s “bad actors.” These calculated numbers can be fine tuned with sophisticated programming. However, this is not required. Let’s examine one last situation regarding MTBF calculated as an average for each piece of equipment. If we assume that a piece of equipment is 10 years old and has had four failures, then the MTBF as an average would be 2.5 years. However, what if all of the failures occurred within the first year and the root cause of the problem was corrected? The actual MTBF of the equipment would be nine years. This example underscores the need for a reliability engineer to scrutinize the key indicators he or she is given.

Horsepower Availability Measuring horsepower outage is an excellent indicator for the performance of rotating equipment reliability (Table 2). One definition of horsepower availability is to measure the amount of horsepower unavailable to run as compared to the total available plant horsepower. An example: If a 100 hp pump and motor train were out of service for five days, then the total horsepower-days lost would be 500 hp days. If we assume that the plant had a total connected horsepower of 10,000 hp, then the horsepower availability for a 30-day month The Pump Handbook Series

would be: [1-(500hp Days/ 10,000hp)] x 30 days = 0.9983 or 99.83%. “Out of service” is defined as any equipment that is unable to be run by the operating department at the end of a shift. It does not include planned unit outages. There are numerous reasons why this is an excellent indicator of rotating equipment reliability. 1. If a plant has too many failures for its work force to repair in a timely manner, then it will be unable to make the necessary repairs in a single shift. 2. If predictive maintenance tools such as vibration analysis, oil analysis and thermography are not utilized or the data not acted on, then unplanned failures will occur. These will not always happen in the morning, so overnight outages will occur. 3. If a plant does not have correct spares or if it has inadequate spare parts in inventory, then equipment will remain out of service until parts can be received. Some equipment can often be used as an emergency spare while parts are expedited. Quickly removing equipment from service without checking on parts can reduce availability. 4. Proper scheduling with operations is imperative to increased horsepower availability. This includes working with operations supervisors to insure that equipment is locked out, tagged and decontaminated for removal prior to the maintenance shift. If it takes a half a day for equipment to be made ready for removal, it will probably be impossible to repair and return it to service the same day. Certain concerns can arise when using key indicators such as horsepower availability. Returning equipment to service in a short time frame is not always desirable. If extra time is required to repair the equipment properly (such as sleeving bearing housings for bearing fits), then by all means the equipment should remain out of service until it is in good shape. In the same manner, if expediting parts or working overtime will increase the cost of repair on a non-critical pump, then this equipment, too, should remain

these problems, site specific measurements should be made. 0.16

Mechanical Seals

0.14

VIBRATION (IN/S)

0.12 0.10 0.08

Average Plant Vibration to Date 0.109

0.06 0.04 0.02 0

Apr’95 Jun’95 Aug’95 Oct’95 Dec’95 Feb’96 Apr’96 Jun’96 Aug’96 Oct’96 Dec’96 GOAL

MOTOR

PUMP

OVERALL

Figure 1. Example of tracking key indicator - plant vibration

out of service so that the most cost effective repair can be made. The reliability engineer must use multiple key indicators to determine how effective the plant reliability program is. No single indicator will give a perfect measurement.

Plant Vibration Many plants do an excellent job monitoring vibration to determine the condition of equipment needing repair. They also trend equipment history very well. What about trending the plant’s overall vibration? One method is to average every data point measured with velocity (in/s or mm/s). Is this not an excellent indicator of plantwide equipment performance? This idea can be expanded to tracking overall motor vibration or overall pump vibration to identify areas that need attention. To further the options, the bearing frequency ranges could be trended as an average. This would indicate how well the plant lubrication and grease program is working.

Component Specific Indicators What if a plant has specific problems in a few areas such as mechanical seal failures or electric motor repairs? When presented with

For years mechanical seal failures and associated cost have been presented at trade shows. But do generic charts and measurements meet every one’s needs? One small plant for years was never given the capital to upgrade mechanical seals. Managers began to track mechanical seal costs each month. After years of reviewing this data, they saw the cost drop from $10,000 a month and then begin to level out. It never went below $3,000 a month on average. However, there was no decrease in seal failures. As money became available, the seal cost increased significantly as old technology seals were replaced with state of the art cartridge designs. The old cost report, though, was still being utilized to measure mechanical seal performance. This report indicated the increased cost without any reference to the increased reliability associated with the newer seals. The point here is that many times indicators need to be reviewed to determine if they are telling the complete story. This particular plant now measures both seal cost and number of failures to give the best picture of mechanical seal performance.

Electric Motors Electric motors are often sent to outside vendors for repair. Is the success or failure of these vendors measured? One facility sent motors to a repair shop that guaranteed a 24-hour turnaround. The result of this guarantee was higher repair costs and reduced reliability. After discussing the advantages of comprehensive repairs, the plant man-

The Pump Handbook Series

agers switched vendors and reduced the number of motor failures. Although the average cost of repair increased by 20%, the total costs decreased by 40%. This success may have never been achieved if key indicators had not been measured.

How Do I Get Started? To achieve maximum benefit from key indicators, it is best that managers and mechanics discuss items that need attention. It is important to select only a few to begin with so that incoming information will not be overwhelming. Keep in mind that starting the measurements from scratch will require at least a year’s worth of data before any action areas will be recognized. Although it is labor intensive, it is recommended to go back as far as possible using information recorded by hand or by computer based maintenance systems. Reliability professionals should take time to measure their performance. To gain maximum benefit of your time and resources, it is important to use key indicators to pinpoint problem areas. Reviewing the data before plotting or calculating, although tedious, can yield more accurate results and save time. Utilizing input from management as well as craftsman can produce valuable insights as to which indicators are most important.■ Eddie Mechelay is Mechanical Supervisor at Total Petroleum’s Denver Refinery and frequent contributor to Pumps and Systems.

219

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R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

New Coatings Make Their Mark Thermal spray materials can extend pump life and improve performance. By Larry F. Grimenstein s the usage of thermal sprayed materials has gained acceptance, the application requirements have increased. Now pump users are asking for special coatings for severe service situations. These requests have led our company to a search for newer and better spray materials. When we began this search, we looked at what materials were available from thermal spray companies. Our search uncovered a limited supply of spray materials from the thermal spray companies. However, there were many materials available in the market from other sources. Those other sources, of course, were welding wire suppliers. Some of their wires have special properties for welding, but we have discovered that they also have special properties that are useful in wire spraying. This article covers only the area of the twin wire arc. There are three basic reasons for all of our developmental work on twin wire arc. One is economics. Wire spray is one of the lowest cost processes available. The second reason for interest in this area is the excellent quality of the spray. The third is that if a welding wire is a specialty material, chances are it will not be found in a powder. This precludes the usage of plasma. Thus, we found that welding wires and twin wire arc are a perfect system for doing developmental work for these new pump applications. One application with which we have been working is a large pump housing approximately 28 inches in diameter. It is used in a steel mill. This pump sees a water environment

A

220

TAFA - 35 MXC® Coating Physical Properties Deposit Efficiency

90%

Melting Point

2500°F

Bond Strength Tensile Strength

8,420 psi 38,000 psi

Hardness

Rc 55-60

Cost

<$10/lb.

*Information furnished by TAFA, Inc. Photo 1. Erosion areas can be seen along the bottom side of the housing.

that is not really corrosive, but the excessive wear is by erosion. The erosion areas can be seen along the bottom side of the housing (Photo 1). A wear plate goes into the base of this pump. Turbulence coming down the casing side to the wear plate, spinning out and coming back up in other areas is eroding these locations. We had to come up with a system that was not only erosion resistant but also economical. We located a material that had excellent abrasion resistance but no corrosion resistance. This wire was high carbon steel micro matrix-composite wire™ - 35 MXC®. Table 1 lists the physical parameters of the sprayed coating. Table 2 shows the results of a wear test where this substance outperformed a special wear resistance material. This wire from TAFA was originally sold for use as an anti-skid coating. With a change in spray parameters, however, it makes an excelThe Pump Handbook Series

Table 1. Physical properties of TAFA 35MXC® sprayed coating

Photo 2. Sprayed wear plate

lent low cost wear resistant coating. Photo 2 shows the sprayed wear plate. It is about 28 inches in diameter and goes down in the bottom of the pump. This was a brand new wear plate that we refinished. We wanted to put on a very wear resistant coating, but there was another problem. This pump receives very large aggregates as well as water. The

*ASTMG65-85 ABRASION TEST RESULTS Coating Thickness

Panel

Th.

Weight Loss

.275

1.20300 g

.025

.0185

0.52024 g

.025

.0105

0.36306 g

.025

.020

0.57223 g

Std 420 SS 35

MXC®

Work Hardening

*Testing done by Certified Lab. Procedure B used

Photo 4. Impeller with worn outside diameter

Table 2. ASTMG65-85 abrasion test results

aggregates are pieces of burned off steel. When they burn off the slabs, the water for the processing gets chunks of steel in it. The pump and wear plate see an environment similar to a wet grit blasting of their surfaces. That causes excessive wear. So we had a problem not only of knowing what type of coating would work, but answering the question of whether it would be good enough. Anytime one of these pumps is started up, the aggregates in the water almost blast the coatings off. We used a coating system in which neither material would work as well separately as when used together. We used a primer system to make sure that we obtained maximum adhesion. We then gave it a polyurethane topcoat. The edges have been coated because we wanted to use the coating to help seal the sides of the plate (Photo 3). This would eliminate the abrasive water going down along the side of the wear plate and eating away at the bottom of the casings. Unfortunately,

Photo 3. Sprayed wear plate with primer system underneath for maximum adhesion and coating on the sides for sealing

there were a couple of areas where the user had to remove some polyurethane coating, and we were told that it was almost impossible to get it off. When we did this, we found out by accident that thermal spray materials are excellent bonding base substances for polyurethane. In fact, customer representatives informed us that in the future they will make sure to mark exactly where they do not want the coating to go because they almost cannot get the coating off once we used this process. This pump, which normally gets pulled at about six months, has been in service almost a year now, and it is still running. We did a complete spray assessment for this pump problem including the impeller (Photo 4). The abrasive media has worn the outside diameter down on this impeller. You can see the little fish hook where the leading edge is worn out. There were also wear marks up inside where cavitation had worn the pump. We used the same wear resistant steel coating. Photo 5 shows the impeller that has been completely rebuilt. This coating is very easy to work with and it is grindable; we hand ground the reworked areas. Also, the leading edges have been rebuilt and squared. The diameter now is actually to the full size of the pump. There was a 10 to 20 thousandths clearance on this diameter, and we were able to spray it to size so it would not have to be machined. The seal area was okay, so we were able to mask it. You can also notice that inside we were able to fill the The Pump Handbook Series

Photo 5. The rebuilt impeller

Photo 6. Housing and impeller of a recirculation pump pulled from service

worn areas and rework them. Once these were smooth, we reblasted and resprayed new material to achieve a smooth, blended surface. Consequently, the repairs would cause no new turbulence. This sprayed pump is better than new, and it has now outlasted a new pump. The sprayed wear resistant steel (Rockwell 5560C) is much harder than any casting material from which an impeller this large could be made. The next application to be discussed is a part that is really unusual. If I were marketing thermal spray materials or wire, this would be the kind of application I would like to

221

show a customer. Photo 6 shows a pump housing and the impeller in the back. Unfortunately, this is after it came out of service. This is a recirculation pump used in an Ecoat paint system. The pump housing was used. We had sprayed and rebuilt the inside with a very wear resistant coating that was also corrosion resistant. The material we chose was called Armacor from Amorphous Technologies International. Table 3 lists the materials this company produces and their physical data. The wire used for this application was Armacor M™, which has good erosion resistance, moderate corrosion protection and a low coefficient of friction. This last property was a factor in keeping this casing from failure. The customer decided not to spray the old impeller. The impeller that went into this pump was a brand new stainless steel impeller. The pump was rebuilt and put in service. When the pump was installed and in operation, something went wrong. It started to generate an end thrust. Photo 7 shows the impeller that came out of the pump when they finally shut it down due to poor flow performance. The impeller had actually ground itself down against the sprayed pump casing about three quarters of an inch. If that had been the original cast iron pump housing, the impeller would probably have gone through the casing and drained the paint tank. The sprayed area inside of the pump casing had a few scratches and one nick in it. We would have never told a customer that a thermal sprayed coating would hold up in that kind of failure mode. In another severe application a grinder pump was submerged in a 20 to 30-foot pit and used to pump shale leachate from a landfill. Chemical analysis of the leachate showed the pH at 1.9. The water also had many active materials such as chlorine and high level of oxygen. Combined, these formed a leachate that was able to dissolve the iron pump casing and distribution housing. Since the pump was iron, a coating could easily be applied to the inside to give it abrasion and corrosion resistance. We chose Inconel

222

Photo 7. The stainless steel impeller ground down against the sprayed pump casing about 3/4”.

Photo 8. Distribution housing sprayed with Inconel 625 and coated with sealer.

625 as the wire spray material for protecting the inside of the pump. Since sprayed coatings are porous, we would have also used a sealer to keep the coatings interface from reacting with the leachate. There was an additional concern that the outside of the pump should be protected with something better than paint. The pump can travel up and down a set of rails to the distribution housing. Because of handling and usage, we decided to use a chemically resistant thermoplastic on the outside of the pump and components. This material was Envelon, an EAA based powder produced by Dow Chemical Co. It can be thermal sprayed and has excellent adhesion and chemical resistance. Table 4 shows a comparison between Envelon and other plastic coatings. The pump was disassembled, and the insides of the casings and the distribution housing (Photo 8) were The Pump Handbook Series

sprayed with Inconel 625 and then coated with a sealer. The original paint was removed, and the outsides of all the components were sprayed with the Envelon coating. The pump was reassembled and shipped for installation. It has been in service for more than a year and is still running fine. These are only three wire sprayed applications in which we have been working to create combinations of materials to solve difficult problems at reasonable cost. At present we are working on additional applications using Haynes Hastelloy® C-22 as a corrosion-resistant material for pumps used in severely corrosive environments. Another new wire material is NOREM 02, which is an E.P.R.I. developed material for applications in the nuclear field where the absence of cobalt is required. This material, which seems to have excellent cavitationerosion wear properties, could be a future pump coating. With the usage of new wire materials in combination with polymeric coating, systems to coat pumps can, in many cases, outlast and outperform original pumps. This is apparent when users send new pumps and spare parts to be coated before installation. The extended life and performance of the pump make the life cost of sprayed coated parts cost effective.■ Larry F. Grimenstein is President of Nation Coating Systems in Franklin, OH. He has been involved in the coatings industry for more than three decades.

ARMACOR PRODUCT SUMMARY

PRODUCT

BOND STRENGTH KSI

THERMAL ARC M H CW DUOCOR

5.5-6.5 5.5-6.5

DEP. EFF. %

DEP. RATE #/HR

>70 >70

30-35 30-35 25-35

SURFACE MICROHARDNESS HV100 GM

BULK HARDNESS Rc (CONV)

1000-1300 1050-1300 800-960 1100-1400

(69-72) (69-72) (58-65) (69-73)

SURFACE* FINISH MICRO-IN

<20 <20

*GRIND & FINAL POLISH WITH 800 GRIT SiC PAPER **Data Provided by Amorphous Technologies International Table 3.

THERMOPLASTIC POWDER COATINGS PROPERTY

VINYLS

POLYESTER

EAA (ENVELON)

Primer Required

Yes

No

No

Melting Point, °C

130-150

160-170

90-99

Pre-Heat/Post-Heat, °C

290-230

300-250

200-175

1.20-1.35

1.30-1.40

0.94-0.97

G-E

E

E

Surface Appearance

Smooth

Slight OP

Slight OP

60° Gloss (Gardner)

40-90

60-95

50-80

Shore D Hardness

30-55

75-85

52-55

E

F-G

E

impact

E

G-E

E

salt spray

G

G

E

weathering

G

E

E

acid (inorganic)

E

G

E

alkali (dilute)

E

G

E

solvent (nonC1)

F

F

G

Specific Gravity, g/cc Adhesion

Flexibility Resistance

*Information provided by Dow Plastics - The Dow Chemical Company Table 4. The Pump Handbook Series

223

✓

R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

Rim and Face Alignment Procedures for Direct Motor Driven Equipment By Ken Hawkins illwright work in alignment of rotating equipment is practiced in highly individualistic ways. The following method is for working on motors and centrifugal machinery located by frictionless bearings. The coupling is in place so that driver and the driven component rotate together.

M

Equipment Required Tools required include dial indicator(s), clamps and extenders necessary to reach from motor to pump with the flexible coupling in place. Horizontal jacking bolts are helpful on each motor foot for horizontal plane corrections. Marked, pre-cut shims of varying thicknesses are suggested, as well as a pair of 4 point alignment levels.

Preliminary Because the dial indicators stay at the same location, eccentricities of shaft and coupling are not being measured. Alignment corrections are restricted to imaginary lines drawn through the centers of the motor and pump bearings. All readings are kept in “total indicator runout” (T.I.R.). Shaft and coupling eccentricities as well as coupling balance should be addressed separately. This is the old story of doing one thing at a time prevents confusion. A few more comments: 1. In an old installation it may be practical to clean the base-to-motor faces and discard old shims first. 2. “Cold” alignment gives accurate results for start-up, but operating temperature of either driver or driven

224

Enclosed Coupling P

A A

P

or a

A(a) Angular Readings P- Parallel Readings

Open Coupling

Driven Shaft

Motor

Take readings at 12,3,6,9. Repeat 12 as a check. Arrow denotes equipment rotation

Figure 1. Alignment readings

components can cause differences. Equipment manufacturers will sometimes furnish the approximate change, but it will be up to the millwright to make a “hot” check. 3. Ratios of motor corrections to alignment changes can be determined by using the following information. There is, however, a catch. Either selfdetermined or computer generated ratios assume “zero” flexible coupling interference, but this is only approached with finished alignment. Misalignment generates forces that shorten the bearing life of operating machines. The same forces are also generated under the static conditions present during alignment. Shaft deflections will amplify “angular” reading – the stiffer the coupling, the greater the amplification. Rotating motor and pump shafts with minimum misalignment strain will save time. Partial flexible coupling assembly or engagement can help, even complete disengagement with rotational control.

Procedure 1. Check foundation and pump bolts for tightness. The Pump Handbook Series

2. Make initial changes. a. With motor bolts loose, rough align visually. b. If motor movement horizontally or vertically is limited, correct before proceeding. 3. Choose an applicable jig and mount the dial indicators, one rim (parallel) and one face (angular). 4. Make a temporary mount of the “rim” indicator jig on a short piece of stiff material (pipe). Invert and record the reading change (sag). Final alignment readings in this vertical plane should include 50% of this sag. That is, if the sag is .004”, subtract .002” from the top center readings and add .002” to the bottom center readings. 5. Check for structural strain. Relieving and tightening the suction/discharge flanges or a single foot bolt on the pump or motor should result in nomore than .002” radial shaft movement of the “rim” indicator. 6. Multiplier determinations (only three needed) a. All parallel (rim) determinations are a 1:2 ratio. That is radial corrections will be 50% of diametric result (T.I.R.). b. Angular (face) ratio determinations: +<——A——>+ <——B——> .} C + + • Span of motor feet bolts, inboard to outboard. • Span, inboard motor feet overhang to “rim” dial indicator

• Diameter or circle span described by the “face” dial indicator. c. Vertical plane angular correction is the vertical face T.I.R. times the motor feet span (A) divided by the “face” indicator span (C). This is the shim thickness to apply under the low end of the motor – or to subtract from the high end. d. For vertical plane parallel alignment correction, add or delete the shims under all four feet, using the 1:2 ratio, until the vertical plane parallel alignment (rim) readings are equal. e. For horizontal plane angular alignment corrections, first reduce the parallel alignment (rim) dial indicator readings to nearly equal in the horizontal plane by rotating the motor around a snug “rear” bolt (Figure 2).

pump is sloped down, and the motor cannot be lowered. 6. The foundation is inadequate for base mounting, or the base is not fixed to the floor.

Figure 4. Horizontal plane parallel correction

the parallel (rim) dial indicator readings until they are equal (Figure 4). 7. Recheck the motor feet for structural strain (twisting). 8. Re-work both vertical a horizontal corrections by trim aligning. 9. Bring the equipment to full operating temperature and “hot” check the alignment for final corrections. 10. These are principles to follow in determining how close to align? a. The closer the running alignment, the longer the coupling and bearing life will be. b. High usage (hours per year) and high rpm will increase the annual cost of misalignment.

Motor Pump

Figure 2. Horizontal plane parallel alignment and correction Mot

or

Pump

Figure 3. Angular correction in horizontal plane

f. Next, continue rotating the motor around the snug “rear” bolt. Indicator measurements taken on an inboard motor foot will be the horizontal “face” T.I.R. times the “rim” indicator overhang (B) divided by “face” indicator span (C) (Figure 3). A little time can be saved by using the rim-mounted indicator readings. They will be the desired inboard foot movement times the ratio (A+B) divided by A. g. Next, using a snug “front” foot, rotate the motor and again reduce

c. I believe that furnishing the millwright with applicable tools (dial indicators, lasers, etc.) – and allowing for the time they will need to achieve the tool capabilities – will give results easily amortized by maintenance reductions. d. Check alignment at 5,000 to 10,000 operational hours. The preceding seems reasonable for alignment with a minimum of trial moves. Now, however, your job is to replace a repaired motor and horizontal split case pump. Here are the conditions you face: 1. There are no piping connections of a type to relieve strain on the pump. 2. Previous alignments were done with a straight edge and feeler gage. Also, workers made use of tin can lid shims. 3. There are no jacking bolts. 4. Bolt holes allow for a bare 1/16” of movement. 5. Frame bound -the axis of the The Pump Handbook Series

7. Soft foot twisting strain between equipment and base (the four legged table problem) is complicated by motor feet that are not parallel to the base. 8. RPM is 3550, and misalignment was the initial reason for failure. 9. You are told that two hours should be a reasonable completion time. To these parameters can be added 90ºF + room temperatures, 90 decibel levels and fatigue that will cause your greatest problem – unwarranted hurry, resulting in time consuming errors. For some possible answers, let’s take one situation at a time, beginning with the time factor. (Item #9) This is the most important single item, and it involves a language problem. Engineering is basic physics, mathematics and time. Management almost always talks money. They can be cooperative if engineers convert their physics, math and time verbiage to “money speak.” A poor pump installation can be very expensive per unit of time. The added cost of a good installation can be quickly amortized, but the truth of such a statement must be demonstrated and the demonstration tran-slated into money speak. If the communication is done correctly, the two sides become cooperating partners in a profitable enterprise. Another side of the coin can be the contracted millwright. Long term incentives are missing, so the machine owner may request that the indicating equipment not be re-moved until he can record the final readings. Regarding #8: misalignment damage is one of the common causes of unit failure. #7: In checking a motor or pump for frame twisting, a motor foot may require a half shim because the base and motor foot are not parallel. #6: The pump and motor base should be firmly mounted to a

225

rigid foundation that will not be subject to cracking. Note that grouting will increase rigidity, and springing a foundation block will decrease rigidity. The result, however, must not include any natural frequencies that match forces in the driver or driven parts. #5: Frame binding will require rotating the pump (with shims under the inboard feet). It may be necessary to loosen the suction and discharge to prevent pipe and equipment strain. #4: Limited horizontal movement will require moving both motor and pump, increasing the motor foot hole size, reducing the bolt size by grinding off unneeded threads or all of the above. #3: Horizontal jacking bolts may not be a necessary addition to a motor that can be moved with the heel of a hand, but the blow required to move a large motor will be inaccurate as well as jog the indicators. #2: The use of a straight edge and feeler gage is rough alignment. It is practical to use them in the preliminaries of mounting prior to attaching dial indicators. Pre-cut and marked shims should be used because they are less expensive than the labor of field shaping. Also, pre-cuts are more accurate, and fewer will be

226

required. After all, multiple layers become spongy. #1: Pump equipment should not be subject to pipe strain. There are many relatively inexpensive devices to prevent the condition. Use will be very important where temperature change is an added factor. The following is a sample of vertical and horizontal corrections done with easily divisible numbers (which you’ll never experience). A. Motor feet span, inboard to outboard: 20”. B. Overhang span, inboard motor feet to rim dial indicator: 10”. C. Face indicator circle span: 5”. Face indicator readings Vertical Plan .005” T.I.R. Horizontal Plane .005” T.I.R. Vertical Plane Corrections “Angular” 20 / 5=4 4 x .005”=.020” Add .020” of shims under the low end of the motor. “Parallel” If the rim indicator shows the motor to be .030” T.I.R. low, add .015” of shim under each motor foot. Horizontal Plane Corrections “Angular” 10 / 5=2 2 x .005”=.010” Rotate the motor around an “outboard” foot and bring the

The Pump Handbook Series

inboard foot in line with the driven unit by moving the inboard foot .010”. The rim indicator can also be used for this change as in (20+10) / 20=1.5 1.5 x .010”=.015” rim indicator movement. “Parallel” Zero the rim indicator readings by rotating the motor around an “inboard” foot. Trim align. Trying for single movements to completion will usually be disappointing. Make each move less than the multipliers indicate. This will be particularly true of large movements. When a single plane is nearly finished, return to the previous plane. It will often have changed. Given the number of motors and pumps in medium and large size plants, a thorough understanding of rim and face alignment procedures can mean an appreciable savings in time and labor. I hope this article has provided you with a good foundation for this basic method.■ Ken Hawkins has more than three decades of practical experience in the pump industry. He currently provides part-time consulting work, specializing in vibration amplification problems and solutions, and is a frequent contributor to Pumps and Systems.

✓

R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

“Best-of-Class” Lubrication for Pumps and Drivers Automatic grease lubrication sytems have helped Scandinavian paper mills become “Best-of-Class” facilities. By Heinz P. Bloch

I

and are always eager to study the merits of competing work processes, procedures, organizational philosophies and equipment options.

Winners We recently visited several Scandinavian “Best-of-Class” paper mills to better understand their experience with plant-wide automatic grease lubrication systems. One of these, Enso Oy’s Kaukopaa mill in Finland, was able to provide detailed productivity improvement data typical of

affiliated and competing pulp and paper facilities in Northern Europe. Their experience with pump lubrication dates back to 1977 when hundreds of lubrication points were automated on centrifugal pumps and electric motors. At the Kaukopaa location there have been no lubrication-related bearing failures for several years and management has gone on record attributing the bulk of their improvements in maintenance and machine up-time performance to the implementation of trouble-free

(source: Safematic Company)

t is hardly necessary to explain that industry is forever searching for ways to reduce downtime and maintenance expenses. However, different industries, and often different plants, are prone to attempt to increase profitability in vastly different ways. There are plants that will blindly “restructure,” or “re-engineer” by reducing manpower or hiring contractors. Many will do so without having a clear idea as to the ultimate results of, say, switching to a less well-trained workforce, or subscribing to the “flavor-of-the-month” maintenance method. We have seen the outcome – more restructuring, more experimentation. “Best-Of-Class” facilities, on the other hand, are thoughtful and consistent in their pursuit of wellresearched opportunities to improve equipment reliability, and thus, bottom-line profits. Among these latter companies we find plants which have clearly defined the job functions and responsibilities of their reliability professionals. These engineers and technicians are engaged in life-cycle cost and profitability studies encompassing many equipment specification, procurement, inspection, installation, protection, operating, surveillance and maintenance issues. They are familiar with stateof-art components and auxiliaries,

Figure 1. Automatic grease lubrication system The Pump Handbook Series

227

(5) and lubrication dosing modules (6), and also interconnects dosing modules and points to be lubricated. Different size dosing modules are used to optimally serve bearings of varying configurations and dimensions. The dosing modules, themselves, are individually adjustable to provide an exact amount of lubricant and to, thus, avoid overlubrication. A pressure sensFigure 2. Production increases, labor savings ing switch (Item 7) completes the installation. Enso Oy has docuEnso Oy - Kaukopaa Mill mented the production Production, Lubrication Labor, Maintenance Down-Time, Automatic Lubrication Systems increases, labor savings and downtime reductions shown in Figure. 2. Downtime hours for a total of 31 process units encompassing over 7,500 lubrication points are illustrated in Figure 3. Here, the Kaukopaa mill documented the 11-year trend from 9,700 hours of downtime in 1985, to approximately 280 hours in 1995. In the same Figure 3. Downtime hours for 31 process units time period, production at Kaukopaa Mill encompassing more than went from 620,000 tons 7,500 lubrication points (1985) to 950,000 tons (1995). From 1990 until Enso Oy - Kaukopaa Mill Maintenance Costs per Production 1995, total maintenance expenditures decreased 26%, and maintenance costs per unit of production were reduced by 46%. These stunning achievements are documented in Figure 4. Needless to say, Enso Oy has realized millions of dollars in extra profits from the timely introduction of engineered automatic lube systems. Figure 4. Maintenance costs per unit of They have included these production systems in the mandatory scope of every new project automatic grease lubrication methand Enso Oy’s mill standard (“EGO”) ods. defining automatic lubrication sysThe newer systems are configtems has been adopted as a National ured as shown in Figure 1. Modular Industrial Standard in Finland, a in design and easily expandable, they “high-tech” country in every sense of consist of a single or multi-channel the word. control center (Item 1), one or more Enso Oy - Kaukopaa Mill

Maintenance Down Time vs. Production

1000

9000

900

8000

800

7000

700

6000

600

5000

500

4000

400

3000

300

Shut-Down Hours

2000

200

Paper & Board Production

1000

100

0

0

1985

1986

1987

1988

1989

1990

1991

1992

1993

1994

1995

6000

20000

5000

15000

4000

3000

10000

2000

5000

Production Lubrication Labor Maintenance Down Time Automatic Lubrication Systems

Maintenance Down-Time (hours), Automatic Lubrication System

7000

25000

Production ('oo tns), Lubrication Labor (hours)

Production '00 tns

Shut-Down Hours

10000

1000

0

0

1991

1992

1993

1994

1995

100

90

Manintenance Costs, 1990 = 100

80

70

60

50

40

30

20

Total Maintenance Cost

Maintenance Cost / Production

10

0

1990

1991

1992

1993

1994

pumping stations (Item 2), appropriate supply lines (Item 3), tubing (4), which links a remote shut-off valve

228

1995

And Losers On the opposite side of the profThe Pump Handbook Series

Figure 5. Magnetic bearing housing seal

itability spectrum we found a refinery 6,000 miles from Finland. When we visited this facility in December of 1996, we realized that 700 of their 1,000 installed pumps required repairs in any given year. Storehouse records showed this refinery to have replaced 1,420 bearings in the prior 12 months; $9,100,000 went into pump repairs each year. Their meantime-between repairs (MTBR) of 1.4 years compares to average MTBR’s of 3 to 4 years in refineries elsewhere. There is obviously considerable room for improvement at this plant site. So why wasn’t an experienced group of reliability engineers able to identify the root causes of their problem and why did they find it so difficult to narrow the MTBR gap? There are several reasons. Their geographic location is in a hot and humid environment. Energy balance considerations mandate that roughly one-half of their pumps be steam turbine-driven, and most of the motordriven standby pumps at this refinery operate infrequently. Without state-of-art bearing housing seals (Figure 5), moisture contamination of the lubricant is inevitable on operating pumps. Unfortunately, the same is true for the standby pumps. Here, the vapor in the space on the top of the oil level in the bearing housing is expanding when ambient temperatures increase, and contracting as the temperature drops. Using cooling water on the bearing housings further promotes moisture condensation and the bearings experience rapid degradation.

Solution Options Three different and equally acceptable solutions would prevent these moisture-related failure events. The first option was spelled out above. It requires the use of automatic grease lubrication provisions and I hope that the effectiveness of these systems has been demonstrated to the reader’s satisfaction. Option two, using face-type bearing housing seals, has been highly effective as long as the equipment owner proceeded to hermetically seal the entire bearing housing. This would include replacing the housing vent with an inexpensive expansion chamber, and replacing the typical oiler bottle with a bulls-eye sightglass. The third option involves the use of dry-sump oil mist lubrication. Here, the oil mist would exist in a bearing housing at pressures slightly above atmospheric. The oil mist would, of course, provide proven bearing lubrication. However, an additional, and, in this particular case, even more important feature is the prevention of ingress of atmospheric moisture into the bearing space of standby equipment. The merits of removing cooling water from rolling element bearings include reducing the risk of promot-

ing moisture condensation and causing a closing-in of bearing-internal clearances. These facts should be well known – they have been documented in books and articles dating back to 1977.

– that’s terrific!” If, finally, you pay attention to this last paragraph and accept it to be the truth, you’re well on your way to becoming a “Best-ofClass” company.■

Conclusion

Bibliography

That removal of cooling water is still news to many reliability professionals is part of our final message. We must clarify and strengthen the role of the reliability professional in industry. Colleges and institutions of higher learning should take the lead in accomplishing this clarification and training task. A reliability professional must thoroughly understand and analyze “Best-Of-Class” practices. He or she must engage in life cycle cost analyses every step of the way, and display both the skill and motivation to lobby aggressively for improvement opportunities with rapid and potentially high return on investment. Assisted by the insight and analytical skills of informed reliability professionals, we must abandon such shortsighted and increasingly inappropriate views as “we can always upgrade,” lubrication is old technology “if bearings fail, it must be the bearing’s fault,” and “we used to have 1,000 bearing failures per month and now we only have 70

1. Bloch, H.P. Improving Machinery Reliability, Gulf Publishing Company, Second Edition (1986) 2. Bloch, H.P. Oil Mist Lubrication Handbook, Gulf Publishing Company,(1987) 3. Bloch, H.P. How And Why Centrifugal Pumps Continue To Fail, P/PM Technology, June, 1994 4. Bloch, H.P. Bearing Housing Seals Ward Off Failures,” Chemical Engineering, March, 1993 5. Bloch, H.P. Eliminating Cooling Water From General Purpose Pumps and Drives, Hydrocarbon Processing, January 1977

The Pump Handbook Series

Heinz Bloch is a consulting engineer in Montgomery, Texas who has more than 34 years of experience with process machinery. He retired from Exxon in 1986 and, since then, has specialized in machinery reliability improvement and maintenance cost reduction consulting on all six continents.

229

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R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

Universal Drive Shaft Maintenance By Will E. Johns III and David M. Cline Chemical an polymer gear uneven wear and can lead to premapumps can operate for several years ture failure, especially in needle without downtime. To keep the sysbearing designs. tem running for this period, it is Universal shafts should be sized imperative that the bearings be lubriwith a 2.8 SF for maximum momencated on a regular schedule. Until tary loads for start-up and solidificarecently, however, lubrication was tion of the plastic material. Roller not possible while systems were in bearings are the preferred bearing operation because the bearings are design because they perform better revolving. Premature failure of the in carrying high loads. A rating of bearings or coupling would often 50,000 B-10 life hours or 5.7 years is result. typical in polymer gear pump serFour point lubrication is recomvice. mended to insure that grease reaches In typical chemical/polymer all four bearing caps on the joints, pump applications the Side View of Pump System even if there may be some solidified pump can reach temMotor Gear Reducer Coupling Chemical Pump Universal Shaft grease in the lube channels. And peratures exceeding (guard removed for clarity) now a patented automatic lubrica600ºC. This heat is tion device (manufactured by the often transmitted to Cline Company) allows the operator the coupling driving to lubricate the universal shaft or a the pump and then to gear coupling while it is rotating (Figthe gear reducer. The ure 2). This innovative device allows heat transfer from the Figure 1.Side view of pump system the operator to inject grease into all pump shaft or flange to eight bearings. It can be used on the mating coupling right angle and parallel shaft type can actually liquefy lubricants (norTo keep the universal shaft rungear reducers and on any size univermally grease) and reduce effectivening optimally: sal shaft and nearly any size pump. It ness, especially in bearings or is especially advantageous for use on couplings. Oil can separate from the • be sure the shaft is offset proplarger, expensive pumps and shaft carrier so the carrier, which has virerly assemblies. Using a heat disk and an tually zero lubricity, does not proper• check the shaft alignment periautomatic lubrication device can ly lubricate the roller bearings or the odically increase longevity of the shaft coupling. • lubricate the shaft regularly assembly and bearings to five or With a universal shaft assembly, after careful selection of the more years without downtime. a heat disk can be installed to isolate lubricant the pump from the Top View of Pump System driven universal joint The universal shaft compensates Reducer Coupling Motor assembly. The heat for thermal growth in the high temdisk absorbs heat perature environment of the polyPolymer Pump Four Point Automatic Lubrication Device and releases it into mer system. Optimal offset for the atmosphere. It proper operation of the shaft is 5% to also reduces heat insure that universal joints on the transfer to the matshaft get adequate lubrication. Offing flange. This prosets of less than three degrees can longs the life of the cause the bearings in the joints to bearing assembly. rotate only partially. This causes Figure 2. Top view ependability is a high priority for chemical and polymer manufacturers considering pump systems. Keeping the pump (the heart of the system) running smoothly and continuously requires a drive that is properly designed and maintained. A vital link between the pump and drive system is the universal shaft (Figure 1). The shaft allows for thermal growth of vessels and piping, and it isolates mechanical devices from high temperatures.

D

230

The Pump Handbook Series

The patented lubrication system can be used to lubricate automatically the drive shaft assembly during operation on a timed frequency – hourly, daily, or at longer intervals if required. The system can also be operated manually by depressing provided pump handles several times to lubricate the bearings. The grease enters all eight bearings at the same time. Each injection forces a small, me-tered amount of grease into each bearing per cycle to insure that new lubricant enters the bearings evenly at all times. This system assures bearing longevity through the entire life cycle of the pump. Lubricant selection is crucial. The grease needs to withstand high temperatures to be effective. It also should be able to handle extreme pressure and continuous wearing conditions. A lubricant that will be shear stable for long periods of time, such as one with a lithium base, should be selected. An example would be Lubriplate 1242. Proper torquing of the flange bolts is also imperative to prevent flange face slipping because of the high heat transfers. The torque values of Grade 10.9 bolts will be recommended by the manufacturer of the shaft assembly. Assembling the proper components and committing to correct and regular maintenance will greatly decrease downtime and prolong the life of a chemical or polymer gear pump system.■

FLANGE YOKE BEARING CAP BEARING CAP SCREW BEARING BRUSH COMPL. GREASE NIPPLE JOURNAL CROSS GREASE NIPPLE END CAP HEXAGON NUT WASHER PACKING SET DUST CAP SLIP STUB SHAFT

SLEEVE YOKE BEARING CAP BEARING CAP SCREW

BALANCE WEIGHT END CAP JOURNAL CROSS GREASE NIPPLE BEARING BUSH THRUST WASHER ROLLING ELEMENT SEALING RING RETAINING RING

BEARING CAP SCREW BEARING CAP FLANGE YOKE

BEARING CAP SCREW BEARING CAP TUBE YOKE TUBE CENTERING RING

Figure 3. Exploded view of typical universal shaft – design may vary with different manufacturers

Will Johns III has a B.S. in Mechanical Engineering from Clemson University with twelve years experience in rotating equipment with LCI Corporation and Ingersoll-Rand. He is currently Regional Sales Manager for LCI in the Southeast USA and is involved with polymer/chemical projects throughout the world. David Cline has been involved for 30 years in the manufacturing, design and fabrication of universal drive shaft assemblies, gear couplings, clutches and brakes for the pulp and paper, steel and chemical industries.

The Pump Handbook Series

231

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R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

Optimizing High Energy Pump Operation Reduce energy costs and increase capacity without adding more pumps. By Derek Bailey

T

In many cases pump modifications to reduce energy costs will automatically result in reduced maintenance costs. Whenever possible in the article, examples have been used from specific installations. This is intended to be a practical approach, assuming that in most situations pumps will not be taken out of service other than at planned maintenance intervals or for mechanical failure. Most of the examples used are relatively quick fixes that can be combined with scheduled shutdowns. The first tip is almost certainly the most important. Without acting on this, the effectiveness of all measures to optimize pump operation will be significantly reduced.

6000 ACTUAL OPERATING POINT (NOTE EXCESS HEAD)

232

SPECIFIED DESIGN POINT

11 INCHES

5000

60

65

70

73 76 78 OPTIMUM FLOW 79

10 INCHES

78

4000

76

9 INCHES

3000

2000

1000

0

Tip #1 Keep good records of operating conditions and all pump problems. Some of the most important points to record are:

• product temperature • power absorbed • dates, duration and reasons for pump outages

• pump flow rates • pump suction pressure • pump discharge pressure • product density

HEAD (FEET)

his article is targeted at medium and high energy pump operations and maintenance engineers. The objective is to explain how costs of operating pumps can be significantly reduced by maintaining good records and working closely with manufacturers and good quality repair shops. The article covers three main areas: • reducing energy costs • reducing downtime and maintenance costs • increasing capacity without adding more pumps

0

200

400

600

800 1000 CAPACITY (US gpm)

DAVID BROWN PUMPS, INC.

4X6X10.5C

Shreveport, LA. Project:

Minimum Diameter

Figure 1. The Pump Handbook Series

9 inches

1200

DB34 10STG

1400

1600 Speed (R.P.M.)

Pump

Maximum Diameter

11 inches

3560 Curve No.

346M911

20 0 NPSHR - FT

6 0 M

81

79

77

68

58

48

NPSHR WATER

800

73

240

FEET

METERS

CHARACTERISTIC CURVE SHEET EFFICIENCIES BASED ON API CLEARANCES (BELOW 500°F) 27" DIA

220

82

700 24.5" DIA

ACTUAL HEAD

HEAD

79

REVISED HEAD

600

77

180

81

200

22" DIA

160 500

00 17

140

BH P

120

400

13

90

0

00

BH

BH

P

P

0

1000

0 0

2000

3000

500 1000

2000

4000

5000

1000 3000

4000

6000

7000

1500 5000

CAPACITY

8000

9000

2000

7000

8000

9000

10000

IGPM

2500

M3/HR

11000

USGPM

6000

DAVID BROWN

Max Dia Min Dia

Impeller

27" 22"

Eye area sq ins

12 x 16 x 27A Impeller

Supersedes Reference

PUMP

1760 RPM

H5260 Volute or Bowl

166"

DB 21 Curve No.

H5251

6M622

Figure 2.

It is important that values for flow, pressure, density and power be recorded at the same time and that recordings be made often enough to identify the full operating range of the pump. Most importantly, the minimum and maximum capacity and the minimum and maximum differential head relative to flows (calculated from suction pressure, discharge pressure and density) must be identified. Having collected this data and determined that the range of duties meets all of the anticipated pump set requirements, the next step is to obtain a performance curve. If the pump is relatively new and you can guarantee that it is operating as it was supplied, then the original test curve will suffice. Experience indicates, however, that in many cases pumps have been modified in some way (often for the worse) and the original performance curve is no longer valid.

Tip #2 Have your pump supplier or repair shop carry out a performance and NPSH test to API 610 standards to determine the duty of the pump accurately. At a minimum, this test should be made after your next maintenance repair. If possible, the test

should be made before the repair if you suspect the pump is running well away from its optimum operating conditions. Optimum operating conditions means the flow is close to peak efficiency and producing the required head without throttling on the discharge valve.

Reducing Energy Cost To appreciate fully the potential for wasted energy and corresponding pump problems (especially in older pumps), it is necessary to look at how many pumps were initially selected by the engineering companies.

Example 1: Refinery Pump Process flow requirement is 800 gpm. The process engineer adds 10% to its calculated flow as the design of the pump is not finalized (880 gpm). The equipment engineer sends the pump data sheet to the supplier with the duty condition of 880 gpm and the design duty 10% above the duty capacity. The flow requirement is now up to 968 gpm. API 610 6th edition and prior editions required the design duty to be to the left of peak efficiency (this is no longer the case). A selection 10% back from peak would have been seen as a good one by the pump supplier and contractor, and in many cases this selecThe Pump Handbook Series

tion could be 20% to 30% back from peak efficiency. For illustration purposes let’s pick a pump that is 10% back from peak efficiency. As such, the flow becomes 1075 gpm. Looking back at the actual requirement of 800 gpm, we have a pump with a duty requirement of 800 gpm with its optimum running capacity at 1075 gpm – i.e., 30% above the duty required. This is typical of many pumps purchased in the mid-1980s and before, and it still happens today although not usually to the same extent. Figure 1 shows a typical pump selection for a true required capacity of 800 gpm for a unit supplied in the mid-1980s. Note the efficiency of 72% compared with the peak efficiency of 78%. If we assume the pump is still running at the original required head and a capacity of 800 gpm (many will be running below this value), by making some general assumptions we can identify the value of wasted energy. Assumption: Product S.G. 0.8 Power Cost: 4 cents per kW-h Operating: 24 hours per day for 330 days per year BHP= Quantity (gpm) x Head (ft) x S.G. 3960 x Efficiency Pump as purchased BHP = 800 x 3820 x 0.8 = 857 BHP 3960 x .72 Pump modified with duty on peak BHP = 800 x 3609 x 0.8 = 752 BHP 3960 x .775 Savings = 105 hp (78 kW) = a 12% power reduction For a pump running 24 hours per day for 330 days per year, this gives the following savings based on 4 cents per kW-hr. Savings = 78 kW-h x .04 dollar x 24 hrs x 330 days= $24,710 per year The above shows only the energy savings; however, if a pump is running at or near its optimum duty, it will also be running at its optimum mechanical design conditions, with a resulting improvement in mechanical reliability in most cases. This example shows the amount of power

233

wasted by a pump running away from optimum design.

Tip #3 Provide your repair shop with your newly identified operating range and ask for proposals to modify the pump to run at, or near, peak efficiency. Insist on a performance test to demonstrate the success of the modification. A typical solution to the example in Figure 1 would be to install lower capacity impellers in the pump, pulling the peak efficiency to between 800-900 gpm and trimming to the exact duty head. This will produce the type of saving shown in Example 1. If we assume the only items required in addition to a standard overhaul were 10 low-capacity impellers, the total additional cost including testing would be roughly $35,000. This provides a 1 1/2 year pay-back.

Example 2: Pump Generating Excess Head Another common waste of energy (with a simple fix) is a pump set generating more head than that required by the current operation. There can be several reasons for this. The original head specifications might have been too high. The pump could have been accepted when tested with excess head. And the duty requirement may have changed since the pump was specified/installed. Again, it is important to identify the exact duty requirements for the pump and to have an accurate pump curve with which to compare the results, then and now. Figure 2 shows a pump operating with 10% excess head at the duty points. In this instance, a simple lowcost solution is available. All that is typically required is an adjustment to speed on a variable speed set, or more likely, an impeller trim (for constant speed). Do not forgo the test to verify the results even though the solution is simple. The example in Figure 2 shows a large two-stage radially split refinery pump with the following required duty:

234

Revised Duty Capacity Head SP GR Power cost Operating

8000 gpm 550 feet 0.8 4 cents per 12 kW-h 24 hours per day, 330 days per year

Actual Pump Duty Capacity 8000 gpm Head 605 feet Actual BHP=18000 x 605 x 0.8=1207 BHP 3960 x .805 Revised Duty BHP=8000 x 550 x 0.8=1104 BHP 3960 x .805 Savings = 1207 - 1104 = 103 BHP = 77 kW For a pump running 24 hours per day, 330 days per year, the saving is: 77 x .04 x 330 = $24,300 In Example 2, assuming the solution to be an impeller trim and test, the cost, in addition to a normal overhaul, is likely to be no more than $8,000-$10,000. This corresponds to a six month pay-back.

Reducing Downtime and Maintenance Costs Other pump problems that are not so easy to quantify in terms of savings – but which should be readily identified if good records of pump outages are kept – are those associated with premature pump failure. Listed below are target periods of operation to aim for between major overhauls on high and medium energy pumps. Recognize that it is impossible to estimate the standard life expectancy of a pump because service conditions, product quality, installation and maintenance can have significant impacts. However, in my experience working with end users over many years identifying and solving problems, properly specified and designed equipment that is correctly installed and maintained should be capable of the following: Water Injection Pumps (axially split barrel 5000 hp, 6000 rpm) In the early 1970s, the expected The Pump Handbook Series

life of 1.5 years was normal. This has been increased to 8 years by working closely with end users over the span of many years. Crude Oil and Product Pipeline Pumps (axially split single and multistage, powers from 500 to 5,000 hp) On relatively clean product, an 8 year-plus life is regularly achieved in this service. Vertical Bowl Pumps on Product Temperature Down to –296ºF In general, products such as propane, butane, ethylene, ammonia, and liquid natural gas are relatively clean, and even on large 24-stage vertical pumps up to 1000 hp at 2 pole speeds, a life expectancy between major overhauls of 8 years is being achieved. Large Axial Flow Pumps on Polypropylene and Polyethylene Life expectancy of the internal bearings was historically bad on these pumps. The average life between failures was 6 months. By working closely with end users, this has been increased to more than 3 years. Thus, unplanned shut-downs can be avoided. As advised earlier, it is impossible to say that the life of every pump can be extended to 8 years or more between major overhauls; however, it is a good achievable target in many installations. What is certain is that a high percentage of pumps in many cases can have their life between shut-downs extended significantly, just by recording and sharing information about the history, duty and past problems with the repair shop or pump supplier. As a starting point, focus on any pump that fails regularly within a 12 month period (most people have them). Expand the search outwardly as pump life is improved.

Increasing Capacity Without Adding More Pump Sets In many cases, it is possible to increase the pipeline capacity significantly without major new equipment expenditures. Compare Figures 3 and 4. Figure 3 shows the curve for a high speed axially split multistage volute type water injection pump as originally

TOTAL HEAD (ft) 10000

TOTAL HEAD (ft) 10000

9000

9000

NEW DUTY

ORIGINAL DUTY 8000

8000

7000

7000

6000

6000

5000

5000

4000

4000

3000

3000

2000

2000

1000

1000

0 100

0 100

90

90

80

80

70

70

60

60

EFFICIENCY 50 (%)

Water POWER (HP)

EFFICIENCY (%)

50

40

40

30

30

20

20

10

10

0 4000 3500 3000 2500 2000 1500 1000 500 0

r at SG=1.027

Powe

0

300

600

900

1200

1500

1800

Water POWER (HP)

2100

2400

0 6000 5000 4000 3000 2000 1000 0

=1.027

Power at SG

0

300

600

CAPACITY (US gpm)

DAVID BROWN Pumps, Inc. Shreveport, LA. Project

4x6x10.5B Speed (R.P.M.)

5684

Figure 3.

supplied in the mid-1970s. Figure 4 shows the same pump unit operationally today. The main points to note are the following increases in pump performance: Capacity increase 30% Differential head increase 8% Efficiency improvement 8% The increase in capacity was carried out in phases to match the water injection requirement for the field. This avoided the need and cost for an additional pump set on an offshore platform. It also saved space, which is at a premium in such operations. This particular application required some concessions on the motor and gearbox sizing. However, by working with the end user, it was possible, with improved efficiencies, to meet all the new design parameters.

Pumps With Varying Duty Requirements

900

1200

1500

1800

2100

2400

CAPACITY (US gpm)

DB34

Pump

Max/Trim Dia.(in)

11.1/10.25

7

Curve No.

DAVID BROWN Pumps, Inc. Shreveport, LA.

8340514C

Project

Stages

4x6x10.5B Speed (R.P.M.)

5663

DB34

Pump

Max/Trim Dia.(in)

11.1/10.25

7

Stages

Curve No.

B772846

Figure 4.

This subject is far too complex to cover in this article. Requirements vary widely and solutions vary with them. Broadly speaking, varying duty situations tend to fall into two main groups. Pumps with seasonal demands – e.g., high demand in the winter and low demand in the spring and autumn – or pipelines with an increasing demand over the years. The solution to this problem can often be to have both low and high capacity elements for the same pump. Pumps with a constantly varying demand – in these instances, the solution is more complex and expensive, and it is likely to require variable speed drivers or variable speed couplings Again, if a requirement exists for vastly varying duties, talk to the repair shop or pump supplier specifying exact requirements. In the hope of picking up modifiThe Pump Handbook Series

cation work, a good supplier will be happy to make a number of proposals without charge for improving your pump operation. Again, a test of the final solution is crucial.

Conclusion This article was not intended to solve your problems, but to make you aware of the importance of keeping good records of pump operating conditions and problems, and of sharing such information with your supplier or repair shop. A good shop will welcome this and work with you to reduce your total pump operating cost.■ Derek Bailey has nearly 40 years of experience in the pump industry and is currently President of David Brown Pumps, USA. Prior to this he worked for the U.K.-based company in various engineering and marketing positions.

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R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

Managing Vertical Pumps in a Changing World Prepare your pump for anticipated service changes before they cause problems. By Terry Wold

hen preparing the material for this article, I decided to review current problems on my desk, so if some of this information seems familiar, you know why. Most of today’s problems stem from the use of the same pump for several services – with different pressures, speeds, viscosities and temperatures. Or a change in plant personnel occurs, and a pump is operated in a different manner than for what it was designed. This is good in the long term because it means industry is trying to utilize existing resources rather than expend capital money to make

(Photo courtesy Afton Pumps)

W

Photo 1. Vertical turbine pumps in pipeline service

236

changes to their plants – a practice that ultimately increases the cost of products you and I use. Such changes, however, can cause several problems F F for pumping operations that must be dealt with, preferably in the initial evaluation, or B B C C before a pump is changed over to a new service. Once this service is decided upon, the pump must be installed in a well thought out manner if you’re going to have a reliable system. This includes foundation, venting, alignment, environmental concerns and, most A A D important, operation and maintenance. I will discuss axial thrust, T = [Pdx(A–B–C)] + (PsxD) + (PsealxF) minimum flow, variable speed Figure 1. Unbalanced impeller drivers, installation, alignment imperative that we know the magniand pump operation. I will also comtude of these factors. ment on the role mechanical seals play in the operation of a vertical Thrust – Single Stage pump. We’ll explain what the typical Overhung Pumps vertical pump problems are and Downthrust is generated by what can be done to help solve them. pressure acting on the top of the Axial Thrust – What are the impeller between the outside diameVariables? ter and the shaft diameter (Figure 1). Upthrust is generated by pressure Thrust is the sum of the areas acting on the bottom side of the multiplied by the pressures acting on impeller multiplied by the area the plane perpendicular to the shaft. between the outside diameter of the Axial thrust, speed, temperature and impeller and the bottom wear ring lubrication are what ultimately diameter, added to the area of the determine bearing load and life of a inside diameter of the impeller sucpump and driver. Therefore, it is The Pump Handbook Series

tion, and multiplied by the suction pressure. Since this configuration has no balance holes or top wear ring, stuffing box pressure is almost the same as pressure at the top side of the impeller; therefore, an additional upthrust generated by the seal diameter (seal area minus shaft area multiplied by the differential pressure) must be added. This is considered an unbalanced impeller design.

ORIFICE

ORIFICE

ORIFICE

FROM PUMP DISCHARGE

TO PUMP SUCTION

API PLAN 11

API PLAN 14

RECIRCULATION FROM PUMP DISCHARGE THROUGH ORIFICE & SEAL CHAMBER (NOT RECOMMENDED)

RECIRCULATION THROUGH THE SEAL CHAMBER BACK TO THE PUMP SUCTION WITH PRESSURE BLEED-OFF BACK TO SUCTION

Balanced Impeller

TO SUCTION

CLEAN FLUID

ORIFICE As shown in Figure 2, a balanced design in-corORIFICE porates two impeller wear rings (top and bottom) FROM DISCHARGE and balance holes. If the TO wear rings have the same PUMP SOLIDS TO diameter, then the up and SUCTION TO SUCTION SUCTION down thrusts are equal, or API PLAN 31 balanced, from the RECIRCULATION FROM API PLAN 13 DISCHARGE THROUGH impeller OD to the nomiCYCLONE SEPARATOR - SOLIDS RECIRCULATION nal wear ring diameter. BACK TO SUCTION AND CLEAN THROUGH THE SEAL FLUID TO THE SEAL CHAMBER CHAMBER BACK TO THE Also, the balance holes WITH PRESSURE BLEED-OFF PUMP SUCTION BACK TO SUCTION are sized such that the pressures on either side of Figure 3. Typical seal flush plans for vertical pumps the impeller between the unbalanced, you would have marginacting on the end of the shaft times wear ring and the shaft diameter are al pressure differential in the stuffing the shaft area, and the stuffing box essentially equal. Only two areas are box and virtually no flow across the pressure acting on the seal area not balanced – the suction pressure seal faces. Plan 11 is used almost minus the shaft diameter. This exclusively with balance holes and is a thrust balanced impeller wear rings so that a pressure differdesign. ential exists. If significant suction Thrust balance can be pressure is present, the thrust will altered by increasing or likely be upward. decreasing the wear ring F F API Plan 13 is used when there diameter, thus modifying the are no wear rings on the back of the thrust balance. The balance impeller (unbalanced design). The hole diameter can also be piping runs from the stuffing box (or changed. This will increase or B B C C gland) to the suction of the pump, decrease the stuffing box presresulting in a pressure differential. sure and alter the thrust. An orifice in the piping, or a drilled Another benefit of decreasing hole in the gland, is used to control the stuffing box pressure is the flow across the seal faces. This is longer seal life. the most popular approach for a verStuffing box pressure can tical pump because it has the lowest also be modified by the type of stuffing box pressure and best presseal flush plan used (Figure 3). E A D D A sure differential. The pump will usuWith API Plan 11, the seal ally be operating with a downthrust flush is piped directly from that creates a more stable system. the pump discharge to the seal Another option used primarily faces. An orifice is usually in T = [Pdx(A–B)] + [Psx(D+E)] – (dPxC) + (PsealxF) with vertical turbines is to incorpothe piping to regulate the flow Figure 2. Balanced impeller rate a bleed-off bushing in the stuffto the seal. If the impeller is

The Pump Handbook Series

237

ing box bearing. This pressure bleedoff is piped back to the suction of the pump (Figure 3: API Plans 14 and 31). It enables you to take the pressure off the discharge, and this generates a pressure differential across the seal face. Thrust in vertical turbine designs is basically the same as in other vertical pumps except that there are more stages, the shaft upthrust acts only on the first stage, and the stuffing box is exposed to discharge pressure. Instead of calculating all the areas and pressures in vertical pump stages, a “K” factor is obtained from a test conducted by the pump manufacturer. It is used to calculate the hydraulic thrust per stage. There will be different factors for balanced, unbalanced and special first stages. Thrust must be evaluated from no flow to open valve flow, and you must also consider maximum and minimum suction pressures, specific gravity, viscosities, mechanical seal size and seal flush plan. If a pump with high suction pressure must be idled, this factor must be considered when sizing the driver’s motor bearings. (Both up and down thrust must be taken into account.) The motor manufacturer needs to know the maximum and minimum thrust capacities (can be upthrust and/or downthrust), along with the normal operating thrust conditions. The motor manufacturer may state a minimum thrust condition, and the pump manufacturer will produce a thrust verses flow curve. The customer will be notified that he or she must control the operation of the pump between certain flows so as not to go below minimum thrust or above maximum thrust. These limitations are based on bearing life calculations as well as other conditions such as loading and skidding. A number of bearing designs for high thrust conditions are utilized in the motor industry. Some can take equal thrust in either direction, while others take only momentary (startup only) upthrust. If you have the latter, a problem will likely develop if the pump is exposed to high suction pressure while idle. Motor bearings must be able to handle both static and dynamic thrust conditions.

238

Vertical turbines can have very high thrust requirements resulting in large bearings limited by speed. So it may become necessary to balance stages. Impeller balancing in this context is similar to the way it is done for inline pumps except there are more stages. The “K” factor is different from that of unbalanced impellers. If required to meet thrust conditions, you can use balanced and unbalanced impellers in the same pump. However, it is recommended that the unbalanced stages be at the bottom to help stabilize the shaft. If a continuous upthrust condition exists, it is advisable to install a shaft locking device in the pump half coupling and shrink fit the motor half coupling to help stabilize the pump shaft. If the pump has a column and is operated in continuous upthrust, an additional sleeve bearing can be installed to help stabilize the shaft and avoid high shaft deflections or vibration. Another vertical pump challenge is finding a suitable driver. Only a few major motor manufacturers in the United States supply a high thrust vertical motor. Because of this, pump designs are available that incorporate thrust bearings. The thrust bearing option is available for all vertical pumps – both inline and vertical turbines. Bearing sizes are usually based on a 25,000 or 45,000 hour L10 bearing life. Caution! Motor manufacturers are not required to meet the same standards as the pump industry. They will comply if asked in the quotation stage, but be prepared to pay extra for special tolerances and/or bearings with a longer bearing life. Thrust bearings should be two 40º angular contact bearings mounted in a back-to-back configuration. Sometimes three bearings will be used – two in the direction of thrust, and one in the opposite direction for momentary or static thrust conditions. You should always determine the capacity of the bearings in your pump or motor. You should know their limitations, as well. Do this whether you are buying new equipment or repairing existing equipment that is going into a new service. You may opt to balance or unbalance more stages if the driver cannot be The Pump Handbook Series

changed.

Minimum Flow There are several things to examine when determining the continuous minimum flow of vertical pumps. Minimum flow should be evaluated on an individual installation basis. 1. Thrust. This can be the limiting factor when the maximum thrust rating of the motor is exceeded at closed valve, or continuous upthrust cannot be tolerated by the driver bearings. 2. Temperature rise. To avoid vaporization, you must know the specific heat of the fluid pumped and the maximum allowable temperature rise. Typically, a 15º temperature rise is the maximum allowed. A rule of thumb is that 35 gpm/100 bhp = 15º temperature rise. 3. Recirculation. Much has been written on this topic, and to date no one has really come up with an easy way to determine the onset of recirculation. F.H. Fraser has written a couple of papers that present a way to determine suction and discharge recirculation. The best thing to do is to discuss recirculation with your pump manufacturer and follow his suggestions for your application. 4. NPSH. Net positive suction head is measured in feet absolute. You must have adequate NPSH for your pump to run without vaporizing the fluid. The requirements are printed on your pump performance curves. Don’t forget to check NPSHA vs. NPSHR for all operating conditions and fluids pumped. 5. Minimum stable flow. Some pumps, especially those with a large eye diameter or high suction specific speed, tend to have a “hump” in the performance curve near shutoff. This represents an area where the pump may try to shuttle between two flows with the same differential head. This is a problem only when two pumps are run in parallel, or when a continuously rising curve is required for system control purposes. A hump is almost always present when a pump has a specific speed greater than 1000 because pumps designed in this range (for optimum efficiency) will have a large eye area,

creating recirculation at low flows. Hence, a hump in the curve at shutoff occurs. The hump can be designed out of the curve, but there will be a sacrifice in pump efficiency, and the NPSHR will increase.

Variable Speed Drivers There are several good reasons to use variable speed controls, most of which have to do with energy conservation and reduced operating costs. However, some serious drawbacks exist when they’re used with vertical turbine pumps, gears or engines. Vertical pumps have flexible shafts, and therefore, at some point during startup, the pump will pass through a critical frequency. Continuous operation at this speed would be catastrophic. The first thing to do is determine the economic benefits associated with using a variable speed drive in your application. If the economic benefits are significant, the following must be considered: 1) Critical speeds based on a damped lateral analysis. To perform this analysis, you must know the WR2 of all rotating components (motor rotor, coupling, impellers and wear rings); center of gravity distances between the bearings, impellers, wear rings, bushings, couplings and motor); bearing stiffness characteristics and mass elastic data of the motor, gear, and/or engine. In addition, if a gear is used, the gear ratio and torsional characteristics of the shafts need to be known. If you’re working with a used engine, you need to know the inertia of each cylinder, flywheel, clutch and crankshaft as well as the number of cylinders and firing sequence. Once the critical frequencies are known, operation cannot be within 25% of those speeds. Once the safe speed range is known, go back to the original design considerations to insure that the pump can be operated safely and efficiently. 2) If, after the above, you decide to continue with a variable speed drive, the next thing to consider is the structural analysis. The equipment manufacturers with whom you’re working must be made aware of the intended operational speeds, and a

resonance test should be specified. All resonant frequencies should be at least 10% outside of the operating speeds. 3) The last, and possibly the most important, consideration for a VSD is who will operate it and how it will be used in service. The variable speed controller should be locked so that the equipment will not be run at a critical speed. The manual override should be locked to all operating personnel at all times. This includes the “night shift.”

Installation

OUTER BRL FLG. (SOLE PLATE) INNER BRL. FLG.

PUMP C L

DISCH. HEAD FLG. TO INNER BRL STUDS INNER TO OUTER BRL (SOLE PLATE) STUDS ANCHOR BOLT OUTER BRL FLG. SOLE PLATE GROUT

4D 5D DISCHARGE HEAD FLG.

2.5D CONCRETE

4D

PLAN C L

OUTER BRL. INNER BRL SECTION A-A WITH FABRICATED STEEL OUTER BARREL SUCTION

DISCHARGE PUMP C L

VENT F/SPACE BETWEEN INNER BRL & OUTER BRL

DISCH. HEAD FLG. TO INNER BRL STUDS ANCHOR BOLT

4DGROUT 6D

2.5D 4D

CONCRETE Figure 4 depicts some typical installaCONCRETE PIPE OUTER BARREL tions. Note the relationINNER BARREL SECTION WITH CONCRETE OUTER BARREL ship of the foundation studs and anchor bolt Figure 4. Typical installation of vertical turbine pump size. The top surface of changes in the field after the equipthe grouted soleplate, barrel flange or ment is delivered. outer barrel must be leveled above If tandem or double mechanical grade. If this top surface is out of seals are used, there will likely be a square by as little as 1/2º, there plan 52 or plan 53 reservoir with varcould be mechanical interference ious instrumentation located on a when installing the pump and/or barstand and piped to the pump. Some rel, depending on the pump length. wiring of alarms, pressure and level The barrel flange or soleplate must switches will also be required. If you cover all open area for proper suphave a tandem seal, the reservoir port. If the area between the outer may be of substantial size because barrel and the inner barrel is to be the secondary seal is designed to carvented, specify a tap in the side of ry the pump safe working pressure if the soleplate or outer barrel flange the primary seal fails. In normal for that purpose. If a purge line for operation the seal only sees the presthe barrel is required, be sure to sure in the reservoir, but the seal specify the size and type of drain faces are loaded to carry the maxirequired. The drain may be internal mum pressure. This will cause the to the barrel or external. It can termisecondary seal to produce a lot of nate in a flange or a pipe tap and be heat that must be carried away and located on center or in the corner – dissipated to keep the seal cavity between the foundation bolt and from vaporizing (causing dry seal head flange. If it is to be placed in the faces and eminent failure). If you use corner, be sure to notch the foundaa double seal, the pressure in the tion or outer barrel for its entire reservoir must be maintained above length. If it is to be on center, specify the stuffing box pressure to keep that an oversized barrel flange be product from leaking into the reserused. voir. An outside source – such as a Close attention to the pump nitrogen purge – should maintain manufacturer’s outline drawing is this pressure. required. If something is not clear, it should be clarified as soon as possiAlignment ble to avoid embarrassing and costly After the pump has been

The Pump Handbook Series

239

those called for by NEMA. An inspection of all the motor dimensions and runouts should be done before attempting to align the motor, unless you know the pump manufacturer has already complied with special tolerances. If any face runout exists, reface the motor flange prior to mounting the motor. (This applies to both new and repaired equipment.) Allow .001”TIR for the motor shaft. The face and rabbet should be .000”FIR. Do not accept a motor with a rabbet fit exceeding .004”TIR or a face runout exceeding .003” runout prior to facing. Maximum endplay should not exceed .010”. (This insures proper bearing installation.) Every motor should get a complete inspection before being mounted on any vertical pump. To finish getting the pump ready to operate: 1) Tighten the drive collar setscrews and remove the spacer (which sets the mechanical seal) from underneath the drive collar. 2) Fill the pump with product and bleed off all vapors. 3) Don’t forget to bleed the mechanical seal and seal flush piping. 4) The vent connection in the head should be piped to flare or back to the suction vessel. The head DISCHARGE HEAD TO MOTOR FLANGE must be vented to prevent vapor locking.

installed and piped, and the motor wired, remove the coupling spacer and check the rotation of the motor. If the pump has a 2-piece shaft connected by a threaded coupling, make sure the motor has a non-reversing ratchet installed to prevent shaft uncoupling. The key to smooth running and long life for a vertical pump of any kind is quality pump to driver alignment. There are 7 steps required to align the pump to the driver successfully (Figure 5): 1) Indicate the motor shaft to the pump head to within .001”TIR. 2) Install motor coupling hub; indicate the hub face to .0005”TIR. 3) Install the pump hub and adjusting nut. 4) Install coupling spacer with bolts. Indicate the flange diameter to less than .002” - 003”TIR. 5) Turn the adjusting nut up to the pump setting. 6) Install the coupling bolts to lift the rotating assembly to the running position. 7) Shaft runout below the coupling and above the drive collar should be < .002”TIR. These steps may not be achievable unless the motor has been ordered with tolerances tighter than

“A”

MOTOR COUPLING HUB

“B” COUPLING ADJUSTING NUT

COUPLING SPACER “D”

“E” “C” IMPELLER SETTING PUMP COUPLING HUB

SEAL GLAND

DRIVE COLLAR

Figure 5. Alignment for vertical turbine pumps

240

Pump Operation Most of this discussion has been of a mechanical nature. Now we have the pump installed and ready to run. Let’s discuss operation. 1) Design capacity. Every impeller and case or bowl has a certain capacity for which it is designed. The pump will be most efficient at this capacity and head. If it is operated anywhere else on the pump curve, the efficiency will be less because the relative vane angles and hydraulic areas will be too large or too small. 2) Pump at open valve. A pump operating at open valve has the least system resistance and will operate as far out on the curve as the system will The Pump Handbook Series

allow. The limiting factors include NPSHA, case capacity, impeller capacity and downstream piping. If the pump exceeds design capacity, it is likely that there will be vibration, surging and problems with seals, bearings or couplings. Also, the shaft may break. 3) Pump at low flow. Efficiency is reduced for the same reasons that apply when flow is too high. If flow approaches the minimum flow condition, there will be an increase in vibration, recirculation may become damaging, and NPSHR will increase. 4) The Hydraulics Institute defines two areas of the pump curve for operation: POR - Preferred Operating Range, usually 70-120% of BEP AOR - Allowable Operating Range, based on: 1. life of components - bearings, shaft, driver, etc. 2. vibration and noise Each of these considerations has an influence on the AOR. AOR is determined by the pump manufacturer and should be evaluated on an application basis. According to Hydraulic Institute standards, pump life for the AOR must exceed two years. The bottom line for successful pump operation is to evaluate how the pump will be run, consult with the pump manufacturer concerning off duty conditions that may exist, and prepare for proper installation. If the pump does not operate as expected, have all pertinent information available. This includes suction and discharge pressure, flow rate, specific gravity, speed, temperature, vapor pressure, viscosity, vibration data and any other information that apply. Consult with the pump manufacturer as soon as possible. I hope this presentation has provided some ideas on how to address vertical pump concerns before they become problems.■ Terry Wold is Chief Engineer at Afton Pumps, Inc., and a frequent contributor to Pumps and Systems. He has also spoken at ASME meetings in the Gulf Coast region on numerous occasions, and was a discussion group leader at the 14th International Pump Users Symposium in Houston.

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R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

Improving Progressing Cavity Pump Reliability Common sense and a few rules of thumb can help keep your pumps running smoothly. By Larry Shanley, Jr., Shanley Pump & Equipment, Inc.

rogressing cavity pumps are a unique, single screw positive displacement design. The pumping elements are a helixshaped metallic rotor that turns eccentrically in an elastomeric stator (Figure 1). Progressing cavity, or “PC” pumps as they are often called, are problem solvers. They are sometimes called the “pump of last resort” because they can get the job done in applications where classic centrifugal designs and all metal positive displacement pumps would not provide satisfactory service. PC pumps are characterized by their accurate metered flow, high discharge pressure limits and good suction lift capabilities. They also provide gentle handling of shear sensitive fluids and are most widely used in services in which solids handling and resistance to wear are required when pumping abrasive fluids. A common PC pump application is pumping (liquid explosives) ammonium nitrate slurries. The slurry is an abrasive and viscous mixture pumped from trucks to mining sites through hoses at high discharge pres-

(photo courtesy of Shanley Pump & Equipment)

P

Photo 1. A PC pump in service on a liquid explosives truck

sures (Photo 1). Because volumes are tracked, accurate flow is important, and gentle handling of the volatile substance is critical. Maintaining acceptable levels of pump reliability in these services can be challenging. But reliability can be significantly improved and maintained by following simple guidelines: application, selection, installation, monitoring, maintenance and training.

Start with the Application Accurate information on actual operating conditions of a particular service is invaluable. Reliability diffi-

Figure 1. Typical progressing cavity pump showing rotor, stator connecting rod and drive shaft. The Pump Handbook Series

culties with PC pumps (as with other designs) are often the result of relying on incorrect or incomplete operating data. Correct information is vital for making a proper pump selection. The minimum required data should include pump flow and discharge pressure, the chemical nature of the fluid being pumped, percentage and size of solids, and fluid temperatures at startup and normal operating ranges as well as any upset conditions. Chemical compatibility is especially important for pump stator and rotor selections. Information on preferred materials of construction, especially if it is based on previous experience with the fluid, is vital. Even information about pipe configurations can help application engineers with pump selection, so piping dimensions should be passed on to your pump vendor to help in proper equipment selection.

Make the Right Selection – Pump Speed Is Vital! One feature of the progressing cavity design that makes it so durable in difficult services is its design to run at reduced speeds, not at direct motor speeds. This reduction in speed is typically from 1750 rpm to 100-400 rpm. Such a decrease and the way the fluid travels axially through the pump dramatically lowers internal fluid velocities and thereby reduces wear. Gear reducers installed between the motor and pump are typically used for speed reduction. Constant torque variable

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frequency drives are used in variable flow applications to vary the pump speed. When sizing a PC pump for an application, speed has a tremendous influence on pump reliability and service life. Unfortunately, it can also increase the initial pump investment. Reducing pump speed increases the size of the pump. The user who installs a pump at 400 or 500 rpm or higher has a smaller investment, but the complete life cycle lost will be dramatically higher than a pump installed with a speed of 250-300 rpm. PC pump parts subject to abrasion will wear out about four times as fast at 500 rpm as at 250 rpm. This not only means four times the amount of spare parts, but it also results in additional costs for repairs and down times. The misapplication of progressing cavity pumps at inappropriately selected high speeds is what causes market misconceptions about their life and overall cost. “Biting the bullet” and making the investment in the next larger unit running at lower rpm will result in a reliable, longer lasting pump.

Make the Right Stator Selection The most often replaced part on PC pumps is the stator. The wrong stator elastomer selection can be an expensive and time consuming mistake to correct. The key factors in stator selection are chemical compatibility, temperature, pressure ratings and operating temperatures. Different elastomers have varying maximum temperature limits. Check with your manufacturer because there are many elastomer materials to choose from today. Chemical compatibility lists are readily available to aid in the elastomer selection. The pressure rating of the stator also has an important influence on its life (Figure 2) and varies with pump design and manufacturer. The higher the pressure rating of the stator, the better the volumetric efficiency. Volumetric efficiency has a large impact on stator life due to back flow or “slip.” Slip is what generates abrasion and wear to the stator surface as the solids “slip” back between the rotor

242

Figure 2. PC pump stator rated at 175 psi - note self gasketing overlap end design

and stator under discharge pressure. Again, to maximize pump reliability, a savvy PC pump user will choose a stator with a higher pressure rating.

Installation and Startup One secret to the innovative design of the progressing cavity pump is the interference fit between the rotor and stator. This sealing line enables the rotor/stator to move the fluid in sealed cavities. However, tight tolerances make the pump susceptible to damage from dry running, which can result in a burned out stator and pump failure. Always make sure the pump is filled with fluid prior to startup. And as with other positive displacement designs, a PC pump should not be dead-headed. If it is, stator failure will result unless the motor overloads first. As always, the pump should be installed free of any pipe stresses. PC pump suppliers have instruction manuals for installation, startup and troubleshooting techniques. It’s a common reminder, but ...be sure to study the manual.

Monitor Your Pump Service The service life of your PC pump is unique to your installation. There have been some installations in which stators lasted one week, and there are others in which stators are still in service after five years. Understanding the requirements of your system and tracking pump maintenance and service life are key in improving progressing cavity pump reliability. Tracking service hours and pump performance in the first few months following installation and startup is particularly useful for effectively planning preventative maintenance service schedules in the future.

The Pump Handbook Series

Figure 3. Universal joint with wear bushings

Preventive Maintenance of the Pump Drive Train is Essential The drive train, which consists of the shaft, bearings, shaft wear sleeve, universal joints and coupling rod, should be monitored as part of a preventive maintenance schedule. Suggested scheduling for inspection and service are provided in suppliers’ instruction manuals. Use these schedules as a starting guideline. Monitoring your pump in actual service will indicate when parts need to be examined, serviced and replaced. Following a sound preventive maintenance program will result in minimal replacement of wear parts. It will also protect big dollar items such as shafts and coupling rods. Most important, it will help reduce the chances of catastrophic failure. Most progressing cavity pumps manufactured today incorporate design features that enable the user to isolate wear to small and less expensive parts such as shaft wear sleeves and wear bushings (Figure 3). Replacing these as part of a regular service schedule – before they wear out – will save dramatically on overall pump repair costs and down time.

Proper Training is Vital Proper training on the assembly and disassembly of particular pumps will result in more efficient repairs and lower overall maintenance costs. Training will also help service personnel be less intimidated by unusual design aspects and unique parts. Designs vary greatly from one manufacturer to another, and knowing the idiosyncrasies of specific

pumps is invaluable. Most suppliers today will provide training to their customers at no charge.

Keep Inventory of Parts Based on Needs Spare parts also impact your pump service. Delivery times from suppliers and down time limits should determine what parts to keep on hand. The bare minimum requirement for getting a PC unit back in service quickly is usually a replacement stator and a shaft seal. However, if your supplier stocks the required parts, you may not need to carry these costs too. This is another important role the supplier can play in reducing service costs. Check it out. If a supplier stocks it, you don’t need to – as long as delivery time is quick enough.

Improving Reliability of Existing Installations Let’s suppose your pump has been selected and installed correctly and is now in service. But you need it to operate longer, and you need to reduce downtime even more. You can take further steps to improve your situation. One is to add a dry run protection device that shuts down the pump before the stator runs dry. Several designs are available for this protection, depending

Figure 4. Slotted adjustable stator provides extended service in severe duties.

on the pumped fluid and service. If your application requires an expensive stator – Viton™ , for example – then a dry run protection device could pay for itself quickly. All PC pump suppliers have these devices and can provide one best suited for a particular application. They can be retrofitted in the field as well. Another important way to increase rotor/stator life is to upgrade materials of construction. In some cases, simply changing to a softer stator material will dramatically increase stator life. The stator will flex instead of abrading when solid particles pass by. Another helpful innovation is an adjustable stator that enables the user to reintroduce the interference fit and thereby increase stator life by up to four times (Figure 4). Of course, the rotor can be supplied with even better coatings and

The Pump Handbook Series

hardness properties to extend service life. Another easy fix is to replace an existing stator with one that has a higher pressure rating. This can be done by using different stator designs or by installing longer rotors and stators. Retrofitting to a higher pressure rating has an effect similar to reducing pump speed. Doubling the pressure rating can make stator service life up to four times as long. By following the basic guidelines listed above, you can improve the reliability of your PC pumps regardless of service. Conservative pump selections at lower rpms and higher pressure stator designs will result in longer mean times between failures (MTBF) and lower overall service costs. In addition, an investment in rotor and stator design upgrade can increase service life in existing installations. And last, thorough personnel training coupled with a preventative maintenance program will keep your pumps running better and longer. It all adds up.■ Larry Shanley, Jr. is the National Sales Manager for Shanley Pump & Equipment (Arlington Heights, IL), the exclusive agent for Allweiler products in the United States. A graduate of Loyola University Chicago, he has more than 15 years of experience in the pump market.

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R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

Water-Lubricated Fluid Film Bearings Can Be Trouble-Free Proper design and the right materials are keys to success. By Stanley Abramovitz, Consultant

ow viscosity, corrosion and poor boundary lubrication are the three drawbacks to water and other similar process liquids used as bearing lubricants. All three difficulties, however, can be mitigated by choice of materials and proper fluid film bearing design in which a film of lubricant completely separates the bearing surfaces during operation. Although crude water lubricated bearings have been around since ancient times, reliable and troublefree bearing life in modern pumps and turbines can be achieved by using fluid film bearing design that has been proven by test and field experience. A bearing that “works” for 15 or 30 minutes and then fails never “works” at all – it fails at the start and just takes time to destroy itself. A bearing that is monitored for expected wear is, at best, operating with a partial fluid film in which bearing materials are critical for achieving continuous operation. Thus, bearing life is unpredictable and impacted by any adverse operating conditions such as overload, vibrations and shock. From a practical standpoint, if the bearing life is acceptable, then so is the design. If not, a full fluid film bearing design is needed.

L

Protecting Against Liquid Weaknesses In comparison with oil lubri-

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During long shutdowns, for example, cants, water and similar liquids have a bearing may need to be run periodthree important weaknesses that ically to flush away oxide particles. must be addressed in bearing design. 1. Low Viscosity. Because the vis3. Boundary Lubrication. Oil coats cosity of a conventional oil can be as surfaces with a few layers of lubricant much as 100 times that of water, the molecules. This boundary lubrication thickness of the fluid film separating reduces friction when the bearing is the bearing surfaces in a water lubristarted (before a significant hydrodycated bearing can be as little as 1/10 namic fluid film is developed), when the thickness of an oil film. Thereit comes to rest and during short durafore, bearing performance prediction overloads, shock loads and peritions are critical, and design analysis ods of significant vibration. Although should be based on proven methods no similar boundary lubrication can along with practical field experience. be expected by liquids such as water, With the expected thin films, the special bearing materials and, in bearing should, in most cases, be some cases, additives to the lubricant, designed to compensate sensitively can provide this needed condition. for misalignments and deformations. Also, bearing surface finish, precise The Fluid Film Bearing surface contour and good filtration of the liquid lubricant are important Fluid film lubrication exists considerations. when a fluid film completely sepa2. Corrosion. For bearing materirates the two bearing surfaces. This als immersed in corrosive liquids, can be achieved only if the fluid in corrosion is a limiting factor. It is not the clearance space is pressurized to so much that it constrains structural components as Pressure that corrosion products Film Distribution W can produce harmful dge changes in the condition wE tflo Ou U of the bearing surfaces Q–Qs h2 Q h1 and contaminate the Qs /2 lubricant. Contaminant dge (one side) wE products can cause seiInflo Tilting-Pad Bearing Three-Dimensional Pressure Profile zure through scoring and wear. Also, oxides can Figure 1. The wedge shaped fluid film showing the build up when bearing fluid flowing in and out of the film space, variation surfaces are in static con- in the film thickness and the three dimensional tact for a period of time. pressure profile that supports the load. The Pump Handbook Series

the extent that the fluid pressure forces balance the bearing load and maintain equilibrium. This means that the fluid must be continuously introduced into and pressurized in the film space. Figure 1 shows the fluid configuration and pressure profile that occurs with a tilting-pad thrust or journal bearing in the most common mode – self-acting (hydrodynamic). The film pressure is selfinduced by the relative motion between the two bearing surfaces. Straight-sleeve-type journal bearings also form a natural geometric film wedge when the journal is displaced in its bearing. In the second bearing mode, an external pump is used to induce flow and pressure. This is known as a hydrostatic bearing. Although it will not be discussed in this presentation, it is a very useful design concept for low speed, high load bearing applications. However, it does require an extensive external lubrication system.

Bearing Design In the early 1950s, I was involved in a program to design and develop water lubricated bearings for a circulating water pump in the primary sealed water flow loop of a nuclear pressurized water reactor (PWR). The large vertical “canned” pump (Figure 2) was about 10 feet high and used the circulating water for lubrication of the thrust and jour-

Figure 2. A canned motor water circulating pump used in a pressurized water reactor in which the thrust and journal bearings are lubricated with the circulating water.

nal bearings. In this sealed water loop, zero leakage and bearing reliability with long trouble-free life were essential. Valuable information concerning materials and design concepts for fluid film bearings using low viscosity corrosive liquids was gained from this sophisticated bearing application. Fundamentals have not changed since then, and those strict requirements were the basis for reliable future designs meeting different bearing sizes, conditions and pump and turbine requirements. Also, excellent non-metallic low friction bearing materials developed over the years have become useful for starting and stopping conditions, short duration overloads and shock. Because of the relatively thin liquid films separating moving bearing surfaces, it is usually essential that both journal and thrust bearings be self-aligning. And because the forces that the bearing can reliably support are relatively small compared to a design that uses oil, the self-aligning components must be sensitivity free and designed to move and articulate in order to be effective. However, when cost is important and you’re dealing with small size bearings with low expected deflections and distortions, self-aligning features may not always be necessary (although at possibly some sacrifice in life and reliability).

Thrust Bearing Figure 3 shows the basic components of a tilting-pad thrust bearing. There are a number of segmented pads, or shoes, each resting on a pivot so that it is free to tilt to the optimum film-wedge angle. Pivots can range in complexity from jack screws (for adjusting each pad to a common plane) to a simple knife edge step machined into the back of the pad. For maximum bearing load capacity each pad should support an equal share of the total load. With thin films we can expect a certain amount of misalignment and distortion. The tilting-pad equalized bearing (Figure 4) is a sensiThe Pump Handbook Series

Sliding Surface or Runner

WT

N

W

Pivot

Pads

Figure 3. A tilting-pad thrust bearing with segmented pads mounted on pivots.

tive and reliable design that incorporates the desired features, although at the most cost. Equalizers leave each pad free to move up or down for shouldering an equal share of the load regardless of perpendicularity between thrust-collar face and shaft. The design depicted in Figure 4 has hardened pin pivots. Also shown is a bearing material that is loosely held in a pad retainer. This scheme is intended for materials that have low strength and resistance to shock loads. An alternative method is to simplify the pad design and use weak material as a thrust runner in the form of an annular ring shrunk into a metal disk. However, nonmetallic bearing materials that can be bonded to the flat metal pad surface result in a simpler design. The choice depends on the bearing size, operating conditions and application.

Pad pivot A

Bearing surface

Inserted material Pad Equalizers

Cage

Pin

A

Equalizer pivot

Section A-A

Figure 4. A fully equalized tilting-pad thrust bearing in which each pad supports an equal share of the total bearing load – designed for free and sensitive movement of the articulating members and shown with an inserted bearing material, although a bonded non-metallic could be used if applicable.

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Material removed

Bearing shell Inserted bearing material

Flat

Taper

Retaining pin

Figure 5. A tapered-land thrust bearing with no self-aligning features and a limited load capacity

A tapered-land thrust bearing (Figure 5) is less complex and, with care, can be used for relatively small size bearings where loading and misalignment is not excessive. Supporting the ring on a spherical seat adds some amount of self-alignment, as it would for an unequalized tilting-pad bearing. Flatness must be accurate, and the optimum amount of taper is critical – it should be about the same amount as the expected film thickness. Depending on the material, some wear during start-ups and shutdowns could reduce the taper and limit performance. The simplest thrust bearing is a flat washer similar to a seal face, or one with radial grooves that are either open ended or dead ended with chamfer reliefs for controlled radial groove flow. This bearing has a limited load capacity, particularly with low viscosity lubricants, but it can be used as a guide when loads are relatively light.

Journal Bearing Here again, because of the thin films we can expect, a self-aligning design is essential if the journal diameters are large and there is a long span between bearings (although such a design may not be needed with small bearings). Figure 6 shows a sensitive self-aligning 360º journal bearing with a hardened toroidal pivot. Low strength bearing materials should have a shrink fit in the bearing shell to place them under compressive stress. With inserted materials or bonded non-metallics, spiral or straight axial grooves are usually needed for lubricant flow. Because the expected bearing power loss is small when using a low viscosity fluid, high bearing speeds are practical. However the combination of high speed and low load tends

246

Toroidal surface (hardened)

End plate Housing

Shaft (journal)

Figure 6. A self-aligning journal bearing with a low strength bearing material insert Journal center Journal

Inserted bearing material

End plate

Pad pivot (hardened)

Figure 7. Tilting pad journal bearing

to create hydrodynamic instability (whip and whirl), in which the journal center rotates about the axis of the bearing. If metal to metal contact occurs, the bearing can destroy itself in a short time. A tilting-pad journal bearing (Figure 7) will prevent instability in most cases, but a stability analysis should be made. Other advantages of this type of bearing are that it can easily be made self-aligning; it has better load capacity because of efficient cooling of the lubricant between pads, and, with radially adjustable pivots, the diametral clearance can be closely controlled and easily changed to correct excessive clearance that may be harmful. The curvature of the pad surface of this bearing is an important design factor. As related to the bearing clearance, it is termed “preload.” Here, as with the thrust pads, the use of deposited bearing material or bonded non-metallics can simplify the design.

Bearing Materials Major factors in the selection of bearing surface materials for use with low viscosity corrosive lubricants are compatibility (friction, wear, tendency to score and gall), conformability and corrosion. BabThe Pump Handbook Series

bitts and straight carbon steels have poor resistance to corrosion. Corrosion also limits the use of copperbased alloys, particularly at elevated temperatures. If the bearing is designed properly with a reliable film thickness value, material compatibility comes into play during starts and stops when there is an insufficient film separation of the bearing surfaces, or when there are short duration bearing overloads, shock loads and vibrations. Conformability can be important when a bearing surface must accept and “wear-in” some amount off misalignment without surface destruction. There is a wealth of material information in the literature, and only a brief overview will be given here. Carbon graphite (Graphitar™ U.S. Graphite Co.) is compatible with many materials, provides its own solid boundary lubricant, can be machined easily, and conforms (by wearing) without serious damage to the journal or thrust runner surface. However, because of low tensile strength it must be well supported and retained. Tests have shown good capability with Stellites, hardened 17-4 stainless steel and hard chromeplated stainless steel. Ceramic and sintered carbides, such as aluminum oxide (against itself), tungsten carbide (against boron carbide), and silicon carbide (against itself), have good wear characteristics and are generally free from galling. However, they are more sensitive than carbon graphite to contaminants in the lubricant, and good filtration of 2 to 4 microns is desirable. Deposition of some of these materials on a base material would eliminate the need for retaining those that are relatively brittle. Rubber, elastomers, laminated phenolics, thermoplastics and special proprietary non-metallic materials that can be bonded to a metal substrate or used as staves or solid inserts have excellent bearing characteristics for use with these lubricants. Lignum vitae, a wood that has a long and successful history, is corrosion resistant and has high resistance to wear and compression. In addition, its high resin content provides selflubricating properties.

Bearing Performance Hydrodynamically, low viscosity liquids have the same characteristics as oil. Viscosity decreases with increasing temperature, and these liquids can be considered incompressible. However, at atmospheric pressure water has a relatively low boiling point, and cavitation may be encountered when operating the bearing at elevated temperatures. Also, since water and other low viscosity liquids have a low kinematic viscosity, it is necessary (even at low speeds) to analyze for turbulence in the fluid film. Such factors as surface finish and contour and usable bearing surface (determined by alignment) influence the maximum load a bearing will support with a practical minimum film thickness. For bi-directional shaft rotation with centrally pivoted thrust bearing pads, a convex pad surface is needed. The amount of contour is an important design variable when dealing with low viscosity liquid lubricants. Although water-lubricated bearings have worked with a calculated film thickness of only 0.0001”, calculated film thicknesses should be on the order of 0.0003” to 0.0005”, depending on the bearing size and operating conditions. It is dangerous to generalize about acceptable unit loading since variables such as surface velocity and fluid temperature significantly influence performance. As a guide, however, 25 psi would be a conservative value, with a maximum of about 40psi. With all of the variables involved, including practical considerations such as alignment and deformations, a proven and reliable

10

0.01

8

0.001

Water at 80F

Water at

0o E2 SA

at 100F

Horsepower

S A E 2 0 oil

Film thickness, in.

For structural materials, bearing supports can be made of copper alloys or stainless steel. The 300 series stainless (Cr-Ni) is the choice for maximum corrosion resistance, and it is also used for bearing components. Bearing pivots should be as hard as possible, where the hardenable 400 series (Cr) stainless (440-C) or the 300 series 17-4 PH are hardenable. Hardfacing (Stellite), “nitriding” and solid hard materials (tungsten carbide) can also be used.

il a

0F t 10

6

4

2 00 F

2

0.0001

Water at 80F Air at 8 0F

Air 0 0

20

40

60

80

100

120

Load, psi 0

20

40

60

80

100

120

Load, psi

Figure 8. Comparison of film thickness for different lubricants as a function of bearing load

bearing analysis is essential when using low viscosity lubricants. As a qualitative representation, unit loads for various film thicknesses are listed in Figure 8 for a bearing operating with oil, water at two temperatures and air. Film values of this order are what can be normally expected in fluid film bearings. Reliable and safe operation can be obtained with a correct bearing design for the operating conditions, good bearing surface finish and clean lubricant.

Power Loss and Temperature Rise Power loss will be lower with low viscosity liquids than with oil. Reduction in power loss with a low viscosity liquid is somewhat offset by the small film thickness, which is inversely proportional to power. Nevertheless, the reduction in power loss is significant, particularly when operating at a relatively low unit load. Figure 9 qualitatively compares power loss with air, oil and water lubricants in the same bearing. In general, temperature rise of the fluid as it passes through the bearing is usually quite low, as little as a 5 to 10ºF or less. In the case of water, its relatively high specific heat (twice that of oil) further limits the rise of the fluid temperature, although where high operating temperatures can be expected there must be cooling or sufficient pressure to prevent it from flashing into steam.

The Pump Handbook Series

Figure 9. Friction horsepower loss as a function of bearing load for different lubricants

I hope this presentation has provided the reader with practical information helpful to both designers and users of pumps and turbines whose bearings are, or may sometime be, lubricated with low viscosity corrosive fluids.■ References: Abramovitz, Stanley Fluid Film Bearings: Fundamentals and Design Criteria and Pitfalls. Proceedings of the 6th Turbomachinery Symposium (1977) Fuller, D. D. Theory and Practice of Lubrication for Engineers. Second Edition, John Wiley & Sons (1984) Abramovitz, Stanley TroubleShooting Bearing Problems in Large Hydroturbine Generators. Hydro Review, Volume X, No.6 (October 1991) Abramovitz, Stanley Using Tilting-Pad Journal Bearings in Hydroturbine Generators. Hydro-Review, Volume XV, No.5, (August 1996) Stanley Abramovitz has for more than 40 years, been a consultant involved in bearing design and the solution of bearing problems for a wide variety of applications in practical field conditions. He has a Master of Science degree in mechanical engineering from the University of Pennsylvania, has authored more than 300 papers and lectured extensively at national conferences and company seminars. He can be contacted at P.O. Box 811945, Boca Raton, FL 33481, phone/fax (561)4513470

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R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

Oil Lubrication for Process Pumps and Related Equipment Conservative design and proper specifications minimize failure possibility and maximize life of rotating equipment. By Vinod P. Patel, M. W. Kellog Company, and Donald G. Coppins, Lube-Power, Inc. il lubrication systems are the life insurance policy for processes. If a lube system should fail to perform, the results are bearing failures in pumps, gears or drivers, and ultimately a shutdown of the process. Conservative design and proper specifications minimize the possibility for failure and maximize the potential for long, trouble-free service life of rotating equipment. Understanding a system’s needs, optimizing component selections, allowing for operational maintenance and proper application of the system are the keys to reliability. The pump user’s objective is to maximize revenue from his or her process. To meet this objective, it is vital to have safe and reliable operation of pumps and auxiliaries. Larger process pumps and drivers often use hydrodynamic bearings because of their noted reliability over anti-friction bearings at higher bearing surface velocity. The oil film that forms in hydrodynamic bearings provides the needed separation between shaft and bearing. This continuously replenished oil film removes bearing heat and wear particles. Oil film properties depend on many factors including oil viscosity, bearing/shaft clearance and bearing load. A constant oil supply is

O

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PSLL

PSL B

PSL A

PI

PDI

Drive

Pump

Duplex filter

TI TI

TI FI

TI FI

FI

Shaftdriven oil pump

TI Heat Exchanger

FI

PI

Motor-driven auxiliary oil pump

Oil reservoir

FI TI PI PDI PSLA PSLB PSLL

Flow Indicator Temperature Indicator Pressure Indicator Pressure differential indicator Low-pressure switch (auxiliary pump start) Low-pressure switch (alarm) Low-pressure switch (trip)

Instrument (letters indicate function) Gate valve Relief valve Line strainer Pressure control valve Check valve Block and bleed valve Reflex-type level indicator

Note: This illustration is a typical schematic and does not constitute any specific design, nor does it include all details (for example vents and drain).

Figure 1. API 610, 8th Edition Lube Oil System Schematic (reprinted, by permission, from the American Petroleum Institute) The Pump Handbook Series

required to keep the bearing system in balance. A loss of oil supply means a loss of cooling, loss of oil film properties, and eventually the loss of a bearing and possibly even more. Figure 1 shows an API 610, 8th Edition Lube Oil System schematic. The base orientation requires the main lube oil pump to be shaft driven off the process pump. A single auxiliary lube oil pump is electric motor driven. Balance of system includes reservoir, single shell and tube cooler and duplex filter. Each pump is equipped with its own relief and check valves. Instrumentation includes three pressure switches, one pressure differential indicator, two pressure gauges, two thermometers and a sight glass on the reservoir.

oil and provide for rundown capacities. • Provide for temperature fluctuations, expansion volumes, location for heating and oil purifier connections.

FILLER/BREATHER MANWAY (TYP)

OIL RETURN FROM TANK TOP

SUCTION CONNECTION TO INTERNALLY MOUNTED PUMP SUCTION OF TANK TOP MOUNTED PUMP

OIL RETURN

SUCTION LOSS SUCTION CONNECTION HORIZONTAL MOUNTED

PUMP OR FLOODED SUCTION DRAIN WITH API 610, 8th Edition PLUG 3.0" MIN. 2.0" and other good design STRAINER requirements include: 2.0" 1.0" MIN. • Materials–Austentic AIR GAP RUNDOWN LEVEL stainless steel RUNDOWN CAPACITY CL MAXIMUM OPERATING LEVEL • Retention–3 minutes OIL RETURN MINIMUM OPERATING LEVEL retention time is defined as 1.0" WORKING CAPACITY SIGHT GLASS RETENTION the total capacity below CAPACITY SUCTION LOSS the minimum operating level of the reservoir. Normal system flow rate x 3 = Figure 3. Reservoir suction and return line orientation/reservoir capacities retention capacity. • Armored sight glass– ler breather cap should be located on Internal chamber is the same materiTypical Oil Lubrication a riser to prevent water from running al as the reservoir. The sight glass System into reservoir. Fill cap should have a should extend from one inch above Components, their purpose, API 40 micron breathing element and a center line of the oil return to one 610 requirements, details of applica60 mesh strainer to prevent foreign inch below minimum operating levtion and maintenance consideraairborne particles and objects found el. tions are evaluated in the following. in new drum oil from being ingested • Baffle–The function of the bafinto the system. fle is to separate pump suctions from Lube Oil System Reservoir. The • Mounting pads–Mounting pads oil returns. Oil returns include those oil reservoir has these basic purposes: should be used for any attachments from process pumps and auxiliaries, • Dissipate or settle contamito the reservoir. Side mounted lube oil pump relief valves, pressure nants. Air is dissipated by means of devices require a pad to prevent the control valves, filter vents, control proper baffling and adequate resifoot of the device from piercing the valve head and instrument vents. dence time. Particulate matter is reservoir skin. This is especially Baffles should have an air passage at allowed to settle in the low end of the important during transportation. Top the tank top equal to three times the reservoir. Residence time and flow mounted devices should be bolted to area of the auxiliary and main pump rates in the reservoir determine parpads. To prevent water ingress into suction lines combined. A baffled ticulate disposition in the reservoir. the reservoir, holes for top mounting passage is strategically located 2 Water is heavier than oil. The low components should not penetrate the inches above the tank bottom. This end of the reservoir must be top of the reservoir. baffle should have the greatest possidesigned for water drainage. • Manway–A gasketed manway ble distance between the main pump • Store a prescribed amount of and riser should be supplied so inlet and return line to allow oil access is provided to all compartto pass from the return side to A ments of the reservoir. the suction side of the reservoir. • Return line–Return lines shoTo provide the maximum OIL RETURN uld terminate below the oil level to amount of time for settling of FLOW prevent foaming. They should be contaminants, we try to make STRAINER equipped with end baffles or difthe oil travel as long a path as fusers or be angle cut at 45 degrees. possible from the time it OIL RETURN MAIN PUMP Return lines should discharge away returns to the reservoir until it SUCTION from the pump suction and the reserTYPICAL RETURN LINES reaches the pump suction (FigA voir bottom. ure 2). AIR PASSAGE • Pump suction lines–Pump suc• Sloping bottom–The retion lines should be straight pipe servoir bottom should slope a SIZED 3 TIMES AREA with a minimum number of elbows minimum of 1/4 inch per foot OF COMBINED PUMP SUCTION LINES to avoid accumulation of air and away from the pump suctions. result in smooth pump transfers. The low end of reservoir should be equipped with a drain. VIEW A-A OIL PASSAGE Factors involved in reservoir • Reservoir fill cap–The filsize and selection are: Figure 2. Reservoir baffle orientation/design

The Pump Handbook Series

249

1. System flow rate 2. Retention time 3. Height of return line from process equipment 4. Required working capacity 5. Rundown capacity 6. Location of auxiliary pump (internal or external of reservoir) * See Figure 2 for baffle design and location. Figure 3 shows suction and return locations in the reservoir as well as reservoir capacities. Assume a total lube oil flow rate of 20 gallons per minute. We are lubricating the process pump, a gear and a turbine driver. 20 x 3 = 60 60= retention capacity (total capacity below alarm level) Working capacity is defined as the capacity between minimum operating level and suction loss level. Rundown capacity is the amount of oil expected to return to the reservoir when the system is shut down. (If this is unknown, use 10% of the total capacity of the reservoir. Refer to Figure 3 for details.) Other design concerns, issues and maintenance considerations • Heavy snow or rain locations– Equipment mounting pads should be continuously seal welded around the outside and inside edge of pad. Manways and reservoir fill caps should have a minimum 1-1/2 inch riser. Tank top mounted pumps should have gaskets where the flange of the bell housing meets the tank top and where the pump mounting flange meets the bell housing. Frost or rain guards should be considered for sight glasses, and you should consider oil purifier connections. • Extreme ambient conditions, such as below 20ºF and above 112ºF. Consider oil purifier connections since condensation is likely to result in free water in the reservoir. Additionally, increase slope and water storage area in reservoir, and consider installation of insulation and/or reservoir heater. *Caution–Always add a level switch to the reservoir to shut down the heater on low oil level. When the oil level becomes tangent with the heating element, a risk of fire is pre-

250

sent. • Dirty/Desert atmosphere–Add oil purifier to purify oil and remove condensation. Also add oil sampling connection in reservoir or return line piping and an oil sampling connection in the supply line downstream of the duplex filter. • Avoid: 1. Reservoir end covers with a single bolt mounted on the side of the reservoir. When these leak, and they all do, the tendency for maintenance personnel is to tighten the bolt. This distorts the cover and results in a larger leak. 2. Horizontal heaters that require draining the tank to change the elements. Open coil heaters installed in atmospheric thermowells will lower the element watt density, prevent oil coking and allow for element changes without having to drain the oil. 3. Locating level switch floats in or near oil flow paths. This can lead to bent stems or the float falling off the stem. In both cases the alarm doesn’t function. It is recommended that reservoirs be washed with a safety solvent prior to filling with oil. Use lint-free rags and solvent that doesn’t leave a residue.

Lube Oil Pumps The purpose of the main shaft driven lube oil pump is to provide a flow rate to the process pump and other equipment in the drive train any time the process pump is operating. The purpose of the auxiliary lube oil pump is to provide a lube oil flow rate to process pump bearings and other equipment in the drive train prior to start-up of the process equipment and any time the lube oil supply pressure drops below the specified minimum. Many styles of pumps are available for lube oil system applications. The API community has favored two types: gear and screw pumps. The gear pump is most viable for applications at 1200 to 1800 rpm input shaft speeds, with oil viscosity ranges 100 to 500 SSU, and where lube oil pump discharge pressures are greater than 150 psi. Screw pumps are best for applications with low noise requireThe Pump Handbook Series

ments, where steel cast pumps are required and where pump input speed exceeds 1500 rpm and high viscosity fluids are predominant. Screw pumps are also preferred when pump discharge pressure is 150 PSI or less. (Note: screw pumps are available for higher pressure applications; however, gear pumps tend to be more economical and more efficient in higher pressure applications.) Pump sizing. The process pump manufacturer as prime contractor must determine total flow rates and heat loads and specify this requirement to the lube oil system supplier. Assuming the train requires 15 gpm, the actual minimum required pump flow is calculated as 15 gpm = 17.64 gpm .85 .85 is used as the minimum acceptable value added to maintain head in the lube oil supply header. A minimum of 1.5 gpm extra must be available on systems in the 5 to 12 gpm range for proper valve sizing and for maintaining head. The pressure control valve is sized along with the pump, and it should be sized so that the excess flow of 2.64 gpm will pass over it when open between 15 and 20%. The extra oil passes over the pressure control valve back to the reservoir It is most important to size the main pump and auxiliary pump displacements as close to identical as possible. In any event, major deviations in flow between the two pumps will result in a change in pressure when the auxiliary pump runs alone, as compared to when the main pumps on line are operating alone. Further, consideration must be given to the pressure control valve sizing so that the valve is only 80% open when the excess flow of the auxiliary pump and the full flow of the main pump pass over the pressure control valve. This situation occurs whenever the pumps operate simultaneously. Material Requirements. API 610 requires that the pump housing material be cast steel when the pump is located outside of the reservoir. Pumps located inside the reservoir can be of any suitable material. Installation. API 610 does not

RESERVOIR

BELL HOUSING

MANWAY COVER MANWAY GASKET IMO PUMP

MANWAY RISER

RESERVOIR

STRAINER

Figure 4. Vertical pump mounting assembly

address the issue of spacer couplings for the auxiliary pump. Therefore, auxiliary pump seals typically cannot be changed during operation of the main pump unless the auxiliary pump is removed from its sole plate. Pumps that are mounted vertically in the tank should be mounted so the coupling is above the tank top (Figure 4). Mounting of the coupling below the tank top can result in aeration of the oil due to aerodynamic effects of the coupling. Pump motor bell housings are available line bored to .002 TIR. This tolerance eliminates the need for field alignments at the time of initial start up or when replacing a pump. Shimming is not required. We recommend the pump installation be vertical on a bell housing with C face motor per Figure 4. The bell housing insures a permanent alignment. The vertical pump reduces skid size (real estate is valuable) and the installation is far cleaner and the most economical. Horizontal pump installation can be made either as foot mounts or as C face bell housing mounts. Foot mounting requires shims, sole plates and field alignments. Horizontal pump applications typically include a Y strainer mounted outside of the reservoir in the pump suction line. Maintenance Considerations. For horizontal pump mounts equipped with proper sole plate and a spacer coupling, a field seal change can be made on larger screw pumps. Note, however, that not all screw pump designs and sizes can have their seal changed from the shaft

The Pump Handbook Series

Y-STRAINER

MOTOR IMO PUMP SHIM MOUNTING PLATE

SHIM COUPLING

Figure 5. Horizontal pump mounting assembly CONNECT TO ELEMENT HOUSING

F I L T E R

BALANCE LINE GATE VALVE 2 POSITION 6-WAY TRANSFER VALVE OIL OUT

OIL IN

FO VENT GLOBE VALVE WITH PLUG

F I L T E R DRAIN GATE VALVE WITH PLUG

Figure 6. Duplex filter with transfer valve, vent, drain and balance lines Oil Out

Normally open

FO

Filter

COUPLING

COUPLING GUARD

Filter

MOTOR

end. Y strainers are located outside the reservoir to allow for easy cleaning of the screen. This application requires additional components such as a suction isolation valve for the pump (Figure 5). The strainer typically is 100 mesh, intended to keep only coarse particles out of the pump. It is our belief that if this strainer ever becomes clogged; it is time you entered the tank and cleaned it thoroughly. Consequently, vertical pump mounts with internal strainers make the most economical design since the extra pump suction isolation valve is eliminated. Vertically mounted pumps (with pumps installed inside the reservoir) require that the tank be opened if a pump change is necessary. Opening the reservoir presents the possibility of contamination. Due to improved auxiliary pump suction conditions and the pump’s proximity to the reservoir heater, a vertically mounted pump will be available sooner for cold starting than a horizontally mounted pump. Driver Sizing and Effects of Viscosity/Design Impacts. Total pump discharge is calculated as the nominal output flow rate required by the equipment divided by .85. Our test case is 15 gpm nominal flow with the pump producing 17.64 gpm minimum. Main pump speed is 3000 rpm since this application is 50 Hz, so the electric motor driven pump must run at 3000 rpm (Figures 6 & 7). Figure 6 shows this pump is only acceptable at discharge pressures at or below 50 psi. Further, oil viscosity must be at least

See Note 1

PDI

PDSH

See Note 1

See Note 2 TI TSL

TIC

See Notes 3 and 4 Bypass

TIC

TSH Alternative arrangement (see Note 5)

TCV

Cooler

TCV

FC

See Note 1

See Note 1

Cooler

Bypass Normally FO open Notes: 1. Option A-17a: The purchaser may specify tight shutoff requiring spectacle blinds. 2. Option A-17b: The purchaser may specify a hightemperature switch (TSH) and/or a low-temperature switch (TSL). 3. Option A17c: The purchaser may specify a bypass oil line and a constant-temperature control valve.

TI Oil In

4. Option A-17d: If the fail-close (FC) feature of the direct-acting temperature control valve is not acceptable, the purchaser may specify a valve with a faillocked (FL) feature. 5. Option A-17e: The purchaser may specify a constant-temperature three-way control valve as an alternative

Figure 7. Twin oil coolers and filters with separate continuous-flow transfer valves (reprinted, by permission, from the American Petroleum Institute)

251

150 SSU to expect a pump discharge above our required minimum flow rate. Applications in the process pump industry generally use ISO VG 32 or VG 46 oil. The pump draws its oil supply from the reservoir, which is upstream of the heat exchanger, so this is hot oil that has just rinsed heat out of the pump bearings and into the tank. At this point oil temperatures approach 160ºF, and oil viscosity is approximately 70 to 90 SSU. Therefore pump 3E-118 is too small for this application (Figure 6). A 3E143J pump (Figure 7) will produce the proper flow rate at low viscosities. Many applications require the lube oil pressure to be higher than the cooling water pressure (to prevent water from entering the oil from a leaky tube). Typically our pump discharge pressure is around 100 psi. At 100 psi and 100 SSU our pump produces 22.1 gpm and requires a brake horsepower of 2.0 to drive it. If we have a cold start condition, the brake horsepower becomes 3.2 hp when the oil viscosity is 1000 SSU. Oil temperature of 45ºF relates to 1000 SSU. The customer needs to specify if a cold start up is required, and if so, either a heater should be installed or the pump driver and coupling must be sized for maximum load. Higher pressures also affect the selections of the relief valves and pressure control valves.

Filters The requirements of API 610, 8th Edition include the following: • Type: Duplex • Material: Housing: Cast steel Elements: Cartridge materials should be corrosion resistant. • Filtration: 25 micron or finer • Non bypass, stainless steel valve spool in transfer valve. Here are some additional suggested requirements: In the past filters were rated by micron size. Subsequently we changed filter ratings to nominal versus absolute micron ratings. Today, filter elements are rated with micron and beta ratios. In the API process pump

252

industry the typical duplex filter is 10 micron nominal. This means that some particles in excess of 10 micron regularly pass through the element. Nowadays, filter manufacturers typically rate their elements at 10 micron, with a filtration ratio of 200 (beta 200). This means that 1 particle in 200 of a size greater than 10 micron will pass through the element. ANSI T.3.10.8.8. Standard was developed for testing and verification of effectiveness of the filter. Many OEM’s have determined that there is a direct relationship between the quality (level) of filtration on the lube oil system and the life of the bearings. The higher the quality of filtration, the longer the service life. ISO 4406 and NAS (National Aerospace Standards) 1638 define cleanliness of critical rotating machinery (NAS 1638 Table 4). Our process pump is critical to the reliability of the process, and therefore oil clarity should be NAS 1638 Class 5. API 614 cleanliness standard, which has been adopted by API 610, 8th Edition, allows for a defined number of particles to be caught at the screen of 250 microns maximum. Note that this relates to a NAS 1638 Class 11. The typical quality 10 micron beta 75 element will produce NAS 1638 Class 5 results. We recommend that customers require their lube oil systems be flushed at the point of manufacture to NAS 1638 Class 5 or finer with oil at 160ºF for a one hour duration. During testing the oil system is mechanically agitated. Have the manufacturer provide a test report from a certified particle counting device as evidence of clarity. Filters should be sized so that a clean element passes the entire pump discharge flow rate at 5 psi pressure drop across the elements, filter housing and transfer valve at design conditions. Therefore (for our test case), we should design the filter to pass the entire flow rate at viscosities between 70 SSU and 650 SSU if the minimum ambient temperature is 50ºF and the unit is not equipped with a heater. Given that API 610 requires the non by-pass filter and only a pressure differential indicator is provided to determine filter condition visually, The Pump Handbook Series

specific considerations should be made in the filter selection process. Collapse pressure of the element should exceed the pump’s relief settings. If the element gets very dirty, oil will pass over the pump relief, and the lube oil supply pressure will decrease until oil pressure at the header drops, causing an alarm condition. This alerts the operator to look at filter condition and transfer to the clean element. A bypass design or a low collapse pressure element will allow dirt to flow into the bearing when the element clogs and the bypass works or the element fails. Bypass design elements are equipped with a relief valve poppet. Any time the differential across the element exceeds the bypass valve setting, the bypass opens and allows unfiltered oil downstream. A better solution is the addition of a pressure differential switch across the filter housing to alarm at a dirty filter condition. Presetting the alarm at 25 to 30 psid will give the operator plenty of time to schedule the element change. A filter element that is properly sized will be contaminated/clogged 60 to 70 percent of its capacity at a pressure drop of 30 psid. ASME code stamped filters should be considered based on size, pressure, volume and geographic locations where code vessels are mandated. In addition, filter transfer valves should be open in the neutral position. The inlet port of the transfer valve should be open to the inlet of both elements. The outlet of the transfer valve should be open to the outlets of both elements. This will guarantee continuous flow when transferring from one filter to another.

Maintenance Considerations The single component we expect to maintain frequently is the filter element. The old saying “cleanliness is next to godliness” certainly applies here. Replacing the element regularly increases bearing life. In addition, consider locating the filter at the outside edge of the skid for ease of maintenance. Equip the filter with

individual canister vents and drain valves, and supply a balance/prefill line and appropriately sized orifice. For the above, allow the serviced element to be prefilled and vented of air prior to placing it in service. The orifice is used to restrict flow into the serviced element canister to prevent an unwanted loss of system pressure. It is also used to balance the pressure on both sides of the transfer valve to facilitate a valve shift. Drain valves allow the operator to drain hot oil from canisters prior to removing it from the system. The operator can also verify that the leakage rate of the transfer valve is acceptable prior to opening the housing. Excessive leakage in the transfer valve will cause a system shutdown. Drain valves are useful for oil sampling, particle counting and prevention of oil spills. In next month’s conclusion to this article, we’ll examine the role of heat exchangers, instrument isolation valves and root valves, lube oil piping, and instrumentation in a properly designed oil lube system. We will

also show typical API 610 schematics and detail the requirements for obtaining a proposal.

Part 2 Oil lubrication systems are life insurance policies for processes. If a lube system should fail to perform, the results are bearing failures in pumps, gears or drivers, and ultimately a shutdown of the process. Conservative design and proper specifications minimize the possibility for failure and maximize the potential for long, trouble-free service life of rotating equipment. Understanding a system’s needs, optimizing component selections, allowing for operational maintenance and proper application of the system are the keys to reliability. The conclusion of our two-part article continues with a review of lubrication system components.

Viscosity ssu

Theoretical: 3000 rpm

Maintenance Considerations

API 610, 8th Edition requirements are:

• Removable bundles add ability

PUMP MODEL B

.0065 g/rev Flow Rate - gpm

Pressure - psi 75 100

150

Viscosity ssu

13.3 15.1 15.9 16.6 18.0 18.3 18.9 19.0

12.8 14.7 15.6 16.3 17.9 18.2 18.9 19.0

33 65 100 150 650 1000 5000 10000

50

16.7 17.5 17.8 18.1 18.8 18.9 19.2 19.2

15.6 16.7 17.2 17.6 18.5 18.7 19.1 19.2

Viscosity ssu

.25

BHP - hp Pressure - psi 50 75 100

125

150

33 65 100 150 650 1000 5000 10000

.6 .6 .6 .6 1.0 1.2 2.9 4.3

.9 .9 .9 .9 1.3 1.5 3.2 4.5

1.7 1.7 1.7 1.7 2.1 2.4 4.0 5.4

2.0 2.0 2.0 2.0 2.4 2.7 4.3 5.7

14.7 16.0 16.7 17.2 18.3 18.5 19.0 19.1

1.1 1.1 1.1 1.1 1.6 1.8 3.4 4.8

14.0 15.5 16.3 16.9 18.2 18.4 19.0 19.1

1.4 1.4 1.4 1.4 1.9 2.1 3.7 5.1

Speed:

125

25

33 65 100 150 650 1000 5000 10000

• Velocity in tubes: 5 ft/sec minimum • Fouling factor tubes: 0.002 hr x ft2 x ºF/BTU • Fouling factor shell: 0.002 hr x ft2 x ºF/BTU • Maximum oil temperature out of cooler: 120ºF • Maximum water temperature rise: 10ºF • Maximum pressure drop: 5 psi • Capacity: Maximum heat load of pump, gear and driver plus the horsepower of the lube oil pump driver. The entire lube oil system input horsepower goes to heat.

Heat Exchangers

PUMP MODEL A Speed:

• Type: Shell and tube • Material: Tubes inhibited admiralty Shell steel (pressure retaining) • Orientation: Cooling water is on the tube side. Other design considerations:

Table 1. Screw pump discharge capacity and input power chart (0.0065 g/rev) (pump model A)

Theoretical: 3000 rpm

.0086 g/rev Flow Rate - gpm

25

50

75

22.6 23.5 23.9 24.3 25.1 25.2 25.5 25.6

21.2 22.5 23.2 23.6 24.8 25.0 25.4 25.5

20.2 21.8 22.6 23.2 24.5 24.8 25.3 25.5

Viscosity ssu 33 65 100 150 650 1000 5000 10000

Pressure - psi 100 125 150 18.6 20.6 21.6 22.4 24.2 24.5 25.2 25.4

17.9 20.1 21.2 22.1 24.0 24.4 25.2 25.3

.25

BHP - hp Pressure - psi 50 75 100

125

150

.9 .9 .9 .9 1.7 2.1 5.0 7.4

1.3 1.3 1.3 1.3 2.1 2.5 5.4 7.8

2.4 2.4 2.4 2.4 3.2 3.6 6.5 8.9

2.8 2.8 2.8 2.8 3.6 4.0 6.9 9.3

1.7 1.7 1.7 1.7 2.4 2.8 5.7 8.2

19.3 21.2 22.1 22.8 24.3 24.6 25.3 25.4

2.0 2.0 2.0 2.0 2.8 3.2 6.1 8.6

Table 2. Screw pump discharge capacity and input power chart (0.0086 g/rev) (pump model B)

The Pump Handbook Series

253

to change bundle without removing shell. • Consider adding vents and drains to facilitate maintenance. • Duplex heat exchanger arrangements are necessary on unspared equipment. In the case of duplexed heat exchangers, a balance valve, vents and drain valves on the shell, as well as vent and drain valves on the cooling water chamber, should be added to facilitate the change-out of heat exchangers during operation. Air-to-Oil Coolers • Air-to-oil coolers can be considered where water is not available. *Note: Cooler tube and headers for the economical air-to-oil coolers are usually manufactured from aluminum or copper. If carbon steel or stainless steel tubes are used, the price of the cooler increases as well as the size. TEMPERATURE CONTROL VALVE

M2

Figure 8. Air-to-oil cooler with mixing action temperature control valve

• For installations where ambient conditions fall below 60ºF, we recommend the addition of a temperature control valve. This valve should be a mixing type rather than the bypass type (Figure 1). A mixing type temperature control valve mingles hot and cold oil at its outlet. This allows a continuous flow through the cooler, keeping the oil warm and available. The lube oil supply temperature rises at a steady rate, thus avoiding thermal shock. The by-pass type temperature control valve opens

254

at a set point. This can cause a momentary loss of supply oil as the cold oil must be pushed out of the cooler. This cold oil can cause a thermal shock in the system. Figure 8 shows a thermostatic bypass valve applied in a mixing application mode with an air- to-oil heat exchanger.

Relief Valves API 610, 8th Edition requirements are: • Valve shall have a carbon steel body with stainless steel trim. • Valves shall be located down stream of each pump discharge. (See API 610 Lube Oil schematic for location.) The types of valves typically used: Direct acting – Full flow valves are direct acting. A spring retains a piston in a bore. Oil working against the piston overcomes the spring and pushes the piston back, allowing oil to pass to the tank port. As the flow rate increases, the pressure at the inlet increases, i.e., the pressure drop across the valve increases. Pilot operated – In this case oil works against the area of a small control poppet. When oil pressure overcomes the spring, oil passes to the tank. This allows a pressure drop to occur at the main/slave spool, which backs off. One advantage of pilot operated valves is that within the suggested flow range, an increase in flow does not have an effect on the set pressure. Other Considerations • Relief valve oil return lines should always terminate below the oil level in the reservoir. • Do not install isolation valves either immediately upstream or downstream of the relief valve unless your company follows a lockout policy or you specify car seal open valves. • Conduct site tests to verify that valves do not leak.

Pressure Control Valves API 610, 8th Edition requirements are: • Material: Cast steel bodies with The Pump Handbook Series

stainless steel trim. • Location: Downstream of filter and immediately upstream of lube oil supply connection. • Size: Must handle the excess flow of the auxiliary pump and the entire flow rate of the main pump simultaneously at an 80% open condition. Design Considerations • Flow rate – The pressure control valve must be sized so that it passes excessive main lube oil pump flow when the valve is 15 - 20% open. The valve must also pass the entire flow of the auxiliary lube oil pump and the excess flow of the main lube oil pump at an 80% open condition. This situation occurs any time both pumps are operated simultaneously. • System designers are required to determine if an integral or external pilot is required. Valves at 1 inch port size and below work well with integral pilots. Valves 1-1/2 inch and above typically are far more responsive with external pilots. The reaction time of the valve must be fast enough to maintain a pressure between set point and 10% accumulation. Some externally piloted valves will reduce system pressure by overstroking when the second pump starts. • External oil pilots vs. pneumatic pilots: 1. The key factor to be considered is time of actuation. In small API 610 packages, response time is acceptable when using oil pilots. Oil pilots are less expensive and are easily maintained. • External pilot valves should have needle valves to isolate the pilot, and the actuator of the control valve should be drained to the reservoir below the oil level (Figure 9). This is a requirement of API 614 and is especially important in cold applications. The orifice allows warm oil to maintain temperature at the valve head. Since this line is now a dynamic or flowing line, the operation is much smoother and the actuator level is repeatable. Maintenance Considerations • If continuous duty is required or the process pump is unspared, the

design should include isolation bypass and drain valves as shown in Figure 10. * Note that for reducing valves, two isolation valves are required.

Instrument Isolation Valves/Root Valves API 610, 8th Edition requirements are: • Type: Block and bleed valves, valves can be combination style. • Materials: Carbon steel bodies with stainless steel trim. • Size: 1/2” NPT minimum. • Connection: Each pressure instrument must have its own pressure tap.

LUBE OIL SUPPLY

FO TOP OF RESERVOIR

Figure 9. Pressure control valve with external pilot vented to reservoir FO

PRESSURE REDUCING VALVE

LUBE OIL SUPPLY

PRESSURE CONTROL VALVE FO

Figure 10. Pressure control and pressure reducing valves with isolation and

PI

Design Considerations • Low temperature applications Suggested approach is per Figure 11. Allowing a continuous flow through the instrument tubing will result in a more accurate and repeatable set point. • Install a test port as shown in Figure 11b so the instrument can be checked and calibrated during operation. • Do not install an isolation valve on PSLL switch. If isolated, the process pump will be unprotected from alarms due to loss of lube oil pressure. Maintenance Considerations • Always install a union at the electrical connection of the instru-

BLEED TYPICAL 2-VALVE MANIFOLD

BLOCK

a PSLA

TEST PORT

1

BLEED

BLOCK

b PSLA

TEST PORT

1

LUBE OIL SUPPLY

BLEED

LUBE OIL SUPPLY

PSLL 1

FO

BLOCK

ment so it can be removed without cutting the conduit. • Always install a pipe union or tube union between the isolation valve and the instrument so the switch can be removed without having to shut the system down. • Consider the use of stainless steel for gauge boards; less galvanic action occurs. • The data reflected in Figure 11c are optimum for low ambient applications.

Lube Oil Piping API 610, 8th Edition requirements are: • Type of construction, fittings: socketweld upstream of filter, buttweld downstream of filter. • Flanges: 150# RF, socketweld or slip-on upstream of filter, 150# RF, slip-on or buttweld downstream of filter. • Bolts and studs: ANSI-A193-B7 studs or hex head bolts and ANSIA194-2H nuts. • Bending is preferred over welded joints. • Other requirements: ANSI B31.3 Piping Code with 5% radiography for buttwelds. • Materials per Table 3 (Table 3-4 of API 610, 8th Edition). • Flush per API 614. • Hydro per API 614. • Long shank plugs only. • Bushings are not allowed. Design and Maintenance Considerations • Always use pipe support clamps of the same material as the piping, otherwise isolation pads and nonmetallic dampers should be considered. • Always install flanges so they are located immediately down stream of the valve and are in line with the valve body. This allows a flange to be tightened in 1/4 inch turn increments if a threaded valve is used. • Provide vent and drain valves for maintenance purposes in piping.

ROOT VALVE

c

LUBE OIL SUPPLY

Figure 11. Instrument Valving The Pump Handbook Series

255

System

Auxiliary Process Fluid Flammable/ Nonflammable/ Hazardous Nonhazardous

Steam

Cooling Water >500kPa (>75psig)

Standard

Optional

≤ 1NPS

≥ 1 1/2 NPS

Seamless

Seamless

Seamless

—

ASTM A53 Type F Schedule 40 galvanized to ASTM A153

—

ASTM A312 Type 316L Stainless steelb

—

ASTM A269 seamless Type 316 c Stainless steel

—

Carbon steel Class 800

Carbon steel Class 800 flanged

Bolted bonnet and gland

Bolted bonnet and gland

Type 316L Stainless steel

Type 316 Stainless steel

Manufacturer’s standard

—

Pipe

Seamless

Tubing

ASTM A269 seamless Type 316 c Stainless steel

ASTM A269 seamless Type 316 c Stainless steel

ASTM A269 seamless Type 316 c Stainless steel

ASTM A269 seamless Type 316 c Stainless steel

ASTM A269 seamless Type 316 c Stainless steel

All valves

Class 800

Class 800

Class 800

Class 800

Class 200 bronze

Gate and globe valve

Bolted bonnet and gland

Bolted bonnet and gland

Bolted bonnet and gland

Bolted bonnet and gland

—

Pipe fittings and unions

Forged Class 3000

Forged Class 3000

Forged Class 3000

Forged Class 3000

ASTM A338 and A197 Class 150 malleable iron galvanized to ASTM A153 Manufacturer’s standard

Tube fittings

a

Lubricating Oils

≤500kPa (≤ 75 psig)

a

a

a

Manufacturer’s Manufacturer’s Manufacturer’s Manufacturer’s standard standard standard standard

—

—

Fabricated joints ≤ 1 NPS

Threaded

Socket welded

Threaded

Socket welded

Threaded

Threaded

—

Fabricated joints ≥ 1 1/2 NPS

—

—

—

—

Purchaser to specify

—

Weldedd (see 3.5.5.5)

Gaskets

—

Type 316 Stainless steel spiral wound

—

Type 316 Stainless steel spiral wound

—

—

—

Type 316 Stainless steel spiral wound

Flange bolting

—

ASTM A193 Grade B7 ASTM A194 Grade 2H

—

ASTM A193 Grade B7 ASTM A194 Grade 2H

—

—

—

ASTM A193 Grade B7 ASTM A194 Grade 2H

Note: Carbon steel piping shall conform to ASTM A106, Grade B; ASTM A524; or API Specification 5L, Grade A or B. Carbon steel fittings, valves, and flanged components shall conform to ASTM A105 and A181. Stainless steel piping shall conform to ASTM A312, Type 316L. (See Appendix A for corresponding international materials.) a Schedule 80 for diameters from 1/2 NPS to 1 1/2 NPS. Schedule 40 for diameters 2 NPS and larger. b Schedule 40 for a diameter of 1 1/2 NPS. Schedule 10 for diameters of 2 NPS and larger. c Acceptable tubing sizes are as follows (refer to ISO 4200): 12.7 mm diameter x 1.66 mm wall (1/2 in. diameter x 0.065 in. wall) 19 mm diameter x 2.6 mm wall (3/4 in. diameter x 0.095 in. wall ) 25 mm diameter x 2.9 mm wall (1 in. diameter x 0.109 in wall). d Carbon steel slip-on flanges are permitted.

Table 3. Requirements for piping materials (American Petroleum Institute)

Instruments API 610, 8th Edition requirements are: • Devices located per API 610 Lube Oil schematic. • Pressure connections: 1/2” NPT. • Wetted materials: 316 stainless steel. • Pressure Dials: 4 1/2 inch dial with safety back. • Thermometers: 5 inch dial with a thermowell. Design Considerations • Low ambient conditions as shown in Figure 11 (c). • Applications with a test port are shown in Figures 11b & c. • When selecting pressure tem-

256

perature switches, always verify the switch dead band. Be sure you select your switches so the dead band of one switch does not overlap a set point of another switch function. If permissive pressure switch is set at 30 psi increasing and the alarm for low oil pressure switch is set at 25 psi decreasing, be sure the permissive switch resets on decreasing before reaching 25 psi. Failure to look at dead band results in system instrumentation indicating both an OK pressure and low pressure signal simultaneously. • Pump suction line sizes should be designed at 4 ft/sec maximum. Return lines should slope towards reservoirs and run half full as a maximum. Pressure lines should be The Pump Handbook Series

designed for 10 ft/sec maximum velocities. • Instrument sense tubing is seamless, .065 wall 1/2 inch O.D. • 3 inch pipe size is required as a minimum when thermowells are required in the piping. Function of pressure switches during start-up, alarm – lube oil pump transfer and trip of process pump • The lube oil system in the API 610 schematic includes three pressure switches. The instrument legends for these switches are PSLA, PSLB and PSLL. The functions of these switches are as follows: 1. PSLA: Low pressure switch – starts auxiliary lube oil pump. Typ-

ically set at 15 psi decreasing.

40 PDSH

2. PSLB: Low pressure switch – alarms low lube oil pressure. Typically set at 15 psi decreasing. 3. PSLL: Low pressure switch – trips process equipment. Typically set at 12 psi decreasing. The lube oil supply header pressure is controlled by the pressure control valve, which is normally set at 20 psi. • Typical operating sequence for start-up/alarm pump transfer/trip function are: Start-up – In the shelf or low pressure state all switch elements are open. The customer starts the auxiliary lube oil pump, and as pressure rises above 12 psi, switch PSLL closes. Pressure continues to rise to 15 psi, where switch PSLA and PSLB close. This indicates lube oil supply pressure is adequate, and from a lube oil supply standpoint it is okay to start the process pump. Pressure continues to rise to 20 psi, which is the set point for the pressure control valve. If all other process permissives have been satisfied, the customer starts the process pump. As the process pump comes up to speed, a flow rate is discharged from the main lube oil pump into the lube oil system. At this point, both the main lube oil pump and the auxiliary lube oil pump are operating simultaneously. The pressure control valve must open to pass the excess flow rate of the main pump and the total flow rate of the auxiliary pump back to the reservoir. System pressure with both pumps running may equal the original PCV set point, or may increase a maximum of 10%. We refer to this 10% increase as 10% accumulation. The customer verifies by means of a pressure gauge and pressure switch signal that the lube oil pressure is adequate, and he shuts down the auxiliary lube oil pump. • Low pressure alarm and lube oil pump transfer – Should lube oil pressure drop to 15 psi, the PSLA switch contact opens, and the auxiliary lube oil pump starts. The PSLB switch contact opens simultaneously at 15

PDI

17

6

36 F I L T E R

6 21

FO-1

12

.5"T

11 15 29

TSH

PCV-1 SET @ 100 PSIG

16

2.0" 3

EI 13 (E) 2.0"

1.5"

LPI TO BACKWELD FLANGES TO PCV VALVE

31 A TCV-1 SET @ 100°F B C 1.5"

1.5"

SET @ 10 PSIG

36 .5"T

1.5" 18

PI-2

7

6

24

PSL

7

PSLL

6

6 .5"T

.5"T

.5"T1.5"

LUBE OIL SUPPLY 30 GPM @ 10 PSIG

36

36 37

F I L T E R

FO-2

36 10

COOLING MEDIUM SUPPLY 34.5 GPM @ 80∞F

6

4 COOLING MEDIUM RETURN 1.5"

35

PSH

6

1.5" PSV-1 SET @ 150 PSI 8

JUNCTION BOX

1.5"

1.5" 9

9 PSV-2 SEET @ 150 PSI 8 1.5"

1.5"

1.5" 36 .5"T .5"T PREFILL VENT

33 34 25

1.5"

MANWAY

LSL

FILL CAP

PI-1

1.5" 1.5"

LUBE OIL RETURN

14

V1 (V1) 1

32

TI-1 11

LI-1 BAFFLE

2

28 38

12

2.0"

41

2

2.0" 38

14

MAIN LUBE OIL PUMP MOUNTED & SUPPLIED PUMP MANUFACTURER 2.0"

M1

5

3 AUX. L.O. PUMP

Figure 12. API 610 schematic with typical options (American Petroleum Institute)

psi decreasing and sounds the low lube oil pressure alarm. Normally the addition of the auxiliary lube oil pump flow to the circuit raises the lube oil supply pressure. Pressure switches PSLA and PSLB will close. At this point the electrical control circuit should not shut down the auxiliary lube oil pump, as pressure will again drop off to the low pressure alarm setting. The operator should be forced to acknowledge the low lube oil pressure alarm and determine the root cause for it. If the operator cannot determine the cause, the auxiliary lube oil pump is left operating until the spare process pump can be started. Items as simple as a dirty filter can result in a low supply header pressure. • Trip – Should the lube oil supply pressure drop to 12 psi during operation, the pressure switch PSLL will open. This is the trip signal for low lube oil supply pressure, and the process pump should be shutdown immediately. During shutdown, the auxiliary lube oil pump continues to run.

Typical Schematics The API 610 schematic in Figure 12 shows a system with some stanThe Pump Handbook Series

dard options as listed below. • Balance valves for filters • Temperature control valve • Reservoir heater • Level switch • Pressure differential switch • Lube oil pressure higher than cooling water pressure • Oil over temperature switch • Prefill line for main lube oil pump

Requirements for Obtaining a Proposal The following information is required as a minimum to secure an accurate proposal. • Total oil flow rate required – verify flows to each piece of equipment. • Pressure required at each piece of equipment • Heat load for each piece of equipment • Location of skid – including site elevation, indoors or outdoors, ambient temperatures • Electrical area classification: class, group and division • API requirements/options • Voltage/frequency/phase for motors, heaters and instrumentation • Cooling water supply and return temperatures, maximum al-

257

lowed cooling water pressure drop, maximum cooling water pressure. • Sound level limitations if any • Type of piping and fitting – API 610 or others • Code stamp requirement for cooler and or filter • Purchasing specifications and data sheet (As a guide, complete the purchaser “General Purpose Lube Oil System Data Sheet” in Figure13). The more detail provided by the purchaser, the more accurate the proposal. “A picture is worth a thousand words.” The hydraulic flow schematic is a picture that includes the quantity and orientation of components. If a component is not shown on the schematic, then it will not be in the system. Schematics, data sheets and customer specifications are all necessary information to formulate a design for the lube oil system.

Conclusion The North American Reliability

258

Council (NARC) determined that turbine bearing and lube system failures were the leading cause of forced outages in the turbine related failures. Your lube oil system is the life insurance policy for your process. If you accurately define the scope of your supply, specify your requirements and complete relevant data sheets, you will have an optimized lube oil system design. Safe and reliable pump operation with shafts riding on a properly conditioned oil film will yield maximized revenues to the process pump user.■

REFERENCES 1. API 610, 8th Edition, American Petroleum Institute 2. API 614, 3rd Edition, American Petroleum Institute Vinod P. Patel is a Senior Principal Machinery Engineer for the M. W. Kellogg Company in Houston. He is responsible for the preparation and auditing of specifications, equipment evaluation, engineering coordination

The Pump Handbook Series

and testing and installation/startup of rotating and special equipment. Donald G. Coppins is President of Lube-Power, Inc. His company designs and manufactures oil lubrication systems, seal oil systems, seal buffer systems, heat exchangers and hydraulic systems for the petrochemical, refining, power generating and other process industries. He has 29 years of experience in the oil hydraulic and lubrication systems business. Editor’s Note: This article has been reproduced with permission from the Turbomachinery Laboratory. Edited from Proceedings of the Fourteenth International Pump Users Symposium, Turbomachinery Laboratory, Texas A&M University, College Station, TX, pp. 129-139, Copyright 1997.■

Rev. No.

0

1

2

3

Job No.

Date

Client

Originator

Location

Reviewed

Unit

Approved

Item No.

Page of

Service Requisition No.

1 Design, manufacture, inspection and testing shall conform to specification: ❑ Manufacturer ❍ Purchaser/Manufacturer 2 Information to be completed by: ❍ Purchaser 3 Manufacturer: Model: Serial No.: No. Req’d: Operating and Construction Features 4 Reservoir: 5 Pumps ❑ Mfr. ❍ Electric Heater ❍ Steam Heater 6 ❍ Item No. Standby: ❑ Heater Mfr: 7 ❍ Pump Type: Main: Standby: ❍ Internal Coating: ❍ Oil Purifying Conn.: 8 ❍ Driver Type: Main: Discharge Press (psig) ❑ Working Cap. (gal.): ❑ Retention (gal.): 9 ❑ Rated Cap. (GPM): ❑ Lubricant Type/Initial Fill (gal.): / 10 ❑ RV Set point (psig): ❑ Horizontal ❑ Vertical Lines and Connections: 11 Casing Mount: ❍ Tubing For Instruments: ❍ Min Size: 12 ❍ Coupling Type/Mfr: ❍ Pipe: ❍ Butt Welded ❍ Socket Welded ❍ Threaded 13 ❍ Seal Type/Mfr. ❍ Flanged ❍ Customer Conns.: On skid edge. 14 Materials of Construction: ❍ Reservoir: Instrumentation: ❍ Electrical Area Class: 15 ❍ Pump Case: ❍ Element: Temperature Indicators: ❑ Mfr. 16 ❍ Filter: Housing: ❍ Tubes: ❍ Reservoir 17 ❍ Cooler: Shell: / ❍ Cooler In/Out ❍ Bearing Drain 18 ❍ Piping: Upstrm Filter: / Pressure Indicators: ❑ Mfr. 19 ❍ Valves: Body/rim ❍ Instr. Tubing ❍ Pump Disch. 20 ❍ Accumulator: ❑ Mfr. ❍ Filter Inlet ❍ At switches/transmitters 21 Filters: Item No. ❍ Filter Outlet ❍ Filter Diff. Press 22 ❍ Single ❍ Dual ❍ Code Design/Stamp Level Indicator: ❑ Mfr. 23 ❑ Design Pressure/Hydro (psig): Clean: Allowable: ❍ Reservoir 24 ❑ Diff. Press. (psi): Slight Flow Indicators: ❑ Mfr. 25 ❍ Filtration Particle Size (microns): ❑ Mfr.: ❍ Bearing Drain ❍ Relief Valve Outlet 26 Three Way Transfer Valves: ❍ Filter ❍ Cooler Switch/Transmitter Mfg. (Note 1): 27 ❍ Filter/Cooler ❑ Switches/Trans.: Alarm Shutdown Setpoint 28 Coolers: Item No. ❍ Type: 29 ❑ Mfr. ❍ Lo L.O. Press. psig 30 ❍ ASME Design/Stamp: ❍ Hi L.O. Temp. deg F ❑ Surface Area (ft2): ❍ Lo L.O. Level: 31 ❑ Duty (BTU/hr): ❑ Mfg.: 32 ❍ Temp. Cont.Vlv.Req’d ❍Hi Filt. DP: psi ❍ Dual ❍ TEMA C ❍ TEMA R 33 ❍ Single Turbine: ❑ Mfr: ❍ Item No.: 34 ❑ Design/Hydro Press. (psig): Shell: / Tubes: / ❍ Stm. Conds. In/Out: / (psig,deg F) Shell OD: 35 ❑ Dimensions (inches): ❍ Rating (HP): ❍ Speed (RPM) Tubes OD: BWG: 36 Motor: ❑ Mfr: ❑ Temp. In/out (F): / 37 ❑ Water Flow (gpm): ❍ Voltage/Phase/Hz: / / Speed: ❍ Tube Vel. (ft/sec): 38 ❍ Fouling Factor: ❑ Rating (hp): ❍ Service Factor: 39 ❍ Max delta P (psi): Inspect. & Testing: Required Witness Test Log ❑ Capacity (gal): 40 Accumulator: ❑ Mfr. ❍ Cleanliness ❑ ASME Code Design/Stamp: 41 ❍ Required: ❍ 1 Hour Test Run ❑ Design/Hydro Press. (psig): / 42 ❑ Type: ❍ Check Controls 43 Mounting: ❍ On equipment skid ❍ Separate baseplate ❍ Changeovers ft. ❑ Weight: lbs. ❍ 1&2 Pump Op. 44 ❑ Dimensions: 45 Notes: 46 1) Switches shall be SPDT and rated for the specified electrical 47 area classification and be de-energize to alarm and shutdown. 48 2) Instruments shall have isolation valves. Figure 13. General purpose lube oil system data sheet

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259

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R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

Injectable Packing Compounds - An Alternative Sealing Technology Aided by an innovative stabilizing device, injectables offer real value at Weyerhaeuser’s container board facility. By Dennis L. Weehunt, Weyerhaeuser Corporation igh volume process pumping technology has changed relatively little over the past century. Only in recent years have environmental and economic pressures encouraged industries worldwide to modernize their thinking on fluid movement strategies. Technological advances have stimulated design improvements, resulting in better pumping efficiency and equipment reliability. Computers and computer-aided design software have empowered design engineers to optimize pump and drive motor designs. Laser alignment equipment has enabled maintenance personnel to reduce dynamic stresses on bearings and shafts, resulting in longer equipment life and reduced process downtime. There have been improvements in the field of sealing technology. However, gains have come slowly and at a price. While the average cost of conventional braided packing is $55 to $65 (in 1997 dollars) to seal a typical three-inch shaft process pump, this packing can be expected to last only for a short time before adjustments are needed or repacking becomes necessary. Using conventional braided packing, a typical three inch shaft process pump will consume anywhere from three to five gallons of flush water per minute to keep the shaft sleeve/stuffing box area cool and lubricated (Photo 1). This can add up to more than 1.5 million gallons of water per year for

H

260

Photo 1. A typical three inch shaft process pump using conventional braided packing consumes three to five gallons of flush water per minute to keep the shaft sleeve/stuffing box area cool and lubricated. Shown is a Worthington FRBH-142 pump.

each pump that is configured with this sealing method. Cartridge mechanical seals, slurry seals and split seals are a significant improvement over conventional braided packing. However, the cost for these styles of seals ranges from $600 to $1000 per inch of shaft diameter per seal, depending on the seal type. Although these seals can improve maintenance cycles by two to four years (depending on pump service demands), they still require some amount of flush water for seal face cooling and to wash solids away from the seal faces and out of the stuffing box. To avoid excessive water consumption in mechanical seals, seal cooling flush water sysThe Pump Handbook Series

tems are typically fitted with flow meters at $300 to $500 each. These must also use fresh clean water to remain functional. Injectable packing sealant compounds are a relatively new technology that shows promise and adds to the choices for shaft sealing strategies in many applications. These highly lubricated fiber-reinforced composites have been developed by several companies in various formulations. A.W. Chesterton, Kem-A-Trix and Tom-Pak are three companies that produce injectable packing sealant compounds. The materials are designed to be injected into the pump stuffing box area using the flush water port and a high-pressure hydraulic injection gun. Contained between conventional braided packing rings, two 1/8 inch washer rings, or a taper ring set and two gasket washer rings, these compounds cling to the surface of the shaft and the stuffing box, providing a tenacious leak barrier even with a rotating or reciprocating shaft. The injected material moves over itself molecularly under laminar flow conditions as the shaft moves, providing a sealing continuum in the stuffing box and ensuring leak-free operation without additional heat generation. This eliminates the need for lubricating and cooling flush water. Sealing properties are maintained at shaft surface speeds of 1400 - 2200 feet per minute. Initial results for all brands of

these compounds have been mixed, however. Some have performed better than others, but all have exhibited similar failures. When failures were analyzed, it was evident that the rings containing the material were not staying in position and were opening at the skive or butt cuts. Extrusion of the compound past the inboard ring into the process stream, and past the outboard ring onto the floor was a recurring problem. Once extrusion began to occur, leaks would develop from a lack of stuffing box integrity. It was important to find the root cause to be able to correct these failures. Experiences at our North Bend (Oregon) Weyerhaeuser container board facility revealed that process or equipment disturbances could also lead to failures with the injectables. Sudden pressure or flow fluctuations, cavitation, shaft deflection, vibrations and worn shafts were all found to contribute to extrusion failures. Once extrusion occurred, the retaining rings would lose their ability to hold any additional compound in place. Early expectations were that maintenance personnel would be able to reinject a leaking stuffing box Description North Effluent Pump** South Effluent Pump** Foreword Cleaners Pump West Pulper Pump Machine Chest Pump East Pulper Pump #5 White Water Pump North Effluent Pump South Effluent Pump North Effluent to Lagoon South Effluent to Lagoon W. Saveall Seal Pit Pump E. Saveall Seal Pit Pump S. Clarified Water Pump Sec. Cleaners Pump #1 West W.W. Chest Pump Couch Pit Pump Press Pit Pulper Pump Sec. Cleaners Rejects Pump Sec. Cleaners Pump

with added compound while the process was running to stop developing leaks, but reinjection would seldom correct the leak. Once an extrusion path was developed, it provided a way out for any additional material forced into the stuffing box, perpetuating the initial failure. Some companies that have developed these compounds have attempted to correct the extrusion problem by adapting different stuffing box end containment alternatives. Although the stuffing box ends are where extrusion ultimately takes place, the stability, spacing and conformability of these containment rings is essential. My success using injectables has been the Photo 2. WeeSeal Cage™ stablizes the result of using conforming types of injectable packing compound. times failures have been the direct end containment rings, such as low result of process fluctuations at disfriction braided packing, which are tant points from the equipment. An stabilized and held in place. We’ve excellent example of this is when a learned that stability, spacing and the major leak developed on a Worthingcontainment ring’s material properton FRBH-142 pump. Methodical ties are the magic bullets for effective troubleshooting revealed that a injectable packing sealing. process level detecting sensor in a In figuring out why the chest upstream from the pump was injectable packing method failed, we inoperative, and a low level in the learned that a dedicated approach to chest was causing pump cavitation. troubleshooting both the process and The level detecting sensor was the equipment was essential. SomeBrand & Size

Worthington 6 FRBH 142 Worthington 6 FRBH 142 Worthington 6 FRBH 142 Worthington 6 FRBH 142 Worhtington 6 FRBH 152 Worthington 6 FRBH 142 Worthington 6 FRBH 142 Worthington 6 FRBH 142 Worthington 6 FRBH 142 Worthington 6 FRBH 142 Worthington 6 FRBH 142 Worthington CNG 84 Worthington CNG 84 Worthington CNG 104 Worthington 6 FRBH 142 Worthington 6 FR 172 Worthington 6 FRBH 152 Worthington 6 FRBH 142 Worthington 2 FRBH 111 Worthington 6 FRBH 142

Imp 14.75” 14.75” 12.00” 14.25” 13.00” 14.1” 13.75” 14.75” 14.75” 14.75” 14.75”

13.75” 13.00” 13.75” 14.75”

Shaft Size 3.00” 3.00” 3.00” 3.00” 3.00” 3.00” 3.00” 3.00” 3.00” 3.00” 3.00” 2.125” 2.125” 2.125” 3.00” 3.375” 3.00” 3.00” 2.125” 3.00”

Area

RPM

Date Installed

Waste Water Waste Water Clippings Clippings Paper Mach. Clippings Paper Mach. Waste Water Wates Water Waste Water Waste Water Paper Mach. Paper Mach. Paper Mach. Paper Mach. Clippings Paper Mach. Paper Mach. Paper Mach. Paper Mach.

1200 1200 1200 1800 1800 1800 1800 1200 1200 1800 1800 1200 1200 1800 1200 1800 Variable Variable 1200 1200

1/1/95 1/1/95 10/11/95 10/11/95 11/1/95 12/8/95 1/15/96 2/1/96 2/1/96 2/2/96 2/2/96 2/26/96 2/27/96 3/14/96 4/1/96 4/10/96 4/13/96 4/1/96 3/4/97 3/4/96

** On 02-01-96 both North & South Effluent Pumps were changed out, because the wet end of the pumps were worn out. The injectable was still working and had required very little maintenance, but was changed with the new pump. Table 1. Pumps using injectables at Weyerhaeuser’s North Bend Plant

The Pump Handbook Series

261

cleared, the chest refilled and the cavitation was eliminated. Having a stabilizing device in place enabled successful reinjection of the sealant compound, and that stopped the pump from leaking. In this example, as in many others, it would have been easy to place blame for the failure on the injectable compound and miss the true cause. In pulp and paper plants, as well as other industrial shaft sealing applications in North America, the economical injectable sealant compounds are being used with a good degree of success. Our efforts at Weyerhaeuser have been aided by the use of a WeeSeal Cage™ stabilizing device in more than one third of our process pumps (Photo 2). With this component and the injectable compound in place for more than two years, our plant has a nearly

leak-free operation without the use of flush water (see Table 1 for a list of pumps and applications using injectables). Savings have been realized in a number of ways: reduced water consumption, lower effluent treatment loads, reduced product dilution, improved housekeeping, and improved safety from the elimination of slippery floors around pumping areas. Recent successes using injectable packing expand the options for shaft sealing. Maintenance personnel and facility managers – particularly in older plants – have an additional tool for lowering costs and improving process and plant efficiencies. Of course, not all applications are necessarily candidates for any one sealing technology. Some situations require more complex mechanical sealing systems. In

others conventional braided packing is the best answer. Many applications, however, will work extremely well with the economical injectable packing sealants. Most process applications in the pulp and paper industry are well-suited to these materials when they are used with a good quality stabilizing device. This adds to a facility maintenance manager’s arsenal of tools to help bring down costs and improve equipment reliability and efficiencies.■ Dennis L. Weehunt is a Weyerhaeuser pump rebuilding technician who has more than 15 years of handson experience with all types of shaft sealing technologies. He has spoken at the Oregon Pulp and Paper Workers Health and Safety Conference, sharing his expertise on pump rebuilding. He invented the WeeSeal Cage™.

PUMP SEALING SYSTEMS COST COMPARISON STABLIZED STANDARD PACKING THREE-INCH SHAFT An average pump will need to be repacked two times a year, and the packing sleeve

INJECTABLE PACKING THREE-INCH SHAFT Three packing rings =$40.98

MECHANICAL SEAL THREE-INCH SHAFT Cartridge Seal =$2,200.00

Injectable packing =$100.00

replaced once a year. Five packing ring sets =$70.00

Stablizing device =$85.00

One packing sleeve =$231.45 Four hours is required once a year to change the packing, install a new sleeve and check the alignment. 6 hours to install & check alignment

2 hours to install

6 hours to install & check alignment

TOTAL MATERIAL COST:

TOTAL MATERIAL COST:

TOTAL MATERIAL COST:

$301.45

$225.98

$2,200.00

PLUS 6 HOURS LABOR

PLUS 2 HOURS LABOR

PLUS 6 HOURS LABOR

Water used in a packed pump:

Water used in a mechanical seal:

3 gallons per minute

NO

180 gallons per hour

3 gallons per minute

FLUSH

180 gallons per hour

4,320 gallons per day

WATER

4,320 gallons per day

129,600 gallons per 30 days

NEEDED

129,600 gallons per 30 days

1,576,800 gallons used

1,576,800 gallons used

per pump, per year.

per pump per year.

Note: The amount of flush water used in a packed pump or one using a mechanical seal can be reduced with the use of a flow meter. Flow meters will have a tendency to plug if the flush water is hard or not clean.

Table 2. Cost comparison of pump sealing systems

262

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R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

Multistage Pumps in the CPI Improve reliability of your pumps with this review of selection, operation and maintenance factors. By John Lawler, Ingersoll Dresser Pump Company

he objective of this article is to review the features and operation of a multistage pump in chemical processing services and show how reliability can be improved. Although only one specific pump type will be discussed – a heavy duty, axially split case volutetype pump with opposed impeller construction – many features of this popular pump type are found in other multistage pumps used in the chemical industry. Figure 1 shows a cross section of a high pressure dual volute multistage pump typical of those produced by several global pump manufacturers.

T

4

2

7

1 6

8 5 8 6

The main features and advantages of this pump type are: • Heavy wall casing casting to allow for high working pressures • Axial split to simplify inspection and assembly • Dual volute and opposed impeller design to minimize bearing loads • Large diameter shaft and minimum bearing span to reduce deflection Typical hydraulic coverage includes heads to 7000 feet and flows to 20,000 gpm; however, the more popular pump sizes have flows less than 5000 gpm. Design working pressures to 4000 psig are achievable. Temperatures are generally limited to 400ºF for axially split pumps with unconfined gaskets, but higher ranges are possible depending on the specific application and experience

3 7

1. Heavy-walled, dual-volute casing to allow for high working pressures and nozzle pipeloads. Capnuts are seated on the top half of the axially split casing to simplify assembly and disassembly. Removal of the rotor is possible without disturbing the motor, piping or requiring rotor disassembly. 2. Dynamically balanced impellers of one-piece construction are shrunk on the shaft, keyed and axially secured by split rings. Impellers are mounted in an opposed configuration to balance axial thrust. Shaft is stepped by 5/1000 inch at each impeller fit to facilitate removal. 3. Renewable casing rings and impeller rings control interstage leakage. Optional laser hardened rings are available. 4. Channel rings including center bushing, are horizontally split to facili-

tate replacement and simplify dynamic balancing of rotor without dismantling. 5. Large-diameter shaft and minimum bearing span reduce deflection for longer bearing, mechanical seal and wear ring life. 6. Labyrinth flingers at each end of the bearing housing protect the lubrication against contamination. 7. Standard ring oil-lubricated bearings assure complete oil penetration into the bearings without foaming, for increased bearing life. Optional bearing arrangements and lubrication systems allow a “customized” fit to meet the requirements of the application. 8. Large stuffing boxed, integrally cast with the casing, can handle either packing or single, tandem or double mechanical seals.

Figure 1. Cross section of a horizontal, split case, dual volute, high pressure multistage pump with opposed impellers The Pump Handbook Series

263

Pump Selection An important factor for reliability is to select the proper pump, materials of construction and features at the proposal stage. It is essential to have a good understanding of the fluid(s) to be pumped and the hydraulic conditions to which the pump will be subjected. Several software packages are available from pump manufacturers that will automatically select models and features for specific service conditions. For hydraulic selections, API 610 specifies “preferred” and “allowable” operating regions. The criteria for defining these operating regions are based on vibration characteristics of the pump (Figure 2). Care must be taken not to overstate the conditions of service such that the pump is oversized for normal operating flows and heads. It is well known that actual operating conditions can change drastically from initially expected values once the equipment is installed in field. Operation in the allowable operating region, by definition, may be at higher levels of vibration caused to a great extent by hydraulic mismatch for the flow rate; ultimately higher vibration levels can directly impact pump reliability and seal life. It is important that the system curve be accurately defined in the proposal stage of pump selection. The multistage pump, especially a design with volute construction, offers hydraulic flexibility that can compensate for many hydraulic mismatch situations. A typical pump size

264

Head

may have different capaciAllowable operating ty impellers that can be region interchanged utilizing the Preferred operating same casing. Often the region volutes in the casing can also be modified to match a different set of impellers. The pump can also be destaged or stages can be added to meet new or changing head requireMaximum allowable ments, especially if such @ flow limits changes can be anticipated at the time of initial pump Typical vibration selection. Always consult characteristic Basic limit the pump manufacturer BEP prior to destaging or Flow adding stages to a multi- Figure 2. Relationship between flow and vibration stage pump to ensure that the proper stages are BYPASS CYLINDER SHAFT PROTECTION SLEEVE removed or added and that the proper additional parts are supplied. For destaging, a bypass cylinder is generally used to blank off the unused volute passages, and a sleeve is positioned over the shaft to protect it in case removed impellers are reinstalled later. Balancing devices (or features) may also need to be changed to control the residual axial thrust of the pump (Figure 3). API 610 also specifies material recommendations for pump components DEACTIVATED STAGE based primarily on the corrosiveness of the fluid (API Figure 3. Typical destage arrangement 610, Table H-1) . Carbon Most manufacturers can provide steel, 12% chrome, and various hardened wear surfaces through grades of stainless steel are coatings or special hardening described, and casing and impeller processes while maintaining a duccastings can also be produced in tile base material to resist stress corDuplex stainless steel, Alloy 20 and rosion cracking in service. Monel materials. Other special imNonmetallic wear parts with excelpeller materials are available to resist lent wear properties can also be prothe effects of cavitation. Close runvided. The most common ning clearance components, such as nonmetallic materials used for this impeller and casing wear rings and pump type are carbon-graphite (such bushings, control internal leakage as Graphalloy) and PEEK (polyand provide rotor stability. Material etheretherketone) (Ref.2). Not only selection for these critical wear comwill these materials improve operatponents must not only be compatible ing reliability, but because of their with the pumping fluid, but must self-lubricating properties, the interalso be able to resist galling and nal clearances in the pump may be accelerated wear for the pump to be reduced to improve pump efficiency, reliable. Vibration

of the manufacturer. The chemical processing industry (CPI) is more familiar with ANSI/ASME single stage overhung pumps than with high pressure multistage pumps, which are more popular in the hydrocarbon processing industry (HPI). In the HPI, most pumps are required to comply with American Petroleum Industry (API) standards; for centrifugal pumps, this means compliance with API Standard 610 (Ref. 1). Therefore, this standard can be utilized as a guideline or starting point when selecting multistage pumps, and it will be referred to throughout this article.

The Pump Handbook Series

Preventive Maintenance

Item Instructions

Frequency

Suction Strainer (When Used)

Check pressure differential between the gauges located on each side of the strainer.

Daily

Pump Suction and Discharge Flow Rates

Check suction and discharge pressure gauges for proper pump operation.

Daily

Mechanical Seal

Inspect Visually.

Daily

Instrumentation

Check all related pressure gauges, temperature detectors, etc. to detect any abnormalities.

Daily

Shaft Rotation (Down Periods Only)

During extended down periods rotate the shaft by hand 1-1/4 times to ensure bearing lubrication and to prevent shaft binding.

Weekly

Auxiliary Piping

Check for leakage around connections, etc.

Weekly

Shaft/Casing Vibration

Review all vibration data for any abnormalities and/or sudden changes in levels.

Weekly

Bolting Tightness

Check all external bolting for proper tightness.

Monthly

Cleanliness

General clean-up soiled areas.

Quarterly

Oil System

Refer to Lubrication System section of instruction manual.

Periodically

Figure 4. Preventive maintenance inspections

which equates to lower power requirements. It is important to realize, however, that most nonmetallic wear components have a higher wear rate when liquids are not clean, and they may have lower operating temperature limits than corresponding metallic wear materials. These factors must be considered when materials selections are made. Pump manufacturers can assist in this selection process. Heavy duty multistage pumps usually have one of three different bearing combinations: rolling element radial and thrust (which consists of a rolling element – or “ball”– bearing at each end of the rotor to accept both radial and thrust loads); hydrodynamic radial and rolling element thrust (which consists of a sleeve-type journal bearing at each end of the rotor and a ball bearing mounted in conjunction with the sleeve bearing on one end arranged to carry axial thrust loads only); and hydrodynamic radial and thrust. Hydrodynamic thrust bearings consist of multiple tilting pads designed to carry thrust loads in both directions. Rolling element bearings, and hydrodynamic bearings used in con-

junction with rolling element bearings, are usually ring-oil lubricated, whereas hydrodynamic radial and thrust bearing sets require a pressurized – or forced feed – lubrication system. Lubricant cooling may also be required, depending on the pump operating temperature and conditions. Table 2-7 in API 610 offers a good guideline for bearing selection based on bearing life, energy density and rolling element speed. Many manufacturers can provide mechanical shaft seals that meet the requirements of API 610 or API 682 (Ref. 3). Single and dual cartridge mechanical seals are available in either pusher or non-pusher (bellows) designs. For high pumping temperatures (temperatures in excess of the allowable operating temperature for the elastomers used in most pusher seals), metal bellows seals are also available. Selection and implementation of the right seal flush piping is essential to successful operation of the mechanical seal. It is critical that flush piping be properly sized to ensure sufficient flush flow to the seals to prevent temperature and particulate buildup. Normally the pump manufacturer will provide the piping plan(s), but there must be communi-

The Pump Handbook Series

cation among the manufacturer, seal vendor and the ultimate user to ensure that the sealing system is selected and designed properly for the intended service. A lower cost seal or seal system usually results in higher maintenance costs and reduced pump reliability and seal life. One major cause of field vibration in pumps is the pulsation that occurs when the impeller vanes pass the stationary casing volute tips (usually referred to as the volute cutwaters). This pulsation can be amplified in a multistage pump if the impellers are mounted such that the vanes of several impellers pass cutwaters simultaneously. The purchaser or specifier of multistage pumps should always insist that the impellers are staggered (or “clocked”) on the rotor and that there is sufficient clearance between the rotating impeller vanes and stationary casing vanes so pulsations at vane passing frequency are not amplified. API 610 offers guidelines for vane tip clearance. If required, the impeller and casing vanes can be cut on a bias to reduce the magnitude of vane pass pressure pulsations. Nondestructive testing can be

265

performed on the major pump components to ensure that they meet intended quality levels. All pressurecontaining (pressure boundary) components are routinely hydrotested for leakage at a pressure at least 150% higher than design pressure. Additional testing may be considered for pumps in critical or hazardous services. As a final assurance that the pump will meet all the intended performance requirements, it can be performance tested at the manufacturer’s plant.

Installation Proper installation is critical for the successful operation of any pump. The pump is provided with an installation manual, written specifically for that pump type, which should contain information on storage, transporting, handling and installation of the equipment. An important additional reference document for installation and installation design is API Recommended Practice 686 (4), which provides proven procedures, practices and checklists for successful machinery installation and precommissioning. A typical installation checklist may look like this: 1. Level the baseplate 2. Do preliminary alignment 3. Grout the baseplate 4. Align the pump and driver shafts (coupling alignment) 5. Install the piping 6. Check coupling alignment 7. Dowel the pump and driver The purpose of grouting is to make the pump baseplate and the foundation monolithic. Epoxy grout is preferred by many machinery manufacturers because it has good bonding and low shrinkage characteristics. The underside of the baseplate should be designed to lock into the grout positively, and the surfaces of the baseplate in contact with the grout must be prepared to promote good grout adhesion. The matching foundation surface also needs to be prepared for adhesion. Improper grouting techniques can result in repair cost and downtime that could greatly exceed the time and money

266

spent on the initial pump base installation. The initial cold pump-to-driver alignment may have to include an offset to adjust for anticipated differential thermal movement of the pump and driver as they are operated. After the pump and driver have been in operation long enough to reach equilibrium operating temperature, a hot alignment check should be performed to ensure that the pump and driver shafts are within allowable parallel and offset limits specified in the installation manual. Adjustments may be required by adding or removing shims under the feet of the driver. Adjustments should never be made to the mounting of the pump; the pump should be mounted directly to its machined mounting pads, using no shims. Only the driver should have shims under its mounting feet to facilitate alignment with the pump; the pump should always remain firmly bolted in place. On hot pump applications (pumping temperatures > 200ºF), multistage pumps are usually doweled at the coupling feet (drive end) only, and the outboard (non-drive end) feet should be free to move in an axial direction away from the coupling. Guide blocks are normally provided to keep the pump movement (thermal growth) in a precise axial direction only. Because of close internal running clearances, rotors on multistage pumps should be rotated with great care when they are not filled with liquid (pumpage). It is realized that during initial alignment the rotors must be turned, but this movement should be carefully performed and kept to a minimum. Manufacturers may apply an anti-seize compound to the wear rings to minimize damage during slow turning the shaft. One of the major causes of pump problems during operation is pumpdriver misalignment. This condition can be aggravated by a non-level installation, excessive piping loads or simply shaft misalignment beyond the ability of the coupling to compensate. Misalignment could also happen because thermal growth or piping loads were inaccurately estiThe Pump Handbook Series

mated, a hot alignment check was not performed, or events have caused relative pump-driver movement after the hot alignment check was performed and/or alignment corrections made. During pump-driver alignment, every effort should be made to achieve near-perfect alignment. The tolerances given in the instruction manual regarding shaft alignment – or misalignment – should be reserved for alignment inaccuracies that will naturally occur as the equipment starts up and progresses to full operating conditions. A poor suction and discharge piping arrangement can adversely affect both the mechanical and the hydraulic performance of the pump. Most manufacturers offer guidelines for the installation of piping connected directly to the pump. Unfortunately, good piping practices are often sacrificed to meet the physical limitations of a site. Manufacturers will gladly review installation piping drawings and offer expert recommendations to correct poor piping arrangements before installation and before piping-related operational deficiencies can occur. The pump-baseplate assembly can accept certain loads transmitted from the piping to the pump suction and discharge nozzles. Manufacturers provide users with allowable nozzle loads for the equipment in either the general arrangement notes or pump instruction manuals. These allowable loads are the basis for the pump and baseplate design and represent maximum values for satisfactory, reliable service. (API 610 also provides recommended maximum allowable nozzle load values in Table 2-1.) For hot pumping applications, thermal movements of the nozzles must also be considered when calculating nozzle loading. On multistage pumps with large numbers of stages, the nozzles may be located far from the doweling at the coupling end of the pump, and thermal movement can therefore be substantial. The manufacturer can provide the calculated thermal movement of the nozzles. It is important to realize that allowable nozzle loads are provided

to guide the piping designer in determining how the suction and discharge piping needs to be anchored and restrained in order to eliminate damaging loads on the pump during operation. For optimum pump operation, nozzle loads should be as near zero as possible at operating temperature.

Operation The manufacturer should provide a start-up procedure for the pump. A pre-operational checklist may look like this: 1. Suction valve is open 2. Minimum flow line is open 3. Pump is filled with liquid and properly vented 4. There is adequate liquid supply 5. Bearing housings have the proper oil level 6. Rotor turns freely 7. Pump is adequately warmed With a pressurized lubricating system, the starting procedure should include the auxiliary lube oil pump starting first and a permissive start signal given to the main driver when adequate lube oil pressure is obtained. The oil system should remain running when the pump is not in operation to assure that the bearings remain free from dirt and rust. For a ring-oil system, constant level oilers are usually supplied to maintain the proper oil level in the bearing housings. For a dirty environment, bearing housing seals and a pressurized oil mist system can be supplied to provide a positive pressure in the housings that helps keep contaminants out. For multistage pumps in hot applications, preheating is recommended prior to starting. Warm liquid should be circulated through the entire pump. To ensure uniform warming, the liquid should flow continuously from the suction to the discharge nozzle or vice-versa. Warming through the vent or drain lines most likely will not adequately warm the pump and may result in casing and/or rotor distortion. The outside casing skin temperature should be within a certain range of the pumped

fluid (usually 100ºF), and the rate of warming should not exceed the recommended rate (usually this is approximately 100 degrees per hour). Since the thickness of the casing casting may vary, the casing warming rate may vary at different sections. Checks in several areas of the casing should be made to assure that the temperature during the warm-up process is uniform. Pump manufacturers can size and supply warm-up orifices for each particular pumping application. To confine the heat of the fluid to the inside of the pump during operation, and for safety reasons, thermal blankets are available. To prevent debris and contaminants in the suction piping system from entering the pump upon initial start-up, a suction strainer is recommended. A cone strainer with 1/8 inch diameter holes (perforated plate) is usually preferred. The open area of the strainer should be at least three times greater than the area of the pump suction to minimize restriction. Typically, for new startup where the flow rates are low (below 100 gpm), a 100 mesh screen can also be used. For final operation, after the residual dirt has been removed from the piping system, a 20 mesh or larger screen is be recommended, although for relatively clean processes and systems, it may be desirable to remove all strainers after start-up cleaning. Across the strainer there will be a pressure drop that is proportional to the amount of dirt in the strainer. This pressure drop relates directly to a reduction in the suction pressure – or net positive suction head available (NPSHA) to the pump – and it should always be monitored. Depending on the net positive head requirements (NPSHR) of the pump, the following are guidelines for monitoring the suction strainer: Clean strainer at a differential pressure (Delta P) = 5 psi Alarm strainer at a differential pressure (Delta P) = 10 psi Operating the pump within its preferred flow region is crucial for long term pump reliability. The manufacturer will provide the minimum The Pump Handbook Series

and maximum continuous flow rates for the pump selected. These flows can be determined from the vibration criteria set forth by the API 610 standard (Figure 2). Pump operation at low flows results in pump horsepower heating the liquid being pumped. The pump manufacturer should also calculate and provide the thermal minimum continuous flow based on a worse case operating scenario. The limiting factor for the minimum operating flow may be determined by temperature rise rather than from vibration considerations. The critical point in the pump for temperature rise is usually the throttle sleeve and bushing used to balance the axial thrust. Here, the liquid temperature is increased by heat generated by impeller power, or inefficiency, at lower flow rates. A portion of this heated fluid is transported back to the pump suction area (at suction pressure) through a pressure balance line. If the vapor pressure of the fluid being throttled back to suction is greater than the suction pressure, flashing can occur in the balance line and at the end of the pump with the close clearance throttle bushing and sleeve. This can result in serious damage to these critical components. To avoid this, the balance line can be routed back to the pump suction source (below liquid level), and an orifice can be added to raise the pressure at the throttle bushing area high enough above vapor pressure to prevent flashing. When two 100% pumps are installed (one running and one standby), it is normal for operators to try to equalize running time between pumps. However, this practice is not necessarily best for pump reliability. Starting and stopping is hard on the pumping unit. For long-term pump reliability, it is preferable that pumps be run continuously. Also, if one of the pumps should fail, there is a higher probability that standby pump will also fail if there is equal run time on the pumps.

Maintenance Although heavy duty multistage

267

pumps are designed for extended, trouble free service, certain preventive maintenance measures should be performed on a regular basis to ensure optimum performance. A well-planned program of routine maintenance is the best assurance of dependable operation. The preventive maintenance inspections shown in Figure 4 are suggested as a minimum, and they can be supplemented by the experience of operating personnel. Axially split multistage pumps are designed to facilitate installed inspection and the performance of certain maintenance work at the job site. Removal of the upper half casing and rotor can be accomplished without removing the pump from its baseplate or disconnecting the suction and discharge piping. Some designs also offer axially split wear parts to facilitate inspection of the rotor without the removal of the impellers. However, it is important not to mix the halves of any split parts, including the casing and bearing housings. When two parts are machined as one, they become unique and cannot be interchanged with other similar parts. The pieces are usually match-marked by the manufacturer to avoid this problem. Impellers are individually balanced and the rotor balanced as an assembly. If any changes are made to the rotor or the impellers, the balance should be rechecked. Clearances and rotor runout should be checked prior to installing a modified or rebuilt rotor. Manufacturers can provide acceptance criteria for the critical balance (allowable imbalance) and runout values. An important feature for the reliability of axially split pumps is the sealing of the joint between the upper and lower halves of the casing, sometimes referred to as the parting flange joint. A gasket should never be re-used after the joint is opened for maintenance. Machined surfaces should be visually inspected after the gasket is removed to ensure that there are no imperfections on the flange that would affect sealing. Many manufacturers of axially split pumps taper one or both sides of the

268

parting flange to ensure that the gasket is properly compressed near the bore of the casing. A line may be observed on the casing where such a taper begins; this is normal for a casing with a tapered parting flange. A new gasket of the specified thickness, length and material is required. The casing was machined for a predetermined gasket thickness, and deviation from this gasket thickness or material will seriously affect pump reliability and operability. The gasket must be continuous across the entire surface of the parting flange to ensure positive sealing. Never use two pieces of gasket on a single sealing surface. At each seal chamber, a tab of the gasket should be extended beyond the face. After the upper half casing has been assembled, bolted together and properly tightened (to specified torque values), these tabs should be carefully trimmed even with the seal chamber face. This trimming operation is important as the mechanical seal will have an o-ring or spiral-wounded gasket that also must seal on this surface. There will be a “T” joint formed by the two gaskets, and it must be created precisely to ensure a positive, leak-free joint. Correct torquing technique must be used to ensure proper parting flange sealing. High pressure multistage pumps are typically designed to use capnuts for the main parting flange bolting. The advantage of using capnuts is that the flange bolts can be spaced much closer together than with conventional studs and hexagon nuts, thus creating a much higher gasket sealing force. The pump manufacturer will provide proper torque values and tightening (torquing) sequence for the main flange bolting to ensure an optimum assembly. Note: torque values in excess of 10,000 ft-lbs may be required for larger pumps. For these torque levels, hydraulic torque wrenches are recommended to provide accurate and uniform tightening.

Monitoring On-line monitoring and diagnostic techniques (Ref.5) provide a good The Pump Handbook Series

way to avoid costly pump problems. Multistage pumps can be equipped with a variety of vibration, temperature and pressure monitoring devices. Pump performance can be monitored over time with abnormal instances and trends reported. One example of commercially available monitoring equipment is the Bentley Nevada Trendmaster 2000, which can perform many monitoring and diagnostic functions. Also, many manufacturers can supply custom monitoring equipment designed exclusively for their equipment and tailored for specific pumping applications. Performance can be monitored by the user, or the original equipment manufacturer can track the performance and provide reports. Here is an example of where monitoring can prove beneficial. The effects of internal running clearances increase with service (running) time. When internal running clearances (wear ring and bushing clearances) increase, elevated bearing vibration levels and loss of original hydraulic performance (efficiency) can result. The rotor stiffening characteristics of the “internal bearings” (the Lomakin effect at close-clearance wear rings) will deteriorate as running clearances increase. The increased internal recirculation caused by such ring wear will affect pump hydraulic performance. By trending these two parameters, pump wear will be sensed and maintenance can be conveniently planned and scheduled to restore efficient pump performance before a forced outage occurs.

Final Thoughts Pump reliability is never a “done deal.” It should be regarded as part of a “Continuous Process Improvement” program. This article highlights some of the reliability factors of heavy duty multistage centrifugal pumps (particularly of the axially split, opposed impeller volute type), but it is by no means all inclusive. The most outstanding issue revealed in the reference material used to prepare this article is the importance of the relationship between pump manufacturers (vendors) and pump users. The pump engineer knows the

design of the pump and its operating requirements, and the process engineer knows the operating system and system requirements; but the two seem to rarely meet. An open relationship between the pump engineer and the process/system engineer would seem to have the most value and effect on pump reliability. The pump supplier needs to know more than simply what resides between the suction and discharge flanges of the pump.■

References: 1. API Standard 610, “Centrifugal Pumps for Refinery, Heavy Duty Chemical and Gas Industry Services,” 8th edition, August, 1995, American Petroleum Institute, Washington, DC. 2. Pumps and Systems Magazine, May, 1997 (Graphalloy). 3. API Standard 682, “Shaft Sealing Systems for Centrifugal and Rotary Pumps,” 1st edition, October, 1994, American Petroleum Institute, Washington, DC.

The Pump Handbook Series

4. API Recommended Practice 686, “Machinery Installation and Installation Design,” 1st edition, February, 1996, American Petroleum Institute, Washington, DC. 5. Pump and Systems Magazine, November, 1996 (On-Line Monitoring). John Lawler is a Design Engineer for Ingersoll-Dresser Pump Company with more than 18 years of experience. He is a graduate of Fairleigh Dickinson University and a Professional Engineer in New Jersey.

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The Role of Grouting in Standardized Installation Practices By Stan Moore, Solutia, Inc., and Scott Bullentini, ITW/Escoweld Epoxy Grout Systems

W

orld-class or benchmark values for machinery reliability and MTBF are relative, to say the least. They vary from company to company, and industry to industry and can even vary within plants. Few of us would, however, disagree that regardless of where we are currently, there is room for improvement. How do we improve? Spend money making wrong installations right? Or spend with the idea of getting it right and building reliability into our installations and systems from the start? By installing equipment correctly with reliability in mind from the beginning, one can eliminate many of the problems associated with equipment startups. It is interesting to note that many avoidable problems typically associated with equipment startups are considered the norm at more than a few facilities. In other words, the attitude is “Accept the problems, deal with them or fix them later.” When companies recognize the need to install equipment correctly and then follow through, future maintenance costs can be significantly reduced. This helps a plant striving for “best practice” achieve one of the many goals, namely the lowest possible operating costs. Plants that do this are recognizing that properly managed up front costs (with emphasis on reliability, availability, and maintainability) are necessary to attain reduced operating and mainte-

270

nance costs down the road. Often overlooked but of singular importance is the installation and grouting of baseplates and soleplates. In many instances, the root cause of pump/motor failures is poor or improper installation of baseplates or soleplates. Using a standardized installation specification for pumps and drivers is one method of eliminating this common problem. Improved baseplate installation and grouting procedures or standards can result is the following benefits: • Fewer bearing and seal failures • Easier and longer lasting mechanical alignment • Reduced vibration Our experience at Solutia, Inc. (formerly Monsanto) continues to demonstrate the cost effectiveness of using epoxy grout. Many “generic” systems have been installed over the years with cementitious grout. This practice will work for a period of time, but we have observed that maintenance dollars increase as the grout fails. Standard ANSI pump installations run reasonably well, with typical vibration levels that range from about 0.1 IPS to around 0.25 IPS on startup. However, these levels increase as the base and grout loosens over time. More frequent overhauls are performed, with attention given to the symptoms rather than the cause. For example, high capacity, maxi-fill or “filling-notch” bearings are often installed in the

The Pump Handbook Series

Comprehensive Strength* Tensile Bond to Steel Tensile Strength*

6000 psi

14000 psi

0

2,000 psi

700 psi

2,000 psi

Table 1: Cement-base grouts vs. epoxy grouts

hope that longer bearing life can be attained. More expensive seal designs are often employed. Various coupling configurations are tried in an effort to “handle” the misalignment. Yet the underlying cause is often inadequate installation of the baseplate or soleplate. It is obvious that larger pumps and/or more expensive rotating equipment should be grouted with an epoxy grout. Why, then, is it advantageous to utilize more expensive epoxy grout on smaller, less expensive centrifugal pumps or other smaller horsepower rotating equipment? This question can be answered by drawing conclusions from a comparison of two important physical characteristics common to both cementitious and epoxy grouts – compressive strength and tensile strength as shown in Table 1. (*Compressive strengths and tensile strengths are expressed as an average in a comparison of five different grout manufacturers.) It is clear that the major difference is the greater tensile compres-

sive strengths of epoxy grouts versus those of cementitious grouts. The objective of grouting is to create a single block monolith, one that will act as an effective vibration dampener. Properly installed, an epoxy grout will bond to a concrete foundation with a tensile strength greater than that of the concrete foundation. More importantly, the bond strength of an epoxy to a properly prepared baseplate or soleplate will be, on average, 2000 psi, whereas a cementitious grout will effectively have no tensile bond to even the best prepared baseplate or soleplate. In many instances, the use of a cementitious grout leads to voids because it cannot provide a perfect non-shrink bond to steel. Because of settlement, cementitious grout leaves small deposits of water which, over time, will create rust deposits on the underside of the baseplate or soleplate. In addition, baseplate resonance problems can exist when voids are present. With time, the integrity of the installation will be compromised, and seal and alignment problems will begin to appear. The tensile bonding properties of epoxies provide a method of sealing the concrete foundation and underside of the baseplate that prevents damage from moisture and oil.

Installation Method Critical We’ve established one of the major benefits of using an epoxy grout instead of a cement-based grout. The next step is to select a method of installation so the potential of the epoxy grout can be maximized. The first step is to identify the need for a proper installation specification. This specification should comprehensively detail all aspects of the grouting procedures. The installation specification should include the following sections: 1. Scope 2. Material and testing requirements 3. Baseplate or soleplate design 4. Equipment and material storage 5. Foundation preparation and anchor bolts

Sloping Drip Rim Shown in partial for clarity

open 4”-5” grout opening per section Typical motor supports. And horizontal jack screws to facilitate alignment Radius all edges in grout areas

Min 1⁄2” vent holes @ each intersection or joint of support members

Fabricate support to match pump foot print

Cutouts for rout floor. Min 11⁄2 x 11⁄2 on 12” centers

Min 1” holes for anchor support Min 1⁄2” leveling screws @ each anchor bolt Channel, I-beam, or min 1 ⁄2” plate for support members

Radius all edges

Figure 1. A fabricated steel baseplate designed for epoxy grouting 6. Baseplate or soleplate preparation 7. Installation of baseplate or soleplate on foundation 8. Forming the foundation 9. Grout installation 10. Testing of grout samples There are a number of good installation standards for epoxy grouting. Many of these are available from grout manufacturers, industry trade journals such as Pumps and Systems Magazine and rotating equipment reliability consultants. The key to creating a successful installation specification is an attention to detail extended throughout the installation standard or procedure. We will explain some of the basic steps.

Baseplate Design Fabricated steel baseplates are replacing older installations incorporating cast iron baseplates. Periodic vibration monitoring and analysis is utilized as one tool to define the scope of work and also to provide economic justification for the change from cast iron baseplates to fabricated steel units. Fabricated steel baseplates are also being specified for use with new The Pump Handbook Series

pumps. The reason for the additional cost of a fabricated steel baseplate is simple: it is more versatile. By utilizing jack screws and anchor bolts for adjustment, the baseplate can be twisted and contorted until the machined pads, where the pump and motor mount, are level and coplanar. A properly leveled baseplate is the first step in the elimination of softfoot problems. We typically strive for a levelness of 0.0005”/ft. The installation of the baseplate and placement of the grout is best facilitated with equipment removed from the baseplate. Re-install the motor and pump after the grout has been poured and cured. Fabricated baseplates are available from pump manufacturers, and many of them are ready to be installed with epoxy grout (Figure 1).

Photo 1. Old cast iron installation

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Grout filled holes 4-6” diameter Baseplate underside coated with compatible primer

1

⁄2” vent holes

45” wooden chamfer

Wood forms- wax coated to prevent sticking.

Epoxy Grout Concrete coated with epoxy resin

3 ⁄8” vent holes. Spaced every 812 inches.

Sleeves filled with pliable material

Seal wood forms below concrete and grout interface.

Wrap bolts to protect from grout

Concrete Foundation

Photo 2. Newly epoxy grouted installation

Original: P. Monroe Redrawn: S. Moore

Figure 2. Preparation of concrete for epoxy grouting (original artwork courtesy of Perry Monroe) Lin

ips None

Peak

Mag

1.

.8

.6

.4 Vibration levels dropped after the base was grouted and unit was aligned. .2

0.

5/19/96 9/12/96 1/6/97 X:2/26/96 2:09:30 PM ACRILAN 60% SOLVENT PUMPS A 29m/c MFH Speed3550. Position:1Direction: Horizontal Latest

5/2/97

8/25/97

Figure 3. Vibration levels before and after epoxy grouting They can also be fabricated in local shops. We have done this on several occasions when upgrading existing equipment. The economics of local fabrication versus utilizing an OEM baseplate obviously varies and should be evaluated on a case-bycase basis.

Case History #1 Four deionized water pumps were originally installed on cast iron

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baseplates with no grout (Photo 1). Maintenance costs over the years had risen to about $8000/pump/year. Failure data was gathered from the CMMS system and justification established to replace the baseplates. (To some outside observers, justification for such an upgrade would appear to be a “no-brainer.” However, status quo maintenance activities and methodologies are sometimes a difficult nut to crack.) The pumps The Pump Handbook Series

Photo 3. Small basic baseplates can be epoxy grouted with success

were relocated and installed on fabricated steel baseplates. New concrete pads were poured and prepared for epoxy grouting. Proper preparation of the concrete is essential to realize the full potential of an epoxy grout. Normally, the top 1-2” of a finished concrete pad will contain less aggregate and more sand. This portion of the concrete, often called laitance, is weaker and needs to be removed. This is usually accomplished with a small chipping hammer. The idea is to “rough up” the surface and expose the aggregate. Figure 2 depicts a typical pad that has been prepared for grouting. Overall vibration levels for the installation were reduced from 0.50.7 IPS to less than 0.15 IPS overall. The pumps have been in service for about 24 months (Photo 2). One seal has been replaced in that time. The initial alignment was easily accomplished, with each unit taking about 45 minutes. (Level pads that are coplanar significantly reduce the amount of time it takes for an alignment job.) The total cost, which included the relocation, piping modifications, pump overhauls and grouting, was approximately $40,000. The simple pay-back is just over 12 months. MTBF has increased from 3 months to 8 years and counting.

Case History #2 Two 15 hp circulating pumps shown in Photo 3 were re-grouted to correct alignment and vibration problems that resulted in premature seal and bearing failures. The two existing baseplates were a simple fabricated steel design that we reused. Cementitious grout was used in the original installation. The grout had come loose over time, allowing the baseplate to resonate and move. Periodic vibration surveys were made on these pumps. This information was used in supporting the decision to re-grout. The importance of historical vibration data cannot be overstated in a successful reliability improvement program. Data shown in Figure 3 is typical of other trends observed on similar pump installations with poor grouting. In addition

to reducing the overall vibration on these two units, seal replacements that were once routine have been eliminated.

Conclusion As evidenced from the case histories and information contained in this article, it is necessary to consider a number of factors effecting the final outcome of a pump installation. It should be noted, however, that by utilizing a sound specification for grouting of all rotating equipment, one variable in the overall troubleshooting process is eliminated when considering retrofit applications or questions of equipment reliability. The same holds true for new installations. As stated, there are a number of excellent resources for information

The Pump Handbook Series

relating to procedures for proper installation of pumps and drivers. Epoxy grout is part of the “winning equation” that makes up a sound installation specification for pumps and drivers and any other type of rotating equipment for which proper alignment is crucial.■ Stan Moore, P.E., is a Reliability Group Leader at Solutia, Inc. in Decatur, AL. He has more than 15 years of experience in machinery maintenance with em-phasis on alignment, condition-based monitoring/ analysis and equipment reliability improvements. Scott Bullentini, a Project Manager at ITW/Philadelphia Resins Corporation, has more than 10 years of experience in the field of industrial construction and management.

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Variable Frequency Drives for a Vacuum Pump System By Kevin Skelton, Graham Manufacturing Co., Inc. Editor’s Note: Responding to requests from readers – many of whom work with vacuum pumping systems – Pumps and Systems is introducing this new column with the January issue. Look for it periodically. n many process plant applications, using a Variable Frequency Drive (VFD) to control the capacity of a liquid ring vacuum pump (LRVP) is an innovation. Previous system designs consisted of sizing the LRVP to operate efficiently for a certain portion of the process cycle and then using a recycle control valve system to maintain the desired LRVP suction pressure. This usually meant a loss of system efficiency for most of the operation cycle. A liquid ring pump can be used in vessel evacuation or batch processes where the load volume is high at startup and diminishes as the vessel is emptied and the subsequent vacuum level increases. At the end of the cycle the usual intent is to hold as high a vacuum level as possible for a given period of time. This high vacuum level is a function of the properties of the service liquid, which boils or vaporizes as specific conditions exist. The liquid ring pump handles the least amount of net capacity at this high vacuum. The majority of the capacity at this point is the vaporized service liquid. The LRVP operating at full load rpm is oversized for the vacuum end point. By slowing down the pump, the operator can

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maintain the desired vacuum, but pumping capacity is decreased, lowering the brake horsepower. An additional feature is that the drive could be set up to operate at a speed that stays within the horsepower rating of the motor and automatically adjusts as the LRVP power demand decreases. A given frame size pump has a predetermined impeller tip speed range that would be programmed into operational parameters in the control room. The impeller tip speed is a function of the impeller diameter. (The tip speed or tangential velocity is the rate of the outer tip of the impeller.) The pump would then run in this tip speed range while staying below the maximum horsepower rating of the motor. At the holding point, the pump is slowed to the minimum speed low enough to hold the desired vacuum. This saves power and wear on the pump. Current system design usually includes a provision for over-capacity at the holding point. The device may be a vacuum relief valve or recycle valve system. This design approach can waste energy. An energy audit of a particular facility may be required to determine the feasibility of installing VFD’s on the current LRVP system. The system design engineers should take a serious look at incorporating this approach in future design specifications. Useful LRVP Calculations (approx. The Pump Handbook Series

relationships): rpm1/rpm2 = (hp1/hp2)2 Capacity1 / Capacity2 = rpm1/rpm2 As an example, the capacity of a 25 horsepower LRVP can be adjusted while operating from about 1000 rpm up to 1750 rpm. This represents an impeller tip speed range of 40 to 70 ft/sec. At 1750 rpm and at atmospheric suction pressure, the capacity of this pump is approximately 300 Actual Cubic Feet per Minute (ACFM). Depending on user specific conditions, the brake horsepower (BHP) may exceed 25 briefly during an evacuation cycle. The motor rpm can be reduced to keep the brake horsepower below 25 during this period. When the system is totally evacuated and the vacuum level is at its maximum, the pump is then slowed to a speed at which it operates adequately and does not cavitate. A word of caution – there is a minimum impeller tip speed which is a function of the pump’s compression ratio or vacuum level. If the pump speed is reduced below this minimum impeller speed, the liquid ring will collapse, causing system instability or “upset.” The minimum impeller speed can be determined to be the non-condensable load quantity (which sets the pump suction pressure) as a function of the compression ratio desired. More energy is required in the liquid

Brake Horsepower 50

40

30

20

10 900 43 ft/sec

1000

1100

1200

1300

1400

1500

1600

1700

Pump RPM

1800 83 ft/sec

LRVP Varied Speed Performance 4.7 PPH Dry Air Load, Suction Pressure 29.2" Hg Vacuum Figure 1. Typical 50 hp LRVP operating at a high vacuum level – note that horsepower requirement drops dramatically as the rpm is reduced while the suction pressure remains constant.

ring to achieve a high vacuum than a low vacuum, so as the compression ratio increases, the pump must rotate faster than it does at lower compression ratios. On the other hand, at low vacuum levels it is beneficial to maintain as high an impeller speed as possible in order to maximize the pump’s ACFM capacity. Figure 1 illustrates a typical 50 horsepower LRVP operating at a high vacuum level. As you can see, the horsepower requirement drops dramatically as the rpm of the pump is reduced while the suction pressure remains constant. To illustrate the energy savings over a year, the assumption was made that the pump was operating continuously but at the holding vacuum level only for 60% of the time. From Figure 1: Operating hp @ 1750 rpm = 46 hp Operating hp @ 900 rpm = 13 hp hp saved 33 hp 33HP(.746 kW / 1HP)(8760 hours / year)($0.06 / kW hour )(0.6) = $7884 This is the savings for one 50 hp liquid ring vacuum pump for one year. Power generating facilities use much larger LRVP systems to maintain vacuum level on steam turbine condensers. These pumps operate

over 90 percent of the time at the holding vacuum level, utilizing a vacuum relief valve to introduce air load to the LRVP in order to maintain a vacuum which is above the cavitation point. This larger sized LRVP operating @ 720 rpm with an impeller tip speed of about 65 ft/sec has a bhp of 82. Slowing the pump down to 475 rpm will result in a bhp of 36 a difference of 46hp. 46HP(.746 kW/1hp)(8760hr/yr)($0.06/kW hr)(.9) = $16,233 per year for each condenser vacuum pump. Electrical savings can quickly be calculated because the hp requirement is the square of the ratio of the speed difference. This calculation is fairly accurate and a good “rule of thumb” method of determining power requirements. Traditionally, the capacity of a LRVP has been determined as a direct ratio of the speed difference. This method holds true until the pump is at or near the end point, or no-load operating range. The capacity doesn’t change due to the reduction in the pump rpm at high vacuum levels. This is because the majority of the load to the pump at this point is the vaporized service liquid, not net load. Specific testing has been conThe Pump Handbook Series

ducted to determine an LRVP’s minimum operating speed. At no load, which would be the highest vacuum level that the service liquid would allow, the 50 hp pump became unstable at 1000 rpm (32 ft/sec impeller tip speed). If an air load was introduced to the pump, the suction pressure rose accordingly, and the pump could be slowed down even more, perhaps even to 800 rpm at a 27”Hg vacuum. In this application, the set point would want to approach the 1000 rpm point as this is the minimum required for stable operation. A 25 hp pump could be slowed down to 810 rpm at no load and still maintain the desired vacuum, but it would be wise to include a safety factor in the equation. The costly approach of recycle lines and control valves, or unreliable vacuum relief valves can be eliminated as a means of maintaining vacuum system stability. The total yearly rotations of the equipment are reduced drastically, lowering maintenance costs. Life cycles of bearings, mechanical seals and rotating assemblies will be extended. Bring your vacuum producing equipment up to current technology standards by specifying a VFD control. ■ Kevin Skelton is a Research and Development Engineer for Graham Manufacturing Co, Inc. He has more than 20 years of experience with liquid ring vacuum pump applications and liquid service. He currently holds a patent for operational improvement of the design.

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Turnaround Time! Is it that time of year at your plant? Here’s some advice on turnarounds and spares from one of our regular contributors. By Eddie Mechelay s turnaround planning becomes more sophisticated, plant management personnel are questioning the dollars spent during this time frame. Performing pump repairs during turnaround time is coming under closer scrutiny. With many facilities having fully spared equipment and mean time between failure increasing, the need to work on these pumps during a complete plant shutdown is minimized. There are three philosophies on whether or not to work on pumps during a plant turnaround. The first of these philosophies is to tear down equipment that characteristically has given the plant reliability problems. The trouble with this method is that it is often based on heart strings and very seldom on predictive maintenance data or other documented information. Often operations personnel are disappointed or even angry if certain pumps are not overhauled or inspected during an outage. Their justification can be no more than, “We’ve always looked at that pump during turnaround!” Actually, as maintenance personnel, if we have not done a thorough job of recording our data through the years, who can blame the operator? This often unnecessary work can significantly increase turnaround costs. Basic pump overhauls can cost as much as $5,000 to $15,000, after the price of the bearings and seals are included. The cost of performing this work is greater at turnaround time because of additional employee overtime and possible turnaround overhead. So, if you don’t want extra work during turnarounds, you should track the reliability and con-

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dition of the equipment throughout every year. Another concept is to schedule work on critical pumps or pumps that experience high wear. There are benefits to this plan if it is carried out correctly. For example, slurry pumps can experience wear at an undetermined rate. With improvements in metallurgy and internal geometrical designs, the actual wear rate is constantly changing. Turnaround is an excellent opportunity to assess the condition of these or any other high wear pumps. If spare parts are reviewed and local machine shops warned of potential upcoming work, these pumps can usually be repaired in a timely and cost effective manner. As an example, one refinery typically replaced impellers and/or volutes on their FCC bottoms slurry pumps after six months because of impeller and volute wear. After re-engineering the pump material and implementing boron infusion as a resistance to wear, this six-month time frame of failure was increased to years. In fact, the pumps had not been inspected for wear for three years and were performing satisfactorily. During a recent turnaround, these pumps were not inspected until they were pulled due to a seal leak at startup. During the seal replacement process, inspection revealed that the volute casing was 1/4” away from developing a leak in one isolated area. The complete pump was replaced with an inventory spare on an overtime basis with mechanics tired from a month long turnaround schedule. Had this equipment been inspected as a part of the turnaround schedule, the work The Pump Handbook Series

could have been done in a timely manner with fresh mechanics. Although this concept of scheduling critical pumps has benefits, caution must be taken not to identify all pumps as “critical.” This discussion on high wear pumps brings in a question of operating spared equipment. I’m often asked, “How often should I run my spare equipment?” or “Should I run my spare equipment on a routine schedule?” Without question, all equipment should be run on a timebased schedule. If pumps are operating in a non-wear or slow wear service, I recommend each pump in a two-pump fully spared set be operated for a week at a time. In situations similar to the slurry pumps discussed previously and other high wear applications, I recommend the spare pump be run for 24 hours once a week. This recommendation also holds for applications that have common spares. A common spare pump should be run for 24 hours once a week to maintain bearing lubrication. This also prevents contamination buildup in the stuffing box and around seal faces, a process that can cause immediate failure on startup. An added benefit to a time-based switching schedule is the confidence in the equipment that operators will have as they routinely switch pumps with satisfactory results. The third and increasingly popular philosophy of pump repair at turnarounds centers on an ultimate goal. In this scenario, predictive and preventive maintenance tools are used to their fullest capacity. All nonspared equipment that requires preventive work is scheduled at

turnaround time with all parts ordered ahead of schedule on a routine basis. Predictive maintenance reports are reviewed and critical equipment scheduled for repair. Non-critical equipment can also be scheduled for repair during turnaround if it will be more cost effective. Other non-critical items can be scheduled at this time to further reduce reliability risk on pump services that have a low mean time between failure. Many companies use an accrual system in the maintenance budget to finance turnarounds. With this method the department is charged a fixed amount each month to be applied to a future turnaround. Obviously this dollar amount needs to be maximized at turnaround to gain the greatest number of plant repairs. In the past, facilities charged any and

all work that occurred during a turnaround to the same budget. Now maintenance managers are asking planners to charge to the turnaround budget only those items that cannot be done unless a plant shutdown occurs. Effectively, this removes any spared pump from the accrual budget. However, with proper documentation on pumps with high wear and pump sets with lower MTBF, you may be able to justify these charges to the turnaround budget. Removing equipment for repairs at turnaround because “we always have” is a costly way to do business. Determining a few critical or high wear pumps to work on during shutdowns can decrease the repair cost because it necessitates proper planning. Incorporating predictive maintenance techniques along with recommended preventive mainte-

The Pump Handbook Series

nance items can assist in determining what equipment must be scheduled during plant shutdowns. The turnaround budget “free for all” is a thing of the past!■ Eddie Mechelay is a former mechanical equipment engineer at Total Petroleum who is currently providing consulting services for pump and seal reliability.

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The Critical “First Hour” Startup of a TurbineDriven Boiler Feed Pump By John H. Burgner was introduced to the pump business in 1960 and since then have worked as an employee of OEMs, distributors, engineering contractors and end users. I’ve worked in 15 countries, handling almost every type of liquid, and have watched and participated in the evolution of pumping equipment. Certainly, significant improvements have occurred. In some areas, however, I believe advancements are not noticeable. In fact, there is some indication that the reverse is happening. The following article applies to an area where dependence on technology is failing to produce an acceptable result. This being boiler feed pump startups. Specifically, turbine driven, new boiler feed pump startups in the medium hp range. I don’t want to come across as some old coot who considers the doorbell cutting edge electronics, but in more than a few installations, I have seen instrumentation override common sense with disastrous results. This article may help someone avoid getting into a position in which common sense is considered optional equipment when the valves are opened. As a final introductory note, I would like to add a disclaimer. My responsibilities have usually related to equipment failures. I can never remember being asked to inspect and/or comment on a pump that was running fine. For this reason, this writing respects Murphy’s Law and, as a result, may seem unduly alarmist to some.

I

Design Stage Other than all normal equipment selection and piping layout

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standard precautions, two areas should be given special attention: 1. The suction strainer should be a “witch’s hat” design pointed upstream with openings of about 1/2 the pump impeller maximum sphere size and a total open area of about 5x suction pipe cross sections. Provision should be made to cover this strainer on the upstream side with 100 mesh screen during plant startup, and it should easily be removed for cleaning or inspection. In other words, mount it in a spool. There should be a P.I. tap just upstream and downstream of the strainer. No significant obstruction such as a check valve should be in place between downstream gage and pump suction. 2. The recycle system should not be designed for the pump manufacturer’s stated minimum flow. Depending upon the reliability of the vendor’s minimum flow specifications, the system should be designed for at least 125% of required minimum flow. For pumps in the 500 to 5000 hp range, 20% of the flow at best efficiency can be considered a guide. I personally like to see a definite pressure drop between shut off and recycle.

Additional Recommendations In addition to these major concerns, the following suggestions should be considered. Do not buy a pump with clearances under API 610. Most major manufacturers have good rugged designs such that when a pump is new it can be assembled free from internal contact with clearances slightly under API. This, however, The Pump Handbook Series

requires all future rebuilds to maintain factory tolerances throughout – something that real world conditions do not allow. (All this is for the sake of a point or two of efficiency on a factory test for a unit that will be driven by a turbine.) Make the recycle system idiot proof. If there is any possible way to misoperate or misunderstand it, that will happen, and most systems do not allow an on-site observer to confirm that the recycle is in operation. The recycle should carry as far back into the system as is practically possible. In theory, recycle time should be limited, but operators do not like to shut down and restart, so they do not necessarily observe optimum recycle procedure.

Construction Stage I consider turbine piping a nightmare primarily because few people read and follow the manufacturer’s manual. The turbine manufacturer has provided good basic information, and sometimes (not always) the piping designer has incorporated it. Unfortunately, a lot gets lost during construction. The following are common examples: 1. Drain and bleed-off lines typically should not be manifolded. In fact most bleed-off lines should not be valved. Follow OEM manual instructions. They spell out what is required. 2. Spring cans, spring hangers and skid pads are used to allow thermal growth of turbine and piping without incurring pipe strain. For them to obtain this desired result, they must be installed and adjusted

correctly. Here a little common sense will go a long way. Keep in mind that all system components (turbine, pump, piping, valves, even supports) are going to grow but at different rates and amounts. Preferably before, but during, and always after growth, the piping must be supported and positioned correctly by something other than the flanges of the rotating equipment. 3. Pipe fit-up has now advanced to the point that construction specifications usually state that the piping should be made up from the rotating equipment and not to the R.E. Nowhere is this more important than steam turbine inlet and exhaust connections. High temperature pumps and compressors are usually somewhat symmetrical. There is not much symmetrical about a turbine. Pipe strain will drive you crazy unless each piping step is taken “by the book.” Almost all manuals give a specific vertical offset with the turbine shaft (cold) below the pump shaft to allow for thermal growth. Manuals usually also specify that a hot alignment check be made. Although the offset numbers given today are a lot better than those years ago, they are still suspect, and I have yet to see a practical hot alignment check procedure published in a manual. I have the following specific suggestions. 1. If you want an exact alignment that can be performed by crafts people without expensive equipment and engineering supervision, install Essinger balls for measurement of thermal growth. There is also a computer system available that allows craftsman level personnel to measure thermal growth with a high degree of accuracy. 2. Train one or more of your people in reverse indicator alignment (with or without use of a computer) and require procedural exactness. 3. If you want to go up in technology and expense, use one of the new laser alignment systems available. Properly applied, they replace both the R.I. and E.B. equipment and procedures. 4. If, for any reason, you are willing to accept less than an exact operating “hot” alignment, I suggest the following:

a. Proceed through construction with special attention to points 2 and 3. b. Set cold alignment exactly as specified in the OEM manual. c. Fix a dial indicator firmly to the pump coupling hub with the button just clear (approximately .050”) of the turbine hub O.D. d. Run the turbine uncoupled at design rpm for about 4 hours. If possible, circulate design temperature water through the pump during this time. If you are able to circulate the water, you must turn the pump shaft by hand 90 degrees every 30 minutes or so. Caution! If your pump is foot mounted (and this should be only in the 100ºC range), the thermal growth will be straight up, relatively minor, specified, or easily calculated. If your pump is centerline mounted, there should be relative little vertical growth, and the axial growth should be allowed for by movement of the outboard feet. These considerations make hot alignment somewhat simpler for the pump. However, when pumping temperatures exceed, say, 120°C, a problem relating to alignment, becomes a concern. If, when hot water is circulating through the pump and the pump becomes difficult to turn, you have internal thermal distortion. This problem is not a major factor below 120°C, or for alignment per se. However, it is a factor for startup, operation, hot standby and slow roll. If your pump does not turn freely after circulation warm up, carefully review your insulation, fluid circulation path and standby and startup procedures to insure that the entire unit will be at uniform temperature during all stages of operation. Do not attempt to align (or run) a hot pump that won’t turn freely.

To Recap: 1. Pumping temperature 120°C or below: a. Centerline mounted pump, horizontal or vertical split – Don’t worry about hot alignment growth of pump or internal thermal distortion. b. Foot mounted pump – Calculate thermal rise of pump shaft, align accordingly and forget. 2. Pumping temperature above The Pump Handbook Series

120°C: a. Foot mounted pumps generally should not be applied. Consult your OEM. b. Centerline mounted pumps. Align cold or align hot if the pump turns freely. Don’t worry about vertical thermal growth. If the pump does not turn freely after warm up, do not start up until problems which usually relate to items mentioned above and will carry over into operation unless understood and corrected, are addressed. If you have internal thermal distortion due to insulation, vent and drain configuration it is better to locate and correct it now and not during startup. However, please refer to the warm up procedure described in the following startup section before attempting to correct any existing problem. 5. Shut the turbine down and quickly move the indicator button to contact, and read .000” at top center of turbine hub. Take an additional three 90 degree readings. You now have a fair idea of whether your hot alignment is acceptable. However, the following must be considered: 1. You only have one set of readings. An exact alignment position requires two. It is possible, but very unlikely, that you have misalignment even with 4 – .000” @ 90 degrees. 2. The turbine, unloaded, probably did not reach full operating temperature. Should you wish to continue one step farther (and you should), proceed as follows quickly so as not to allow cool down. 1. Shift the indicator button so it just clears the face of the turbine coupling hub (approximately .025”) near the O.D. 2. Restart the turbine and run for approximately 1 hour (assuming your shut down was 15 minutes or less). 3. Shut down and quickly bring the button into contact and reading .000” at top center of coupling hub face near O.D. 4. Take an additional three 90 degree readings. (Keep pump rotor thrusted toward turbine.) 5. Repeat several times quickly

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to confirm accuracy. You now have enough information to say with reasonable assurance whether your alignment is satisfactory. Keep in mind that hot alignment takes pipe strain as well as thermal growth into consideration. If the face readings look bad (they should be very close to 4-000”) check to confirm that the problem is not with the face. The T.I.R. of a new coupling face properly installed should be close to .000”. If it’s not, runout must be taken into consideration, which makes this procedure more difficult. By referring to hub O.D. readings you can determine parallel misalignment, and by referring to face readings you can determine angular misalignment. Before making extensive corrections you may want to reconfirm cold, final, readings after a minimum 8-hour cool down. These readings should repeat those cold readings previously taken. The advantage of the last suggestion (4) is its practicality. The measure can be taken with people and equipment that should be available at every job site, and when performed carefully it is acceptably accurate. Because most B.F. pumps are run on factory test and not all seal chambers and cases are thoroughly drained, a little job site storage protection is usually in order. Caution! You can jeopardize warranty unless this procedure is performed by factory representatives. Refer to the manual as to what you, the customer, can and should do and what the OEM insists be done under vendor supervision. As a minimum you should confirm that the bearing housings are free from contamination. As a secondary consideration you should attempt to drain and flush the seal chambers with rust preventative. Anything else would be the same steps and procedures applying to all rotating equipment pre-commissioning.

Startup It is impossible to overemphasize the care that should be given to the startup of a turbine driven boiler feed pump. Compared to other rotat-

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ing equipment the pump and turbine are usually expensive and, because hot water has very low lubricity, pumps are easily damaged. In addition, NPSHA will usually be limited and suction level information, during start up, contradictory. For these reasons you need an experienced startup operator who understands the problems involved. Because this article is intended as an overview, I will cover only a few areas of major concern. 1. The strainer arrangement covered in the design section should be in place. After startup watch for a pressure drop across the strainer. When the drop increases by 3 psi, shut the pump down and clean the strainer. 2. You should know what the discharge pressure should be at zero flow on recycle. During the first hour or so of operation watch the discharge gauge for any unexplained drop (or rise) in pressure. If this occurs, shut the pump down until the change is explained. 3. Very carefully vent the pump prior to startup. 4. Be careful of slow roll. Usually it is desirable, but some sleeve bearings will not develop an oil film at very low rpms, and some recirculation systems will not bypass until a pressure above design is reached. 5. Warm the pump gradually. On all startups, and most importantly on the initial startup, the pump should be brought slowly and uniformly up to operating temperature. In our section on hot alignment we have covered considerations of high temperature pumps. We have also discussed internal thermal distortion. To cover warm up specifically prior to startup, please consider the following four situations. (Alignment is now assumed to be complete.) 1. Pumping temperature at 120°C or below – hot circulation water available. Crack open recycle or warm up line. Crack open all drains. Crack open suction block valve and vents – allowing about one hour. By feel you should be able to determine relatively uniform temperature. The pump should turn freely. You may have to The Pump Handbook Series

make some valve adjustments, but after about two hours of warm-up time, all should be ready. (Obviously, consider pump size.) 2. Pumping temperature at 120°C or below – no hot recirculation water available. After going through normal pre-startup (venting, etc.), start the pump as you normally would but monitor vibration and avoid sudden (or even rapid) temperature changes. 3. Temperature above 120°C – hot circulation water available. Go through the same procedure described in situation 1, but use more time and care. For example, if your pumping temperature is 250°C, at some point during warm up you may find that the pump is very difficult (or impossible) to turn. Don’t force it! As long as it turns freely when it is fully heated, all is okay. Note! The exact procedure used for initial warm up should be recorded for use by operations in the future. 4. Temperature above 120°C – no hot circulation water available. This is a repeat of situation 2, but it also requires more time and care. If the pump is properly insulated, mounted and piped, you should have no problem, but for each 50°C rise above 100°C, you should allow 1 hour of operation and be careful to monitor for vibration. Caution! Initially starting a pump in this manner could conceal a problem in the warm up procedure. I don’t want to beat this subject to death, but please consider a large, high temperature feed pump on hot standby. If the insulation, warm up lines and slow roll procedure do not produce a uniformly hot pump, you will have thermal distortion. This will probably be evident in that the pump will be difficult or impossible to turn. As stated elsewhere in this article, if the pump doesn’t turn freely, don’t start it. Some time in the commissioning you should confirm that the operational warm up procedure you intend to use will produce a uniformly hot pump that rotates freely. At all times you should keep in mind that the pump should be at a uniform temperature. And with all major rotating equipment startups,

you should be alert to any vibration, noise level change and temperature rise. Use the manual trip to shut the unit down at least once during start up proceedings. All parties involved in a turbine’s startup work should be aware of the location and operation of the manual trip. If you have followed all, or even most, of the above guidelines, you should now have your unit past the most critical period in its life: the first hour.■ John H. Burgner has been involved in the pump industry for more than 35 years. He has worked as a sales and service specialist for Afton Pumps, and since 1985, has been an independent pump consultant specializing in centrifugal pump troubleshooting for clients around the world.

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Methodologies for Calculating MTBF – Part 1 Learn how to generate a reliable Mean Time Between Failure value and use it to improve identification of problems. By Ken Noble, Petro-Canada,Oakville Refinery ean time between failure (MTBF). Put a group of maintenance personnel or equipment specialists together in a room, and inevitably a discussion of who has what value of MTBF will ensue. However, when the meeting is over, more questions will have arisen than have been answered. How can four simple letters be so confusing and frustrating? Part I of this article will explain how an MTBF value can be generated and show why confusion can easily arise. Part 2 will discuss enhancements that can be utilized to supplement the generation of equipment MTBF values.

ent pump services. MTBF is defined as the mean time between successive component failures. However, even this simple definition must be refined. Two successive component failures could be defined in three ways as follows:

Basic Methods

Equation 1:

M

To illustrate different methods of calculating the MTBF, we will use the simple model shown in Figure 1. This equipment consists of 3 differContinuous Duty Service Only 1 Pump Operates at a time

1. Two successive failures on the site 2. Two successive failures on the same piece of equipment 3. Two successive failures in the same operating service At a site, where there are many pieces of equipment, the interval between successive site repairs can be calculated using Equation 1.

= ∑

1

MTBF for the site

Continuous Duty Service

1

MTBF for each piece of equipment

Figure 2 shows how the equipment is modeled to calculate the MTBF for successive site repairs. In this example, one pump will need repair every 3.8 months. However, we often want to determine the statistical interval of time between successive repairs for each piece of equipment. In this case, we determine the number of repairs that occur in a specified time period and use Equation 2 to calculate the MTBF for each piece of equipment. Equation 2: MTBF = Total Number of Pieces of Equipment [ Number of Repairs] / [Time Interval] In the examples provided so far, no consideration has been given to the effect of pump operating time. An MTBF has been generated utiliz-

50% Duty Pump

Service No. 1 Pump “A”

Service No. 1 Pump “B”

Service No. 2 Pump “C”

Service No. 3 Pump “D”

Repair: June 1/94 Repair: July 1/94 Repair: Dec 1/95 Repair: June 1/96

Repair: Apr 1/94 Repair: Apr 1/95 Repair: Oct 1/96

Repair: Jan 1/94 Repair: Nov 1/96

Repair: Dec 1/93 Repair: Jan 1/97

Service No. 1 Pump “A”

Service No. 1 Pump “B”

Service No. 2 Pump “C”

Service No. 3 Pump “D”

MTBF (A) = 24 mths./3 rep = 8 mths./repair

MTBF (B) = 30 mths./2 rep = 15 mths./repair

MTBF (c) = 22 mths./1 rep = 22 mths./repair

MTBF (D) = 37 mths./1 rep = 37 mths./repair

1 =1+1+1+1 MTBF 8 15 22 37 MTBF = 1 = 3.8 mths. between repairs 0.26

Figure 1. Simplified example of site equipment

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Figure 2. Calculation of MTBF for a site The Pump Handbook Series

ing calendar time rather than operating time. Calendar time is often used as the time measurement because most equipment is not metered for operational time. In an attempt to evaluate the MTBF based on operating time, some facilities calculate separate MTBF values for continuous and intermittent operating services. Others multiply the total number of equipment pieces on site by a factor less than one to normalize intermittent operation as shown in Equation 3. Equation 3: MTBF= [A] x [Total Pieces of Equipment] [No. of Repairs for all Pieces of Equipment/ [Time Interval] Photo 1. Product storage pumps in continuous and intermittent operation handle different products at Petro-Canada’s Oakville, Ontario refinery

Where A is a factor less than 1 For example, a site may contain: 1) 150 pumps in continuous operation 2) 50 pumps in operation 50% of the time In this example, the value of the factor A would be: A = 150 x 1.0 + 50 x 0.50 = 0.875 (150 + 50) Utilizing this concept of normalizing the operating time for each piece of equipment, we can construct Table 1 and calculate an MTBF using the same repairs and time interval shown in Figure No. 3.

Photo 2. Gasoline storage tank transfer pumps operate with flow requirements ranging from 25 to 1200 gpm at the Oakville facility Time Interval = 36 mnths.

1993

1994

1995

1996 Time

Dec

Dec

Dec

Dec

Repair Number of Pumps =4 Number of Services =3 Number of Repairs in time interval =9 4 = 16 mths. / repair / pump or Pump MTBF = [9/36] Service MTBF =

3 = 12 mths. / repair / service [9/36]

Figure 3. Example of MTBF for individual equipment The Pump Handbook Series

Which Method to Select Table 2 summarizes the MTBF values that have been calculated so far. All four are based upon the same time interval and repair dates. Which methodology should be selected? That depends on the type of equipment and plant operation. The site MTBF is the most mathematically intensive of the four examples. An MTBF for each pump must be determined, which in turn must be placed into Equation 1. Since most facilities have various continuous improvement initiatives, the MTBF for each pump must be re-evaluated on a periodic basis so that the site MTBF reflects the results of improvements. This can be done, for example, by determining the MTBF for

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each pump every three months or six months, utilizing the last three or four repairs. The site MTBF concept originates from a reliability model related to a group of components – i.e., a computer hard disk drive. Based on the MTBF of each of the components (servo motor, arm linkages, etc.), an MTBF for the complete hard drive can be calculated. If additional components are added to the design, the MTBF for the assembly will decrease due to the fact there are more components that could fail. For example, if all items had an MTBF of 100,000 hours and a new design was developed with twice as many components, each having the same 100,000-hour MTBF, the MTBF of the entire unit would be reduced by half because twice as many components could fail. This is analogous to the situation of two different plants in which all pumps have the same MTBF, but one plant has twice as many pumps as the other. The plant with more pumps would have an MTBF 50% lower than the other plant. This makes it extremely difficult then to compare MTBF values from one site to another. If a group of new pumps is added to the site population, this also tends to decrease the MTBF. So this method is best used when no outside comparisons are to be performed, and at plants where the equipment population is stable. Calculating a pump or service MTBF is mathematically simpler and more intuitive to most people. In addition, since the calculations involve the number of repairs per unit time, the effects of continuous improvements are easily captured. This is because, due to the improvements, the number of repairs should decrease with respect to previous time intervals. Whether an MTBF should be calculated for individual pieces of equipment or services should be determined by a facility’s equipment configuration. If the majority of pumps do not have installed spares or are operated on a continuous basis in parallel, then the MTBF should be calculated on a per pump basis. This is because the operating philosophy will be centered around a single pump. If a pump is out of service, certain operational issues will arise – either production will be lost or reduced, or

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Pump "A" Only 1 pump operational at any time

Pump "B" 4

130

2

90

270

= Repair

110

Pump "A" Pump "B"

Time (Days) MTBF for Pump A =472/2=236 days MTBF for Pump B =494/3=165 days MTBF for Service =606/6=101 days

Figure 4. Repairs to a typical fully spared pump service

products placed off specification. When the majority of production equipment has installed spares and few pumps operate continuously in parallel, the MTBF should be calculated on a per service basis. Figure 4 illustrates a typical process arrangement that has an installed spare. In this case the plant production philosophy centers on the assumption that the service is always operational to meet production requirements. Whether pump “A” or pump “B” is operational is a question of little consequence as long a one is always operating. As we can see from Figure 4, calculating the MTBF on a service basis better reflects the operational impact of these pumps than calculating the MTBF on a per pump basis. Regardless of whether MTBF is calculated on a per pump or per service basis, a decision must be made on how to handle intermittent duty pumps or services. If MTBF is calculated only for the purpose of site trending, then there is no need to compensate for intermittent duty. The MTBF will trend proportionally to the number of repairs. If the number of repairs is reduced by 50%, the MTBF will double. If MTBF is to be compared from one site to another, it is important that a common basis be used to allow meaningful comparisons. One site may have a pump population in which half of the units operate continuously and the rest operate 25% of the time. Another site may have The Pump Handbook Series

75% of the pumps operating continuously and the remaining pumps operational 50% of the time. How can a meaningful comparison be made? One method is to calculate an MTBF for continuously operating pumps. This concept is similar to statistical polls. The MTBF of continuously running pumps is used as a snapshot to gauge the MTBF of each site. When this approach is taken, a list of such pumps needs to be generated and maintained. Another method is to determine an equivalent number of pumps for each site by normalizing each site pump population to an equivalent number of 100% duty pumps as shown in Table 1. This is not an exact science because the operating personnel must judge how often a pump is utilized. It does, however, provide a more consistent operating basis for comparison purposes. Both approaches to intermittently operating pumps can be applied to calculating an MTBF for either individual pumps or pump services.

Defining a Valid Repair Having developed criteria for determining the pump population, let’s look at how we can define a valid repair. Maintenance work can be divided into two categories: 1) restorative activities (replacing components) and 2) nonrestorative activities (changing oil, greasing bearings or checking alignment, etc.). Some plants use both restorative and nonrestorative maintenance

Typical Restorative Repairs

Typical Non-Restorative Repairs

(Components are Replaced)

(Consumables are Replaced or Components Adjusted)

Replacing Seals Replacing Bearings Replacing Coupling Replacing Wear Rings Replacing Cover Gasket Replacing Bowls or columns Replacing Diaphragms Replacing Rotors or Stators Replacing Shaft Replacing Packing Replacing Worn Lip Seals etc.

Grease Bearings Changing Oil Greasing Couplings Checking Coupling Alignment Re-alignment of Equipment Adjusting Impeller Clearance Tightening Packing Clean Suction Screen Adjusting Belt Tension etc.

Table 3. Examples of restorative and nonrestorative maintenance activities

Work Order 111234 111265 111290 111374 111456 111531

Issue No

Pump

01-Jan-97 23-Jan-97 02-Feb-97 15-Feb-97 01-Mar-97 12-Mar-97

28-P1003-B 29-P2001-A 16-P5003-B 31-P7021-B 28-P1003-B 28-P1011-A

Work Order Description

Work Order Cost

Replace leaking mechanical seal Change lube oil and grease coupling Check pump, low flow Change brgs, hi vibration levels Replace leaking seal Clean suction screen

4,375.00 375.00 125.00 2357.00 4175.00 407.00

Table 4. Typical results of a CMMS work order query

Work Order 111234 111374 111456

Issue No

Pump

01-Jan-97 15-Feb-97 01-Mar-97

28-P1003-B 31-P7021-B 28-P1003-B

Work Order Description Date Cost

Work Order Cost

Replace leaking mechanical seal Change brgs, hi vibration levels Replace leaking seal

4,375.00 2357.00 4175.00

Table 5. Results of filtered CMMS work order query

work as a valid repair for MTBF calculations. Others consider only restorative work as valid repairs. It is important to remember that MTBF includes the word “failure,” so it may be better to restrict the definition of a “valid repair” to restorative maintenance activities. Table 3 lists typical restorative and nonrestorative maintenance activities.

Collecting Data In many facilities, the data used for calculating MTBF can be retrieved from a Computerized Maintenance Management System (CMMS). Usually these systems contain equipment history, work orders and work order costing. The accuracy of equipment history records can vary widely from site to site. However, the work order system usually contains the most accurate and con-

sistent records pertaining to work order issue dates, costs and initial problem identification. Using the work order system, a computer report can be developed and stored to capture the work order number, equipment identification, work order description, and work order cost. It is important to retrieve both the work order description and work order cost to help validate the job as a valid repair. Table 4 lists typical results of a search from a CMMS report. Items such as repacking block valves, changing oil and cleaning suction screens can usually be automatically filtered from the report if they are not considered valid restorative tasks. Also, some work orders entered into a system often show zero or minimal costs. These low cost jobs are typically minor maintenace The Pump Handbook Series

work or work requests to investigate a problem. As such, a minimal cost of 1 or 2 hours of work time may be charged to the order. When reviewing the results of the CMMS report, these minimal cost jobs can usually be filtered out by applying the stipulation that the total work order cost should exceed a specific threshold value of $200 or $300 for the repair to be considered valid. Table 5 illustrates the results of including filters in the CMMS query to reject specific work orders listed in Table 4. These query filters have been entered into the computer program to exclude the work orders that contain one or more of the following: 1) Costs less than $200 2) The phrase “change lube oil” 3) The word “screen” Since many people may enter work order descriptions, various combinations of the same phrase may need to be entered as filters to automatically exclude specific types of work in the automated search for valid repairs. For example, “change lube oil” may need the following variations: 1) “Chge lube oil” 2) “Change l.o.” 3) “Change l/o” Once the CMMS report has been developed, minimal updating is required except to add any new work order description phrases that may be required as filters. Although this automated approach may not capture 100% of the valid repairs, a well constructed query should collect at least 95% of the repairs, which is perfectly acceptable. The concept of MTBF is based upon statistical sampling. Being able to trend and manipulate data based upon 95% of the population is statistically significant.

Determining the Repair Rate Regardless of how a valid repair is defined, a repair rate must be determined. It is possible to use the number of repairs each month to calculate the MTBF value. However, Figure 5 shows how the monthly repair rate can fluctuate considerably. Because of this normal variation, many sites use a rolling average approach – i.e., calculate an average number of repairs per month for a

285

50 45 40

Average =30

35

Average =25 Average =22

Number of Repairs

30

Average =19 25 Average =15 20 15 10

Oct-96

Jul-96

Apr-96

Jan-96

Oct-95

Jul-95

Apr-95

Jan-95

Oct-94

Jul-94

Apr-94

Jan-94

5

Month

Figure 5. Monthly number of repairs for a pump population of 300

period of 12 months. This approach tends to smooth out the repair rate, and it minimizes the effect of plant shutdowns or major plant incidents that may skew the repair rate. Using the average repair rates shown in Figure 5, an MTBF graph for the 300-pump population can be generated at six-month intervals (Figure 6). By simply generating a rolling 12-month average repair rate (current month and 11 preceding months), a monthly MTBF can be generated with minimal effort. Which method should be used to calculate an MTBF? As in many other business questions, the answer depends on how the results are to be used. Calculating and trending equipment MTBF should, as a miniMTBF (months)

25 20 15 10 5 0 Dec-94

Jun-95

Dec-95

Jun-96

Dec-96

Month

Figure 6. MTBF for 300 pump population

mum, enable site managers to monitor the benefits of time and money spent to improve reliability. In companies with similar manufacturing or production facilities, MTBF can be used for internal benchmarking to help identify best in-house practices. However, this necessitates that all plant sites agree on a set of consistent

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definitions for determining the equipment population and valid repairs. Table 2 illustrates how four different but valid MTBF values can be generated using the same data. These values could also be significantly different if they were calculated at four different sites, with different definitions of a “valid repair.” Whichever method is utilized, a CMMS system and the associated relational ability of the database should be utilized to maximum advantage. It should also be recognized that comparing MTBF values with those at other sites or in other companies can be a frustrating exercise, similar to comparing apples and oranges. If comparisons of MTBF from different plants are to be made, it is important to understand the methodologies used. In next month’s conclusion of our two-part article we will review how the selected MTBF methodology can be enhanced to identify improvement opportunities and perform internal benchmarking.■ Ken Noble has held various maintenance positions during the past 15 years, including rotating equipment specialist. He is currently the Preventive and Predictive Maintenance Supervisor at the Petro-Canada Oakville, Ontario refinery and is a registered professional engineer in the province of Ontario. The Pump Handbook Series

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R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

Methodologies for Calculating MTBF – Part 2 In this month’s conclusion to our two-part article, we will review how the selected MTBF methodology can be enhanced to identify improvement opportunities and internal benchmarking. By Ken Noble, Petro-Canada, Oakville Refinery continue to increase. Figure 2 is the continued plot of Figure 1. It shows that the MTBF decreased substantially after April 1994 and started to increase again after April 1995. Will the MTBF continue to increase after December 1996? By plotting the number of repairs used to calculate the MTBSR in Figures 1 and 2 on a Statistical Process Control Chart (SPC) as shown in Figure 3, we could have predicted that the MTBF would not continue to increase or plateau in Figure 1. There 20 were no “out of 15 control” trends from January 10 1992 until April 5 1994. However, 0

Will this trend continue?

Upper Control Limit

Jul-95

Jan-95

Jul-94

Jan-94

Jul-93

Mean

Jan-93

Jul-92

Number of Repairs

Jan-92

12 10 8 6 4 2 0

Figure 3. Statistical Process Control (SPC) chart shows that the MTBF would not continue to increase or plateau.

Jul-96

Jan-96

Jul-95

Jan-95

Jul-94

Jan-94

Jul-93

Pump "A"

Jan-93

MTBF

C

Jul-96

harting the MTBF value displays a history of changes in plant reliability resulting from improvement programs or equipment upgrades. The chart does not indicate if the trend is sustainable. As an example, Figure 1 has a continuously increasing MTBF and may suggest that the MTBF will

from May 1996 to December 1996 we have an “out of control” trend. For eight consecutive months the monthly number of repairs is below the mean. This indicates that a significant shift has occurred (SPC rule of 7), and thus the MTBSR should either continue to increase or plateau around the December 1996 value. Recognizing that this particular population of equipment appears to be continually improving for the years 1994 to 1996, both the mean and upper control limit have been

Jan-96

Part 1 of this article discussed different methods to calculate an MTBF value and how the information required to do so can be extracted from a Computerized Maintenance Management System (CMMS). Part 2 will discuss how the information collected to generate the MTBF can be reorganized to make identification of problems more precise.

Only 1 pump operational at any time

Month

Pump "B"

Figure 1. Plot of MTBF for 50 Pumps in Process Unit No. 1. 4

12

Will this trend continue?

130

2 90

270

MTBF

10

110

= Repair Pump "A"

8 6

Pump "B"

4 2 Time (Days)

Jul-96

Jan-96

Jul-95

Jan-95

Jul-94

Jan-94

Jul-93

Jan-93

0

MTBF for Pump A=472/2=236 days MTBF for Pump B=494/3=165 days MTBF for Service=606/6=101 days

Month

Figure 2. Continued plot of Figure 1.

Figure 4. Repairs to a typical fully spared pump service. The Pump Handbook Series

287

recalculated, ensuring that at least 24 values have been included. Many facilities have 100% capacity pumps installed as spares to prevent business interruption during maintenance repairs. In these situations it is important to make sure that one pump is always operational. Referring to Figure 4, we see that the minimum MTBF for this example is 101 days. Although this may not be a great value, it does not show how really unreliable this service is in that two consecutive repairs occurred in close proximity on two separate occasions – a 2-day separation and a 4-day separation. The effect of these close proximity failures should not be trivialized. From a plant production prospective, each pump can have an excellent MTBF. However, operating personnel are not concerned about performance of the individual piece of equipment. Instead, they worry about how reliable the service is. If a service has failures in close proximity, operating personnel are usually reluctant to allow a pump to be unavailable for an extended period of time. This is due to the increased probability that the second pump will fail while the first pump is being repaired. This can have a profound influence on both repair costs (work may be completed on a overtime basis) and scope of repair (insufficient time to perform the full repair). If the equipment is not properly repaired, it will typically be even less reliable. Decreased reliability can increase future repair costs and reinforce operational concerns, and it may strain relations between the maintenance and operations departments. Table 1 shows how the data retrieved from the CMMS that is

288

This could indicate that targeted solutions have either not been implemented or not been successful or that new problems have developed in REPAIR 16-P-10B 4289 REPAIR SEALS ON 16-P-10B 1157 the plant. A significant reduction in CHECK EFFICIENCY 16-P-10A 1121 the number of service repairs of less REPLACE PUMP CASE 16-P-10A 12399 than 1 month is shown in Figure 7. CHK/RPR CASE/SEAL CON. 16P10B 7510 Although the MTBF in Figure 7 16-P-10B LOST EFFICIENCY 5670 is starting to approach a fair value, it REPLACE SEAL ON 16-P-11A 6174 does not show that 25 out of 82 ser16-P-11A REPAIR SEAL 5271 vice repairs (30%) occur within 3 REPAIR SEAL 16-P-11A 3715 months of each other. This does not CHECK THRUST BEARING 16-P-11A 3382 mean that the same pump failed REPAIR SEAL 16-P-11A 5824 within 3 months. It does indicate that having performed a repair on one pump in a spared service, 30% of the used to calculate the MTBF can be time either the same pump or the downloaded to a spreadsheet proother pump will have failed within 3 gram and grouped by service. Repair months. This can have a significant intervals can be generated for each impact on the risk an operations service once the services have been department will assume relative to grouped and the repair dates sorted. repair time and repair scope. Note (160 Pumps) 20 that repair interNo. Repairs in 1994 = 103 15 1994 MTBF = 160/103 = 1.55 yrs val information cannot be deter10 mined by plotting 5 a graph of MTBF 0 vs. time. Generat0 5 10 15 20 25 30 35 40 45 ing service repair Service Repair Intervals (Months) intervals on Figure 5. 1994 pump service repair interval distribution unspared equip(160 Pumps) ment will indicate 20 Number of Repairs in 1995 = 82 how many pumps 15 1995 MTBF = 160/82 = 1.95 yrs have short repair 10 intervals. 5 A progression of service repair 0 0 5 10 15 20 25 30 35 40 45 histograms (FigService Repair Intervals (Months) ures 5 to 7) clearly quantifies the Figure 6. 1995 pump service repair interval distribution fact that improve(160 Pumps) 20 ment programs or No. Repairs in 1996 = 58 15 point solutions 1996 MTBF = 160/58 = 2.75 yrs are being targeted 10 to maximum ben5 efit by reducing 0 the number of 0 5 10 15 20 25 30 35 40 45 50 Service Repair Intervals (Months) short repair intervals. This cannot Figure 7. 1996 pump service repair interval distribution be shown using an MTBF trend chart. As illustrated These repair intervals can then be in Figure 6, the MTBF increased, but extracted for a given time frame (12 the number of service repairs lasting months calendar year, for example) less than 1 month did not decrease. and plotted using a histogram funcThe calculation of repair intervals tion. can also identify which pumps or Figures 5 to 7 show the repair services have frequent repairs. Those intervals plotted in a histogram forservices with frequent short repair mat on a periodic basis. Although the intervals can be turned into a “bad MTBF in Figure 6 has increased comactors list.” pared to Figure 5, there has been no Although additional time is reduction in the number of service required to calculate the repair interrepairs lasting less than 1 month. Work Order Description

Number of Occurrences

Work Order Repair Order Interval Issue Date (mnths) 16-P-10 16-P-10B 16-Oct-91 16-P-10 16-P-10B 17-Oct-91 0.03 16-P-10 16-P-10A 19-Nov-92 13.30 16-p-10 16-p-10A 2-Feb-93 2.50 16-P-10 16-P-10B 7-Apr-93 2.13 16-P-10 16-P-10B 4-Jan-95 21.23 16-P-11 16-P-11A 31-May-93 16-P-11 16-P-11A 24-Aug-94 15.00 16-P-11 16-P-11A 17-Jun-96 22.10 16-P-11 16-P-11A 22-Jul-96 1.17 16-P-11 16-P-11A 12-Aug-96 0.70 Table 1. Service repair intervals

Number of Occurrences

Service Code

Number of Occurrences

Service

The Pump Handbook Series

Order Cost

1. Cold Pump MTBF: Site MTBF query + [operating temp greater than “T1” and less than “T2”]

Site or Plant MTBF

Category “A” MTBF

Category “B” MTBF

Category “C” MTBF

Figure 9. Subdivision of site or plant MTBF

For example, adding the following conditions to the overall MTBF query will allow specific category MTBFs to be generated:

25

xx xx xxx x x x x x x x x x x xx xxxxxxx xxxx x x 15 x xx xxxxxxxxxxx xxxx Target equipment in this category to increase site

20

Jul-96

x Cold Pumps Hot& Cold Pumps Hot Pumps

Oct-96

Apr-96

Jan-96

Jul-95

Oct-95

Apr-95

Jan-95

Jul-94

Oct-94

Apr-94

Jan-94

0

Oct-93

+ + ++ 10 +++++++++ + ++++++++++++ + + ++ + +++ 5 + ++++++++++++ +++ Jul-93

vals, there is no need to generate these histograms frequently. The graphs can be calculated every 6 or 12 months because of the time required to identify the problem, implement a solution and demonstrate a measurable improvement. Often a plant will concentrate only on its bad actors, and as each bad situation is resolved, a new one is added to the list. However, from Figure 8 we can see that a significant number of site repairs may still be associated with pumps not identified as bad actors. The quickest way to increase site reliability and decrease maintenance costs would be to improve pumps both on and off the list. One way to identify where upgrades or improvements should be targeted for equipment not on the bad actor list is to subdivide the overall site MTBF into various categories. Depending on the plant, the basis for subdividing can be process parameters or operating units. Figure 9 illustrates how the site or plant MTBF can be sectioned. Like the calculation of the overall MTBF, the data used to calculate the category MTBF can be obtained using a Computerized Maintenance Management System (CMMS). Adding additional criteria to the query used to generate the overall site MTBF will usually generate the data.

Apr-93

Figure 8. Typical relationship of “bad actors” and total site population

3. Area MTBF: Site MTBF query + [Unit = “XX” + Unit = “YY”+Unit = “ZZ”] Figure 10 illustrates how an overall site MTBF can be subdivided based on temperatures. For the sake of discussion, we will assume that the bad actors are distributed evenly between hot and cold pumps and have the same effect on both MTBF values. If the site MTBF had not been subdivided, the problems with the hot pumps would most likely not have been identified.

Jan-93

Silent Majority (80% of Population) (60%-40% of Repairs)

MTBF (months)

Bad Actors (20% of Population) (40%-60% of Repairs)

2. Low Gravity MTBF: Site MTBF query +[specific gravity greater than “X” and less than “Y”]

Month

Figure 10. Comparison of overall MTBF and category MTBF values 30 MTBF (months)

Total Site Population (100% of Repairs)

equipment in the category. Solutions, therefore, lend themselves to a program approach whereby the upgrades or improvements are applied to all or a majority of the pumps in the category. These upgrades can be either component changes – different seals, seal flushes, for example – or changes in alignment, startup or operating procedures. Just as with overall site MTBF values, category MTBF values can be compared within a company to provide internal benchmarking and identification of best practices. In addition, they can be used to determine realistic targets for the plant MTBF value. In the example shown in Figure 10, a potential realistic target for plant MTBF value is the cold

Component Changes Initiated

25

20 months or 2 x initial MTBF

20 15 10 5 0

0

5

10

15

20

25

30

Time (months) Figure 11. Effects of category improvement programs on MTBF values

A review of either the equipment repair histories and/or work order descriptions obtained from the CMMS system should be conducted whenever a category MTBF is found to be significantly lower than the plant MTBF or a benchmark MTBF. The review can be used to identify any common failure modes or component failures that may be lowering the MTBF value. Unlike the bad actors, which usually require individual solutions, low category MTBF values can often be attributed to one or two problems common to all The Pump Handbook Series

pump MTBF value. An analysis should be performed to understand the factors that are preventing the hot pumps from achieving or exceeding the cold pump MTBF value. In addition, the difference in MTBF values between the hot and cold pumps can be transformed into realistic maintenance cost savings that can be used to justify equipment changes or upgrades if required. Category MTBF values do not need to be calculated on a frequent basis. As with the services with short repair intervals, it takes time to iden-

289

tify the common problem or problems, implement solutions and see results. Typically these values can be calculated quarterly or semi-annually. If the values are to be used for internal benchmarking, quarterly results may be most beneficial for comparison purposes. Figure 11 illustrates the time lag between initiating a category improvement program and demonstrating results. For this specific category of pumps,a program was instituted to upgrade globally certain components each time a pump was repaired. No pumps were removed from service specifically to install the upgraded components. In this example, it took several months before most pumps were upgraded (original MTBF was 10 months). Once the upgrades were installed, it took about 10 months to influence the category MTBF value. In this month’s part of this article, we have examined how the data obtained to generate a basic MTBF can be further analyzed and expanded to gain increased insight as to where money and efforts are best expended to gain maximum returns.

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Displaying the data using an SPC chart or generating histograms requires only that the data be displayed in a format different from a plant MTBF trend. Additional formats indicate whether the changes in MTBF values are sustainable or not, and if the number of short repair intervals is being significantly reduced. By further utilizing the information in the CMMS system and developing additional relationships involving material, operating conditions and operating areas, one can generate category MTBF values. Comparison of these values can help identify common problems that are not addressed by the bad actor list or identified in plant MTBF trends. The trending of plant MTBF values indicates the effects of past improvements. Trending the plant MTBF is a reactive method to understanding pump repair rates. If it plateaus or trends downwards for a period of time, questions will be asked. The MTBF trend will provide no insight into the factors affecting equipment repairs.

The Pump Handbook Series

However, utilizing a combination of SPC, repair interval histograms of category MTBF values and monitoring the overall MTBF value is a proactive approach to understanding factors that affect pump repairs rates. Additional analysis of this kind ensures that the factors influencing the trend are understood and monitored. When they are justified, additional programs can be implemented to increase plant MTBF further. Internal benchmarking of category MTBF Values can also identify specific opportunities for improvement and aid in justifying additional resources or funding if required. The author wishes to acknowledge the input of Mr. M. Warman in the preparation of this article. ■ Ken Noble has held various maintenance positions during the past 15 years, including rotating equipment specialist. He is currently the Preventive and Predictive Maintenance Supervisor at the Petro-Canada Oakville, Ontario refinery and is a registered professional engineer in the province of Ontario.

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R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

Packing & Rotating Equipment Attention to detail and a simple, effective procedure will dramatically improve the life of your packing. William “Doc” Burke, A. W. Chesterton Company hen you consider the number of fine articles about mechanical seals that have appeared in this magazine, you would have to believe that seal packing is used very little in the industrial world. The fact is that more than half of all pumps in service today, as well as many mixers, agitators and other rotating equipment turning right now, are packed. Packing is still in wide use today. Here are a few key points to consider whenever packing is the sealing device used.

W

Lubrication The life of any packing in service on rotating equipment is determined by the amount of lubrication in the packing itself. The volume of lubricant manufactured into the packing and the material’s ability to keep it in the stuffing box are the major factors in how long the packing will work. End users must strive to keep lubricants in their stuffing boxes. For virtually all synthetic materials, the amount of lubricant that can be introduced into any packing style is somewhat limited. There is a fairly simple reason for this. Many natural fibers that were the materials of construction for packing in the first 85 years of this century were very absorbent. Cotton, jute and hemp were soft and naturally fibrous, allowing lubricants and blocking agents to be drawn into packings during various manufacturing steps. Anyone who has ever

Chesterton 1727 Multi-lon® Packing

pulled apart a cotton ball gets a clear mental picture of how many voids in the material there can be, and these, of course, can become filled with lubricants and blocking agents. By the same token, anyone who has attempted to pull apart a strand of monofilament fishing line has a clear mental picture, and quite possibly a physical scar, from handling this relatively hard string. The very thing that makes the synthetic tough and durable makes it a poor material to absorb substantial amounts of lubrication. Unless synthetic fibers are subjected to a deliberate “texturizing” such as Teflon™, Kevlar™ and fiberglass, their smooth surfaces will not allow a great deal of lubricant to adhere to individual strands of packing fiber. Figure 1 will help you and your vendor determine what packing fibers, lubricants and packing styles you should use. For this reason packing manuThe Pump Handbook Series

facturers pass synthetic fibers through a lubricant “bath,” most commonly a PTFE (polytetrafluorethylene) colloidal or graphite emulsion, prior to the weaving process. These coated fibers are then spun and woven together to get a thicker bundle of fibers that again go through the lubricant bath prior to being braided to create a cross-section of packing. Along the way a number of immersions may be utilized to drive lubricants and blocking agents further into the packing. Packing manufacturers go to great lengths to put this lubricant into their products.

Installation Packing a pump is becoming a lost art – a craft that is leaving with many of the older employees over the years. With the lack of proper installation technique and training, younger workers may well shorten the potential life of packing material. Attention to details and a simple, effective procedure will dramatically improve the life of your packing. Before new packing is installed, the stuffing box on the pump should be carefully inspected and cleaned out, and any loose corrosion and junk should be removed. This includes the lantern ring. Many people are stunned to learn that the lantern ring is designed to be removed, as are the two rings beneath it in the stuffing box! They should be removed every time the pump gets re-packed, not once a cen-

291

tury! Failure to remove the lantern ring and the two rings behind it inevitably leads to the lantern ring becoming situated so that the flush line is not properly lined up with the lantern ring. The resulting flush is very ineffective. Inspect the sleeve and stuffing box areas closely. Start the re-packing process with clean equipment. If the sleeve is badly worn, replace it. If the stuffing box bore is rough and corroded, carefully clean out any loose corrosion. If gland studs are corroded, replace them. Cut individual rings of packing on a properly sized mandrel, one at a time, with a sharp knife; cutting several rings at once assures that some of the rings will be improperly sized, and there will begin a process where cut rings are re-cut again and again before they are finally thrown away. Cutting one ring at a time takes more time but wastes less material and ultimately reduces labor costs. Butt cuts and skive cuts are acceptable for rotating equipment, but butt cuts are easier to make well and consistently. To be effective, packing must be square to the shaft sleeve and both flat and square at the bottom of the stuffing box. Installation of every ring should be done with a simple tamping tool to be certain that every ring is seated against the previously installed ring. Be certain that the joints are properly mated and that each is staggered 90 degrees from the

previous ring’s joint. Getting the first ring properly set is crucial to making the rest of the packing work. Anyone who packs valves for long-term, reliable service uses a tamping tool for every ring in the valve stuffing box, taking time to do as perfect a job as possible. This is how a pump should be packed as well. If the first rings are not properly positioned, none of the following rings can properly transfer load with minimal force. The result is that crafts people are sent out to snug up on the packing. Inevitably the packing becomes over-tightened and dies a terrible and lonely death, hemorrhaging product all over the pump and base. Once the stuffing box is full of the properly sized and seated packing rings, move the gland follower into position and turn the gland nuts to finger tightness. Use no wrenches, please! Packing materials swell once the liquid is introduced into the stuffing box. This is particularly true of Teflon. Having the gland too tight assures that the packing will generate too much heat at break-in/startup.

Flushing Flushing is used on packed pumps/rotating equipment for one of two reasons. The first is to prevent air from entering the system if the stuffing box is under vacuum. The second is to extend packing life

1. Fluid being pumped?___________________ Ph_____________ Temperature_____________ Pressure_____________ 2. RPM___________________ Feet/Minute___________________ 3. Condition of sleeve and pump___________________ 4. Pump Dimensional Data Shaft/Sleeve Diameter___________________ Stuffing Box Bore_____________ Packing Cross-Section_____________ Stuffing Box Depth_____________ Lantern Ring? Yes or No_____ If yes, width___________ 5. What packing are you currently using?__________________________ Materials of construction?_____________ 6. How is that packing working?__________________________ How long between re-packs?__________________________ How often is packing adjusted?__________________________ 7. Are you Flushing? Yes or No_____ Flush Rate (GPM)_____________ Flushing Pressure_____________ Stuffing Box Pressure_____________ 8. How often is the system cleaned?__________________________ With what?__________________________ Figure 1.Packing application questionnaire

292

The Pump Handbook Series

Conventional packed stuffing box with a lantern ring.

because the fluid being pumped has abrasive solids. If flushing is required, start the flush first, being certain that the flush is 15-20 psi higher than the stuffing box pressure. (See Figure 2 for how to calculate stuffing box pressure.) Flushing at high pressure, or with high volume, washes a tremendous amount of the lubricant out of the packing, so reducing flush pressure or flow will probably double the life of your packing. If you can’t properly control the pressure differential, install a flowmeter and control the flushing volume. It is common practice to take a line off the discharge of a pump and back into the stuffing box. This is a poor arrangement from a packing standpoint. Though this environmental control is sometimes used with a mechanical seal, it is used in those applications to keep a product from flashing or vaporizing. You should not be packing those applications. A discharge re-circulation line on a packed pump introduces a higher percentage of solids at higher pressure into the stuffing box; this makes sense only if your stuffing box is under a slight vacuum and you cannot tolerate any product dilution/ contamination from an outside flush source.

Startup Open the suction valve(s), flood the volute and vent the system and stuffing box, open the discharge valve(s) at least 25%, and start the pump, allowing it to leak freely for fifteen minutes. Monitor the leakage and adjust the packing follower one flat of the gland follower nuts at a time, allowing twenty to thirty min-

utes to go by before making each successive adjustment. Continue until the leakage slows to an acceptable rate – about twenty to thirty drops each minute as the minimum. Less leakage than this indicates over-tightening, which will build up heat and glaze the inside diameter of the packing. Once the packing I.D. is glazed, you might as well just unpack and start all over because the packing will never perform properly. This may seem like a lot of leakage, but rotary packing is designed to leak, and the leakage draws away the heat built up in the stuffing box. For higher speed pumps, this heat buildup is even more dramatic and of greater concern. A one drop per second leak would typically result in six ounces of loss per hour. If this much leakage is unacceptable, convert to a mechanical seal. In discussing the break-in phase I mentioned that onehalf this leak rate, about twenty to thirty drops per minute, is an acceptable minimum rate. The reality is that many pumps using packing today are leaking at a rate of a gallon per minute or more. Place a measuring cup under any single stuffing box (observing necessary safety precautions) and keep it there for thirty seconds; double that and you’ll know how much is leaking every minute from that stuffing box.

Injectable Packing In recent years, new products that partially replace conventional braided packings of old with a puttylike material have reached the rotating equipment marketplace. These packings typically have a high content of Teflon or graphite and are made to have a cookie-dough consistency. This material has to be contained within at least two rings of conventional packing. There are two significant benefits to those who can use this injectable material. First, it is designed so that it can be injected while the pump is still in service. When the pump starts to leak at more than an acceptable rate, more material can be added to the stuffing box through an opening installed directly into the lantern ring port without taking anything apart. Second, the amount of leakage required to dissipate heat from the stuffing box is dramatically reduced because

the friction from conventional braided packings is greatly eliminated by the new materials. Do they work? Some of the time. Are they worth evaluating? Where they are applicable, yes. If you have a water-based application that is not going to push the material past its design limitations, you owe yourself an examination of the picking’s performance. It may save you a lot of water/flush and downtime, both of which are increasingly expensive in today’s marketplace. It is not the universal solution some would like, but it can definitely save you money on a long-term basis with a minimum investment. Most companies will charge you for time and materials to try it on as many applications as you would like prior to any investment on your part. It may even help you “limp along” until a turn-around on a badly leaking piece of equipment.

Leakage and Bearings Improved bearing protection over conventional lip seals is imperative if your bearings are going to last. Much of this packing leakage is winding up in your bearing housing. Packing leakage and the old wash down hose are contaminating the pump’s lubrication right through the worn out lip seals. The vast majority of lip seals are worn out in less than

three months of continuous service. There are a number of good vendors who can easily upgrade your equipment to full face seals or labyrinth seals, either of which is a significant improvement over the rubber lip seals, which are also scoring your shafts. Mobil Oil and others have stated for years that 0.002% water cuts the service life of oil 48%**. You can see this oil/water emulsion in your own shop every time a packed pump’s bearing housing is drained. Ask your packing vendor to hold a class in your plant on how packing should be installed, in both rotating equipment and valves. Have him explain the different materials he is selling you and how he can help you extend the life of packing. If he is unwilling or unable to do this for you, you need a better packing vendor.■ ** January 1978 Lubrication Engineering Journal, published by the American Society of Lubrication Engineers, presented at the 31st annual meeting of ASLE in Philadelphia, May 10-13, 1976. William “Doc” Burke is a fluid sealing specialist for the A. W. Chesterton Company in the Chicago area, with nine years of field experience.

STUFFING BOX BORE – SHAFT/SLEEVE O.D. – TOTAL PACKING SIZE TOTAL PACKING SIZE divided by 2= PACKING CROSS SECTION CALCULATING STUFFING BOX PRESSURES: DIFFERENTIAL = DISCHARGE NUMBER – SUCTION NUMBER Pumps with Impellers with Back vanes: 25% of Differential Number + Suction Number Pumps with Impellers with Balance Holes: 10% of Differential Number + Suction Number Split case, single stage pumps: Stuffing Box Number is the same as Suction Number PSI= FEET OF HEAD X SPECIFIC GRAVITY ÷ 2.31 FEET PER MINUTE = RPM X SHAFT/SLEEVE DIAMETER ÷ 4 E=MC2 To obtain Inches, multiply centimeters by 0.3937 to obtain Centimeters, multiply inches by 2.540 Figure 2. Calculating stuffing box pressure The Pump Handbook Series

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Which Alignment Method Is Right for You? By Jack Bolam, Vibra-Con, Inc.

here seems to be a question about alignments. Which is it, shaft alignment or coupling alignment? Let’s talk a little about this. What are we trying to do? Are we not attaching two pieces of rotating machinery so they will operate properly for a long time? We are looking for reliability. One definition of reliability I use is “zero unexpected failures.” In other words, no surprises. What must we do to get reliably operating machinery? For one thing, we must precision align the units so no undue stress is introduced. Precision alignment means that while the units are at design operating temperatures and pressures, there is little to no misalignment at the shaft centerlines. The answer to the question above, therefore, is neither – that is, neither shaft alignment nor coupling alignment in itself. What we are really attempting to achieve is shaft centerline alignment. Do we agree that a precision shaft centerline alignment is needed for reliability? What steps must be taken to obtain precision? I use a number of tools to prepare for alignment. I check the pump while it is still operating. Vibration amplitudes and frequencies, temperatures and general operating and

T

294

ambient conditions all give clues as to what must be done to attain precision alignment. After the unit is shut down, I conduct a new set of checks. The components – coupling hubs, shafts, keys, bolts, nuts and base – are all examined for damage. Cleanliness is also a primary consideration. I use a checklist so I know that everything has been considered. I’m now ready to start the alignment process. Which method should I use? Let’s talk about the choices I am most familiar with – straight edge, rim and face, reverse indicator and laser. With each of these, a certain amount of precision can be obtained if the proper guidelines are strictly followed.

Straight Edge Method In my opinion, the “straight edge” method is the least precise and most commonly used method. It is fast. With the straight edge you are aligning the coupling’s outer surfaces. This method is often successful in removing offset misalignment while totally neglecting angular misalignment. The adjustments made using a straight edge are usually a combination of guesswork and experience. Another version of the straight edge method combines the straight edge with feeler gauges. This tends to The Pump Handbook Series

get the angular misalignment somewhat closer, especially for close-coupled pumps where you can use the feeler gauges to measure the face misalignment directly. The adjustments are still guesswork, however, because the numbers are not plotted on graph paper. You run into a number of other problems when using a straight edge. Irregular surfaces or poorly machined surfaces throw off the readings. The coupling hubs could be bored off-center or crooked, or the machine can have a bent shaft, and you won’t detect either of these problems using the straight edge. Also, trial and error adjustments tend to produce a “that’s good enough” attitude and therefore a lack of precision. The only time I use this method is when I want to obtain a rough alignment after removing one or both machines from their base.

Rim and Face Method A dial indicator is used with the “rim and face” method. A bracket is fastened around one shaft or coupling hub with a rod extending to the other shaft or hub. The dial indicator is mounted on the rod to read from a smooth surface on the hub, the rim. Alignment readings are taken at 90 degree positions around the rim – usually zero at 12 o’clock with read-

ings taken at 3, 6, and 9 o’clock. The indicator is then adjusted so it is reading from the face of the coupling hub. Successive hub readings are taken in the same way as the rim readings. Since the rim reading gives you the offset misalignment and the face reading gives you the angular misalignment, both readings are required. A great deal of error is introduced in the above readings by irregular surfaces. One shaft is turning while the other is stationary. A better version of this method leaves the shafts coupled together with a post mounted to the second shaft so the indicator reads from one spot while turning both shafts. To obtain the most precision with this method, one should make a graphical plot. In most cases, this is not done, so these adjustments are also trial and error. Each approach to alignment has problems that must be accounted for to obtain accurate readings. The rim and face method is no exception. Whenever brackets are used, they must be tight. The bracket for the rim reading sags, and this deflection must be measured and accounted for in the final readings. The sag does not affect the face reading, but shaft location does. For each face reading both shafts must be moved manually to the position they were in for the previous reading. Coupling backlash will also affect the readings. A piece of shim in the coupling will usually solve this problem. I have used this method many times but I only use it when necessity demands because of the uncertainty involved in thrusting the shaft for each reading. I then use

the version of turning the shafts together.

Reverse Indicator Method The “reverse indicator” method uses two dial indicators at the same time for taking rim readings. The indicators are mounted on each side of the coupling. The shafts turn together. The brackets may differ from one to another by the location of the indicators. The basic premise of this method is to locate two points so a straight line can be drawn through them. This line is the actual location of the misaligned shaft centerline. Graphing should be used to determine the amount of corrections needed. This method gives a high degree of precision. Again, certain pitfalls must be avoided. Brackets must be tight. The brackets for both rim readings sag. The sag must be measured and accounted for in the final readings. Any coupling backlash should be corrected. This is the method I use all of the time. I also teach it.

Laser Method Rather than mounting two indicators, the laser is mounted on one side of the coupling, and mirrors are mounted on the other. The laser shoots a beam of light across the coupling to the mirrors, which reflect the beam back to a receiver/photo detector. The photo detector then locates the beam on a fine grid, and the electronics do the rest. Horizontal movements and vertical shim changes appear on the screen. Plotting the laser readings will help avoid erroneous moves.

The Pump Handbook Series

There are also some areas to watch out for while using a laser. Looseness and coupling backlash will cause errors. Heavy dust, steam or extreme heat can bend the light beam. Extreme cold or heat and low batteries will affect the electronics and cause errors. A lack of basic knowledge of alignment will lead to a lack of precision. I have used and taught the use of the laser, but I still prefer the reverse indicator method. Regardless of the method you use, there are certain details that must not be neglected if you hope to attain reliable alignment. A machine cannot be precision aligned if there is soft foot present in the support structure. A machine is going to change between the cold shutdown condition and the design operating condition. Thermal growth or shrink should be taken into consideration. The machinery should be deliberately misaligned to take the known startup changes into consideration. This is what coupling tolerances are for. I hope this discussion has shed a little light on the alignment process. Use the method that meets your needs, but do it with reliability in mind.■ Jack Bolam is President of VibraCon, Inc. (Loveland, CO), a consulting firm specializing in machinery alignment. He is a featured speaker at PumpUsers Expo ‘98 in Cincinnati (October 28-31, 1998) and can be reached at (970) 663-2118, e-mail [email protected]

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R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

Troubleshooting Liquid Ring Pumps By Joe Aliasso, Graham Manufacturing, Inc.

I

f your plant or process has liquid ring vacuum pumps with problems, then these are some basic troubleshooting guidelines to help prevent or locate the difficulty.

940

421

433

210

940.1 230 110 106 922

107

901.1 357 360.1 901.2

Cavitation Cavitation can be detected by a noise that sounds like marbles rolling around inside the pump. A liquid ring vacuum pump cavitates when the vapor pressure of the sealant fluid approaches the suction pressure (vacuum) of the pump. The sealant begins to boil, and tiny bubbles are produced. As the bubbles form and rise through the sealant, they implode (or pop). The voids that were occupied by the bubbles are immediately filled with liquid, producing shock waves. As more and more bubbles form, these waves get stronger. This phenomenon will cause pitting and deterioration of the vacuum pump internals, thus leading to premature wear and loss of vacuum. To prevent cavitation, you should do the following: • Use a colder sealant fluid. This will lower the vapor pressure of the sealant and keep it from boiling. Remember that cavitation occurs when the sealant fluid vapor pressure approaches the suction pressure. • Use a different sealant fluid. If a sealant fluid with a low vapor pressure, such as an oil or ethylene glycol is used, then it will allow the liquid ring pump to operate at low absolute pressures without cavitating. The

296

923 940.1 940 923 922 920 905 903.1 901.2 901.1

360

505

Impeller Key Drive Key Bearing Locknut Impeller Locknut Tie Rod Nut Tie Rod Drain Plug Bearing Cap Bolt Bearing Seal Housing Bolt

400 1 1 1 1 8 4 3 4 8

137.4 505 433 421 400 360.1 360 357 320 230

903.1 Abutment Ring Mechanical Seal Oil Seal Gasket Set (As Required) N.D.E. Bearing Cap D.E. Bearing Cap Bearing Seal Housing Bearing Impeller

137.1

905 920

320

1 2 2 1 1 1 2 2 1

210 Shaft 1 137.4 Discharge Side Plate 1 137.1 Suction Sideplate 1 110 Impeller Casing 1 107 Discharge End Casing 1 106 Suction End Casing 1 VDMA Description Quan.

Figure 1. Cross section of a typical liquid ring vacuum pump

motor size will need to be reviewed when using a sealant other than water. • Increase the pump’s suction pressure (in absolute terms). This can be done by bleeding in an air load. This will keep the suction pressure from approaching the sealant fluid vapor pressure. • Use more sealant flow, gpm, through the pump. This will control or diminish the temperature rise through the pump. This means that the corresponding vapor pressure of The Pump Handbook Series

the sealant will be reduced – perhaps enough to keep it out of the range of cavitation. • Install a booster, such as an ejector or rotary lobe blower. The booster will compress and boost the pump’s operating pressure. This means the suction pressure will not approach the sealant vapor pressure. A sound similar to the sound of cavitation is hydraulic noise. This occurs when the pump’s impeller(s) try to compress the sealant. Liquids are considered to be non-compress-

ible. To check for hydraulic noise, cut back on the sealant fluid flow, but do not exceed minimum flow as recommended by the manufacturer. If the noise goes away or gets quieter, then this is probably the cause. If this does not work, then check for cavitation as explained here or contact the manufacturer for further assistance. To check for cavitation, measure the pump’s suction pressure with an accurate instrument. Bourdon tube gauges lose their accuracy over time and should not typically be used. Also measure the temperature and flow of the sealant. As a rule of thumb, if the sealant is water, then add 10 to 15ºF to the sealant inlet temperature to come up with an average temperature of the liquid ring. This will vary by pump design. Look up in steam tables the vapor pressure of the water at this average liquid ring temperature. If the difference between the suction pressure of the pump and the vapor pressure at the average liquid ring temperature is 5 mmHg or less, then the pump is operating in the cavitation range. To further confirm cavitation, try bleeding in some air. If the noise goes away, then the pump was cavitating, and one of the options explained previously should be chosen to prevent it.

Bearing Problems A liquid ring pump normally produces little vibration. The rotating assembly is mounted off center (eccentric) to the casing whereas the liquid ring is formed concentric to the casing. Typically, the rotating assembly is dynamically balanced, a practice that can minimize any undue vibration. Typical vibration levels for a liquid ring pump are 2 to 3 mils. If the bearing housing is hot to touch, there is a bearing problem. The bearing can be worn from normal use, or sealant could be leaking past the shaft seal and lip seal. If there is corrosion on the bearings or moisture in the grease, then the problem is from a faulty shaft seal (mechanical seal or gland packing) or also from the lip seal (oil seal, flinger, etc.). Make sure that the shaft

seal is being properly lubricated and cooled and that the bearings are properly greased at the recommended maintenance interval. A bad or worn bearing situation should be corrected immediately because this is the shaft support for the pump. Costly repairs can become necessary if the bearings fail (shaft support fails) and the rotating assembly starts rubbing or galling on the pump internals. Another problem that can damage the internals is if the pump is started full of sealant. This puts undue stress on the shaft and can cause serious damage. The pump should be started when the sealant is at the centerline of the pump and also with any discharge valve open. The coupling should be checked for alignment, and the pump should be properly anchored. Mis-alignment and poor anchoring can cause unwanted vibration, which will lead to premature wear and bearing failure.

article can be easily verified or measured in the field. Knowing what causes these problems allows operators to easily identify and correct them before major damage occurs.■ Joe Aliasso is a Product Supervisor, Applications Engineer for Graham Manufacturing, Inc. in Batavia, New York where he has worked for nine years.

Pump Not Producing Correct Vacuum This can be caused by excessive air leakage into the system. Start by measuring the suction pressure of the pump with an accurate instrument. Then measure the sealant temperature and flow. Also verify that the process loading and system design has not changed. If these readings are within design, then excessive air leakage could be the cause of the problem. If possible, measure the air leakage coming out of the vacuum pump with a rotameter. Or perform a vacuum drop leak test on the pump. This will tell how much air load the pump is handling. If excessive air leakage is not the cause and the system design has not changed, then the internals of the pump will need to be examined. If clearances have increased from wear, cavitation, erosion or corrosion, then internal slippage can occur. This means that the gas at the discharge port is allowed to pass back to the inlet port, thus reducing the capacity of the pump, which will affect the vacuum level The problems mentioned in this The Pump Handbook Series

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R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

Replacement of the Boiler Feedpumps at Janschwalde Power Station ..

Fitting new cartridges into existing barrel casings key to successful retrofitting program.

By C. H. Laux, Sulzer Pumpen A.G., Winterthur, Switzerland U. Burchhardt, VEAG Vereinigte Energiewerke A.G.,.. Berlin and U. Kirstein, VEAG Vereinigte Energiewerke A.G., Janschwalde Introduction ..

anschwalde Power Station comprises 6 x 500 MW units that were commissioned between 1981 and 1989. These units, which were equipped with turbo-feedpumps manufactured in Russia, had an average efficiency of only 78% and a mean time between overhauls of 25,000 hours. The challenge was to upgrade these pumps at the lowest possible cost, and to achieve a pay-back time of 4 to 5 years by improving pump efficiency and reducing operating costs. This was accomplished by fitting new cartridges into the existing barrel casings. The cartridges consisted of 6-stages of a well-proven hydraulic with a specific speed of 1700 operating at the nominated turbine speed of 5600 to 5700 rpm. This article describes the processing of the contract within the pump manufacturer’s organization. It also explains the extensive coordination and quality assurance required within a tight production schedule (delivery time of only 7 months before the trial operation of the first pump unit). The efficiency of the modified pumps was determined on site by means of thermodynamic measurement. The specific features of the pumps and the proof of efficiency in the power station by means of thermodynamic measurement are given in detail.

J

principle that two steam generators serve one 500-MW turbine, which is supplied by separate turbo-feedpumps, in so-called duo operation (Figure 1). In 1991 the utility decided to upgrade the units and retrofit flue gas desulphurization systems. Another motivating factor was the requirement to comply with the regulations for large-scale firing systems. Additional measures aimed at improving the efficiency and the availability of these units were also undertaken. These investments mean the station will operate more efficiently for another 20 years.

Existing Feedpump Drawbacks All the units were equipped with PT 850 turbo-feedpumps of barrelcasing design. Although the two 50% turbo-feedpumps for each unit were safeguarded by a parallel connection of 2 x 25% electric feedpumps as stand-by (Figure 1), retrofitting of the

845 t/h 1.84 Mill lbs/h 250 MW Boiler 50% T BFP

25% E BFP #1

..

298

Main Turbine

Generator 500 MW

25% E BFP #2 6 Units each 500MW Block A,B,C,D,E,F

Power Station Configuration Janschwalde Power Station is in the Federal German state of Brandenburg. It’s design is based on the

turbo-feedpumps was considered especially important because of the long average running time of 7500 operating hours per year. In addition, plant managers and operators were seeking a marked increase in pump efficiency, a longer period between overhauls, a dependable supply of spares, and the possible need on an emergency basis to replace the inner casing within a period of 72 hours. An analysis of the existing feedpumps’ unplanned stoppages revealed the following weaknesses: (For principal design features refer to Table 1, Col. 1). a) The disk design of the axial thrust balancing device contributed to excessive wear during transient operation. b) The seal water supply system on the throttle bushing shaft seal was a complicated set-up resulting in high heat losses, and the shaft was prone to bending during transients. c) Thermal problems associated with the shaft seal reduced bearing

50% T BFP

Figure 1. Pump arrangement in the 500 MW Janschwalde Power Station The Pump Handbook Series

support, increased wearing clearances and likely contributed to rotor instability. d) Following several cartridge changes, poor fit of the inner and outer casings continued. With an average period of 20,000 to 25,000 operating hours between overhauls, the utility had to bear the costs of about $150,000 for each overhaul. A further handicap was that for various reasons it became increasingly difficult to obtain original spares from the pump manufacturer. For all of these reasons, the utility decided to investigate various possibilities for retrofitting the feedpumps. The objective was not to change the existing suction and delivery pipe-work or foundations, but to adapt the connecting lines, base frame and supply lines at a minimum cost.

foundations and base frames. In addition, they decided to use a new cartridge with improved hydraulics. The existing barrel casing was to be adapted to the new cartridge (utilizing existing pipe-work connections as well as the base frame of the main pump), and the pipework connecting the booster and main pump would also be altered. The following considerations were pivotal in achieving financial viability for the project: • Investment costs for the new feedpumps • Residual life of the power station installation and the number of operating hours per year • Saving of driver power (converted to the saving on coal) • Overhaul and maintenance costs • Time period for the pump replacements (e.g., in connection with the scheduled stop-pages) • Extra costs associated with the operation of electric feed pumps

Economic Considerations Based on the utility’s assessment criteria and after careful analysis, it was decided not to retrofit the existing main pump, PT 850, and deploy a new booster pump. Instead, managers chose to pursue replacement of the booster and main pump with the latest machine designs that could be adapted to the existing pipework,

BOOSTER PUMP Type No. of stages No. of flows Shaft seals Gear ratio (gear to booster) MAIN PUMP Type Design Cartridge Diffuser Balancing device Shaft seals Bearing span Barrel OD. Barrel ID. Wear ring clearance Impeller diameter D2 Weight of the barrel No. of stages Specific speed nq NSUSA

Under the most favorable conditions, the utility profit statement showed a potential pay-back period of 4 to 5 years. On the basis of this projection, it was decided to implement the retrofit program in April

1 Initial State

2 Retrofitted Variant C (Execution)

HGD-300/2/40 A 2 single flow Stuffing box 1:4

HZB 253-640 1 double flow Mechanical Seals Existing

PT 850-250 Barrel In series with horizontally split case 4-vaned diffuser

HPT 300-315-6s/33A Barrel In series with radially Split case Diffuser with closed channels Piston Mechanical seals

Disc Fixed throttling bushes 7.8 ft (2380 mm) 45.2 in (1150 mm) 28.5 in (725 mm) 170% 100% 11000 lbs (5000 kg) 5 28.8 1500

7.3 ft (2222mm) Existing Existing 100% 94% Existing 6 33 1700

Table 1. Data of existing and retrofitted turbo-boiler feedpumps The Pump Handbook Series

1994.

Contract Conditions Prior to concluding final negotiations, the partners defined minimum technical requirements, scope of guarantees as well as how results would be measured. This included a delivery time of only 10 months to the commissioning of the first feedpumps and further intervals of two months in each case for delivery of two addditional units – a main pump and a booster. In addition, since there was not enough time to dismantle and refit the feedpumps, in order to refurbish existing barrel casings, two casings would have to be available in advance for each installation. To ensure economic viability, an efficiency greater than 84% and an output time of greater than 50,000 hours had to be guaranteed. It was also stipulated that further use of the barrel casing would be acceptable only if three conditions were met: mathematical proof showing a residual life of 20 years; quality assurance steps for material and weld examinations to be carried out in such detail that they would suffice as proof for all the relevant positions (the size of defects and maximum number of indications were to be noted exactly for every critical point), and a new component to be delivered if any casing or pressure cover was rejected.

Pump Selection and Production Logistics In view of the required efficiency and contract stipulations, the pump manufacturer proposed a 6stage cartridge in the existing casing with corresponding adaptations (Photo 1). Although the original pump was a 5-stage design, a 6-stage cartridge could still be fitted in the existing barrel casing (see Table 1, Column 2).

Photo 1. Cartridge of the HPT 300-3156s/33A boiler feedpump

299

Experience with other installations has demonstrated that a hydraulic with a specific speed of NSus = 1700 (≈nq 33) could be employed for the 6-stage pumps. A hydraulic with a higher specific speed than that of the original design was used to increase efficiency. The pump manufacturer was awarded the contract for the retrofit of the 12 turbo-feedpumps in May 1994. The conversion, entrusted to the German Division (SWB), is a good example of international cooperation. SWB was responsible for order processing and commercial aspects of the contract. Since the parent company in Switzerland is responsible for the design and construction of large feedpumps, the draft design work was done there. The pump cartridges were manufactured, assembled and tested by the company’s division in Great Britain. Then they were shipped to the power plant maintenance facility in Germany, where adaptations were made to fit the cartridges to the casing already in place.

PREREQUISITES DESIGN Booster

Main Pump

Flow Qd Head H Speed of Turbine

US GPM (FT) (RPM)

No. of stages Head Speed η Booster D2 No. of stages Head Speed η Main Pump Impeller Dia. D2 Shaft seals Bearing span Clearance sD/nominal cl. Weight

ECONOMY First costs

(FT) (RPM) (%) (IN) (-) (FT) (RPM) (%) (IN) (IN) (%) (LBS)

Initial state Existing pump

Variant A Existing pump

PT 850 (old state)

PT 850 (retrofitted) Sulzer (latest develop.)

Variant C Tender to VEAG Old barrel + new hydraulic

15,000 9,030 5600-5700

15,000 9,030 5600-5700

2 stages/single suction 1 stage/double suction 1 stage/double suction 1 stage/double suction 433 433 433 689 1417 1417 1460 1450 85 85 85 82 25.2 25.2 22.6 22.6 6 6 5 5 8590 8590 8580 8330 5676 5666 5840 5900 85 (guarantee) 85.3 80 78 12.4 12.4 13.2 13.2 Mech. seals Mech. seals Mech. seals Fixed throttl. bushs 87.5 78.7 87.5 87.5 100% 100% 146% 170% 24,230 15,420 24,670 24,670

(%) (%) (%) (%) (%)

Booster-+M.-Pp. Adaptations Variable costs Operating costs Maintenance costs Price for comparison*) Total costs

15,000 9,030

15,000 9,030

Variant B Ideal Choice

30 270 50 350

388 100 488

60 8 12 20 100%

75 18 20 113

Table 2. Possibilities of retrofitting an existing boiler feedpump

Design of Replacement Pumps The new booster pumps (type HZB 253-640) were single-stage radially split units with a double flow impeller to improve the suction conditions. In contrast to the previously installed pumps, which had packingtype stuffing boxes, mechanical seals were employed (Table 1). The internal mechanical seal-water circuit is equipped with two magnetic filters and an intensive cooler integrated around the shaft on the sides of each pump. The main pump (type HPT 300315-6s/33A) is a centrifugal barrel casing design. The suction and delivery branches are arranged vertically facing downwards (Figure 2 ). The bearing span of 7.28 ft (2222 mm) is governed by the existing pump casing – i.e., about 10% more than the span usually needed for a 6-stage hydraulic. From the standpoint of the bearing span, the 6-stage pump corresponds to a 7-stage machine (Figure 3 ) when manufactured to the latest specifications. Consequently, the static sag of the retrofitted pump

300

Figure 2. Cross section of the retrofitted HPT 300-315-6s/33A feedpump

100% L 28% L

32% L

Shaft Seals + Inlet

Stages:

13% L

Impellers

1 ..

2

3

4

Piston

5

27% L Cover + Shaft

6

Figure 3. Bearing span at Janschwalde (new cartridge in existing casing) The Pump Handbook Series

rotor is greater than expected with a new machine. Since the existing casing had been in service for some 15 years, it had to be determined – by means of a residual life analysis – whether the strength of the casing had been damaged by fatigue full load-temperature cycles experienced during previous service. Maximum stress and possible fatigue occurs at the connecting point of the delivery branch. With the aid of Wohler curves for this material, the effect on the fatigue strength of the material was determined from the peak stress – a combination of internal pressure, pipework load and the temperature transients. The analysis indicated that the reduction in fatigue strength was only 4% of the permissible value. In other words, the required life expectancy had barely been affected by the previous operation of the pump. The cartridge design (with internal block consisting of rotor, inner casing, delivery cover, balancing device, shaft seals, radial and axial bearings) makes replacement faster and reduces downtime. The impeller and diffuser were precision-cast with a smooth surface for optimum efficiency. The impellers were shrink fitted to the shaft to guarantee good concentricity and assure sealing of the stage pressure. Hydraulic axial thrust is taken up by a shaft shoulder on the first stage and by split-rings on following stages. Torque transmission is achieved by two opposing keys on each impeller. The axial thrust is largely compensated for by means of a cylindrical, non stepped balance drum, so there can be no touching during transients. Residual thrust is taken up by a tilting pad axial bearing that can be subjected to load on both sides. This ensures sufficient safety, even with maximum running clearances and increased coupling thrust. The piston is shrink-fitted to the shaft and can be easily dismantled with an appropriate hydraulic device. Swirl brakes are employed in the balance bushing of the drum, ensuring good damping of the pump rotor in the event of wear resulting in excess clearance.

A mechanical shaft seal (type 270 F 6.5”) with a material combination SIC paired with syntheticimpregnated carbon has been fitted as shaft seal. The circulation is affected through an externally located cooler and magnetic filter – and without any dosing circuit as is customary for an operation in combined oxygen treatment (feedwater between alkaline and neutral). A cooling chamber is provided to reduce the heat transfer from the hot pump in the direction of the rotating seal ring. The mechanical seals can be dismantled as a complete unit. During operating conditions, the circumferential speed in the middle of the rotating seal ring is about 157 ft/s (48 m/s). Leakage measured in the installation is less than one litre per hour. Following their removal and dismantling, the casing and cover of the old main pump were examined for possible further use, remachined for the fitting of the new cartridge and then subjected to another quality inspection.

Practical Experience during Pump Replacement Only a short time was available for conversion of each pump group. This was linked with and determined primarily by the respective flue gas desulphurization system. Any delay in meeting the deadline was subject to a heavy penalty of 0.1% of the contract value per calendar day. The first pump group had to be ready for operation in the plant on March 1, 1995. Seven months were allocated for design, model-making, the casting process, manufacture, and trial run on site. Quick decisions were therefore required, and needless to say, several of them were based on assumptions. Unfortunately, it soon became apparent that some of those assumptions were incorrect. Two premanufactured barrel casings were always available as forerunners. As is normally the case, it was assumed that all casings and covers were identical and interchangeable. However, during the processing of the contract it was learned that only one cover matched the respective casing in each case. At installation time it became clear that the holes in the cover did not match The Pump Handbook Series

up with the bolt pitch circle. Unplanned adaptations which were not part of the time schedule increased pressure on the project even more and necessitated exceptional efforts by the work team in the service center. Different casings had to be repaired because the ultrasonic and liquid penetrant tests revealed cracks and porous lines in the welds. The cracks were ground out and welded. The delivery branch of one casing had to be cut off and welded anew.

Operating Experience The first pump was commissioned on schedule. It ran smoothly (bearing support-vibration less than 0.08 in/s (2 mm/s) and had “good” vibration behavior in the operating conditions from minimum flow to maximum load – i.e., from 25% up to 100% best efficiency point. After commissioning of the first three pumps, however, pump A2 suffered damage. During the commissioning and functional testing of the drive turbine’s mechanical overspeed safety device, which was conducted in Janschwalde with the pump coupled to the turbine, the unit seized up. This occurred in the run down phase, creating a critical situation in the execution of the contract because the cause of damage was not readily apparent. The pump was dismantled, and scuffing marks were found on the sealing rings of two impellers. An analysis of the cause of damage showed that instead of the proven impeller material G-X5CrNi13.4, the material 17/4 PH (≈G-X7CrNiMoNb 15.5 with 17.5% Cr and 7% Ni) had been substituted for reasons of expediency. This material was paired with the material 1.4138 wnt (soft nitrated)(≈ASTM A743 Gr.CC-50), which was obtained locally and did not meet original specifications. In addition, it was presumed that the scuffing marks had occurred during in-house runs on the test stand. A strip test had not been performed because of the tight time schedule. So the pumps could be quickly put back into service, it was decided to open the clearance of the impeller sealing rings by 20%. Since all the

301

Shaft Vibration µ p/p (2 x smax)

shaft_vi.xis

Limit of Shaft Vibrations Shaft Vibration mils p/p

smax

300

C 200

VDI 2059

8

B 5 4

A

100 80

3 60 50 40

API

2

30

1

20 Veff = VRMs = Xo/p x π x √2 x f

10 1000

2000

Janschwalde HPT 300-315-6s/33A 3000

4000

5000

6000

0.5

7000 10000 Speed (RPM)

Figure 4. Limit of shaft vibration

supplied pumps had been manufactured with the same material pairing, all the subsequent units were provided with corresponding increased clearances. In addition to the larger gap clearances, smooth impeller gaps were employed in these special cases. The larger bearing span conditioned by the use of the existing casing resulted in increased static sag of the shaft, leading to scuffing. Consequently, precise alignment of the impeller in the casing was critical. In operation, pump A2 is now exceptionally quiet. It already has an opened impeller gap clearance. Rather than provide detailed information about the rotordynamic behavior of the pump, suffice to say that the low level of broad-band vibrations indicates that the flow guidance with the pump is optimal. (Figure 4 shows the limit of shaft vibrations.) Other reasons to say this include good casting quality, and the correct choice of the number of impeller and diffuser vanes and all critical dimensions of the impeller. Also, the clearances between the impeller and side walls are within the specified tight tolerances. The high efficiency achieved on test is a further indication of these features. The guarantee values were realized in full. The smooth running of the pump and rotor indicates that

302

continued operation will result in low wear of the clearances, bearings and mechanical seals. Therefore only a slight decrease in efficiency is expected over the operating life of the pump.

Scale-Up Effects of Efficiency The guaranteed efficiencies were based on the efficiencies of other manufactured pumps and a basic hydraulic, which was measured cold in a model machine at a speed of about 1500 rpm. This included the measurement of the clearances, piston leakage and bearing and seal losses. The disk friction of the impellers and the piston can be accurately calculated. From these values, it is possible to determine the hydraulic efficiency. In the full-scale version,

this was revaluated according to the Sulzer formula for impeller diameter D2, temperature and speed (Figure 5). This means that the expected hydraulic efficiency of the full-scale version is known. With the clearances of the full-scale version, and the bearing and shaft seal losses specified by the suppliers, it is possible to determine the attainable efficiency of the full-scale version. Since the pump operates at 95% of its best efficiency point, the actual operating efficiency is slightly lower, and the guaranteed efficiency 85%. In principle, hydraulic efficiency should serve as the basis for all efficiency determination formulas. However, if the effective efficiency measured cold at reduced speed on the test stand is converted with the Karassik formula, the result is usually excessively high values. This is understandable because mechanical and hydraulic losses cannot be converted in the same manner. (The hydraulic efficiency is the efficiency of the stage, i.e., – friction in the impeller and diffuser channels, inlet and outlet losses. It is recalculated by means of the volumetric, mechanical losses and disk friction losses (Figure 5). Figure 6 shows the distribution of the Janschwalde feedpump efficiency, Nsus=1700 (nq=33) at the best point. (Note that 100% minus the sum of the expected losses in efficiency points = the offered efficiency). It is apparent that the hydraulic efficiency of the pump stage is responsible for the greater part. From this it is also clear that a good pump hydraulic can influence overall efficiency the most. For this reason

EFFICIENCY - MAJORATION MODEL - PUMP Basic Perf. Test for Tender D2 = 13.8 inch n = 1500 RPM cold, reduced speed, reduced dia. -Losses η hydraulic

Test Run Full Size (at Manufacturer) Verification of Flow, Head, NPSH D2 = 12.4 inch n = 5584 RPM cold, full speed η = 84% (BEP) calculated +Losses

Thermodynamical Measurement (on Site) Verification of the Efficiency D2 = 12.4 inch n = 5676 RPM t = 338°F site condition η=85% (BEP) calculated +Losses

Majoration of η hydraulic for D2, n and t Figure 5. Method of scale-up effects of efficiency The Pump Handbook Series

Approx. Allocation of Efficiency - Lossess .. Turbine-driven BFP Janschwalde Referred to calculated Efficiency at BEP Qs1=1.23 Points

QS3

PRR 1.48 Points

Hydraulic 9.3 Points

QE 2.05 Points

hydraulic

QE=Balancing Flow

PRR=Disc Friction

Qs1=Large Werar Ring

QS3=Small Wear Ring

Thrust Bearing

2 Mech. Seal

PER=Friction of Piston

2 Radial Bearings (kW)

Figure 6. Approximate allocation of efficiency losses of boiler feedpump

Feedwater Tank

VEAG - Janschwalde

Booster HZB 253-640

Minimum flow line

Minimum Flow Valve

Gear Main Pump HPT 300-315 6s

SteamTurbine

Qmin

+6m

floorlevel di = 307 mm Ps

Td

Ts Balance Line

Pd

Q di = 301 mm

Figure 7. Position of the instruments for the on-site thermodynamic measurement

expensive diffusers with closed channels and spatially curved overflow were employed in this application. Figure 6 also shows that the efficiency is only 25% dependent on the clearances in the labyrinths and the balancing system. The mechanical losses of the bearings and mechanical seal have a rather subordinate influence.

Verification of Efficiency The new cartridge (No. 104) was tested in cold condition at full speed (125ºF, 52ºC), 5676 rpm with normal clearances, on the test bed in Leeds,

Great Britain on December 12th, 1994, just seven months after the work order was received. An efficiency of 84.6% (cold) in the operating point was achieved during the first run. The delivery head was very close to the offered curve within the guaranteed limits. Likewise, the .. NPSH was also maintained exactly. The head with closed valve was slightly higher but still within the guaranteed tolerance of 3%. The main aim of this test was not to determine efficiency but the Q/H/NPSH characteristic of the pump. To be safe, one pump was also The Pump Handbook Series

tested at the Leeds facility under hot conditions (338ºF, 170ºC) and at full speed. A hot water efficiency of 86.2% was determined with the customary measuring instrument tolerance (about 2%). Each aggregate in the power station was further subjected to a separate acceptance measurement that called for each pump to be jointly measured thermodynamically by specialists of the utility and the pump supplier. The purpose of this test was to determine efficiency at operating conditions. The payment of a penalty by the utility is based on this efficiency. To obtain more accurate results, a recorded data collecting system was developed in co-operation with the Innotec research division. It enables continuous measured values to be called up and stored and select certain measurement points for the evaluation (Figure 7). The advantages of this procedure is that all steps are reproducible, and the results can be viewed by all concerned. Data collection took place over a period of about 3 hours, during which all measurements were called up and stored at intervals of 14 seconds (verification according to DIN 1944/I , with a inaccuracy of measurement for efficiency limited to 1%). By plotting the values against the respective period of time, it was possible to see exactly when steady-state conditions occurred. The efficiency was calculated from the average of these values. Despite the increased clearances, the guaranteed efficiency was attained for all pumps, and this was verified by thermodynamic measurement in the plant. It should be noted that the first three pumps (B1, B2, A1) are still operating with the original clearances and are attaining an efficiency of about 85.6%, – i.e., 0.6% above the offered efficiency of 85%. The measured pump efficiency plotted as a function of the wear-ring clearance is shown in Figure 8. The efficiency of the original pump has also been also measured by the thermodynamic method and found to be at only 78%. The measuring tolerance, however, was approximately +/- 2%. By September of 1996, the first new pump had run 10,000 hours. In

303

schwalde Power Station once more for their support.■

HPT 300-315-6s/33A (t = 338°F = 170°C, n = 5676 RPM)

87 86

Cord H. Laux received his degree in Mechanical-Engineering from the Technical University of Braunschweig (Germany) in 1962. Since 1970 he has held various technical and sales related management positions for Sulzer Pump. He is currently a consultant for the company responsible for the selection of pumps in the energy market. Uwe Burchhardt received his degree in Power Station Engineering and in Pipeline Transportation from the Technical University Magdeburg (Germany) in 1990. Since 1988 he has worked for the VEAG Vereinigte Energiewerke AG headquarter in Berlin. He is responsible for pumps and blowers in this plant and process division. Specifically he is responsible for the planning in the above mentioned field for new power plant facilities, including Schwarze Pumpe, Boxberg IV and Lippendorf. Uwe Kirstein received his degree in Thermal Mechanical Engineering from the Technical University Dresden (Germany). Since 1983, he has been associated with the VEAG Vereinigte Energiewerke AG as operation engineer .. in the Janschwalde Power Station. In 1993 he became the head of the machine division there

85 84

Efficiency (%)

83 82

Minimum Efficiency Guaranteed

81 Average Efficiency (Thermodynamically Measured) of the Feed Pumps Installed since 1982 of Russian Manufacture

80 79 78

Efficiency Themodynamically Measured of the Retrofitted Boiler Feed Pumps Installed since 1995

77 Measured Diametral Clearance of the Impeller Wearing Rings sD (% of Normal Clearance)

76 75 90

95

100

105

110

115

120

125

130

135

140

Figure 8. Thermodynamic efficiency measured on site

order to detect a possible efficiency loss during this time, pump A2 (with original clearances) was again measured by the hermodynamic method by experts at the utility. A decrease of only .02% in efficiency was detected. This, obviously, is within the range of the measuring tolerance. If the weighted instrument and measuring tolerance are combined with the tolerances of the values of the steam table (≈ 0.5%), a total inaccuray of the efficiency-measurement with the thermodynamic method results in a tolerance of approximately ± 1%.

Conclusions The customer’s need to retain the existing turbines, fundamental anchorage points and pipework connections was fulfilled by the pump supplier through the delivery of new cartridges for installation in the modified casings. The actual time for upgrading the turbo-driven pump units at Janschwalde was not greater than anticipated, and the power station’s service was not adversely affected by the retrofitting of the feedpumps. The extremely short delivery period from the receipt of order to commissioning (Table 3) called for a high degree of cooperation and involvement among working personnel. Time and again people had to be full of ideas to solve the problems that suddenly occurred. The cooperation of the European divisions of the pump supplier is worthy of special

304

note in this respect. All participants – on both the customer and supplier sides – have contributed their capabilities and showed an absolute willingness to work. Proof was provided of all guaranteed technical parameters. As is shown in Figure 8, the efficiencies guaranteed by the pump supplier have been maintained or even exceeded. In comparison with the existing pumps, the efficiency has been increased by more than 6%. As a result, the investment in replacing the pumps will be recouped within the planned time. The short deadlines and the complete scope of work could only be met and progressed in such a safe manner with the care and attention of good specialists and effective coordination of the work schedule by plant engineers. We would like.. to thank the plant personnel at Jan-

Editor’s Note: This article was reproduced with permission of the Turbomachinery Laboratory from Proceedings of the Fourteenth International Pump Users Symposium, Turbomachinery Laboratory, Texas A&M University, College Station, TX, pp. 15-24, Copyright 1997.

May 94 June 94 July 94 Aug 94 Sept 94 Oct 94 Nov 94 Dec 94 Jan 95 Feb 95 Mar 95 April 95 Award (verbally) Design Patterns Cast + Forged Parts Machining Cover + Casing Machining Assembly Cartridge Test Run Transp. Leeds to Vetschau Assembly in Vetschau Transp. to Janschwalde Connecting Main Pump. B1 Commissioning

7 Months

10 Months

Table 3. Time schedule for the first pump (B1) The Pump Handbook Series

✓

R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

The Nuts and Bolts of Vibration Signature Analysis Learn how it is used to determine the source and location of abnormal machinery vibration. By Cliff Hammock, Industrial Vibration Consultants hen applied correctly, vibration monitoring programs have proven effective in reducing operations and maintenance costs at industrial facilities. The vibration signature produced by industrial machinery provides more mechanical condition information than any other single parameter. Monitoring vibration levels over time can enable maintenance personnel to identify impending failures and resolve the situations prior to catastrophic failure. Problems can also be identified on new installations and during postmaintenance testing. This helps maintenance and production departments not only reduce downtime, but also schedule it in a more cost effective manner. Many maintenance and production people are aware of the benefits of vibration monitoring programs, but have never been exposed to the nuts and bolts of how signatures are interpreted. This article will explain how the vibration signature is used to determine the source and location of abnormal machinery vibration. Several of the more common sources of vibration will be discussed in detail, and real-world case histories will be used as examples.

W

Vibration Sources In my experience, unbalance, misalignment and resonance are the most typical causes of machinery vibration. Others include beat vibration, rolling element bearing defects, electrically induced vibration and

MED - Remelt Homogenizer Fan North R11-F0001 -H4 Fan Outboard Bearing (CW) Max Amp 19.4

P-P Displacement in Mils

Plot Scale 20

After balancing fan 19:16:26

0 Before balancing fan

0

600

1200

1800

2400

3000

Frequency in CPM

17:46:54 Freq: 1260.0 3600 Ordr: 1.000 Sp 2: .899

Figure 1. Vibration at fan shaft speed of 1260 cpm, before and after balancing the fan rotor

belt vibration. Many of these vibration sources produce very similar spectra, so determining the exact cause is usually a process of elimination. Other tools, such as phase analysis, impact testing for natural frequencies, shutdown testing and monitoring of overall vibration levels, are used to help eliminate possible sources of vibration - hopefully leaving a clear impression of the single most probable source. Sometimes more than one possible source persists after much time has been spent on testing. At some point, therefore, it becomes more cost effective to take action than to perform additional tests.

The Pump Handbook Series

Unbalance Unbalance is a leading cause of abnormal vibration. It is characterized by a high level of vibration at shaft speed, which is often referred to as 1X in vibration lingo. Unbalance occurs when an object’s center of mass is different from its center of rotation. Causes are incorrect installation, wear, material buildup and broken parts. The purpose of balancing a rotor is to make the device spin around its center of mass and reduce the load on the bearings. Due to standards imposed by industrial facilities, pump impellers are often balanced prior to installation, resulting in a longer life for both rotating and stationary components

305

of machinery such as bearings, shafts and seals. Industrial facilities can enhance reliability by putting in place a standard for balancing new or rebuilt rotating machinery. It may take some work to get vendors to meet balancing specifications, but it can be done. Unbalance creates a centrifugal force that can be calculated by the following equation, which takes into consideration the weight, radius and speed of the machine. Fc ≈ (.000001775)(U)(N ) 2

where Fc = centrifugal force exerted on bearings U = the unbalance force in oz-inches N = the speed in rpm After balancing a rotating part, such as an impeller, the weight needed to balance the part can be used to calculate the force that was present before the unbalance was removed. If a 5-ounce weight were placed at a radius of 12 inches on a machine operating 1775 rpm, for example, the force would be Fc = (.000001775)(5oz) (12inches))(1775rpm2) = 335 lbs. Balancing a rotating part will remove from the bearings the additional centrifugal force created by that amount of unbalance. This will increase the life of the bearings that support the part. To see how much unbalance affects the life of bearings, you can look at the equation used by many bearing manufacturers to calculate bearing life. L10 Life(hours) ≈(16,666 (Rated Load rpm Unbalance Load)3 where L10 Life = Life expectancy of bearings (hours) rpm = Machine speed Rated Load = The rated load capacity of the bearing (lbs.) Unbalance Load = The sum of the static rotor weight and the unbalance force You can see that the Unbalance

306

Load portion of the equation is in the denominator and is raised to the power of 3. This means that reducing the unbalance load by half would increase the bearing life by 8 times. This is why developing a balance standard is so critical. Figure 1 shows the vibration levels on a fan bearing before and after balancing the fan rotor. The decrease of almost 20 times in vibration will reduce the load on the bearings and add a significant amount of life to the rotating and stationary components.

Misalignment Misalignment occurs when the rotating centerlines of two shafts are either offset or meet at an angle, or both. Misalignment can also occur between two bearings. Misalignment often results in high axial vibration, as well as significant vibration at 1X, 2X or 3X shaft speed. The vibration signature produced often depends upon the type of coupling installed. Studies have shown that a substantial savings in power consumption can be achieved by obtaining a precision alignment. Although precise alignment can be achieved with dial indicators, many plants are standardizing on laser shaft alignment tools because they can often produce precision alignment in less time.

Resonance Resonance occurs when a structural natural frequency is at or near a source of vibration. The sources of vibration are those mentioned in this discussion, as well as others. When resonance occurs, what may be normal vibration can be amplified to very high levels – sometimes as much as 20 times as much or more. There are many tools used to detect resonance problems, including impact testing for natural frequencies, ratio analysis, phase analysis and shutdown testing. Impact testing is performed to determine if a structural natural frequency is at or near a source of vibration. An impact test is performed, with the equipment not running, by striking the equipment with a soft object, such as a rubber hammer, The Pump Handbook Series

and monitoring the natural vibration that occurs. This is similar to striking a bell and letting it ring down. The natural frequencies of vibration will show up as peaks in the spectrum. The natural frequency of a piece of equipment depends on the mass (weight) and the stiffness of the system. The amount that the vibration is amplified by resonance depends on the damping of the equipment. The formula for the natural frequency is as follows: Fn = √k/m, where k = stiffness and m = mass As you can see, increasing the stiffness or decreasing the mass can increase the natural frequency. Or increasing the mass (weight) or decreasing the stiffness can decrease it. In ratio analysis, the amplitude ratio from the vertical to horizontal directions is compared. For most problems that are due to mechanical defects, it would be expected that the vibration from vertical to horizontal would be similar. Often, a slightly higher amplitude can be seen in the horizontal direction due to less stiffness. Having an amplitude ratio from one direction to another of greater than 3:1 is often an indication of a resonance problem. During phase analysis, phase data are taken from two vibration transducers mounted in perpendicular planes (90 degrees apart). If a problem is purely mechanical, the most likely scenario would have the phase from one direction to another approaching 90 degrees. Having a resonance problem allows the phase to shift, and may allow the phase to approach 0º or 180 degrees.

Addressing ResonanceInduced Vibrations Resonance problems can be resolved in several ways. First, the source of vibration can be removed or minimized. Although it is often very impractical to change the speed of a machine, the source can be removed by changing the frequency (speed) of the vibration. By performing a precision balance job, the

MED - Colonial Pipeline-Clg Twr #2 C11-P002 -H01 Motor Outboard Bearing Comparison Spectrum C11-P002 -H01 13-May-97 09:32

764.48

0.05 0.04 0.03

Natural frequency near motor shaft speed

0.01 0 1.8 1.6 1.4

Reference Spectrum C11-P002 -H01 13-May-97 08:48

720.14

PK Velocity in In/Sec

0.02

1.2 1.0

Very high vibration due to resonance problem

0.8 0.6 0.4 0.2 0 0

600

1200 1800 2400 Frequency in CPM

3000

Freq: 720.14 Ordr: 1.000 Sp 1: 1.221

3600

Figure 2. High motor vibration due to resonance problem MED - Colonial Pipeline-Clg Twr #2 C11-P002 -H01 Motor Outboard Bearing

0.12

Comparison Spectrum C11-P002 -H01 14-May-97 15:19

827.14

0.10 0.08

Increase natural frequency from stiffening

0.04 0.02 0 0.14

Reference Spectrum C11-P002 -H01 11-JUL-97 09:25

0.12 713.08

PK Velocity in In/Sec

0.06

0.10 0.08

Reduced motor shaft speed vibration

0.06 0.04 0.02 0 0

600

1200 1800 2400 Frequency in CPM

3000

Freq: 713.084 Ordr: 1.000 Sp 1: .05516

3600

Figure 3. Reduced motor vibration after resolving resonance problem CIW -Boiler Cond RH Pump (B2-A1) B11-P005 -V03 Pump Inboard Bearing

0.5

Analyze Spectrum 30-MAY-96 09:02 PK=.5963 LOAD=100.0 RPM=3573. RPS=59.56

Bearing Defects

58.21

53.36

19.40

14.55

9.70

4.85

0.1

38.81

0.2

48.51

43.66

29.10

0.3

24.25

PK Velocity in In/Sec

33.96

0.4

0 0

10

20

30 40 Frequency in Order

50

source of vibration could be minimized. For example, if a pump that is operating at 1780 RPM has a natural frequency at or near the shaft speed, the vibration may be amplified by a factor of 10. If the normal unbalance is at .1 inches/sec of vibration, the amplification due to resonance would cause the vibration to be at or near 1.0 inches/sec, which would typically be a “fault” level. If the pump were balanced to .02 inches/ sec, the amplification would cause the vibration to be .2 inches/sec, which may be acceptable. The main problem with using the minimization technique is the fact that a small change in vibration will be amplified. So if the pump is allowed to wear and the vibration increases, the increase may be amplified 10 times due to the resonance. Two other methods of resolving resonance include changing the mass (weight) or stiffness of the structure. If a machine is being operated just below a natural frequency, it may be advisable to increase the stiffness in order to move the natural frequency further above the operating speed. If a machine is being operated just above a natural frequency, mass could be added to move the natural frequency further below the operating speed. Figures 2 and 3 are related to a resonance problem that was resolved by changing the stiffness of a structure. The lower plot in Figure 2 shows the abnormally high vibration level on a vertical pump motor bearing, and the upper plot shows the results of an impact test on the same bearing. The impact test results show that the natural frequency for this motor is very close to the shaft speed of 720 rpm. It can be seen in the upper plot of Figure 3 that after stiffening the motor’s structure, the natural frequency was increased to above 825 cpm. The lower plot in Figure 3 shows that the motor vibration dropped over 24 times to .05 inches/sec. – just by changing the natural frequency.

60

Ordr: 4.851 Freq: 288.89 Spec: .04160

Figure 4. Vibration spectrum produced by outer race defects on an MRC 5310 bearing The Pump Handbook Series

Identifying bearing defects is often misunderstood and seen as a form of magic. In reality, identifying bearing defects in a vibration spectrum is actually very straightforward

307

BEARING FREQUENCIES FOR MRC 5310 PHYSICAL DATA Number of Balls/Rollers: Ball/Roller Diameter: Pitch Diameter of Races: Contact Angle (Degrees):

12 .6875 3.150 30.0

HARMONICS SHAFT TRAIN SPEED (FTF)

SPIN OUTER INNER (BSF) (BPFO) (BPFI)

1 2 3 4 5 6 7 8

2.19 4.83 4.39 9.66 6.58 14.49 8.77 19.32 10.96 24.15 13.16 28.98 15.35 33.81 17.54 38.64

.40 .81 1.21 1.61 2.01 2.42 2.82 3.22

The waveform showed that the beat

PK = LOAD = RPM = RPS =

occured over a period of 37.2 seconds

.2427 100.0 577. 9.62

0.4

7.08 14.16 21.25 28.33 35.41 42.49 49.57 56.66

Table 1. Calculated defect frequencies for an MRC 5310 bearing

for a vibration analyst. The most common defects are inner and outer race spalling, and cage or rolling element defects. Most analysts have access to a database of defect frequencies that is provided by either the manufacturer of their vibration equipment or the bearing manufacturer. When a bearing is suspect, the analyst compares the frequencies present in the spectrum to the actual calculated defect frequencies and looks for matches. Defect frequency calculations are based on shaft speed and bearing geometry, which includes the number of rolling elements and their size, the pitch diameter and contact angle. The key in bearing analysis is determining how far along the failure has progressed. One question that analysts hear over and over is “How long will this bearing last?” Typical answers are “I don’t know” or “It depends.” Every bearing can follow one of several different failure patterns. The key to remember is that once a bearing starts to fail, it will never get any better. Bearings do not wear in they wear out. Figure 4 shows the vibration spectrum produced by a bearing with outer race spalling. Table 1 shows that the calculated outer race defect frequencies for an MRC 5310 bearing are at multiples of 4.83 times the shaft speed. Figure 4 confirms that these frequencies are, indeed, present. A visual inspection of the

308

BLO - Vertical Bandmill South -H06 Saw Driven Shaft - East Bearing Waveform Display 14-JAN-95 09:55

0.6

Velocity in IN/SEC

.99 1.99 2.98 3.97 4.96 5.96 6.95 7.94

H11-B002

0.8

0.2

-0.0

-0.2

-0.4

Maximum vibration occurs as the two wheels are in phase.

-0.6

-0.8

0

5

10

Minimum occurs as the two wheels are 180 degrees out of phase.

15 20 25 TIME IN SECONDS

30

35

40

Figure 5. Vibration waveform created by beat frequency vibration

removed bearing revealed significant spalling on the outer race.

Beat Vibration Beat vibration is caused when two closely spaced vibration peaks go in and out of phase. This will actually cause a beat that can often be heard and felt. This occurs when there are two machines in close proximity to each other producing vibration frequencies that are almost equivalent. Examples of this are when two machines are side by side, running at the same speed, or when there are two belt driven sheaves of similar size on one piece of equipment. When beat vibration occurs, it can be significant if it causes large amplitude swings. The beat frequency is the difference between the two closely spaced vibration peaks. It can also be calculated by dividing 1 by the period of vibration. Beat = F2 - F1 or Beat = 1/period Another major problem is that beat vibration can cause vibration amplitudes to vary significantly from month to month, which makes vibration trending very difficult. It is imperative that beat vibration be identified, so that changes in vibration amplitude will not be attributed to mechanical degradation. The Pump Handbook Series

Figure 5 depicts the vibration waveform produced by two rotating shafts on one machine moving in and out of phase. The period of vibration is 37.2 seconds, which creates a beat frequency of 1/37.2 sec. = .026 cycles per second (Hz), or 1.6 cycles per minute (CPM). The actual beat frequency an also be calculated by looking at the spectrum in Figure 6. F1 = 588.3 CPM, F2 = 589.9 CPM, and the beat frequency is Fbeat = F2-F1 = 589.9-588.3 = 1.6 CPM or .026 Hz

Belt Vibration Although it is not very common to see a belt driven pump, there are a few of them out there. Belt vibration is due to worn, loose or mismatched belts, and can also be caused by sheave misalignment, which causes some belts to be tight while the others are loose. Belt frequency vibration is often highest in the direction of tension. Normally, the fundamental belt frequency is not the predominant frequency in the spectrum. In many cases, the 2X or higher multiple of belt frequency vibration is predominant. As a belt leaves a sheave, there may be a standing wave present, which, when it impacts the other sheave, causes significant vibration. This occurs twice per belt revolution, which explains why the

H11-B002

0.24

BLO - Vertical Bandmill South -H06 Saw Driven Shaft - East Bearing Waveform Display 14-JAN-95 09:55 PK = LOAD = RPM = RPS =

0.21

PK Velocity in In/Sec

0.10

.2348 100.0 577. 9.62

0.15

0.12

0.09

0.06

0.03

0 570

576

582 588 Frequency in CPM

594

600

Freq: 588.3 Ordr: 1.020 Spec: .181

Figure 6. Vibration spectrum of two closely spaced peaks that create beat vibration CIW -Boiler Cond LH Pump (B2-A2) B11-P004 - MCI Motor Current Analysis

103

REFERENCE SPECTRUM 12-APR-96 09:00 PK = 104.24 LOAD = 100.0 RPM = 3560. RPS = 59.33

RMS Amplitude in STANDARD

102 60 Hz line frequency at 103 amps

101

Pole pass sidebands at 5.89 amps

100

10-1 50

52

54

56

58 60 62 Frequency in Hz

64

66

68

70

Freq: 57.44 Ordr: .968 Spec: 5.894 Dfrq: 2.540

Figure 7. Motor current spectrum acquired with a spectrum analyzer and clampon amp meter

second multiple of belt frequency is normally predominant. The fundamental belt frequency can be calculated from the equation below. If all of this information is not available during an investigation, a strobe light can be used to determine the speed of the belts. B ≈ (3.142)(P)(N) LB where FB = Fundamental belt F

frequency vibration P = Pitch diameter of sheave N = Speed in rpm LB = Belt length Belt frequency vibration is normally cyclical - in that the amplitudes of the different belt frequency multiples oscillate up and down as the belts go in and out of sync with each other. This causes the overall level of The Pump Handbook Series

vibration to cycle up and down over time. This leads to high cycle fatigue, which is detrimental to all parts of the machine.

Electrical-Induced Vibrations Vibration due to electrical problems can also be identified with a spectrum analyzer. Some of the sources of electrical vibration include broken rotor bars and end rings, loose connections, air gap distortion and rectifier problems on DC Motors. These problems are characterized by frequencies being generated at multiples of line frequency (60 hertz), or at multiples of shaft frequency. It is critical to ensure that enough resolution is present to distinguish between mechanical frequencies and electrical frequencies. For instance, a motor operating at 3580 cpm will have a second multiple at 7160 cpm (119.3 hertz), which is very close to 2 times line frequency (120 hertz). Rotor Bar Pass Frequency (RBPF) is the number of rotor bars times the shaft speed. This indicates a potential rotor bar problem, which causes uneven current flow through the bars, resulting in high heat, which can cause rotor bowing. On some cast rotors, voids are present which result in vibration at RBPF. Vibration at RBPF can vary greatly with load, but it may not be a significant problem unless the amplitudes are increasing under constant load conditions. A clamp-on current meter can be used to look at the motor current spectrum. The power ratio between the 60 hertz line current and the pole pass sidebands (slip x # of poles) should be greater than 40 decibels. As it drops below 40 dB, the chance that a rotor bar problem is present increases. Figure 7 shows the motor current spectrum obtained with a clamp-on amp meter. The power ratio is calculated as follows, based on having 103 amps at line frequency (60 Hz) and 5.89 amps at the pole pass sidebands. 20 log (5.89/103) = 24.8 dB. This is significantly under the 40 dB limit, and a visual inspection of the motor revealed a serious broken rotor bar problem. Another common electrical fre-

309

quency occurs at 2X line frequency, or 120 hertz. This usually signifies that a rotor, stator or air gap problem exists. A motor air gap can be distorted by an eccentric rotor or stator, or when a soft motor foot is tightened down. This can distort the motor frame, a condition that produces a distinct electrical hum. This can often be audibly distinguished from mechanical vibration, and by loosening each motor foot – one at a time, the soft foot can be identified and corrected. On a DC motor with a full wave rectified system, the SCR firing frequencies will show up at six times the line frequency or 360 Hz. There may also be submultiples at 120 or 240 hertz.

Conclusion Vibration monitoring has been employed for decades by plant personnel. An experienced mechanic, who has a history with ‘his’ machinery, can often identify an abnormal vibration, with the touch of the hand. The tools that are available today, however, allow for more indepth analysis and better recall. These tools enable a vibration analyst to not only identify that there is a problem, but also the nature and severity of the problem. Employing these tools is a must for facilities that intend to compete in a global economy.■ Cliff Hammock is the Regional Manager in Georgia for Industrial Vibration Consultants headquartered in Lebanon, Ohio. He is a graduate engineer from Mercer University and a Certified Vibration Specialist II. Mr. Hammock can be contacted by phone (912) 474-0463, or fax (912) 474-0612.

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R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

Energy Savings Pay for Reliability Improvements at Chevron A joint effort meets California’s “Pay for Performance” energy reduction mandate and upgrades plant processes. By Gary C. Koelbl, Planergy Services, Art Mares and Mike Lubcyik, Chevron

Introduction his article is a case study of five energy efficiency measures recently implemented at the Chevron Refinery in Richmond, California. The results are improved equipment reliability, improved process control, $750,000 annual power savings, and a simple payback period of less than 2 years for a $1.2 million investment. Secondary benefits are improved equipment reliability and process control. It is noteworthy that the project was accomplished without major endcustomer funding for capital equipment. The Richmond facility is the single largest producer of petroleum fuels and lubricants in the San Francisco Bay area.

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Demand Side Management Program In the mid-1990s, the California Public Utilities Commission determined that it was more economical to reduce electrical power consumption than to build new power plants. Some Demand Side Management (DSM) programs, such as energy rebates, did not provide the measured and recorded energy savings the CPUC desired. Therefore, California changed to a “Pay for Performance” approach. Pacific Gas & Electric, the Northern California Utility, used an open bidding forum to select implementers of this program, and Planergy, an Energy Services Company (ESCO), won a portion of the bid based on its plan to reduce industrial energy consump-

tion. The concept was to pay the implementers for measured savings through a 7year term. Planergy decided refineries and petrochemical plants were excellent targets because pumps and compressors run 24 hours a day and many consume 100 bhp or more per piece of Photo 1. Ground view showing the pump to the left equipment. Planergy (behind the valve wheel), motor (green box) in the midused its contract with dle and power recovery turbine (PRT) on the right end PG&E to develop the energy savings meadesign flows increased pump vibrasures identified by Chevron. tions and reduced hydraulic efficienPlanergy is a nation-wide energy cy. Chevron reliability personnel service company headquartered in evaluated several proposals to Texas and founded in 1977. Its improve the situation, but project Northern California office specializes return-on-investment hurdles, budin energy efficiency for industrial get and manpower priorities precludprocess companies. Planergy Sered any immediate action. vices, Inc. has demand-side contracts Unaware of these difficulties, with Pacific Gas & Electric, Houston Planergy approached Chevron as a Lighting and Power, and Texas Utilipotential candidate for the PG&E ties that subsidize the installation of DSM program. When the Chevron energy efficient equipment for comEnergy Coordinator and others panies in those utilities’ serving learned they could get pump rerates areas. Planergy provides variable and variable speed drives at no cost, frequency drives and pump upgrades they immediately started the process to industrial users at no capital cost, to upgrade the plant equipment. The where the energy savings are suffiinitial program included the followcient to amortize the capital costs on ing projects: a shared savings basis. • Install a medium voltage variable speed drive on a 2250 hp “DHT” Chevron Plant Change first stage charge pump and one on a In 1993, the Richmond refinery 700 hp product transfer pump changed its Vacuum Gas Oil (VGO) (reducing power loss through flow plant to a Diesel Hydro-Treater control valves). (DHT) plant, decreasing the feed rate • Replace the internal element on by 60 percent. Lower than original a 2250 hp second stage charge pump The Pump Handbook Series

311

• Evaluation factors: Since this was a diffuser style pump, the easiest approach to accommodate the 40% reduction in flow at the same TDH was to replace the internal element. • Solution: Both the nozzle (diffuser) and runner (impeller) flow passages were reduced in volume by using new patterns. Width of flow passages was narrowed, and vane angles were re-designed, but the impeller diameter stayed the same to accommodate the TDH. • Results: Electrical energy sav-

• Size and model: 5HMTA, 4 after project implementation stage, split case turbine • Problems: Operated below the zero-horsepower line due to revised conditions of service. Plant limit 1300 The turbine had not run in several years.

3000

Reliability and availability data provided the focus for equipment evaluation. Through a joint study, Chevron, Planergy and AllenBradley engineers developed a work scope for each project. Energy savings, improved process control and

2800

E-1601 DHT Hydraulic Power Recovery Turbine

2500

Evaluation and Implementation of Energy Savings Measures

2400

1800

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1400

1200

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Plant limit 1300 GPM

1000

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0

Normal flow rate 875-1020 GPM

200

• Size and model: 6X12 CB, 8 stage, double case barrel pump • Problems: Low hydraulic efficiency; average process flow was 40% of BEP • Evaluation factors: Existing impeller trim precluded the need for a new internal element, and process records 4500 indicated a need for flow Photo 2: Close up of P-1601 first stage DHT 4000 variances. charge pump • Solution: Installed an 3500 and a 400 hp hydraulic power recovAllen-Bradley variable speed 3000 ery turbine (improved pump efficiendrive, Model 1557, 2250 hp, 2500 cy). 2300 VAC 2000 • Change operating procedures • Results: Energy savings of 1500 for the main 5500 hp and backup 2.4 MkWh/yr; easier process 4000 hp charge pumps in the ISOcontrol; reduced vibration 1000 500 MAX unit (operate the most efficient amplitudes from 0.3 ips to 0.12 0 pump based on plant feed rate). ips; no bearing or seal problems since start-up.

2200

P-1601 First Stage DHT Charge Pump

2000

reliability were the key selection factors for the measures depicted in the following case studies.

HEAD Before HEAD After

Figure 1. P-1671A pump head before and 80

70

60

50

40

Normal flow rate 875-1020 30

20

NormEfficiency increase at 1020 GPM = 17 points

10

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EFF Before EFF After

Figure 2. P-1671A efficiency before and after 2000 1800 1600

Horsepower savings at 1020 GPM 475 HP = $185,000/Yr.

1400 1200 1000 800 600 400

Plant limit 1300 GPM

Normal Flow Rate 875-1020 GPM

200

HP Before HP After

Photo 3. Ground view from the back side of the installation. The PRT is to the left, receiving the pipe elbow.

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The Pump Handbook Series

Figure 3. P-1671A horsepower before and after

3000

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ings of 1.9 Mk Wh/yr; rerated power recovery turbine (PRT) produces a 300-hp assist to the P-1601 motor driver at variable speed. No seal or bearing problems since start-up. Vibration levels are only 0.10 ips.

P-1671 Second Stage DHT Charge Pump • Size and model: 6X12 CB, 8 stage double case barrel pump • Problems: Low hydraulic efficiency; high vibration; revised flow is 40% BEP • Evaluation factors: This pump had maximum impeller diameters, and the process flow was more constant than that of P-1601. The energy economics evaluation indicated that a new internal element was the proper selection. • Solution: The pump OEM designed new impeller and diffuser patterns with narrower passages and corrected vane development. Additional design changes included: a. Split ring shaft arrangement and shrink fit impellers b. New thrust bearing housing with Kingsbury LEG thrust bearing c. Pressure dam journal bearings for proper bearing loading d. Larger seal chambers for current seal designs e. Metal-to-metal head fit for easier maintenance and compliance to API-610. • Results: Energy savings: 2.1 MkWh/yr; reduced vibration amplitudes from 0.45 ips to 0.05 ips, with no bearing or seal problems since start-up.

P-1665 Diesel Transfer Pump • Size and model: H-8X14 DSTHE • Problems: Pump excessively oversized for flow, resulting in high vibration, reduced bearing and seal life, and low hydraulic efficiency. This pump transfers diesel fuel to several tanks of different elevations and distances and was constantly running back on its curve. • Evaluation factors: System head curves from field measurements, and usage profiles from the operations

department provided data for the energy savings analysis. Due to the various flows and pressures, a variable speed drive was selected over a pump rerate. • Solution: Installed an AllenBradley VSD, Model 1557, 700 hp, 2300 VAC. • Results: Energy savings: 2.0 MkWh/yr. Pump speed range is 3000 to 3500 rpm, with vibration reduced by a factor of 3 at the slower speeds.

P-401 and P-401B, ISO-MAX Charge Pumps • Size and model: 8” IJ-11, and 6” RHM IJ-11 • Problems: The large pump, with a 5500 hp motor, ran almost all year even though much of that time it was running back on its curve. When the pump exhibited high vibration levels from minimum flow, operators switched to the small pump. • Solution: The energy analysis included calculations for five scenarios from worse case to best case. Ultimately, the conclusion suggested running the small pump until it reached its maximum capacity, then switch to the large pump. Usage profiles indicated that the small pump could be used about 85% of the operating year. • Results: Energy savings: 7.0 MkWh/yr. Vibration levels for both pumps remain low.

Engineering Considerations Variable Speed Drives Allen-Bradley worked directly with the Chevron engineers to ensure a smooth integration into the operating system. Several independent studies were performed. Rotor dynamics studies, such as torsional and lateral analysis, were completed for all machines. P-1665 was acceptable for the 2500 rpm to 3600 rpm speed range, and P-1671, with new components, would be suitable for 3570 rpm. Medium voltage drives require a fair amount of floor space (80 sq ft) and cooling. A separate building is preferred. Chevron had space in the MCC building serving the P-1601 and P-1665 motors, and this location The Pump Handbook Series

minimized the amount of new cable, which is an important cost consideration. Insulation and air conditioning had to be added because the large drives generate significant heat and require both a clean and temperature controlled (32 - 104ºF) environment. The existing area classification was suitable for this electrical installation. Electrical harmonics studies indicated that the standard isolation transformers would sufficiently protect the refinery grid. Power factor correction was not required. The control package had to be integrated into the Chevron process control system so that the operators could use the VSD instead of the control valve for process control. Safety interlocks, motor protection devices and other limit controls were included in the controls package. A comprehensive VSD training program was given to the operators and maintenance technicians to cover normal running control as well as emergency situations.

Equipment Rerates Ingersoll-Dresser worked directly with the Chevron mechanical engineers to make sure the redesigns were accurate. Present and future flow requirements were factored into the component selection process. For both P-1671 and E-1601, the TDH had to remain high while the flow was significantly reduced. Impeller patterns with narrower shroud passages were chosen from existing inventory. Milled-vane diffusers allowed for a more precise flow control geometry. Special selection criteria applied to PRT, E-1601, because it would be connected to the P-1601 train and run at various speeds. Its better to undersize a PRT slightly rather than oversize one. During the detailed engineering work, the factory engineer discovered that the field rotation for P-1671 was clockwise whereas the new element initially selected was counter clockwise. Chevron wanted to keep the clockwise rotation for interchangeability with the spare element, so another rerate element was selected. Actually, this worked to the positive because the second selection

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had higher Ns (1500), which gave an additional 2 points in efficiency, and it had less TDH rise to shut-off to accommodate the pressure rating on the discharge flange. Thoughtful consideration was given to selection of new components with respect to interchangeability with older components. For example, the factory designers preferred larger shaft diameters for better rotor dynamics, but seal and coupling sizes would be affected. Shaft designs were altered to use the existing seals and couplings. Ultimately, Chevron engineers decided which components to keep and which to replace. In most cases, they chose in the direction of easier maintenance. Provisions for vibration and temperature monitors were added. Full laboratory performance tests were conducted to verify the new impeller performance and vibration levels. Certified curves and data sheets were included. The PRT missed the performance on first test but passed on second test after rework of the diffusers. The pump passed on first test.

Operations Procedure Change Initially, P-401 and P-401B seemed to be VSD candidates because their flow changed with different feed stocks and other process variances. Consulting engineers compared present flow conditions to five

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scenarios of future flow requirements, concluding that 80% of the VSD savings could be achieved by merely changing the operator’s procedure to use the smaller pump until it could no longer charge the reactor. Then they should switch to the larger pump. No equipment was purchased.

Conclusions The project rate-of-return based only on energy savings is considered excellent. It had been somewhat difficult for Chevron to achieve competitive rates-of-return for capital intensive, self-funded energy projects before such a program as described in this article became available. Equipment reliability and availability have improved for the DHT plant as evidenced by no repairs on project equipment start-up in June 1997. The variable speed drives provide an overall system reliability improvement. The motors have a soft-start that reduces inrush current and stress to the mechanical components. Operators have better process control, and the pumps run near their Best Efficiency Point. If the drive were to fail, the system reverts to constant synchronous speed as it was before the installation. The equipment rerates provide machines with today’s technology. Improvements may include better materials as well as new design fea-

The Pump Handbook Series

tures promoting longer mean times between failures and easier maintenance routines.

Planergy and End-User Business Roles Planergy is prepared to invest in viable projects. As consideration for project development and funding, Planergy keeps a percentage of the savings generated during the contract period. After this period, 100% of the energy savings belong to the end-user. Also, the end-user is obligated to operate and maintain the equipment during the contract term. Both energy consumption and process flow are continually monitored and verified to provide the energy savings data used to calculate the shared savings amounts.■ Gary Koelbl is Director of Engineering, Mechanical Projects, for Planergy Services, Inc., in Richmond, California. He has 27 years of pump related experience as a sales engineer and test engineer with Dresser Pump Division and Ingersoll-Dresser Pump. Art Mares is a, Reliability Analyst for Chevron’s Richmond Refinery with 12 years of experience in equipment reliability at the facility. Mike Lubcyik is Operating Assistant, Distillation and Reforming Business Unit for Chevron. He was the Refinery Energy Coordinator during the time the projects were implemented.

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R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

Teamwork Key to Pump Reliability Upgrades at Conoco By Dan Gourley

am a member of a reliability team consisting of technicians (craft persons) who identify defects, prioritize them according to profit potential and eliminate them by that formula. Technicians work with engineers and vendors to establish a philosophy of pump reliability that manages equipment, eliminates defects and educates operations people about the relationship between pumps and the units they operate. The area reliability team was established as the brainchild of the Coker/Combo area reliability engineer with a goal of upgrading pump reliability. His vision was to create the role of a dedicated reliability technician who would be involved more in overall pump reliability than a traditional maintenance technician, who would be involved only in pump repair. This concept would combine the experience of the technician with the technical expertise of the reliability engineer. Each would benefit from the other’s special skills and elevate everyone’s understanding of pumping reliability problems. Through the evolution of this concept, the Coker/Combo area’s centrifugal pump MTBR (Mean Time Between Repair, based on a rolling 12-month average) increased from 1.08 in December 1994 to 6.5 by June 1997. Annual pump repair costs have decreased from more than $800,000 to less than $300,000. The one half million dollar annual sav-

I

ings in repair costs can be leveraged through unit uptime (production) for increased revenue of 2.5 to 3.5 million dollars. The initial idea for involvement of crafts people in reliability at our complex was as caretakers of equipment. We developed a set of procedures that defined and laid out best preventive maintenance practices. In these practices a specific set of instructions was given for each type of equipment, and an interval was established for them to be carried out. These practices were a foundation that evolved to include preventive maintenance, predictive maintenance and asset optimization. At first, pump reliability improvements were based on failure evaluation. This was done during the repair process as defect elimination, and if warranted, Root Cause Failure Analysis (RCFA). Evaluation of failed or damaged equipment was done as the technician made the repairs. This determined the affected parts and the manner in which the failure occurred. Vendors were consulted as technical experts for their pumps, and repairs were made based on findings and conclusions of everyone involved.

Investigation of Stripper Feed Pump Failures P-1172 was selected as a candidate for the defect elimination process because of its significant economic impact when out of service. The Pump Handbook Series

The spare pump’s capacity was 18,000 bpd instead of the 23,000 bpd capacity for P-1172. In addition, P1172 had been a chronic reliability problem with an MTBR of 0.725 years. The pump was determined to have a 2 gallon per hour seal leak. Vibration analysis was done prior to the reseal. The results indicated no change in vibration from the last check and OK to run. So the seal was changed. A recent change to tandem seals required the removal of the motor for the seal to be replaced. Vibration analysis at start-up after reseal showed a four fold increase in vibration. After inspection it was determined that coupling match marks were not aligned as they were previously. (Several sets of match marks caused confusion as to the proper set to use). The marks were realigned, and an attempt was made to align the motor to the pump. This effort was not successful with a total indicator run-out of 0.002” on the motor shaft and 0.015” on the pump shaft directly above the mechanical seal. Vibration analysis indicated looseness, misalignment and imbalance but the equipment was still judged O.K. to run. When we discovered a second mechanical seal leak and increased vibration, we decided to do a complete overhaul, with an informal RCFA to complement the rebuild. The following results are from the collaborative effort between Cok-

315

er/Combo and Main Shop personnel.

ment).

Physical Evidence

Motor Stand Condition

Shaft diameter and span between bushings play an important role in shaft stability. Current API 610 specifications for this pump with a 1 3/16” shaft diameter call for a maximum span of 33 inches. This spacing was adequate in all spans. The pump had been converted to nickel impregnated graphite bushings during a prior overhaul. Clearance specifications for nickel impregnated graphite bushings with a 1 3/16” shaft are 0.004” + 0.002/0.000”.

Both ends of the motor stand were out of parallel. Bosses on both ends were off center.

Bushing Clearances A. Throat bushing – 0.020” oval wear 0.004” (indicating misalignment) B. First line shaft bushing – 0.061” oval wear 0.003” C. Second line shaft bushing – 0.107” oval wear 0.013” D. Third line shaft bushing 0.057” (less wear – stabilizing effect because of proximity to closely spaced bowl bushings) oval wear 0.017” E. There are eight bowl bushings with clearances ranging from 0.011” to 0.032” and with oval wear ranging from 0.003” to 0.012”. (Possible source of oval wear pattern discussed in pump well mounting surface condition). F. Suction bushing 0.017” clearance

Coupling Condition The coupling hubs on the motor and pump ends displayed excessive clearance with 0.005” at both ends. This causes two problems: looseness and misalignment. The coupling faces were worn and inconsistent (another source of misalignment).

Discharge Head Condition The pump discharge head on both the well and motor stand mating surfaces was out of flat. Discharge head bosses to the pump base and to the motor stand were off center. Clearance from the stuffing box to the discharge head was 0.012” (causing throat bushing misalign-

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Motor Mount Condition Motor mount surfaces were out of parallel by 0.015”.

Suction Barrel Mounting Surface Condition This was out of flat by 0.153” (This would cause the entire assembly to have uneven bushing (oval) wear).

Shaft Condition

ends were surfaced to make them parallel.

Stuffing Box A new stuffing box was made to tighten up clearances and allow the shaft to be shortened.

Suction Barrel Mounting Surface This was field machined to make it flat.

Shaft The shaft was replaced, but the new shaft was warped and had to be straightened. It was also shortened (see coupling repairs).

The shaft did not show signs of excessive wear at journals for bushings. Threads for coupling the lifting nut were worn, however.

Motor Mount

Bowl Condition

The rotating assembly was balanced, and the bearings were replaced.

Bowls were in good condition, being both parallel and perpendicular.

Actions Taken Or Planned Bushings Nickel impregnate graphite bushings were again used with the addition of spiral grooves on all but the throat bushing. Tolerances were set in accordance with the manufacturer’s specifications.

Coupling A new precision coupling was made and balanced according to specifications. This tightened up clearances and made it both parallel and perpendicular. A longer spool allowed the shaft to be shortened (to facilitate seal replacement without removing the motor).

Discharge Head The discharge head bosses were spot welded and turned down on both ends to establish line bore. Both ends were surfaced to make them parallel.

Motor Stand The motor stand bosses were spot welded and turned down on both ends to establish line bore. Both The Pump Handbook Series

The motor mount was machined to make it parallel.

Motor

Results This investigation draws no single conclusion as to the cause of failure. Rather it reveals a combination of the factors listed above. Total indicator run-out was checked with the motor coupled to the pump. The results were a reading of 0.001” on both the motor and pump shafts. When P-1172 was returned to service, the best vibration analysis results ever recorded indicated that we could expect a significant improvement in reliability.

Reboiler Pump System Assessment Along with the preventive maintenance duties, reliability technicians were given a list of bad actors (based on highest repair costs and most repetitive failures) in their respective areas with the expectation that they would identify and solve the problems that existed. This meant that the technicians had to understand previous failures and the controls and systems in which these bad actors functioned.

Example: The #2 Reformer Stabilizer Reboiler pumps are controlled by the DCS (unit control system). When the DCS detected low pressure on the pump that was running, the spare pump was automatically started. (This happened only during a power blip.) During this time both pumps would operate, causing minimum flow problems that led to repetitive pump failures. This situation had been a problem for years, and by evaluating the computer alarms and start-ups after a failure, the difficulty was identified. The solution was to rewrite the control logic to start the second pump and shut down the one with low flow. Since these changes were made, there have been no failures on these pumps.

Asset Optimization Asset optimization was the next objective for the reliability technician. This took place through equipment assessment. Projects were

evaluated with economics as the primary driving factor. Example: The Delayed Coker unit’s 8 main residuum pumps used API plan 32 seal flush. This system, which introduces a clean cool seal flush to the mechanical seals, was replacing about 850 barrels per day of product throughput, to the unit. The seals were evaluated, and changes are being made that will make it possible to use API Plan 11 (self flush) on these pumps. When this project is completed, the anticipated net earnings gain will be 1.5 to 2 million dollars per year.

Conclusion Attention to detail is the main focus of the reliability technician. The traditional tools of sockets, wrenches and dial indicators have been replaced with a polysonic flow meter and calculator. With the pump curve as a basic tool, the reliability

The Pump Handbook Series

technician is now working toward overall long term reliability instead of focusing on quick repairs. Large strides are being made in pump reliability by breaking down traditional barriers. For this to be effective, an investment must be made in people. This is facilitated when skilled workers trade ideas and experiences with engineers. The training that each receives results in a better understanding of what reliability is, and the “how to” is a culmination of experience for all involve. Acknowledgment: I would like to thank members of the Coker/Combo Area reliability team and the repairman and machinist crafts for their contributions to make this program a success.■ Dan Gourley is a Reliability Repairman for Conoco in Ponca City, Oklahoma and a recent speaker at a Pumps and Systems sponsored conference on pump reliability and maintenance.

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Maximizing Reliability in Sealless Pump Operation Understanding the common requirements and differences among these pumps is critical to achieving a highly reliable installation. By Robert Mulholland, Sundstrand Fluid Handling Corp. he chemical processing industry is focused on managing the total life cycle cost of plant capital equipment, including sealless pumps. Equipment downtime must be regarded as a lost profit opportunity. When added to repair parts and maintenance personnel costs, the cost of downtime can easily dwarf the initial purchase price of a pump. Although sealless pumps are used more and more in a broad range of applications, how to select the right one remains widely misunderstood, as do their operational requirements. There are three major categories of centrifugal sealless pumps: canned motor, metallic magnetic drive and non-metallic, lined magnetic drive. Where is each type best utilized? What are the major advantages and disadvantages of each? Understanding the differences and common requirements among these pumps is critical to achieving a highly reliable installation.

(Courtesy Sundstrand Fluid Handling Corporation)

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Photo 1. A typical canned motor pump installed in a plastics facility

available in a range of configurations that are designed to provide a controlled environment for the process lubricated bearings. The various arrangements can increase the pressure in the bearing area when volatile fluids are being pumped, prevent process solids from entering the bearing area, provide a cool source of lubricant on high temperature services and so forth. These measures mean more control over the critical bearing environment, resulting in higher reliability potential. Canned motor pumps also benefit from a wide range of monitoring options that facilitate preventive maintenance. The principal disadvantage of canned motor pumps is the relatively special nature of the motor. The same metallic liner, or “can,” that provides the primary fluid boundary under

Canned Motor Pumps The single biggest advantage of canned motor pump technology is double containment. If the primary fluid containment boundary ruptures, the external motor stator shell is a true pressure vessel. It is often rated to the pump’s full working pressure, tested to assure integrity and resistant to mechanical contact or “rub-through.” This feature makes canned motor pumps the choice for lethal or extremely hazardous applications. They are also well suited for high pressure operations. Canned motor pumps are usually

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Figure 1. Cutaway of a typical canned motor pump The Pump Handbook Series

tend to have bearings made exclusively from silicon carbide (SiC), due to its high resistance to chemicals. The disadvantages of lined pumps stem from the physical limits inherent in today’s fluoropolymers. Temperatures must be limited to about 250°F (120°C), and the pumps are generally limited to smaller power levels and lower developed head due to limits in bonding strength of the liners on large diameter, high tip speed impellers.

Major Causes of Failures in Sealless Pumps While the specific applications for sealless pumps vary, the primary causes of their failures are similar. In all cases, they result from the failure to control the operating environment of the process fluid lubricated bearings.

Figure 2. Cutaway of a typical magnetic drive pump

normal operation adds to the difficulty of decontaminating and repairing a breached motor. While it is technically feasible to repair these motors, it is often not economical to do so.

Metallic Magnetic Drive Pumps Metallic mag drive pumps are well suited for a wide range of pumped materials, including: • mild acids • solvents • liquids that are sensitive to heat input • fluids containing moderate solids • fluids that present difficult sealing challenges, such as liquids with dissolved solids that tend to precipitate out of solution • high melting point liquids • high temperature services including heat transfer oils Synchronous mag drive pumps add less waste heat (from eddy current losses) to volatile fluids than canned motor pumps. Metallic mag drive pumps are available in sizes as large as 520 hp (400Kw) and utilize standard NEMA and IEC motors The primary disadvantage of these pumps is that they have no secondary containment capability, making them less than ideal for lethal or extremely hazardous applications.

Various secondary control devices are available from suppliers to prevent massive spills if the primary fluid containment shell ruptures.

Dry Running The leading cause of failure in sealless pumps is probably dry running. Operating a pump for even a few seconds without liquid at the bearings will shorten pump life. Much has been written and presented at various chemical industry technical meetings on the subject of dry running bearings for sealless pumps. New materials, including hybrid carbon/SiC designs, “diamond like” coatings on SiC and a wide range of carbon graphite blends, have been tried and tested. But, no material offers a panacea for long term toler-

Non-metallic, Lined Magnetic Drive Pumps The best of these designs feature a metallic outer pump case housing lined with thick, bonded fluoroplastic that combines excellent corrosion resistance with the strength of a metallic casing. Such pumps are ideal for acetic acid, hydrofluoric acid, hydrochloric and sulfuric acids as well as ferric chloride, hydrogen peroxide, sodium hypochlorite and other dangerous liquids. These pumps 800

3500 rpm

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Figure 3. Example of hydraulic performance of magnetic drive pump at 60hz The Pump Handbook Series

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ance of dry running under loaded conditions. Bearings for sealless pumps can be designed for greater tolerance to dry running, but only at the expense of load carrying capacity, wear resistance and run life under normal operation. In general, SiC provides the greatest load carrying capacity of any available material. It also offers excellent corrosion resistance. It can’t tolerate even brief dry runs, however. Carbon graphite, on the other hand, provides a degree of self lubrication under dry run conditions, and it is more tolerant of this abuse. Yet it cannot handle unit loads as high as SiC. Nor can it withstand abrasive particles. It will therefore have a shorter useful life under even ideal conditions. Off-Design Operation & Undisclosed Fluid Properties Because sealless pumps rely on the process liquid for bearing cooling and lubrication, they are application sensitive. When pump suppliers recommend a specific pump configuration and bearing material selection, they do so based on anticipated operating conditions identified by the buyer. Deviations from these design conditions, including variations in pumped liquid characteristics, extreme changes in flow rates, contaminants and changes in temperature or viscosity, can have detrimental effects on the reliability and bearing life of sealless pumps. Pump Oversizing Although a common practice, applying too many “safety factors” to pump rated flows is detrimental to all types of centrifugal pumps. Oversizing results in off-design operation of the pump. It increases radial bearing loads, disrupts the pump’s hydraulic thrust system, and may compromise the motor’s internal cooling capacity by decreasing fluid flow. With some fluids, such as polymers, this last factor can result in the formation of particulates that clog critical internal flow passages. Solids Presence In addition to the secondary formation of solids mentioned above, sealless pumps are no more (or less)

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tolerant of entrained process solids than conventionally sealed pumps. Again, the central issue is bearing lubrication. Bearings made of hard materials, such as SiC, can tolerate the passage of solids more than soft materials like carbon graphite. Additionally, some manufacturers offer pump configurations specifically designed to exclude particulates from the bearing area. This is accomplished through screening or centrifugal separation or by supplying a clean secondary flush and a flush restriction device. It is critical that the nature of the entrained solids is discussed with the manufacturer to assure proper pump selection. Specifics such as particle size, percentage by weight or volume, abrasive hardness and tendencies to agglomerate — are all important details that the manufacturer must know.

Special Considerations for High Vapor Pressure & Volatile Liquids When pumping high vapor pressure liquids, especially at pressures and temperatures close to the bubble point, additional information such as fluid specific heat and, if available, the actual vapor pressure curve, will be useful. Because the process fluid is used in sealless pumps to remove heat due to electrical inefficiency, containment shell eddy current losses, bearing friction and hydraulic inefficiency, the temperature of the fluid within the motor or magnetic coupling section of the pump will increase. A potentially damaging condition exists if the line between liquid and vapor phase has been crossed when a higher temperature fluid is returned to suction pressure. Bearings may be running “dry” (i.e., in vapor rather than liquid.) Depending on the internal flow path of the specific pump design, this vapor may be routed back into the suction of the pump impeller, resulting in cavitation. The presence of vapor in the motor or magnetic coupling area reduces heat rejection and can result in overheating. Be sure to disclose as much detail as possible about the vapor pressure/temperature curve for your specific liquid to your supplier. In the case of liquid mixtures, The Pump Handbook Series

include information on the highest vapor pressure constituent. Gas dissolved in the process liquid will naturally increase the likelihood of flashing within the pump. This fact, too, should be communicated to the manufacturer.

Venting and Start-up Procedures This issue, important to all centrifugal pumps, is often given inadequate attention by operators. Even brief periods of operation with improper venting will cause sealless pump bearings to run dry. In the case of carbon graphite bearings, accelerated wear almost always happens under these conditions. With SiC bearings a rapid rise in bearing temperature occurs. When full venting is finally attained, the bearings are easily shocked thermally. Extremely hard, brittle materials such as SiC can shatter under these conditions. Secondary damage to pump internals, resulting from the passage of the razor sharp fragments of the shattered bearings, is also likely. The following simple procedure can be used to assure complete venting of even hard to vent systems. It assumes the pump is empty of liquid and that both suction and discharge valves are closed. 1. Open suction valve. (Pump fills part way.) 2. Close suction valve. 3. Open discharge valve. (Once the pressure equalizes, air will rise into the discharge piping.) 4. Open suction valve. 5. Start pump. (Additional information on this topic can be obtained from an article by Michael D. Smith in the November 1993 issue of Pumps and Systems magazine.)

Misapplication of Design and Materials When selecting a sealless pump, note that the process of application engineering for this type of pump is in many ways similar to selecting a proper mechanical seal and seal support system. As mentioned, the more detail manufacturers know about the service, fluid properties and intended

(Courtesy Sundstrand Fluid Handling Corporation) (Courtesy Sundstrand Fluid Handling Corporation)

Photo 2. Typical mechanical type bearing monitor that senses axial and radial rotor position for a canned motor pump

Photo 3. The new generation of pump monitors determine precise rotor positioning independently .

operation of the pump, the better they will be able to select a configuration and proper materials of construction. This helps ensure long, reliable service. Be prepared to supply information on not only design point conditions but the full range of intended flows and pressures. Be as exact and comprehensive as possible in identifying fluid properties. Also, don’t forget to describe piping and control system details. Try to detail any upset conditions that the pump may experience. Remember that corrosion rates are accelerated by increases in temperature. A solid understanding of all these factors will help the manufacturer select a pump with the highest potential reliability for your service.

Inadequate Monitoring While the primary ad-vantage of sealless pumps is the complete, hermetic sealing of the process liquid, this can also lead to one of the biggest frustrations with pumps of this design: You can’t see what’s

going on inside. Whereas normal seal wear in an ANSI pump can be detected visually, normal bearing wear occurs with no outward indication. If preventive maintenance is not conducted in a timely manner, more serious and expensive peripheral damage is likely.

Core Issues for Reliable Operation The top causes of unreliable sealless pump operation are: • bearing loads, design and material options • process fluid lubrication • cooling and “the other NPSH” value Proper control of bearing loads is basically within the control of the pump designer. Look for thrust balancing designs that can reduce the hydraulic loads imposed on the bearings. The more conservative the load placed on the bearing, the better. No single bearing material is right for all liquids and pump operating scenarios. As stated, softer materials are more tolerant of the abuses of liquid flashing or intermittent dry running, but they provide a shorter useful life under normal operation. Harder materials such as SiC provide increased wear resistance to abrasive particles, and will yield longer life under ideal conditions. However, these materials are intolerant of dry running and may result in more extensive peripheral damage if they are run to failure. Process fluid lubrication of the bearings is, by definition, dependent on the properties of the fluid pumped. Bearings require two things for reliable operation: liquid (not vapor) between the bearing surfaces, and that this liquid be relatively clean and non-abrasive. Even with this obvious constraint, much can be done to create an improved bearing environment. For fluids containing abrasive solids, a secondary clean fluid flush can be injected into the bearing area. With volatile liquids, pre-cooling or internal flow schemes that boost and maintain the liquid pressure above the bubble point are available. Cooling is a relative term as applied to sealless pumps. Modern rare earth magnetic drive pumps function well without external coolThe Pump Handbook Series

ing to approximately 250°F. Torque ring design mag drive pumps can tolerate temperatures to 750°F without external cooling. Canned motor pumps featuring ceramic motor insulation systems can operate up to 850°F without supplemental cooling. Bear in mind, however, that these temperatures are based solely upon the technical capability of the individual pump design. The requirement of limiting the fluid temperature within the pump may be based on maintaining the characteristics of the process fluid required for reliable bearing life: liquid (not vapor) and no abrasive solids. Therefore, in the case of certain fluids such as polymers, cooling may be required to prevent localized hot spot polymerization within the bearings. When pumping volatile liquids such as chlorine, cooling may be required to ensure that the liquid does not flash off to a gas within the bearings. With high temperature applications such as heat transfer oils, canned motor pumps with conventional organic resin insulation systems may require cooling to protect the motor stator core from overheating. The “other NPSH value” is a widely misunderstood phenomenon. In sealless pumps the process fluid is circulated within the motor (or mag coupling) and through the bearings to remove heat. The amount of heat generated varies widely from design to design and is generally less in nonmetallic lined mag drives due to the absence of eddy current generation across the non-metallic containment shroud. Metallic mag drives rank second in heat generation. The heat from electrical inefficiency of the drive motor is removed through air cooling rather than by the process liquid. Canned motor pumps generally add the most heat to the pumpage. As noted earlier, there are special considerations for volatile liquids. It is not solely the added heat that is of concern, but the combined effect of heat vs. pressure and the operating point on the fluid vapor pressure curve. The internal flow paths within the pump must also be considered. The most common internal flow scheme used in sealless pumps takes a side stream of liquid from the pump discharge (or at least a point within

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the pump volute where partial discharge pressure has been generated) and circulates it across the rotor, between the bearings and back to suction. It is the pressure differential between discharge and suction that causes the stream to flow. With this setup, liquid at an elevated temperature is being introduced directly into the pump suction. If the liquid is prone to flashing at this lower pressure and higher temperature, it will do so in the pump suction eye. Thus, even though the NPSH of the pump is determined to be a lower value, cavitation can still occur. Be sure to consider this phenomenon when selecting a pump. Alternate internal flow schemes are available to avoid this problem. You can maintain higher internal pressures or return the side stream directly to the suction vessel rather than the impeller eye.

Improving Reliability in Sealless Pumps Look for various configurations to control the bearing environment. Variations in internal flow paths, centrifugal separation, screens or filters, auxiliary heat exchangers, flush injection points, flush restriction devices and auxiliary impellers are available to assure that the bearings operate in as ideal an environment as possible. This will result in long, reliable life. Bearing materials should be determined based on fluid properties and the anticipated “real world” operating range. The hottest place in a sealless pump moving corrosive liquid is the bearings. A wide variety of carbon graphite binders are available to suit the corrosive nature of most process fluids. Avoid building in unnecessary “safety factors” in establishing the design flow rate of the pump. Look for suppliers with a wide range of hydraulic designs to match your required conditions. Some vendors have hydraulic designs for low specific speed (low flow, high head) services that help you avoid selecting too large a pump and operating well back on the curve. These designs help minimize radial loads imposed by off-design operation. Use monitoring devices appropriately. Fortunately, many good products for sealless pumps are avail-

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able. One of the simplest and most economical ways to protect a sealless pump from operational abuse is by using a power control monitor. This measures input Kw and enables the user to detect potentially damaging operation of the pump, such as low or high flow (beyond recommended operating range) and dry running (resulting from improper venting and/or loss of flow). Upsets in fluid viscosities are also easily detected by an increase in Kw. Power control monitors are readily available for use on mag drive pumps, where it may be impractical to use other types of direct sensing bearing wear monitors. Several manufacturers of canned motor pumps offer a variety of bearing monitors, including direct-acting mechanical units that sense axial and radial rotor position (Photo 2), mechanical monitors with electrical switches to provide a remote alarm/ shutdown feature, and electronic units that can provide a progressive indication of the rotor’s position within the motor. The most advanced of these monitors (Photo 3) can determine precise rotor positioning independently at both radial bearings, as well as axial rotor positioning, and they can identify thrust direction, direct detection of two phase (gas/liquid) flow at the bearings, and direction of rotation. Several user interfaces are available, including intuitive local displays using LEDs to indicate rotor position (Figure 4). Automatic, remote monitoring is available through relays and a 4-20 mA signal. Two manufacturers offer RS485 serial port communication links to the user’s DCS system. This enables the operator to access a wide range of data from the monitor’s host software and use it for advanced rotor dynamics analysis including shaft orbital plotting, time waveform analysis and spectrum frequency analysis. All the monitors are noninvasive and maintain the integrity of both the primary and secondary liquid containment boundaries. Given the enclosed “black box” nature of sealless pumps, it is vital that one of these technologies is included in all sealless installations. All sealless pumps should have some sort of monitoring device, if only a power control monitor. These The Pump Handbook Series

Figure 4. Example of an intuitive user interface on advanced pump condition monitor

devices are inexpensive and will pay for themselves quickly by reducing maintenance costs. For critical applications, extremely hazardous liquids, special pump designs or larger, relatively expensive pumps, consider including an electronic diagnostic monitor. Choose from the available technologies of canned motor pumps, non-metallic lined mag drives and metallic mag drives based on application. And avoid “one technology fits all” decision making. You wouldn’t put the same seal on all applications, would you?

Summary As with all mechanical equipment, the results depend on a thorough understanding of the capabilities and requirements of the technology. A firm knowledge of your process, along with an open relationship with a manufacturer who thoroughly understands sealless pump technology, can yield similar results for your company.■ Robert Mulholland is Director of Sealless Products for Sundstrand Fluid Handling Corporation in Arvada, CO. He received his BSBA from Regis University and has been with Sundstrand since 1974. He was the General Manager of Sundstrand’s Milton Roy subsidiary from 1992 - 1996 and has been closely associated with sealless pumps since 1983.

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R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

Laser Shaft Alignment of a “Bark Hog” by Pumps and Systems Staff with Ludeca, Inc. any paper mills increase efficiency by converting waste to energy. One machine critical to this process is the hammermill, commonly referred to in some regions as a “bark hog.” Any wood products that cannot be used in the paper-making production line are often routed to the bark hog for preparation before they are incinerated. Heat generated in the incineration process is used to fire boilers for the creation of steam, which in turn drives turbo-generators, producing electricity for the plant. Hence, bark hogs are very important. They vary in size, configuration and operating speed, but all work on essentially the same principle. Stock is fed into the hammermill (usually motor driven) from above, where it is shredded by rows of swinging metal arms pinned to a rotor. Controlled mayhem is a polite way of describing what goes on inside a bark hog. Proper balance and alignment are critical in protecting bearings and increasing the uptime of these important pieces of equipment. Some bark hogs are easy to work with, while others are quite difficult. Maintenance and problems with bark hogs increase with feed stock size, large improperly sized/mounted hammers and high rotor speed. A recent failure and rebuild of a ham-

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Figure 1: Bark hog alignment dimensions

mermill in the southeast was precipitated by the first two of these conditions. The ensuing alignment reveal-ed unique factors that contributed to the poor performance history of this unit. How poor was the performance? If you view bearing change-outs every eight weeks as acceptable, then the unit had an acceptable failure rate. Unfortunately for the end user, a six to ten week MTBR was unacceptable. Though the unit operated at a mere 600 rpm, production people frequently OK’d its operation with overall axial vibration levels in excess of 20 mils peak to peak. Why? Unfortunately, they did so out of ignorance. Considering that the rolling element bearings used in this application have internal clearances in the range of 5 mils, it is not difficult to surmise the havoc that 20 mil displacement readings can wreak. Spectral data confirmed the existence of a large 2X signal with several lower amplitude harmonics present as well. Conclusion: classic misalignment. The alignment procedure used on this vital piece of equipment was woefully ill-considered. Overall configuration of the machine assembly is shown in Figure 1. A hammermill is coupled to a 2500 hp motor by means of a 64 inch spacer shaft employing gear couplings. Major obstacles to precision alignment are: • Severe backlash generated by a gear coupling (Photo 1). • Lack of a precise method for shaft rotation. It is nearly impossible to stop rotation at predetermined clock positions. • Gross misalignment that exists over a 64 inch spacer. • Elastic deformation/shaft sag due to rotor weight when the unit is The Pump Handbook Series

Photo 1. Gear coupling between the hammermill and the motor

Photo 2. A typical bark hog

Photo 3. Close-up view of a bark hog’s hammers

not being rotated (Photos 2 and 3). • Torsional deformation of the shafts when they begin turning. Measuring rotating shaft centerlines is the superior approach for determining shaft misalignment. Consequently, precision alignment readings are best obtained when the rotor is turned. Eliminating elastic

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equivalent to one mil per inch. Vibration consultants stated that prior to the precision alignment in August 1997, vibration readings in the neighborhood of 20 mils peak to peak were not uncommon for this machine assembly. Post alignment vibration readings were recorded at 3 mils peak to peak. The unit continues to run problem free.■

Figure 2. Alignment results are expressed as angles at each coupling plane. The unit to describe angularity is the mrad or milliradian. One mrad is equal to one mil per inch.

deformation as a factor in alignment readings requires initiating shaft rotation before gathering alignment data. Therefore, the ability of an alignment transducer to collect data “on the fly” is of crucial importance in this application. Dial indicators were employed during previous alignment procedures. Victory was always declared, though alignment was rarely achieved. One problem with dial indicators is that they must be used at clock positions. The crane used for turning the rotor rarely positioned the unit at clock positions. Further, obtaining measurements took too long. Repeatability between readings was not achieved due to elastic deformation of the shaft. Previous alignment attempts using a laser system less capable of handling shaft distortion issues and gross misalignment also failed. Just as with dial indicators, many laser measuring devices require measurements at pre-defined clock positions. Additionally, large initial misalignment involving spacer shafts cannot be measured by many commercially available laser shaft alignment systems because they lack range extension features. After 30 hours of unsuccessful attempts, Ludeca, Inc. was contracted to finish the alignment. Measurements for the latest alignment attempt on this machine assembly were made using the Rotalign laser shaft alignment system. This state-of-the-art measuring device offers five measuring modes capable of handling a variety of alignment applications. The patented continu-

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ous sweep measuring mode proved ideal for overcoming obstacles to precision alignment for this application. The shaft was turned 360 degrees by a crane in preparation. Thus, the ”bow” was removed from the rotor along with its demonstrated corrupting influence to alignment data. Measurement data collection was carried out while the rotor continued to rotate, eliminating the need to stop at clock positions. Rotor sag never had the opportunity to rear its ugly head again. Further, Rotalign is virtually insensitive to coupling backlash, a condition that sometimes affects dial indicators and first generation laser shaft alignment systems. Piping and other obstructions to the laser do not affect the Rotalign because it requires only 75 degrees of rotation to calculate extremely accurate alignment results. Rotalign’s range extension feature was not needed even though initial foot movements of more than 800 mils were measured. The alignment was completed in two moves (one vertical/one horizontal) over two and a half hours. Figure 2 shows final alignment results expressed as angles at each coupling plane of the spacer. The tolerance prescribed for this alignment is an angle of 1.8 mils per inch at each coupling flex plane. In other words, if a dial indicator is installed to generate a face reading across each coupling of the spacer, it should yield an angle value no greater than 1.8 mils per inch in horizontal and vertical orientations. The engineering unit to describe angularity in Figure 2 is the mrad or milliradian. One mrad is The Pump Handbook Series

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R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

Balancing Pump Impellers by Gordon Hines, Hines Industries What Is Balancing? alancing is the procedure by which the mass distribution of a rotor is checked and, if necessary, adjusted in order to ensure that the vibration of the journals and/or forces on the bearings at a frequency corresponding to service speed are within specified limits (Ref. 1). In practical terms, balancing corrects unbalance. There are two types of unbalance, single and two-plane. Single plane unbalance can be pictured by imagining a disc shaped part, such as a bicycle wheel, with a weight taped to the rim. When the bike is lifted off the ground, the wheel rotates and comes to rest with the weight at the bottom. If you were to spin the wheel, the bike would shake as the wheel tried to rotate about the center of the wheel’s mass, which is no longer located at its axle. The center of mass is displaced from the geometric centerline. Another way to illustrate this effect is to place a lump of modeling clay on the inside a Frisbee’s rim and throw it. Instead of flying normally, the disc will wob-

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ble as it spins around the shifted center of mass. Unbalance of this type is sometimes called force or static unbalance. It can be corrected by removing the weight or by adding an equal weight directly opposite (180 degrees from) the unbalance weight. Either measure would move the center of mass back to the centerline of the part. Two-plane unbalance can be pictured by imagining a cylindrical or drum-shaped part, such as an automobile wheel’s rim, with one weight attached at one end of the cylinder and another attached at the other end, but offset 180 degrees from the first weight. Note that the ends of the cylinder are in different planes. If the rim were raised off the ground, it would not rotate as the bicycle wheel did. Spinning the automobile wheel, however, would cause it to wobble as it sought to rotate about the axis of its mass, which is no longer parallel to the geometric axis. Two-plane unbalance is sometimes called couple unbalance. It can be corrected only by adding two correction weights at an axial distance from each other (Figure 1). Unbalance Mass

When both single and two-plane unbalance are present in a rotor, the condition is called dynamic unbalance. To correct this type of unbalance, one must compensate for both eccentricity (caused by static unbalance) and wobble (caused by couple unbalance). In practice, any dynamic unbalance can be corrected by making adjustments in two axially separated planes. However, as the planes get close together, couple correction weights become very large.

Why Balance? The forces due to unbalance increase as the square of rotational speed. Thus, an impeller running at 3600 rpm produces 16 times as much force as it would at 900 rpm, due to the same unbalance. Failure to bring parts into dynamic balance within close tolerances causes rapid bearing wear and requires frequent maintenance. While the cost of rebuilding might be high, most of the loss is in productivity, which can run into thousands of dollars a day. Not only does a well balanced impeller run better and longer, the end user’s perception is that it is running smoothly. In other words, the end product is not only of higher quality, the quality is easily recognized by the user.

Mass Axis

How Do I Balance?

Geometric Axis

Unbalance Mass

Figure 1. Couple unbalance The Pump Handbook Series

As mentioned above, parts that have significant “thickness” (as opposed to disc shaped parts) need to be balanced dynamically. For parts such as pump impellers and blowers, a horizontal overhung balancer can make the balancing process simple and easy. The part is mounted on a horizontal shaft held by a motor-driven spindle. The spindle incorporates a pair of vibration transducers that make it possible to measure dynamic un-balance. The operator mounts tooling that

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will accept the part to be balanced and enters the part’s dimensions (essentially the axial location of the two planes where corrections will be made and the radius at which weight will be added or removed). When the part is spun, the balancer indicates the amount of correction to be made in each plane in terms of weight to be added or removed and the angular location of the correction. The process is computerized, and all measurements and calculations are displayed on a large color CRT. The part is rotated manually until the computer angular position display matches the present angle with the angle of unbalance correction for one of the two planes. The operator then knows precisely where to make the correction for that plane. The ergonomics of the machine allow the user to learn how to operate the machine in a matter of minutes. The machine is also equipped with a brake that makes on-machine correction possible (Figures 2 and 3).

How Well Do I Need to Balance? The International Standards Organization (ISO) has issued guidelines with regard to a number of different kinds of devices. A pump impeller has a recommended “balance grade” of G6.3. In practical terms, that grade

is achieved by balancing a rotor that is to run at 3600 rpm to a value of 0.01 ounce-inches per pound of weight. As an example, assume an impeller weighs 1000 ounces (62 pounds). The tolerance would be 0.62 ounce inches. This means that the center of mass and geometric center must be aligned within 0.00062 inches. If the balance correction radius is 5 inches, the correction weight must be selected accurately to within a tenth of an ounce, about 2.8 grams. Actually, this is a slight simplification. The ISO standards contain detailed methods of calculating different static and couple unbalance tolerances that are dependent on the ratio of the part’s diameter to its length.

Figure 2. Balancing procedure for a horizontal overhung balancer

Repeatability of 0.01 ounce inch would at first glance seem to indicate that a 1 pound impeller could be balanced on this machine. Another limitation, however, is simply a physical one. It was just stated that we are aligning the mass and geometric center of a 62 pound part to within 0.00062 inches. Though this can be accomplished, the part must be mounted on exactly the same center in the final assembly or the balance will not be as good as that achieved on the balancer. Suppose, for example, that the worst case fit of the impeller on the shaft in the final assembly is 0.002” clearance, which will be taken up by one or more setscrews. This moves the part off center by 0.001”, causing a balance change of 1.0 ounce inch due to the mass center shift. Maintaining a pre-

How Well Can I Balance? How well parts can be balanced depends on several limiting factors. Our HO-100 (100 pound capacity) overhung balancer, for example, has a specified repeatability of +/- 0.01 ounce inches. This limitation is one of “pulling the signal out of the noise.” There are two major types of noise: thermal and mechanical. Thermal noise is produced by electronic devices and transducers. Mechanical noise is produced by moving elements such as belts, motors, spindle bearings and windage as the part is spun.

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1 Figure 3. Balancing procedure for a cradle balancer 1. 2. 3. 4. 5. 6. 7.

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adjust stanchion distance adjust stanchion bearings, left and right lay ground arbor in cradle place belt on drive pulley adjust belt tension device adjust shaft hold-down device, left and right position and adjust left end stop

8. locate strobe or pulse angle mechanism 9. put memory mark on shaft 10. load part 11. put locking collar on shaft 12. position and adjust right end stop *13. Test spin and calibration of analyzer * Soft suspension balances only

The Pump Handbook Series

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How Well Can Tooling Center a Part?

Photo 1. A horizontal overhung balancer with grinding correction

Photo 2. A CNC balancing machine equipped with a vertical spindle

cision balance in the final assembly requires precision fits. A precision balance is a wasted effort if the part is not centered as well in the final assembly as it is on the balancing machine.

How well can a part be centered on a balancing machine? We have found collet tooling to be capable of repeated centering within +/0.00002 inches. In the case of the previously discussed 62 pound part, this centering accuracy limits the balance repeatability when the part is removed from the collet and reinstalled to +/- 0.02 ounce inches. If that is added to the basic machine repeatability of +/- 0.01 ounce inches, the uncertainty becomes 0.03 ounce inches. If the part is rather “long,” the ISO tolerance is split in half for each correction plane, i.e., 0.31 ounce inches per plane. Thus, a balancer using standard collet tooling is capable of resolving about 1/10 of the part tolerance, which is adequate to allow rapid balancing to tolerance without “chasing” the balance correction around the part. The advantage of an expanding collet is that it takes up the tolerance in the bore size and still maintains the centering of the part. Collets in the range of an inch diameter can expand 0.010 inches or more. In some applications where an impeller is to be run at low speeds, there is a lower cost, alternative tooling method—solid post tooling. The part is placed on a post and locked down with a drawbolt and cap. The limitation of this type of tooling is that the post must be smaller than the smallest allowed bore. Clearance of 0.0002” to 0.0005” is required. If the tolerance on the bore is 0.002”, a maximum clearance of 0.0025” may be encountered, and repeatability with part removed from and replaced on the tooling can be +/- 1.25 ounce inches! That is, a total shift of 2.5 ounce inches can be observed. At the other end of the range of possibilities are special grease filled arbors. These have a very thin section that can be expanded by tightening a setscrew that compresses the grease. Such devices are capable of two or three times better centering than a standard collet that is expanded by sliding it on a tapered locator. These devices are, however, specialized and expensive.

How Can Unbalance Be Corrected? Pump impellers are generally balanced by removing material, by The Pump Handbook Series

Photo 3. A horizontal overhung balancer with tooling and a computer with a vector display of unbalance in two planes.

drilling, milling or grinding. Horizontal overhung balancers can be fitted with grinding correction (Photo 1). This method spreads the material removal process over a large area with minimum depth. Impellers can also be balanced by drilling or milling, if the material permits. Machines can be supplied with correction devices and automatic correction cycles. On-machine correction is fast and accurate.

Does Machine Configuration Make Any Difference? Items such as pump impellers have been balanced on cradle balancing machines by placing them on a balance mandrel. That is, the part is mounted on a shaft with a pulley. The shaft and part assembly are placed on a pair of bearings on a cradle balancing machine. The bearing spacing needs to be adjusted from part to part. A belt is placed on the drive pulley, and the part is balanced. Distances from bearings to correction planes must be precisely maintained in order to achieve an accurate display of corrections at both planes. On completion, the part must be removed from the mandrel. In addition to the several steps required, there is the drawback that mandrels can become damaged. A bent mandrel causes the part to be improperly balanced. Damage or correction debris at the bearing surfaces can cause false readings or poor repeatability of readings, making balancing much more difficult. The above discussion regarding the limita-

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tions of solid post tooling also applies to a mandrel. We believe that the horizontal overhung balancer is the answer to these concerns. The part is simply slid onto the tooling, the tooling locked, the part unbalance measured and corrected, the tooling unlocked and the part removed. Some smaller impellers can be loaded directly by an unassisted operator and conveniently balanced on a vertical spindle machine. Photo 2 shows a CNC balancing machine with grinding correction that automatically balances impellers from 5 to 25 lbs. to very fine tolerances in one or two minutes

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(mass center to geometric center of 0.00001”). Photo 3 shows an horizontal overhung balancer with tooling and a computer with a vector display of unbalance in two planes.

Conclusion Precision balancing and careful assembly of impellers will result in a smooth and quiet product whose quality is recognized by the end user. A two to five minute operator investment ($5.00) to balance a part to a very fine tolerance can result in many days or even months of added “uptime.” ■

The Pump Handbook Series

Reference 1. International Standard ISO 1925, Balancing (Second Edition). Gordon Hines introduced the HO balancing machine to the pump industry in 1981. It was designed and built for pump companies because of the difficulty of using the cradle balancer to balance impellers. Hines has been involved in the design and manufacture of balancing machines since the early 1960s.

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Revitalizing Vertical Lineshaft Turbines Learn the causes of wear, the ways to scout them out, and how to bring your vertical pump back up to speed. by David LaCombe, American Turbine here are three major reasons to revitalize a vertical turbine pump. One of the most obvious is to prevent an unexpected and complete pump failure, thereby disrupting critical system or plant operations. Secondly, increased clearances between the bowls and enclosed impeller skirts, increased clearances between the bowl shaft and bearings, and break-

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down of the fluid passages reduce the hydraulic efficiency of the unit. Over time, running a badly worn pump can be just as costly as some system shutdowns. Finally, revitalization may be required where operational conditions have changed — for example, alterations in pressure and capacity requirements due to fluid level variations. Revitalizing a vertical lineshaft turbine pump can involve a variety of processes, from adjusting the impeller position on a semi-open impeller pump to replacing the entire bowl assembly. In between these extremes, replacing bearings or shafts, flame-spraying coatings on shafts, applying coatings to flow surfaces, wear ringing close tolerance parts and replacing shaft seals are some of the more common refurbishing processes. Often, these intermediate repairs are more expensive than simply installing a new bowl assembly. Individual parts are costly, and such repair operations normally require considerably longer downtime than rebowling. While some companies offer complete repair facilities, any lineshaft turbine manufacturer will at least have some form of a rebowling program.

Causes of Vertical Turbine Pump Wear Barring any operational problems and normal wear, there are two primary factors that cause a vertical turbine to deteriorate in performance: axial misalignment and pumping a fluid that the system was not designed to handle.

Axial Misalignment Lineshaft turbine pumps are relatively simple and rugged pieces of rotating equipment. Their proper operation and performance, however, depends on delicate axial alignment. Referencing the rotational axis of the unit, all butted, machined surfaces must be perpendicular, and all registers, bushings and bushing housings must be concentric. The consequences of misalignment are vibration, undue wear and compromised performance. Fluid Properties and Suspended Solids The longest equipment operational life for a vertical turbine is generally obtained when pumping cool, clean water. Often is the case, though, that turbines are adapted for other fluids or have contaminants introduced into the fluid media that will degrade the machinery and reduce its useful life. Material erosion is expected in areas where rotating parts have a minimal clearance with stationary parts, such as an impeller rotating inside a bowl or a shaft rotating inside a bearing. However, certain constituents suspended in the fluid also contribute to erosion. Gases such as air, carbon dioxide or water vapor from cavitation can physically destroy the pump when they are entrained back into a liquid state. Suspended solids such as sand can cause the same results as a sand blaster. Finally, corrosion brought on by aggressive fluids deteriorates the wetted parts of the unit, creating additional turbulent flow channels.

Figure 1. A typical vertical turbine pump configuration The Pump Handbook Series

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because they take up the space where the liquid should be. Cavitation can eventually destroy the pump.

Photo 1. A misaligned shaft caused this bearing bore to be “whipped out.”

Photo 2. Scale buildup in a bowl

Monitoring Performance The only way to know that a turbine pump is deteriorating is to monitor its performance and compare the production and power usage to previous levels. A gradual decrease in production, increased vibration levels and increased power usage are all indicators that the pump may be wearing. Operational Problems The loss of production can have several causes, depending on the application. In a well pump, it could mean the water level is dropping. In any application, it can indicate that the bowl assembly is beginning to wear (i.e. the bowl and impeller clearance and/or the bowl shaft and bearing clearance is increasing). If the performance reduction is sudden and dramatic, the causes could range from a reduction in operational speed due to faulty controllers, power supply irregularities or other driver failures. Other causes include an impeller becoming dislodged from the shaft, or holes developing in some part of the flow channel. If the pump never met the expected performance to begin with and the unit uses a semi-open impeller, the lateral may not be properly adjusted. Also, check the application to make sure sufficient NPSH is available to prevent cavitation. The gases produced from cavitation decrease liquid flow rates simply

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Vibration Excessive vibration when the unit is first put into operation likely suggests an alignment or driver problem. Vibration that progresses gradually indicates that close tolerance parts are wearing due to normal operation, the pumping of unclean fluids or slight misalignments. In the case where vibration begins at startup, the first component to check is the driver. Disconnect it from the pump and operate it by itself. If the vibration level is the same, the problem is in the driver. If it operates much smoother, there is a problem in the pump itself. Check the straightness of the shaft extending through the driver and the shaft connections. Any rust inhibitors applied to the top of the discharge head must also be removed. These are some of the most common problem areas, but every point of connection has the potential for misalignment. Power Usage Abnormal power usage indicates problems as well. The lateral may not be set properly to prevent the impellers from dragging in the bowls, misaligned shafts may be forced into contact with the bearings (Photo 1), or the clearances around rotating parts may be too tight because of scale buildup (Photo 2). These situations will add to the driver load and cause it to consume more energy to produce the same amount of work. Scale and other types of surface degradation can also build up in fluid passages. Unless other system changes occur, it is likely that a loss in production and reduced power consumption will be noticed. Debris around the suction screen can have the same effect. If the power consumption stays constant, but production levels drop, there is probably a hole somewhere in the fluid passages. The pump is doing the work; it just doesn’t show up where it’s useful. The energy is being expended pumping some of the fluid back to the source.

Making the Diagnosis Once a pump is disassembled and The Pump Handbook Series

Photo 3. Sand wear on a bowl shaft

Photo 4. The scoop marks in the bowl are signs that gas was in the fluid.

inspected and the pump components appear to be degraded abnormally, it’s time to decide whether the pump is repairable or if it should be replaced. But before making any repairs, pinpoint the cause so that the right problem is corrected. Visual clues often indicate exactly what the problem is. If you cannot determine the cause, contact the pump manufacturer for assistance. Visual Clues Wear patterns do give clues to their causes. Bowl shafts that come out hourglass shaped at the bearing locations usually indicate that the pump has been handling sand (Photo 3). Scoop marks in the bowls similar to a horse’s hoof print in the sand are signs that gas is in the fluid (Photo 4). Impellers that are coated black and cast iron bowls that are soft enough to cut with a knife are evidence of graphitization. Graphitization is reaction of the metals to a catalyst, such as carbonic acid, that removes the iron from the bowls and plates it onto the impeller. If a bowl bearing or the impeller bore at the top of a bowl with enclosed impellers is worn eccentrically and the corresponding wear on the rotating part is concentric, this indicates that there is a problem in a stationary part. Debris, such as small pebbles, can work themselves between the two surfaces. In some cases the faces of mating parts, such as the bowls or pipe, are not assembled or machined parallel. During assembly

recent system maintenance operations or other developments that could possibly affect performance? The more information you give, the more likely it is the manufacturer can help determine the problem and suggest a solution.

Rebowling Photo 5. Worn enclosing tube bearing

all parts should be clean and smooth so that they mate properly. Foreign material between the mating faces produces the same wear signs as faces not being machined parallel or bearing bores and/or registers that are not concentric with the rotational axis. If the wear on stationary parts is concentric and on rotating parts is eccentric, there is a problem in the rotating assembly. The shaft faces must be parallel to each other and the shafts should be straight to within .005” (TIR). The shaft coupling may have been installed into the tapered portion of the threads. The coupling must stay on the straight portion of the threads or it will tilt the shaft. Similarly, the ends of the shaft must be smooth and clear of debris. In both cases, improper assembly will simulate a bend in the shaft. The ends should be faced parallel after being straightened. If the ends are machined prior to straightening and then faced, they will no longer be parallel.

Troubleshooting Information Above are just a few of the things that can happen. The variety of problems seen in one application will be few compared to all the applications that the manufacturer sees. Pump manufacturers can help you troubleshoot. When contacting them for assistance, give them all the data you have available. If a flow meter and pressure gauge are available, record flow and head readings at several points, including shutoff if possible. Take amperage and a voltage reading at those points also. Give some history on the pump’s operation. Did the problem develop suddenly or gradually? Have there been any changes in the application, such as temperature fluctuations, fluid level variations, flow or head requirement changes,

Just the cost of downtime can easily outweigh the cost of rebowling versus repair. If minimizing downtime is critical to operational processes, a spare bowl assembly may be in order. In that instance a bowl assembly can be removed and replaced in one procedure, and the original unit can be repaired for future use. Taking it one step further, if the unit is a short-coupled pump a complete standby replacement may be advisable. When ordering a replacement bowl assembly, all design parameters need to be given to the pump supplier. In some cases the information may not affect the cost, but it will help the manufacturer determine what types of special adaptations will be required. That information can affect production time estimates. The minimum information that should be provided includes, but is not limited to, the following: head, capacity, operating speed, driver rating, fluid temperature, NPSH available, length limitations, width or diameter limitations, materials of construction, and bowl to pipe and shaft connection information. Column connections are the most difficult obstacle in rebowling. Threaded column connections are the simplest to match, because normally they are machined to a pipe thread standard. Flanged and bolted connections, which do not have standards with respect to mating to a bowl assembly, require exact measurements so that adapters can be made. Threaded enclosing tube connections are close behind the flanged pipe in the variations that can exist. Finally, the thread standards on the end of the bowl shaft must be matched. Although industry-wide thread standards exist for the shafts, exact dimensional information is still required to match it to the proper standard. If at all possible, at least the parts to be mated (enclosing tube bearings, shaft couplings, and column adapters) should be sent to the The Pump Handbook Series

pump supplier to ensure that all pieces fit together properly.

Typical Bowl Assembly Repair Operations If it is not the case that rebowling is less expensive than refurbishing, you will find that there are typically four areas within a bowl assembly that are economically repairable: the bowl bearings, the bowl shaft, the bowls (diffusers) and the impellers. Some manufacturers offer both repair and replacement services. Bowl Bearings Our first consideration is the bowl bearings. Most often they are bronze, rubber or a combination of the two. Other materials used include bronze backed rubber bearings (commonly called “marine” bearings), Teflon and carbon composites. The typical shaft/bearing clearance is .010”, except when rubber bearings are used. In that case the clearance can run up to .030”, which makes it a much more forgiving material in applications where abrasive solids are present. Bowl Shafts The bowl shaft is typically made of some grade of stainless steel rather than a carbon steel because of the resistance of stainless to abrasion/corrosion. If abrasive solids are pumped, chrome or other hard coatings are sometimes applied to the shaft. Unless the shaft is an abnormal size or material (making it expensive to replace) a new shaft is usually installed rather than refurbishing or coating the old one. The shafting is the heart of the pump. Make sure all shafts are straight and properly machined before assembly. Bowls and Impellers The third and fourth parts, which are limited to enclosed impeller pumps, are the bowls and impellers. Specifically, the areas around the impeller skirt and the adjacent area in the bowl are closely examined. The clearance between these two parts is typically around .015”. If there is sufficient material available to bore the bowl and/or turn the impellers, wear rings can replace the worn material and bring the parts back into tolerance. In some cases, both parts require wear rings. In oth-

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ers, however, it may be possible to bore or turn one part, while the other accepts a wear ring specifically sized to meet the machined part. Wear rings are sometimes installed as part of the original assembly. Unless the fluid contains some particularly abrasive solids and the wear rings are made of materials to minimize the abrasive erosion, forego the expense of wear rings until the parts are worn. Then, during maintenance or repair operations, the parts can be machined to accept the wear rings. If wear rings are ordered with the original unit for easy replacement later, both bowl and impeller pieces are necessary to return the bowl assembly to its original tolerances. Column Pipe, Enclosing Tubes and Shaft Supports The next section of the pump to consider is the column assembly. The pipe may be pitted or encrusted with scale, or the coating may be damaged. The degree of roughness will be the determining factor as to whether the pipe should be reused, recoated, cleaned or replaced. The exterior of enclosing tubes, when used, take the same consideration as the pipe. The faces of the tubes should be inspected for leakage or damage, and the tubes themselves should be replaced when necessary (Photo 5). Enclosing tube bearings should have approximately the same tolerances as the bowl bearings (.010”). Open line shaft bearings should be replaced when they are visibly damaged. Most often they are rubber and can be purchased directly from the pump manufacturer. The condition of these bearings does not necessarily affect efficiency, but it is vital to supporting the shaft. Shafts This brings us to shafts, which are commonly available in two materials: stainless steel or carbon steel with an alternative wear surface. Some grade of stainless steel is often used when aggressive fluids are handled. In open lineshaft pumps, if the wear is not excessive enough to weaken the shaft and it is threaded the same on both ends, it can be turned over and reused. This places a new wear surface in contact with the column bearings. Care must be taken, though, in examining the faces of the used shafts. Recesses in the ends of the shafts may form ridges in the oppos-

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ing faces that will need to be removed before the unit is reassembled. Under the same conditions, similar methods can be used on the carbon steel shafts, such as placing new sleeves on the opposite end of the shaft or renewing the hardened coatings. Much like the bowl assembly, consideration needs to be given to the cost of simply replacing the parts instead of repairing them.

are specific to vertical turbines and require a trained eye to identify them. Whenever possible, send the complete pump to the repair shop. When it is torn down and ultimately reassembled, there will be no question that the parts fit. If the fits are not checked until they’re in the field, the cost of having personnel and equipment waiting while adjustments are made can be high.

Discharge Heads The most visible part of the pump, the discharge head, generally requires the least maintenance. Most of the repairs are performed on the shaft seal, whether it is a stuffing box or a mechanical seal. Packing in the stuffing box should be replaced to control leakage. Stuffing box or seal housing bearings need to be checked to make sure the shaft is properly supported. Send mechanical seals to the original manufacturer or a local repair expert for inspection.

Why Revitalize?

Drivers Last but not least, the condition of the driver should be reviewed. The thrust bearings should be checked to ensure that they are not excessively worn. Motors upgraded to premium efficiency or at least replaced or rewound can significantly increase overall pump performance. Motors, gear drives and engines should be inspected by specialists to determine any maintenance procedures that will optimize the system’s performance. Other Repair Considerations The fluid passages of the bowl and impeller are often overlooked by maintenance personnel. If they are damaged or pitted or have a particle buildup or eroded coatings, there can be a significant decrease in efficiency. Coatings are normally applied to the interior of the bowl assembly, mainly to enhance efficiency. Although they are applied for protection in some cases, the velocity of the fluids and suspended solids being pumped often makes it difficult to keep the coating intact. Cleaning, recoating or replacing the affected parts are all options, depending on the severity of the problem.

Repair Specialists Who should inspect the unit and perform repairs? Make sure it’s someone who is experienced and qualified to work specifically on vertical turbine pumps. Some problems The Pump Handbook Series

Assume a 12” bowl assembly is in operation and the bowl/impeller clearances change from .015” to .064”(1/16”). The loss in capacity would be in the neighborhood of 10%, and the loss in bowl efficiency would be approximately 2.5 points. Unless the pump has some type of speed controller, such as a variable frequency drive or an engine that can be accelerated, that loss in capacity cannot be regained. In the case that other parts of the system, such as the discharge or column pipe, become irregular from damage or buildup, some capacity can be regained by opening valves. However, more power will be expended to meet the same production rate. A 1000 gpm pump operating at a total dynamic pressure of 100 psi will cost between $200 and $300 more per year to operate per point drop in overall efficiency. That is assuming operations of eight hours per day, every day of the year, with a power cost of $0.10 per KW-h. The bottom line is that the user pays more in operational costs when operating a vertical turbine pump that has deteriorated beyond its original design tolerances. The challenge maintenance personnel are faced with when the pump’s performance is no longer acceptable is to determine the most cost-effective option— either eliminating the cause of the deterioration or minimizing its effects. Don’t wait for a catastrophic failure. Perform regular maintenance repairs when downtimes are acceptable. Otherwise, the pump system will inevitably break down when it is least desirable. ■ David LaCombe is the Engineering and Operations Manager for American Turbine Pump Company. He has been with the firm for 10 years. He is also a graduate petroleum engineer from Texas A&M University.

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Sealing Lime Slurry at Alberta-Pacific By David Djuric n the recausticizing department at Alberta-Pacific Forest Industries, Inc. there are two 8 x 6 x 21 PWO ITT-AC pumps, one in operation and one on stand-by, that are used to pump lime slurry from causticizers to white liquor pressure filters at a rate of 2013 gpm to a 99’ TDH. Solids content is 15-20% with a specific gravity of 1.17. A backwash of the pressure filter is required during system operation to clean the filter socks. The backwash cycle causes cavitation and vibration problems for these pumps — the life of the wet end components is only about 18 months. The cycle can be initiated by several conditions:

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1. volume through the system 2. differential pressure between the body of the pressure filter and the top of the filter socks 3. a timer set by the operations personnel 4. manual backwash initiated by the operator The Backwash Cycle 1. The outlet valve on the pressure filter closes. 2. The pump continues to operate, building pressure in the dome of the white liquor pressure filter to 50 kPa (7 psi). 3. The by-pass valve opens to the causticizers. 4. The infeed valve to the pressure filter closes, and pump speed ramps down for 20 seconds. The pressure filter blows back through the socks in the filter, cleaning them. 5. The outlet valve on the pressure filter opens. 6. The pressure filter’s inlet valve opens. 7. The by-pass valve to the causticizers closes, and the pump speeds up again.

We have experienced problems with the mechanical seals on these pumps since start-up of the mill in 1993. Originally, the seals on these pumps were Chesterton 233s, and the face materials were RSC/CB/ RSC/CB - ethylene propylene Orings. The shaft diameter is 4.999” with a 5.000” mechanical sleeve. The stuffing box was originally specified for packing with a two-bolt gland. To achieve more reliability from these seals, we changed the face materials of the Chesterton 233 seal to TC/SSC/SSC/CB and changed the O-ring material to AFLAS. This improved the seal’s reliability. However, we were now faced with distortion of the stationary seal faces because of the two-bolt gland. We wanted a large bore oversized stuffing box with a four-bolt gland for equal and opposite clamping. A large bore oversized stuffing box was not available for this pump, so we machined back the face of an existing stuffing box to allow a machined flange with a four-bolt pattern to be welded to it. The bore of the stuffing box was opened to the maximum possible, and the bottom of the box was removed to allow the slurry to be flushed out of the box. This prevented solidification of the product in the stuffing box when the pump was shut down and switched over. The modified stuffing box was placed on the pump to check concentricity and squareness to the shaft before the mechanical seal was installed. We chose to install Chesterton’s 280 M (mixer design) heavy duty dual slurry cartridge seal with TC/TC/CB/TC faces and AFLAS Orings. We believe that this is the best solution to the problems we experienced with these mechanical seals — The Pump Handbook Series

Photo 2. PWO ITT-AC B-3 stuffing box modified from two-bolt standard packed box to a four-bolt enlarged bore seal chamber for maximum radial clearance and even clamping.

Photo 1. Chesterton 280 heavy duty seal

at a fraction of the cost of the previous failures.■ Dave Djuric is a Mechanical Maintenance Specialist at AlbertaPacific Forest Industries, Inc. in Boyle, Alberta. He has more than 25 years of experience as a millwright. He will be speaking on “Reliability-Centered Pump and Seal Maintenance” at PumpUsers Expo ‘98.

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AlliedSignal GeismarWorks One team’s comprehensive improvement program reduces failures and saves more than $1 million. by Randall Jones, AlliedSignal Background lliedSignal GeismarWorks, located in Geismar, Louisiana, is the world’s largest producer of hydrofluoric acid. The company also produces environmentally safe fluoro-carbons (HFCs and HCFCs). Its MultiProducts Plant (MPP), which was built in Geismar in May 1994 to produce fluorocarbons, has about 150 centrifugal process pumps including mag- netic-drive, canned motor, multi-stage turbine, sliding vane, horizontal split case and other standard ANSI equipment. The millwright staffing consists of three AlliedSignal millwrights (assigned to the central machine shop, which services both the Hydroflouric Acid Plant (HF) and MPP facilities) and two contract multi-craft mechanics per shift who are assigned exclusively to the MPP. The Multi-Products Plant was built based on technology that was only previously proven in a pilot plant setting. The plant is a continuous operation and can swing production to produce either 120 or 130 series fluorocarbons. Many of the processes involve streams operating with very high vapor pressures, high temperatures, no lubricating properties and high concentrations of solids, all of which ultimately contributed for a time to high pump maintenance costs and plant downtime due to pump failures. In early 1996 we initiated a pump improvement program in the MPP as part of a Total Quality initiative. The purpose of the program was to reduce pump maintenance costs and reduce or eliminate plant downtime due to process pump failures.

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Pump Improvement Program The plant’s Computerized Maintenance Management System

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(CMMS), an older mainframe software package, was not scheduled for upgrade until 1999. It offered little more than cost information for each work order and dates of each failure. We developed a Microsoft Access™ database to manage cost and failure data and loaded one year of existing pump histories (cost and failure dates from the CMMS system). It became apparent when we analyzed the data that 16 pumps out of the total population accounted for two-thirds of the total pump maintenance costs and six pumps were responsible for all of the plant downtime caused by pump failures. Also, magnetically driven and vertical pumps caused most of the problems. We targeted these pump types for improvement and prioritized them according to their safety impact, contribution to plant downtime and maintenance cost. Before the improvement program began, millwrights in the field were repairing most of the pumps. We discontinued this practice (except for minor repairs), and set up a pump shop in the central machine shop with all repairs conducted under the leadership of the lead millwright (Photo 1). The team arranged an area of the machine shop specifically for pump repair and modified the work area to accommodate each type of pump. During each repair, personnel checked and restored all critical dimensions and running clearances in accordance with the manufacturer’s recommendation. The central machine shop was already equipped with a balancing machine, and the team continues to check and correct the balance of each impeller assembly as needed. They also use precision shaft alignment techniques in conjunction with the pump improvement efforts and pay particularly close attention The Pump Handbook Series

Photo 1. R.J. Blount, Millwright and Ray Mullins, Lead Millwright work in the central machine shop at GeismarWorks.

to pumps with frequent bearing and mechanical seal failures. Since the implementation of the improvement program we closely analyze all pumps after each failure to try to determine its cause. To aid in this process, we developed a single page repair form for recording dimensional checks and any observations related to the cause of the failure (Figure 1). The repair form is completed on each failed pump, and this information is entered into the Access™ Database on a monthly basis. The millwrights, the lead millwright, the area maintenance supervisor, process operators and the reliability engineer took a team approach to analyzing failures and solving the problems. We also brought in the pump, bearing and mechanical seal manufacturers as needed to help determine the causes of failures and to recommend solutions. Once the team members understood the mechanical failure modes, they could apply a systematic problem solving approach to determine the root cause of the failure. They identified several opportunities for improving pump performance— including process system design problems, equipment design prob-

lems, operational problems, shaft misalignment and pipe strain problems and improper pump or mechanical seal specification. The team attacked these problems by modifying some of the original equipment designs, implementing process changes and improving the quality of pump installations and repairs. Since the technologies for sealless pumps and mechanical seals were advancing rapidly during this time period, we were able to reduce or eliminate some problems by replacing existing equipment with newer equipment. In cases where extreme process conditions caused the pump failures and modifying the process was not an option,

the team targeted the pumps for more stringent predictive and preventive maintenance. They also added an online spike energy system to a group of problem pumps to help predict oncoming failures. Although the team conducted most of the failure analyses informally in the pump shop during the repair process, recurring magneticdrive pump failures made the use of a failure logic tree necessary to understand and correct the root causes of problems. A cross-functional team from the Maintenance, Operations and Engineering departments developed the tree, and an outside consultant facilitated the

work of the team. A logic tree systematically determines all possible failure modes and then hypothesizes causes. Each hypothesis is then either proved or disproved. This approach is particularly helpful when multiple causes create a single type of failure. It also prevents jumping to conclusions that may prove to be wrong or that may not be complete solutions to a problem. After completing the analysis process, the team incorporated several mechanical and process changes designed to improve the Mean Time Between Failures (MTBF) of the magnetic-drive pumps. AlliedSignal has developed a national alliance with a single manufacturer to provide mechanical seals and technical support for all mechanically sealed process pumps. We now send all failed mechanical seals to the manufacturer’s local repair facility. The AlliedSignal millwrights conduct a cursory inspection of each failed seal before sending it out for repair. When it receives the seal, the manufacturer’s repair facility conducts a complete inspection and failure analysis and faxes a single page report back to the plant. The manufacturer’s expertise is leveraged through the national alliance, which requires the supplier to produce annual improvements in productivity or cost savings. As the pump program matured, we expanded the Access™ database to generate reports listing individual pump and plant average MTBF, as well as individual pump costs, the top five ”bad actors” and total plant pump costs for given time periods. The database can also query failure histories by pump manufacturer, seal type, seal manufacturer or pump service. We use the reliability data from these queries when specifying pumps and mechanical seals for new capital projects or equipment upgrades. A write-protected copy of the database was placed on the plant computer network so the central machine shop and area maintenance supervisors also have access to the information.

Metrics Before we could develop this module of the database, we had to deterFigure 1. Example of a single-page pump inspection and repair form The Pump Handbook Series

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due to process pumps by 92%. By the end of 1997, when many of the improvement efforts were finalized and the repair program institutionalized, the maintenance costs were 52% below the 1996 levels, and no plant downtime occurred in 1997 due to process pumps. Total savings in maintenance costs and plant downtime is more than $1 million to date. The pump improvement program, institutionalized in the MultiProducts Plant, is now being implemented in the Hydrofluoric Acid Plant.

Conclusion

Figure 2. The AlliedSignal GeismarWorks Pump Tracking Program

mine criteria for the MTBF calculation. Most critical process pumps run continuously and have in-line idle spares, so we use a running MTBF calculation for these pumping systems. This calculation treats a spared pumping system as a single pump (counting the failures of both pumps), so that the plant average MTBF calculation is not artificially inflated by the idle time of inactive spares. Failures of valves, piping or instrumentation on a pump, or minor repairs such as packing adjustment are not included in the calculation. If we rebuild a pump and it fails during startup, we treat the initial repair and the rework as a single failure. Whoever does the work notes both the repair costs and the circumstances of the rework in

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the comment section of the database. We use the plant average MTBF metric to gauge the overall health of the plant pump population, and individual pump MTBF to identify problematic pumps. In addition, we also track the total pump maintenance cost and individual pump maintenance costs over time to help identify opportunities for improvement. This information is also helpful when determining life cycle costs for specific pumping systems.

Results At the end of 1996, after 12 months of improvement efforts, the pump maintenance costs were reduced by 20% and plant downtime

The Pump Handbook Series

People are any plant’s most valuable resource, and each work group brings different ideas and expertise to the table. Managing equipment assets with limited personnel and financial resources is often challenging, to say the least. These demands require maintenance professionals to be creative and to optimize the use of all resources. Improvement decisions must be data driven whether you use an elaborate CMMS system, a simple tracking program written from a canned software package, or a paper system. These continuous improvement ideas, when coupled in a team environment, can result in significant improvements in any rotating equipment program. ■ Randall Jones is the Reliability Group Supervisor at GeismarWorks. He has a Bachelor of Science degree in mechanical engineering from Auburn University and is a registered Professional Engineer in Louisiana.

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R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

Well Pump Applications for Mine Dewatering Choosing the right pumps means knowing drainage requirements, dewatering schedules and well construction as well as system and fluid conditions. by Mark List, Miller Sales and Engineering nce the economics of a mineral deposit have been determined as favorable and a decision made to proceed with mine development, significant financial commitment is placed at risk in expectation of a calculated return on investment. Many aspects of mining carry relatively high levels of uncertainty that contribute to the overall degree of risk. One important consideration is groundwater control, should it be a factor, during mining operations. Where mining must take place below the water table in highly permeable geologic material, operations would not be possible without effective control of groundwater. Several mines in the western United States have developed large well fields, up to 70,000 gpm capacity, to intercept inflow and divert groundwater from the workings.

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Mine Dewatering Objectives There are two general objectives to most mine dewatering programs: Keeping the working conditions relatively dry and maintaining the stability of the excavation or opening. Operating costs and production rates are directly influenced by working conditions. Wet floors and working faces create poor ground conditions for heavy loading and hauling equipment. They also increase tire wear, reduce cycle times and impact tonnage factors. Safety becomes a direct issue if elec-

tric powered machinery is used. In the worst case, a submerged working level becomes inaccessible and production is halted. The stability of open pit walls and underground openings is of vital concern for safety and economic reasons. Adequate drainage must occur in order to keep pore pressure at acceptable levels based on geotechnical stability analysis. General Types of Mines and Groundwater Control Methods Mines are either open pit excavations or underground excavations, or both. In some cases large scale open pit mines have succeeded prior episodes of underground mining in the same area. In other situations, underground mines are developed adjacent to or from within existing open pit mines. Driven by metal prices and advancing technology, companies have continued to explore deeper and/or more challenging geologic territory for mineable orebodies. Mine dewatering methods have evolved out of necessity in response to the increasing groundwater control requirements of contemporary mining. Where conditions permit, open pumping from collection sumps is a standard practice. This method is commonly used in open pit mines to control surface water drainage and in both open pit and underground mines where groundwater inflow rates are small enough to be The Pump Handbook Series

Photo 1. In-pit well in service with loading operation on left, blasted material on right, and high wall in background

managed in this manner. Booster stations might be required employing positive displacement pumps or horizontal centrifugal pumps designed for high head dirty water, abrasive solids or slurry service. Depending on the hydrogeologic setting, however, some mines cannot be effectively or safely dewatered using this method alone. Well Field Systems for Dewatering Several mines in northern

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Nevada require the use of wells to intercept and lower groundwater levels around open pit and underground excavations. The orebodies associated with these mines are hosted in fractured bedrock formations along mountain ranges that contact alluvial basins high in groundwater storage. Mining companies drill large diameter deep wells in bedrock fracture zones tested for favorable production yield. Other wells are completed so as to intercept shallow recharge or promote drainage of less permeable zones. Wells are typically located outside the open pit, but drilling sites in the pit are often unavoidable due to local hydraulic compartmentalization. A well field can be comprised of 30 or more individual wells with completion depths up to 2000 feet or greater and production casing diameters up to 24 inches. Well-specific capacities can exceed 60 gpm per foot of drawdown. Vertical turbine line shaft pumps (400 hp) are in service at setting depths of 1020 feet, and 1500 hp vertical turbine line shaft pumps are in service at setting depths of 800 feet. Single 2200 hp submersible units work at setting depths of more than 1800 feet, and at more than 2000 feet deep single 1500 hp submersible units are applied. Submersibles are also installed in series using tandem 1000 hp (2000 hp) units, tandem 1170 hp (2340 hp) units and combination 1500 hp/800 hp (2300 hp) units with the lower setting at depths exceeding 2000 feet. Pit perimeter wells typically discharge to gathering mains at relatively low well head pressures. Inpit well pumps can be staged for the additional head required to discharge to a surface location outside the pit, or to a booster station in the pit. Vertical turbine can pumps and horizontal centrifugal pumps are in service for this application.

Understanding the Application The uncertainties involved in developing an efficient mine dewatering program become much better understood as operations progress. Groundwater flow information

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available at the onset of large scale dewatering can be very complete and supported by sophisticated model simulations, but such information is usually based on field test data that cannot be conducted at a scale proportional to what will actually be undertaken. Granted, pump applications engineers are most comfortable when customers assume all risk by specifying the necessary conditions for pump equipment selection. The outcome is likely to be better for all parties involved if knowledge is shared prior to establishing the conditions for equipment selection. Getting the Right Concept Well field dewatering involves lowering the water pressure level in the mine area to permit safe, efficient excavation. Over the life of the mine, the change in pumping lift from the initial static water level to some future pumping water level can be large—on the order of 1000 feet or more. Individual well production capacities can decrease dramatically depending upon aquifer system characteristics. From a pump applications perspective, this means selecting equipment with initial operating points that best match starting conditions and which can be made to fit, if possible, conditions that are expected to occur as formation dewatering progresses. Dewatering Schedule and Pumping Rates The rate at which a mine is expected to be deepened below the static water level is an important planning factor. It is used to establish a schedule for lowering groundwater heads before excavation begins, and it is a major consideration in predicting the required overall pumping rate. The change in pumping lift over time, indicated by the dewatering schedule, is the variable component required to evaluate intermediate and final TDH conditions for pump selection. Pump capacity range can be estimated assuming that sufficient test or operating data are available to be confident in doing so. The most reliable values for individual well production capacity and

The Pump Handbook Series

Photo 2. Twin 800 hp, 5000 gpm vertical turbine can boosters with 42” discharge main in background

efficiency (well drawdown) are not available until the well is constructed and test pumping has been completed. However, the overall mine development schedule might not allow for the long delivery times that may be needed for special pump engineering or construction. If this potential problem is not addressed during project planning, pump equipment orders can be placed with results that are not costeffective over the long term. Well Construction Well dimensions limit the size and type of pump equipment that can be installed. Although very costly to construct, large diameter deep wells can accommodate the installation of the large diameter four-pole (1800 rpm) 1500 hp and 2200 hp submersible electric motors that are required for high volume deep set applications beyond the practical setting depth limitations of line shaft pumps. Similar deep set applications with smaller casing diameters can require the use of two-pole (3600 rpm) submersible motors with an overall length of 100 feet or longer. In both situations a well must be drilled deep enough to achieve the design pump intake elevation and to accommodate motor equipment length and standard

Photo 3. 400 hp vertical turbine line shaft (1020' setting) with mineral processing plant in background

clearances. In addition to well diameter, well alignment is of critical importance for deep set line shaft pumps. Even the closest attention to construction alignment standards, however, cannot prevent ground movement from adversely affecting well alignment as dewatering progresses. Depending on the severity of ground movement and resulting deflection, shaft vibration can result in shaft and motor bearing failure. System Considerations System conditions vary in response to production well field changes and discharge method modifications. A well head pressure condition can usually be determined for use in staging the well pump, but a conservative approach is often taken to ensure that the desired pumping capacity can be maintained. If required, throttling is used to impose pressure temporarily until system conditions are within pump operating conditions. Fluid Conditions Water temperature and corrosivity are major factors influencing the selection of dewatering well pump equipment. Water temperature can influence the type of construction and materials used in a line shaft pump, but elevated water temperature adds significantly to the cost of submersible electrical equipment and thus can be a limiting factor in selection. At one particular dewatering operation, line shaft pumps are not an option, and submersible motors rated at up to 2200 hp are operating in water temperatures of 140°F. These are oil filled motors of specialized construction sometimes fitted with heat exchangers. Because

the motors are located in the lower reaches of the wells, below the pump intake, shrouds designed for adequate water flow past the motors are usually required for cooling purposes. Unforeseen corrosion damage to pump cases, impellers and column pipe joints can ruin the best efforts in hydraulic applications engineering. Corrosion potential can sometimes be estimated up front by water quality analysis, and should be taken into account, if possible, in the specification of pump equipment materials and coatings. Unfortunately, geologic formation water conditions can and do change during dewatering. Partial aeration can occur with the rapid displacement of groundwater, and this can lead to unanticipated corrosive damage. Bronze and bronze alloys should be considered if conservatism is justifiable. If standard materials are selected, the first pump tear-down will reveal what doesn’t work.

Equipment Selection Making a reasonable attempt to understand the factors that dictate initial conditions and influence future conditions is key to selecting dewatering well pumping equipment that will remain effective under actual operating conditions. Nevertheless, there are limitations, and well yields can eventually decline to the point that pumps must either be operated intermittently or replaced with lower capacity units. Electric Submersible vs. Line Shaft Pumps Vertical turbine oil lubricated line shaft pumps are operating successfully at setting depths of more than 1000 feet. The slower pump and driver speeds (1800 rpm or less) of these units are favored by many operators. Pump damage from abrasive particles or partially aerated formation water is significantly less intense at slower speeds. Electrical problems are much simpler to troubleshoot and correct because motors are located at the surface above the discharge head. On the downside, slower speeds require larger bowls

The Pump Handbook Series

Photo 4. Pump rig installing deep set line shaft pump

and well casing diameters. Line shaft equipment is mechanically complex and requires special engineering and manufacturing for bowl tolerances to accommodate the effects of relative shaft elongation under high thrust loading. Well alignment problems can adversely affect shaft and bearing life or even preclude the use of line shaft equipment. High capacity submersible equipment is available in both 1800 rpm and 3500 rpm classes. Small diameter high yield wells or setting depths greater than 1000 feet generally restrict pump equipment to the submersible type. The limitation here is motor power output. Four-pole 20inch diameter motors (1800 rpm) are available to 2200 hp for 24-inch casing applications. Slim line two-pole motors (3500 rpm) can be coupled in tandem to produce more than 1000 hp. These two-pole motor assemblies are more than 100 feet long, requiring additional well depth. Because the motors are installed below the pumps, electrical faults that occur in the down hole power cable or motor system require retrieval of the equipment string from the well for testing and repair

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to take place. The Pump Curve For a desired initial performance and estimated final performance, there is a simple rule of thumb for dewatering pump selection: start on the right side of the H-Q curve, run back to the left through the Best Efficiency Point, and plan to refit the pump end with additional stages if necessary to conform with estimated future conditions. Another method is to throttle the pump during initial operations if the range of expected conditions indicates that this will eliminate the need to pull and refit the pump end. Throttling is most common in deep set submersible applications involving relative certainty in the drawdown rate and final conditions. Pump Mechanical Considerations Line shaft applications involving deep settings and high thrust require special consideration for relative shaft stretch and bowl endplay requirements to establish adequate lateral impeller clearance under running conditions. Enclosing tube tension design as well as manufacturing tolerances for tube, shaft and column pipe lengths also need to be considered. Surface Equipment Power transformers and switching equipment are available in modular form on skid mounted platforms. Also, individual components can be

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custom assembled on a common skid or placed on individual pads if preferred. Mine power distribution systems often are plagued with swing loads and transients depending on the variety of electric machinery in service. Power factor correction, system protection and motor control requirements vary with the application. Vertical hollow shaft type motors used with line shaft pumps are usually 460 volt or 4160 volt. Submersible motors are typically 460 volt, 2400 volt or 4160 volt. Installation Considerations Line shaft pump installation can be more mechanically involved than submersible pump installation. The oil tube and shaft are usually shipped assembled in lengths of 20 feet and must be individually placed in each piece of column pipe before installation in the well. Because the column, tube and shaft assembly are run in the well casing, three threaded connections must be properly made at each 20-foot interval. The projection dimensions of the tube and shaft, which start at the pump discharge case, must be maintained over the length of the column assembly for proper fit at the discharge head and motor coupling. Depending on the manufacturer of the submersible pump equipment, motor system and pump assembly during installation can be more or less complicated, generally

The Pump Handbook Series

requiring manufacturer’s field service in addition to the installation crew and equipment. However, once the pump and motor equipment are assembled in the well and tested for continuity, column installation is a straightforward process of making one threaded joint per pipe length and securing the power cable to the column. This goes relatively quickly, especially if the pump rig can handle pipe lengths of 40 feet. Operating Considerations It is important to follow up on the performance of a pump after it has been placed in service. The operator will no doubt inform someone associated with the sale of the equipment if a failure has occurred or performance is not as represented; conversely, the operator will be concerned with other matters if equipment performance is acceptable. Either situation involves information that can assist the applications engineer in selecting proper equipment and recommending the most effective modifications. ■ Mark List has more than 25 years of experience in the construction and mining industries, with 10 years of practice in groundwater investigation and mine dewatering equipment design, construction and operation as Chief Dewatering Engineer for a western gold mine. He is a registered engineer and state water rights surveyor in Nevada.

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R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

Hazardous Fluid Pump and Sealing Systems: Reliability Driven Improvements An engineering specialist describes his facility’s successful approch to reliability that involves equipment monitoring, categorized record-keeping, employee incentives and enhanced communication. ver the past several years at a major midwestern chemical plant, a growing emphasis has been placed on gaining production capacity through reliability improvements. There is also a growing effort to reduce operating expenses. We find ourselves dealing with the operation and maintenance of an ever increasing number of pumps but with fewer people available to handle the job. In response to this situation, we have initiated a reliability program to monitor and extend pump service life. We have also improved tracking and communication efforts. Records on pump failures and MTBF are kept according to chemical service, pump type and manufacturer. Also tracked are repair costs, and whether repairs are planned or unplanned. Additionally, our efforts to improve equipment reliability include an employee incentive system. For a given year, key cost

O

reduction targets are selected and tracked. If the targets are met, employees receive an annual payout. For the last several years, pump maintenance reduction efforts have been included as one of the program’s targets for our facility. Consistently hitting the target on pumps required that we extend and improve the working relationships between the reliability/maintenance and operating people. Shop people doing the work and the entire operating team discuss each significant pump repair, either planned or unplanned. Routine repairs are covered in monthly performance reports distributed to the site. At the request of a given manufacturing team, a special review is held. The accompanying chart (Figure 1) shows the progress made in reducing pump maintenance costs from 1991 to 1997. Cost was calculated in constant dollars in order to remove any inflation effects. The overall cost reduction was accom-

1.00

Trend Line

Repair Cost

0.90 0.80

COST INDEX

0.70 0.60 0.50 0.40

Number of pumps on Site

0.30 0.20 0.10 0.00 1991

1992

1993

1994

1995

1996

Figure 1. Annual pump repair cost at the plant The Pump Handbook Series

1997

plished even though there has been a small increase in the total number of pumps on site during this time. Below is a sampling of the areas addressed by the reliability improvement program.

Pump Training One issue to emerge from closer work relationships was the need for training. In 1993 and 1994 pump training was given to manufacturing and maintenance engineers and onsite operating teams. During that same time period, pump mechanics were given the ability to control their stock system. This resulted in improved parts availability, reduced overstock and virtually eliminated the stocking of obsolete parts.

Reactive Chemical Service: The Advent of Canned Motor Pump Use This particular plant produces and uses more than 10 reactive fluids and has well over 200 pumps in this service. Physical properties vary widely for this type of fluid. Normal boiling points, for example, can range from less than 30°C to above 200°C. From a pump sealing standpoint, the reactivity of these chemicals to water presents two difficult challenges. Normal product leakage across a seal face reacts with water to form hydrochloric acid. This will attack all the typical stainless steels and, at a lower rate, even Hastelloy® alloys. This virtually eliminated all the metal components that mechanical seals were being fabricated

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Figure 2. The nitrogen flowing through the grooves of the outer seal reduces contact loading, thus reducing seal generated heat and face wear.

with about 25 years ago. The second problem was development of other unwanted reaction products—either a sticky fluid or an abrasive solid. Build-up of these compounds would overheat seal faces and/or cause seal components to hang up. Either situation led to seal failure. The use of double seals with an oil buffer fluid was never extensively used—mainly because we could not tolerate buffer fluid contamination in the processes. About 25 years ago the company started to use canned motor pumps in reactive chemical service applications up to 20 hp. Mag drive pumps were not readily available at that time. Service life improvements were immediate as these pumps were phased in over the years. Typical life was several months to more than three years. Leaks from severe failures still plagued operations, although they occurred to a lesser extent than with sealed pumps. Several years ago we began to develop vibration monitoring as a tool to predict when maintenance was needed on canned motor pumps. It took about four years to determine what to look for and, more importantly, to gain credibility with the operators and tradespeople that this could be a valuable tool in determining when routine carbon sleeve bearing replacement work is required on canned motor pumps. With further work in vibration monitoring, problems including mild cavitation, unstable axial rotor positioning or rotor thrusting, bad mechanical balance and parts loosening could all be detected. One of the most unusual problems we have diagnosed with vibration monitoring is rotor core slippage on the shaft in a canned motor.

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This causes a downward drift on the running speed vibration peak (which is the largest vibration peak on a properly operating pump), from a typical value of 3550 cycles per minute (cpm) down to 2900 to 3300 cpm. In a slip situation, the impeller is turning a little slower, so a loss of head and perhaps flow will also be noticed. This was the case for one pair of in-line spare canned pumps on a process. Although the pumps were identical, one always performed better than the other. The operators always ran the better pump, and the other pump had undergone several tear-downs in the repair shop, but nothing was ever found to be defective. In several vibration checks this lower cpm peak was always seen, along with a notable absence of the usual running speed peak. The defective rotor finally showed itself when it was run long enough that some looseness of the rotor core— and some chemical contamination— could be noticed at time of repair and inspection. Once the rotor was replaced, the pump operated the same as its counterpart. Vibration monitoring turned out to be a major breakthrough in canned motor pump reliability. A tool was now available that could reduce catastrophic failures to routine bearing replacements in applications where canned motor pumps were correctly applied and properly operated. Our efforts have been focused on providing better feedback to the various operating teams on pump performance. In one case, a canned pump in reactive chemical service in a tank farm was failing every three months. Vibration monitoring was picking up carbon bearing wear, and the pump required bearing change-outs with such consistency that you could set a calendar to it. After about six change-outs, frustrations began to build. At a meeting to discuss the situation, shop people provided information on the type of bearing failure being caused by excessive radial loading. Operations people told how the pump was used: mostly to feed a distillation column at about 15 to 20 gpm, but on occasion to load rail cars at more than 100 gpm. The Pump Handbook Series

Engineering records showed that the pump was sized for about 120 gpm and the best efficiency flow was around 170 gpm. It was never communicated that the pump would spend most of its time running at the lower flow. The solution agreed upon was to install a recycle orifice sized at about 40 gpm. This change increased bearing life to more than 1-1/2 years. A solution to this problem would not have been found, however, if operations, maintenance and engineering people didn’t work together to find it.

Dry Gas Seals In 1994 we saw a need for higher horsepower pumps. Pumping applications were going to increase beyond the options available with canned motor pumps currently used on site at that time. (Canned motor pumps have been and continue to be extensively used in reactive chemical service.) The canned motor choices available came with a substantial jump in capital cost. This led us to begin evaluating dry gas seals, which were then becoming available for centrifugal pumps operating at 1750 and 3500 rpm. A few Goulds chemical process pumps in distillation service were selected as a test. They were initially fit with a contacting dry gas double seal (Durametallic’s GB-200) in a cartridge assembly. The features of dry gas seals are: • Seal faces run together. • Gas is used to keep seal faces clean and cool by reducing face contact loading (Figure 2). Filtered nitrogen, supplied from the plant nitrogen system, was fed in between the two seals as the purge gas (Figure 3). The nitrogen supply pressure to the seal cavity was monitored and alarmed in the control room. Flow was monitored locally at the pump. After a prestart-up review, a bottle back-up system with appropriate check valves was added to the nitrogen supply. Flow would automatically start from the bottle back-up if the nitrogen pressure dipped too low. For applications in which the required seal pressure was higher than the plant nitrogen supply pressure, a gas powered booster pump

N2 Pulsation Dampener Tank F

PG RV

PSL PR RV

Booster Pump F Nitrogen Bottles N2 F PR PSL FT FI PT PG RV

Plant Nitrogen Supply Coalescing Filter Pressure Regulator Low Pressure alarm Flow Transmitter Flow Indicator Pressure Transmitter Pressure Gauge Relief Valve

RV F FT

PT PG Dry Gas Seal

FI

Figure 3. Nitrogen from the plant system was fed in between the two seals as the purge gas.

(piston pump) was used along with a bottle back-up. A small surge tank also had to be installed to reduce booster pump pulsations to an acceptable level for the seal. Initially, seal life was only a few weeks. Failures were occurring mostly on the inner face. This was usually indicated by excessive nitrogen usage. A few failures took place on the outer seal face as well. The Durametallic (now Flowserve) local and factory representatives helped troubleshoot most of the failures. On the inner seal faces, at least some combination of face loading and our service was causing unacceptably high wear. We were never quite sure, however, why we still had occasional outer face failures. After about nine months of effort, we decided to switch to a Durametallic GF-200 non-contacting double seal cartridge assembly. Once the lift-off rpm is reached, a thin gas film separates the seal faces. With no face-to-face contact, there is no wear. If this condition can be maintained, seal life could be several years. Or so we thought. Yet seal life in this pumping application did not substantially change (still only a few months), and most failures were on the outer seal. We suspect that the failure was due to a higher pressure differential across the outer seal. We expected longer seal life. Other users of the GF-200 seal were seeing seal lives of

a year or more in different services. Our experience was also below Durametallic’s expectations. Failures were indicated by a relatively rapid swing in the nitrogen supply flow to the seal. Initially, flow would either rise or fall, but it always ended up with a high flow and lower pressure than a new seal. During the troubleshooting of these failures, efforts began to focus on one of the O-rings in the assembly. This particular O-ring was required by design to shift slightly during a pressure reversal across a given set of seals. This shift was designed into the seal to help flatten the closing force on a given set of faces and allow the required face separation to be maintained. During disassembly and parts inspection, we had noticed the slight swell typically seen in this service. This is normally not a problem in most static seal applications. In this case, however, even a very small amount of swell interferes with proper positioning of the seal faces relative to each other and results in either hang up and an excessive gap or occasional touching of the faces. A small amount of face wear causes excessive gas consumption and renders the seal unusable. In trying to solve this problem, we evaluated several O-ring materials. All showed small amounts of swell. This proved to be a continuing problem for this application. The Pump Handbook Series

Not being able to find a suitable Oring material, we had to abandon efforts to get satisfactory seal performance. All was not lost, though, as our efforts to get the GF-200 to work in other services at our plant did prove successful. This seal system has since been copied to sister applications in other plants. The third seal tried in our pump test group was a John Crane noncontacting dry gas seal, which does not have the same O-ring constraint as the previous seals. Service life to date has approached 21 months with low and steady seal gas usage, so we have expanded use of this seal to other similar pumps. One issue still open here is determining the nitrogen supply pressure to the seal. Manufacturer literature typically recommends that this be set at about 30 psi over fluid pressure, usually taken as pump normal discharge pressure. In our internal reviews we identified other operating pressures, stemming from upsets, which can greatly increase the supply pressure requirement. The greater the supply pressure, the greater the usage of nitrogen gas. While we are still working through this question, we set the supply pressure to the seal to handle a “maximum” operating pressure, even though the design pressure for the mechanical system may be higher still for containment reasons.

Continuing Efforts Future reliability programs will focus on better sets of pump reliability measures and improved communication between operators, repair people and engineers. Part of this effort will be in continuous vibration monitoring. Information on pumps in critical service will be displayed in the operations control, and continuous monitoring will provide the opportunity for trending. A chart will show vibration levels over time. The ability to compare trends of vibration with other key process parameters, such as flow, temperature and pressure, will be a new and powerful troubleshooting tool. With the proper analysis, we can decide whether the appropriate path is to change operating procedures, the system around the pump or the pump itself. These efforts will provide reliability gains for the next several years. ■

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Classifying Chemicals to Ensure Effective Sealing Confused about how to seal different fluids? Take the first step—classify them! by William McNally, Consultant he most common question asked by people who sell seals is “What are you sealing?” This is usually followed by questions about shaft size, product, temperature, speed, stuffing box pressure and any other operating conditions that come to salespersons’ minds. The problem with this approach is that those in sales would have to have a very large data bank of information on hand to be able to reference a particular problem and then be able to make a sensible seal recommendation. There is a much more logical approach to the problem. An alternative approach to the sealing of various chemicals, mixtures and compounds can be divided into three parts: 1. Knowing how to select mechanical seal components that will not corrode or be attacked in any way by the fluid you are sealing, or any other chemicals that might come into contact with the seal as a result of cleaning the system, flushing the stuffing box, using barrier fluids between double seals and quenching behind the seal. 2. Understanding the total range of operating conditions of the equipment, and then selecting seal designs that can handle this range. 3. Devising a method of classifying chemicals that puts them into neat, logical categories. Each category can be sealed by the use of a special seal design and/or environmental controls. It is important to note that the sealing environment will affect the sealing fluid, often preventing the lapped sealing faces from staying in contact.

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This article will concentrate on the classification of chemicals and leave the selection of seal materials, types of seals and use of various environmental controls to another occasion. A fluid can be classified as either a liquid or a gas and can be placed into one of seven categories: 1. Fluids sensitive to changes in temperature and/or pressure 2. Fluids that require two mechanical seals 3. Non-lubricating liquids, gases and solids 4. Slurries, which are classified as solids suspended in liquid (Solids might or might not be abrasive.) 5. Liquids sensitive to agitation 6. Liquids that combine to form a solid 7. Lubricating liquids We will be investigating these categories in detail and learning how they each affect the life of a mechanical seal.

Fluids Sensitive to Changes in Temperature and/or Pressure Corrosive Liquids Most corrosives will double their corrosion rate with an 18°F (10°C) temperature rise. The temperature at the seal face is always higher than the temperature recorded in the stuffing box or seal chamber. Keep in mind that any contact between the rotating shaft and a stationary component will cause high heat, and this will be detected as localized corrosion. Wear rings and throttle bushings are subject to such rubbing. If the equipment is outfitted with a cooling jacket, but it isn’t The Pump Handbook Series

being used, the air inside can act as insulation, increasing the heat in the stuffing box considerably. Liquids that Vaporize Most any liquid will vaporize if it becomes hot enough, or if the stuffing box pressure gets too low. Products with a low specific gravity give us the most trouble. If the product vaporizes between the lapped seal faces, it will separate the faces as the gases expand. When hot water vaporizes, it leaves behind any chemicals that were dissolved in it. Most of these chemicals are left behind in a hard crystal form that will damage the lapped faces. Liquids that Freeze Fluids such as benzene, and others with a low specific gravity, will freeze as they vaporize. If any oil or lubricant was placed on the seal face, it will freeze and possibly damage the lapped faces. Moisture on the outboard side of the seal will also freeze, and this can restrict movement of the sliding or flexing seal components. Liquids that Solidify Some fluids solidify with an increase in temperature, others with a decrease. Solvents vaporize with low pressure, leaving behind any solids they were carrying. Paint is a good example of a product whose solvent will vaporize at or below atmospheric pressure. In most cases you can reference a vapor pressure chart to learn when the solvent or carrier will vaporize. Viscous Products Viscosity usually decreases with an increase in temperature and

increases with a decrease in temperature. Oil is a good example. High viscosity can interfere with free seal movement and cause seal face contact problems. Lowering the viscosity can often increase the seal face wear because the film is not thick enough to keep the surfaces separated. You need a film at least one micron thick to keep lapped seal faces separated. Film-Building Liquids Petroleum products will form a varnish when first heated and then gradually form a layer of coke as the temperature is elevated. These transformations are not reversible, and the resultant hard film restricts the sliding and/or flexing motion of the seal components. Hard water is another example of a film-building fluid. Liquids that Crystallize Sugar and salt solutions are two examples of these fluids. If crystals form between the seal faces, they can destroy the carbon. If they form in the sliding or flexing components, they will open the seal faces as the shaft moves. Any leakage across the seal faces will form a solids build-up on the other side of the seal, causing interference as the seal tries to move when it compensates for wear. Fluids that Produce Magnetite Hot water systems pick up black or reddish magnetite (ferric oxide) from the insides of pipes. Magnetite is an abrasive material that will be attracted by a magnet. It will collect on the seal components and destroy the dynamic O-ring as well as restrict seal movement. This causes the lapped faces to open. Magnetite is a severe problem in new hot water systems. The problem will diminish, however, as the system ages and the protective film stabilizes. The names of these chemicals are not important. If you know how to seal any one of them, you can seal them all. It is just a matter of fitting the particular chemical into the right category and learning how to seal that category. Common sense dictates that the product temperature and/or pressure must be controlled in the seal area to prevent any of the above unwanted actions from occurring. In most cases you should try to avoid using two hard faces in these applications because of the additional heat that will be generated between them. Needless

to say, only hydraulically balanced seals are acceptable for use with any temperature/pressure sensitive fluid.

Fluids that Require Two Mechanical Seals When two seals are installed, they have a circulating barrier fluid that can be a “forced circulation,” or in many cases a convection system with a “pumping ring.” The pressure of the barrier or buffer fluid can be regulated to indicate a failure in either of the mechanical seals, allowing time for a pump shutdown, isolation and, hopefully, no subsequent loss in the pumped fluid. The following categories of liquids usually require two seals: Costly Products Sometimes the product costs so much that you just cannot afford to have a leak. There are plenty of charts that show how much leakage you get from various sized drips or steady streams. The smallest steady stream you can produce will be between 25 and 30 U.S. gallons per day (95 to 115 liters/day). Dangerous Products These fluids are given a special category because even a small amount of leakage is unacceptable. The danger could be from any number of different properties: radioactivity, flammability, explosivity or the presence of bacteria or other biological hazards. The new “right to know” law is having a major effect on how mechanical seals used with these type of products will be repaired. Pollutants Usually there is a “penalty” involved for allowing these fluids to leak, and the bad publicity does no one any good. In this day and age, a responsible company will not let pollutants leak to the atmosphere or to the earth for any reason. Fugitive emission legislation and the values underlying it have increased the need for mechanical seals that prevent product leakage. Products for which Unexpected Seal Failure Would Be Inconvenient Downtime can be very costly in many plants. Using two seals can prevent the unexpected shutdown caused by seal failure. This is espeThe Pump Handbook Series

cially important with batch operations or when no backup pump is installed. On the atomic submarine Nautilus, the backup shaft seal enabled us to get to the surface if a main shaft seal failed while we were submerged.

Non-lubricating Liquids, Gases and Solids Non-lubricating Fluids Solvents and hot water fall into this category. We experience more rapid face wear with these types of fluids. In most cases their film thickness is less than one micron and cannot support a load between two sliding surfaces. Dry Gases Unlike non-lubricating liquids, these fluids will not conduct heat very well and often are dangerous. This is a common problem if you forget to vent the stuffing box of a vertical pump. A top entering mixer is another example of this type of application. Dry Solids These materials can clog the sliding components of a seal, and they provide no lubrication for seal faces. Once the faces are open, the solids penetrate the space between them and usually destroy the lapped surfaces. Pharmaceuticals, freeze dried coffee and cake mix are examples of materials in this category. You can think of many more.

Slurries (Especially Abrasive Slurries) A slurry is defined as a liquid carrying solids that cannot be dissolved by normal control of temperature or pressure. The number of solids or their size is not important. These mixtures can clog the seal components and destroy faces like the dry solids mentioned above. The solids will collect on or in the sliding or flexing components of the seal, causing the faces to open. They then penetrate between the lapped faces and cause leakage and damage. In some designs the springs or bellows (metallic or elastomer) will experience severe wear in a short time. In these designs it is important to rotate the fluid rather than have the bellows component rotate within the abrasive slurry. The list of these products is almost endless.

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Liquids Sensitive to Agitation Dilatants The viscosity of these fluids increases with agitation. This is how cream becomes butter. Some clay slurries have the same properties. The resulting high viscosities will restrict the free movement of a seal. When dealing with dilatants it is important that you do not continually rotate the fluid in the stuffing box area. Thixotropic Fluids These fluids’ viscosity decreases with agitation. They seldom present a problem for mechanical seals, except for an increase in seal face wear. Plastic Fluids These can change their viscosity suddenly. Catsup is a good example. Newtonian Fluids These do not change viscosity with agitation. They present no problem for mechanical seals.

Liquids that Combine to Form a Solid Two examples of these liquids are epoxy and Styrofoam. Epoxy is a combination of a resin and a hardener. Styrofoam is formed by combining several liquids. We seldom have problems with these liquids in pumps because the blending takes place outside of the pump. Problems therefore can occur in mixer applications. You will note that I have not included anaerobic fluids (which solidify in the absence of air) in any of the categories. (“Super glue” is the product that first comes to mind in this regard.)

Lubricating Liquids This is the ideal application for a mechanical seal, but we seldom see it. More often than not we are sealing raw product that falls into one or more of the above categories. Back in the days when packing was used more frequently in pumps, we did not pay too much attention to these categories because we were either prepared to let the product leak on the ground or we would flush in clean liquid and concentrate on sealing the clean flush instead.

A Classification Example Now that leakage is no longer tolerable and product dilution is no longer desirable, you must know

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these categories to approach the job of effective sealing. In most cases the fluid you are sealing will fall into several of the above mentioned categories. Heat transfer oil is an example. It can be placed into all the following categories: • Hot - Normally pumped at 600700°F (315 -370°C), the fluid is too hot for available elastomers. • Film-Building - The product “cokes” at these high temperatures. • Dangerous - You cannot have this high temperature oil leaking out. It is not only a fire hazard, but a danger to people as well. Recent information indicates that some of these oils are also carcinogenic. • Costly - Most transfer oils cost between $12 and $20 per gallon (3.8 L) • Slurry - Because of the coking, solids are always present. To seal heat transfer oil successfully you would have to address all of these problems at the same time. As with all slurry applications, you would also have to recognize the problems with vibration (impeller imbalance), thermal growth and frequent impeller adjustments.

Don’t Forget Operating Conditions In addition to handling various chemicals, we are often faced with extreme or severe operating conditions. These include: Hot Products These are products too hot for one of the seal components, or hot enough that the fluid may change form. Heat transfer oil, for example, will “coke” at elevated temperature. Cryogenic Fluids These present a problem for elastomers and some carbon faces. Liquid nitrogen or oxygen are two examples. High Pressure Defined as stuffing box (not discharge) pressure in excess of 400 psi (28 BAR). Pipeline and boiler circulating pumps can have stuffing box pressures of this magnitude. Hard Vacuum Defined as 10-2 Torr or below. This number is well below most condenser or evaporator applications, but it does come up every once in a while in some industrial applications. The Pump Handbook Series

High Speed This is a situation in which the seal faces are moving at greater than 5000 feet per minute or 25 meters per second. Most process pumps do not approach this speed. The Sundstrand “Sundyne” pump is typical of a high speed application. Excessive Motion Defined as more than 0.005” (0.15 mm) in a radial or axial direction. Mixers, agitators and other specialized equipment have shaft movements up to 1/8” (3 mm). Long shaft vertical pumps and pumps equipped with sleeve or babbitt bearings are another application that can produce excessive motion. Excessive Vibration Unfortunately, there are no reliable numbers for the vibration limits of mechanical seals. Most vibration studies have addressed the bearings. It is important to remember, though, that excessive vibration can: • Open the lapped seal faces • Chip the outside diameter of the carbon face • Break the metal bellows used in some seal designs • Wear the driving mechanism used to transmit torque from the set screws to the seal faces • Loosen drive screws • Shorten bearing life • Damage (fret) expensive sleeves and shafts Some, but not all, designs have built-in vibration dampers to relieve some of these problems.

Conclusion Take the common sense route. Classifying chemicals before you try to seal them can save time and money. And it can help eliminate operational and maintenance problems down the road. Now the mix and match game isn’t one of chance, but of logic and strategy— and one you can win. ■ William McNally has more than 45 years of experience in the seal industry. He currently runs the McNally Institute for Pumps & Mechanical Seals and can be reached at [email protected] or through his web page at www.mcnallyinstitute.com.

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R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

Reliability: Big Opportunity or Big Distraction? Vibration analysis analyzed: It’s no silver bullet, but this predictive maintenance program may be the right choice for your company. By Cliff Hammock and John Schultz, Industrial Vibration Consultants, Inc. he ladder of maintenance strategies has many rungs that lead from reactive work, at the bottom, to total equipment asset management. As the ladder is traveled from bottom to top, other strategies, such as preventive, predictive, reliability-focused and proactive maintenance are encountered. Predictive maintenance programs are the cornerstone of many reliability initiatives. These programs often employ vibration detection and elimination as the lead technology, mainly because vibration signatures contain more mechanical condition information than any other single parameter. Because there is no silver bullet to reliability, this article briefly discusses several advanced maintenance strategies. Also offered are additional details related to both portable and on-line vibration detection and elimination programs, along with real-world case histories. Although they are very important and should be incorporated into a reliability initiative, this article does not discuss other maintenance technologies such as infrared thermography, ultrasonic detection, motor circuit evaluation and oil analysis. Most reliability initiatives begin with the introduction of a ”new” technology into a maintenance organization. While it is true that advanced maintenance strategies do provide big opportunities for savings at industrial facilities, it is also true that big opportunities, without

T

big strategies, produce big distractions. Through product knowledge and belief in the technology, vendors can often lead customers to believe that the technology itself, whether it is a computerized maintenance management system or a vibration detection and elimination program, will have a direct, positive impact on equipment reliability. Time after time this has proven to be wrong. People must embrace technology to obtain results. Often times workers not on the implementation team will not embrace a new technology because they believe it is not needed, will not work or may lead to job cuts. The truth is that change is very difficult for most organizations to handle unless a carefully structured plan has been developed. While developing a structured implementation plan, it is crucial to identify and resolve the cultural barriers to implementation. A simple method of identifying cultural misalignment is to ask members at every level of the organization an identical set of questions. These questions will include defining terms and rating the organization as to how well it is performing in specific areas. Where people differ in their perceptions will become obvious. After identifying these areas, awareness sessions can be held to help the entire organization realize the gaps that occur between where the organization currently operates and where it needs to operate to be The Pump Handbook Series

considered world class. Ensuring cultural alignment is as critical as any other issue to ensuring the successful implementation of a new technology.

No Silver Bullet Successful implementation of a new technology, however, is only the start of a reliability initiative. Manufacturing facilities are finding that there is no simple answer to the reliability question. It often takes major cultural changes, along with fulfillment of several technologies, to improve the reliability of a facility’s asset base. In a recent statistical review of reliability at manufacturing facilities recognized for excellence, the following factors were found to be positively correlated (Ref. 1): • management support and plant culture • organization and communication • performance measurements • training • preventive, predictive and proactive maintenance • inventory and stores practices • overhaul practices, including contract rebuilds Because the synergistic effect of these factors in combination is significant, a structured implementation plan that includes each of them is crucial. Without it the chances of a reliability initiative taking hold in

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an organization and producing the desired results are limited. The only factor in the study that was negatively correlated to uptime and reliability was reactive maintenance. It was found that reactive maintenance typically costs twice as much as planned maintenance. There is, however, a place for reactive maintenance at industrial facilities. If the failure of a machine that is replaced, rather than rebuilt, has no adverse effect on production or personnel safety, this machine may be a good candidate for run-to-failure reactive maintenance.

Funding Reliability Initiatives Limited resources are available at the onset of a reliability initiative. Two methods that many plants use to fund reliability initiatives are the up front expenditure approach and the self-funding approach. With the first approach, capital is provided to ramp up reliability efforts through technology implementation. If the funds are available from the start, an aggressive time line is used to implement technology at a rapid pace. Care should be taken to ensure that technology is not implemented at a pace that the organization cannot effectively handle. After the initial expenditure, the organization can see significant reductions in maintenance spending in the first two to three years. The second approach takes a smaller initial expenditure and allows the program to be funded from savings generated by the program. Decisions are made at the onset of the program as to the types of savings that can be agreed upon. When such savings are achieved, the money is plowed back into the program to fund implementation of other technologies. Savings can result from reductions in several areas including reactive maintenance, spare parts inventory, interval or time-based preventive maintenance tasks and equipment failures, just to mention a few. Selffunding can be slower paced but is nonetheless an effective way to take initiative when all of the needed money is not available up front. Significant bottom line results can

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Asset Health Report Summary September 1, 1998 Technologies Used: # of Machines on Route: # of Machines Monitored this period: % of Machines Monitored this period:

Vibration, Infrared Thermography, Oil Analysis 300 280 93.3%

Code

Baseline 01/01/98

This Period 9/01/98

Target

Red Yellow Green

35% 40% 25%

7% (20) 23% (64) 70% (196)

0-1% 14-19% 80-86%

Figure 1. Sample of an asset health summary report

be achieved after the first three years.

Allocating Limited Resources Because limited resources are available, it is very important that the resources be applied when and where they can have the greatest impact. One quick and dirty method to ensure the optimum use of funds is to rank the assets from most to least critical and apply the funds accordingly. A critical equipment ranking database takes into account the impact that an asset has on safety, the environment, quality, production and maintenance. Weighted factors are applied, along with a probability of failure, to determine the importance of an asset. This approach is not intended to bypass more thorough programs such as Reliability Centered Maintenance, but it is intended to help get a program off to a quick start and pointed in the right direction. Over time, assets should be reranked to see how things have changed. The main factor that can cause an asset to become more or less critical is the probability of failure. As you apply resources and technology to your most critical assets, the probability of failure should decline. That is why it is important to review the rankings on a regular basis, such as annually.

Measuring and Tracking Asset Condition To improve the overall condition of the asset base, the condition must be monitored and tracked. This is

The Pump Handbook Series

where predictive maintenance, or condition monitoring technologies, shine. Vibration and oil analysis, infrared thermography, ultrasonic detection and other condition-monitoring and non-destructive testing technologies provide much needed mechanical and electrical condition information about assets. How can the condition of a facility’s asset base be tracked? One simple method is to use a “red, yellow, green” report that summarizes an asset’s condition based on whether or not it has tripped an alarm set by one or more of the predictive maintenance and non-destructive examination technologies. Red - an asset is code red when it triggers a five standard deviation alarm. This condition means immediate attention must be given to this asset. Yellow - an asset is code yellow when it triggers a three standard deviation alarm. This condition warrants additional attention, such as more frequent monitoring. Green - an asset is code green when it has triggered no statistically based alarms. This condition enables the machine to operate until the next scheduled monitoring interval without action. Typically, the alarms are statistically based on the prior operating history of the asset. If no history is available, then the decision can be made to locate and utilize industry accepted alarms. An example of a statistically-based alarm is a vibration alarm that uses the mean and standard deviation to determine the

action plans can be developed to ensure maximum asset life.

Vibration Programs

Photo 1. A technician collecting vibration data

asset condition. Assets that trigger a mean plus three standard deviations alarm are typically coded yellow. Additional factors, such as type of fault and failure mode, may warrant a machine that has triggered a three standard deviation alarm to be considered red. This nomenclature is familiar to workers at most industrial facilities. It is used in other areas of plant operation, one such area being safety. The condition of assets can be updated monthly and, through the use of a database, can be tracked over time. This method will also help point out bad actors at the facility. If an asset is continuously going from red (alarm exceeded) to green (condition corrected or asset repaired) to red (alarm exceeded again), then the underlying problem, or root cause, has not been identified and corrected. An example of a summary report can be seen in Figure 1. A complete listing of assets sorted by condition would be attached to the summary report. Root Cause Failure Analysis (RCFA) and Failure Modes and Effects Analysis (FMEA) are two very powerful techniques that are a part of all successful reliability initiatives. RCFA makes it possible for a facility to learn from history. Any time the opportunity arises to learn from history and we fail to do so, we are not making the most effective use of resources. RCFA is a tool that helps a facility learn from an unwelcome event, in hopes that action can be taken to keep it from reoccurring. When a thorough understanding of an asset’s functions and modes of failure is achieved, proactive maintenance

As mentioned, vibration detection and elimination is the cornerstone of many reliability initiatives. The remainder of this article discusses vibration programs in more detail. Vibration programs are often grouped into two distinct categories, portable and on-line. The difference between the two is the manner in which information is collected and analyzed. Portable vibration programs are labor intensive and require technicians and engineers to collect data at preset intervals with portable instruments. The data are then analyzed for fault detection and severity. Two problems associated with portable programs are that much time is spent collecting data on smooth running machines while some machines fail between data collection periods. Photo 1 shows a technician collecting vibration data on a pump with a portable data collector. Portable programs are further broken down into contract and inhouse. During the mid 1980s it seemed to make sense that large corporations would utilize in-house personnel to handle vibration programs, and smaller companies, with fewer resources, would hire out their vibration programs. A decade later it seems as though this trend is somewhat reversing. You might wonder why. The answer comes in two words: employee turnover. Many companies are finding it difficult to keep employees with vibration analysis skills. These are skills that have become more and more valuable over the past decade. The following scenario has been repeated over and over throughout the country in every major industry. The names have been changed to protect the innocent. ACME Industrial decides to jump on the vibration bandwagon. It spends $50,000 up front to purchase state-of-the-art vibration analysis hardware and software, along with the necessary accessories and computer equipment. Billy Millwright gets the nod to become the in-house

The Pump Handbook Series

Photo 2. MAARS portable on-line system set up on pump train

vibration analyst, and over a period of two to three years, ACME spends another $20,000 to $40,000 training him in the art and science of vibration analysis. The program gets off to a quick start, and Billy earns a reputation, and maybe even a certification, as a top-notch analyst. A vibration user’s group is formed in Billy’s area, and he starts to attend quarterly meetings to gain more knowledge about vibration analysis and other predictive maintenance skills. MIMOSA Manufacturing decides that it too must pursue the benefits of vibration analysis. The plant engineer hears about a local vibration user’s group meeting and decides to attend to learn more about the technology. At the meeting, Billy Millwright presents a case history on how he put his skills to work by identifying a significant bearing defect on a vital piece of machinery at ACME Industrial. A conservative estimate is that over $100,000 in both opportunity costs and collateral machine damage was saved as a result of the call made by Billy Millwright. Rather than start from scratch, MIMOSA Manufacturing offers Billy Millwright a position as lead vibration analyst at its facility. The position comes with a 20% increase in salary. Billy sees the offer as a challenging opportunity and an excellent chance to advance his career, so he takes the job with MIMOSA. At this point ACME has made a very significant investment and has very little to show for it other than some outdated vibration analysis

349

can handle employee turnover. However, they are not always economical. Contracted vibration analysts typically achieve a much higher monthly ”wrench time” than in-house employees. Several factors account for this including the absence of scheduled breaks, holidays, sick days, meetings, vacations, etc. Contracted analysts often have the incentive to work harder for highFigure 2. Typical layout of an on-line condition monier pay. With all this toring system said, it has been seen that, in many cases, the equipment. When Billy left ACME, best compromise is a vibration prohe took not only his skills with him, gram that combines in-house and but more importantly his knowlcontracted resources. edge of ACME’s equipment and As the price of contracted seroperation. These human assets canvices decreases, the justification for not be replaced by hiring another monitoring more equipment vibration analyst off the street. At a increases. The cost of a typical minimum, the vibration program at vibration analysis contract varies ACME, which was off to a promissignificantly, so let’s use a competiing start, suffers a significant settive price of $3 per bearing per back whether or not management month. This means that the annual admits it. For this very reason, cost of monitoring a typical four many large facilities are divesting bearing motor pump combination is themselves of non-core business $3 per bearing X 4 bearings X 12 activities such as vibration analysis months = $144. As a percentage of and contracting for these services. both opportunity costs and replaceThis is not to say that some comment cost of the equipment, this is panies have not overcome the very low. It is often thought that the employee turnover problem, at least more critical the equipment, the temporarily. Many vibration promore often it should be monitored. grams utilize in-house resources This is not necessarily the case. As effectively, but none is immune to stated earlier, equipment criticality employee turnover problems. I was should be used to determine which personally involved in an effective assets are put on a vibration analyvibration program at a major utility. sis route first when resources are The initial vibration engineer served limited. The failure mode of the more than seven years before movasset is used to determine the freing on to other opportunities. I perquency at which it is monitored. If sonally left after three years, and a machine has a failure mode that since my departure in 1994, two progresses rapidly, it may need to vibration engineers have come and be monitored every two weeks gone, and the third was recently instead of monthly. Unfortunately, hired and is being trained. in some cases the machine has to As with any service provided by fail before the failure mode is an organization, whether in-house understood. or contracted, there is always the On-line vibration systems, often cost-effectiveness consideration. Inreferred to as continuous monitorhouse programs can be effective at ing systems, utilize permanently accomplishing the goal of detecting mounted transducers to collect and eliminating vibration if they

350

The Pump Handbook Series

Figure 3. Trend plot of vibration throughout the pumping cycle

data. This eliminates the manpower requirements for data collection. Typically, the information is transmitted from each transducer through a remote multiplexor unit to a signal processor. It is then routed to a computer through a direct link or a network interface. The computer analyzes the data, based on preprogrammed setups, and presents the end user with a graphical representation of the machine’s condition. The user is alerted when the vibration level reaches a predetermined amplitude, and in-depth analysis can then be performed to determine the source and severity of the problem. A diagram of a typical on-line monitoring system is shown in Figure 2. Two drawbacks associated with on-line systems are the high initial cost and the upkeep and repair of cabling and sensors. A subset of the on-line system is the portable on-line system. These are typically multi-channel signal processing units that can be set up to collect data on a problem machine for a specified time period. These are more for troubleshooting problem equipment. An example of a portable on-line system in use follows. Background On Sept. 10, 1998 three pump trains were monitored with a portable, on-line continuous monitoring system, the Model 5000 from MAARS. Pump seals had been failing on this particular pump train for some time. The suspected problem was cavitation. The pumps ran for only about one hour per charging cycle, and then weren’t run for

ing the quietest part of the cycle.

Figure 4. Modulated vibration waveform caused by cavitation

about two hours. This is a Class 1, Division 2 hazardous environment, and downtime is very expensive for this chemical plant. In this area there were four pump trains, each consisting of two centrifugal pumps in series. The output of the first pump fed the input of the second pump. The material being pumped is proprietary, but it can be said that it is less dense than water. At the end of each cycle, the lines were flushed with water for approximately three to four minutes. The Data Analysis The Model 5000, seen in Photo 2, can monitor all channels simultaneously, and six channels were deployed for this test. Three channels were placed on each pump, in the usual horizontal, vertical and axial positions. Five sets of data were collected in a 10 hour period, with two pumps being monitored on two separate cycles. As can be seen in the trend plot of Figure 3, upon start-up the vibration levels would ramp up and then settle back down to a consistent level during the charging cycle. At the end of every test, the vibration levels would begin to oscillate significantly, also seen in Figure 3. This corresponds to the time when water is being introduced into the pipes for flushing the system. During the routine part of the pumping cycle, the ambient noise from the pumps and piping was very loud, but the vibration energy was relatively low. When the water was introduced, the vibration levels increased, yet the noise levels decreased. In short, the machine train had its greatest vibration dur-

Final Analysis Consultation with several process engineers and operators gave insight as to the relative densities of the two materials being pumped. Because water is much more dense, one theory proposed was that there was now a different load being placed on the pump, thereby causing it to cavitate. This cavitation may have been causing the material being pumped to surge, starving and flooding the first pump. This put stress on the bearings and seals and showed up as a modulated time waveform in the vibration data (Figure 4). The data collected with this portable on-line system are being used to resolve this problem. The question is often asked ”When should an on-line vibration monitoring system be installed?” There are several conditions that warrant the installation of an online system. A critical machine is a good candidate for on-line monitoring if: • It fails between portable monitoring intervals. • It is inaccessible or difficult to monitor. • Process problems can cause the

machine to fail. • It gives very little warning before failure. With the price of on-line technology dropping, it makes sense to install systems on more and more machinery. A few of the candidates for on-line systems include pumps, of course, and: • paper machines - especially the press section • gas and steam turbines • compressors • extruders On-line condition monitoring systems are becoming more integrated with other systems and can input various signals such as: • • • •

vibration and speed pressure and flow current and voltage temperature and load

On-line systems can also output signals for control panel alarms, as well as interface with a CMMS to generate work orders automatically when a degrading condition develops. This makes it possible for continuous integration systems to be developed that bring information

Figure 5. Trend of overall vibration on sawmill chipper related to knife replacement The Pump Handbook Series

351

from the plant floor directly to the plant manager’s office. Typically, portable vibration routes have set frequencies for monitoring, such as monthly, or maybe even bi-weekly on critical equipment. For some equipment, even biweekly data collection is not enough to tell the whole story. Consider for example a sawmill chipper that was being monitored by a Monitoring Technology Corporation MS-2000 on-line system. Chippers are critical to the continuous operation of a sawmill. Unless waste can be processed through a chipper, the mill cannot continue to run normally. A new chipper was installed, and baseline vibration data indicated that it was a relatively smooth running machine. On November 5th, the vibration more than doubled at shaft speed, indicating unbalance. The machine was inspected for broken parts or buildup on the rotor. No obvious problems were identified. Two weeks later the vibration dropped to the baseline level for no apparent reason. Exactly two weeks later the vibration at shaft speed increased again. It was apparent that a trend was developing, as can be seen in Figure 5. An investigation determined that on night shift, new chipper knives were being installed every two weeks. The problem was that only half of the knives were replaced, leaving some worn knives with the newer knives, resulting in significant rotor unbalance. A method of distributing the new and old knives to minimize unbalance was then used to resolve the problem. A portable data collection and analysis program may have never identified the fact that this machine was running at high levels of vibration 50% of the time. If information was collected on a monthly schedule, the increases could have very well gone unnoticed. This could have led to premature failure of a very important piece of machinery. The Last Frontier Over the past several decades manufacturing processes have been

352

significantly optimized for production. Today, capital projects for process optimization are yielding less and less of a return on investment (ROI). Reliability initiatives have proven to yield a significant ROI, ranging from 5 to 1 to as high as 30 to 1. It is crucial that facilities that intend to compete in a global economy take serious the business of reliability, which may be the last frontier. ■

References 1. Moore, Ron. “Reliability and World Class Manufacturing,” Reliability Magazine, December 1997.

John Schultz serves as the Vice President of Operations for the Allied Services Group, Inc., which specializes in the development and implementation of reliability initiatives at industrial facilities. He holds a B.S.M.E. degree, with a minor in Economics, from Rose Hulman Institute of Technology. John also holds certificates in Vibration Analysis, Infrared Thermography, Tribology, and Root Cause Failure Analysis, and is a member of ASME, ASQC, ASNT, STLE, SMRP, and the Vibration Institute. Mr. Schultz can be reached by phone (513) 932-4678, or fax (513) 932-4980.

Cliff Hammock is the Regional Manager in Georgia for Industrial Vibration Consultants, which is headquartered in Lebanon, Ohio. He is a graduate engineer from Mercer University and a Certified Vibration Specialist II. Mr. Hammock can be contacted by phone (912) 474-0463, of fax (912) 474-0612, or e-mail [email protected]

Condition Monitoring Companies* Name Bently Nevada Corporation

Address 1617 Water St. Minden, NV 89423

Phone (800) 227-5514

Fax (702) 782-9337

Website (www.) Bently.com

Computational Systems Inc.

835 Innovation Dr. Knoxville, TN 37932

(423) 675-2110

(423) 675-3100

Compsys.com

Entek IRD International Corp.

1700 Edison Dr. Milford, OH 45150

(513) 576-6151

N/A

Entekird.com

Idax Incorporated

5301 Robin Hood Rd., Ste. 134 Norfolk, VA 23513

(757) 857-4967

(757) 857-1982

Idax.com

Machinery Analysis & Reliability Systems

5201 Kingston Pike, Ste. 6-142 Knoxville, TN 37919

(423) 927-6626

(423) 927-6627

Maars.com

MachineXpert, LLC

507 South Gay St., Ste. 1120

(423) 637-1760

(423) 521-6380

N/A

Monitoring Technology Corp.

2731 Prosperity Ave. Fairfax, VA 22031

(703) 698-5520

(703) 698-5539

N/A

Predict DLI

253 Winslow Way West Bainbridge Island, WA 98110

(800) 654-2844

(206) 842-7667

Predict-dli.com

Prüftechnick Inc.

1848 Memory Circle Gardnerville, NV 89410

(702) 265-6650

(702) 265-6670

Pruftechnik.com

SKF Condition Monitoring

4141 Ruffin Rd. San Diego, CA 92123

(619) 496-3400

(619) 496-3531

Skfcm.com

* This listing may not include all companies involved in the condition monitoring business. No companies were intentionally omitted.

Table 1. Condition monitoring companies The Pump Handbook Series

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R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

Controlling the Seal Environment—A Key to Seal Reliability By Keith Schindler, Ultramar Diamond Shamrock

he mechanical seal was developed to overcome some of the disadvantages of compression packing, such as the amount of leakage required to lubricate the packing. Today, the mechanical seal plays a key role in keeping leakage levels within the environmental standards of local, state and federal regulatory agencies. Seal configurations and material combinations are available that can handle many different environments; however, using a mechanical seal requires a higher level of maintenance personnel training, a larger initial monetary investment and often a custom seal design (Ref. 1). The mechanical seal can provide many years of reliable service if it operates within its design specifications. Controlling these parameters and the environment is as important for proper operation and long service as choosing the proper design and materials. Several factors in the seal environment affect reliability. These can be separated into two categories. The first encompasses process parameters, which include operating temperature, operating pressure, vis-

T

BACK VANES Psc = Ps + 0.25( Pd - Ps ) Seal cavity pressure = suction pressure + 25% of the Differential BALANCE HOLES Psc = Ps + 0.10( Pd - Ps ) Seal cavity pressure = suction pressure + 10% of the Differential DOUBLE SUCTION Psc = Ps Seal cavity pressure = suction pressure Ps = suction pressure to the pump Pd = discharge pressure from the pump Psc = seal cavity pressure

Figure 1. Impeller design and seal cavity pressure

cosity, corrosiveness and vapor pressure. These factors determine the design and construction requirements for the seal. The second category of seal environment considerations is purely mechanical. It includes throat bushing clearance, wear ring clearances, impeller design (balance holes or back vanes), radial clearance around the seal, vibration and runout at the seal area. Proper equipment specifications and repair procedures take care of the majority of the mechanical parameters. However, one often overlooked variable has a big influence on seal life. It is the seal flush. A proper seal flush supplies cooling to control the temperature; it adds lubrication to the sealing faces; it prevents the product from boiling in the seal cavity, and it removes abrasive products from the seal environment. For the flush to be effective, it must be sized properly. The first step in establishing an effective flush is determining the seal cavity pressure. There are several formulas to calculate the seal cavity pressure, depending on equipment design. A few of these formulas for different impeller configurations are shown in Figure 1. These equations are used to calculate the seal cavity pressure at the best efficiency point (BEP) without introducing the seal flush. However, not all equipment operates at the BEP, and not all equipment operates without the seal flush. These formulas will give inaccurate results for equipment operating away from the BEP or using a seal flush. To demonstrate this point, a survey was conducted to determine actual seal cavity pressures. A list of The Pump Handbook Series

operating conditions and equipment used in the study is shown in Table 1. The table includes the calculated seal cavity pressure at BEP, the actual seal cavity pressure and the operating point on the pump curve represented in % of BEP. The seal cavity pressure was measured with a calibrated pressure gauge attached to a port on the flush gland. The readings were taken with the flush on and also momentarily isolated. The results indicate that the calculated seal pressure can differ from the actual pressure by as much as 45%. There are a number of reasons why the formulas do not indicate the actual pressure. Without knowing the equipment design and the clearances, one can only guess at the seal cavity pressure. There are two ways to determine the actual seal cavity pressure: measure it in the field with a pressure gauge attached to a port on the head or the flush glands; or request the seal cavity pressure on the pump specification sheet at the time of purchase. Most pump manufacturers have tested their equipment to determine the seal cavity pressure and can readily provide this information for any given set of operating conditions (Ref. 1). In some instances one really needs to know the seal cavity pressure. In any flush arrangement that incorporates a bypass flush (API Plan 11 flush arrangement, Figure 2), an external flush (API Plan 32 flush arrangement, Figure 3), or a pressurized dual seal arrangement (API Plan 54 flush arrangement, Figure 4), the seal cavity pressure must be known to determine the

353

Equipment Make/Model/Size

Actual Actual Suction Discharge Pressure (psig) Pressure (psig)

Calculated Seal Actual Seal Cavity % of BEP Cavity Pressure Pressure with/without (psig) flush (psig)

Pacific 4x10 WYC wear rings - no balance holes (with a plan 11 flush arrang.)

80

160

80

160/145

95 %

IDP 2x8.5 SVCN-7 wear rings - balance holes (with a plan 11 flush arrang.)

35

110

43

65/-

65%

2x3x6.5 Peerless Pump pumpout vanes balance holes (with a plan 32 flush arrang.)

48

135

70

105/48

30%

Pacific 1.5 RVC-2 wear rings - no balance holes (with a plan 11 flush arrang.)

30

255

30

135/30

98%

a

The Pacific vertical suspended double casing (or can) pump design incorporated a spillback orifice to the suction nozzle after the pressure reducing bearing. The drawings showed the orifice. The pump was shipped with the orifice plugged. a

flush pressure or the buffer pressure. There are also a number of graphs and computer programs available to help calculate the flush rate through the throat bushing (Ref. 2). Every pump configuration involving a bypass flush creates the possibility of the product boiling in the seal cavity because of heat generated at the seal faces. One way to prevent this is to raise the pressure so that it is higher than the vapor pressure of the product, taking into account the increase in temperature due to seal-generated heat. The way to raise the pressure in the seal cavity is to supply enough flush pressure to the seal cavity and let the throat bushing control the flush rate. Maintaining at least 25 psig more than the vapor pressure of the product will prevent flashing (Ref. 1). If the calculated flush rate through the throat bushing is higher than what is required by the seal vendor, the throat bushing can be modified to reduce the amount of flush needed.

A bushing can be machined out of Graphalloy® (a graphite-metal alloy) and pressed into the seal housing throat. The clearance can be adjusted to relatively close tolerances (Ref. 3). Since the bushing is made with graphite, it will not gall with rotating metal elements if there is shaft deflection or whip. A floating bushing arrangement can also be used to reduce the radial clearance. The bushing has a smaller outer diameter than the bore, and this makes radial movement of the bushing possible in the event of shaft deflection. The bushing is held in place by a spring and collar arrangement. Changing the bushing clearance will also help increase the pressure in the seal cavity. Several other throat bushing configurations are available from pump and seal manufacturers. In a vertical pump application there is always a problem with vapors trapped in the seal cavity. These cause the seal to run dry Close Clearance Throat Restriction

momentarily. To eliminate this problem a recirculation line (or an API Plan 13 flush arrangement, Figure 5) can be incorporated with the bypass flush plan. This combination (now called a Plan 14 by API) accomplishes two things. First, it automatically vents vapors out of the seal cavity during start-up. Second, it promotes circulation around the seal for cooling. The flush rate can be controlled by a restrictive orifice in the re-circulation line. The combination of the flush rate through the throat bushing and the re-circulation line can be balanced to achieve the rate recommended by the seal manufacturer. The combination should also maintain the minimum pressure needed in the seal cavity to prevent the product from flashing. If some forethought is put into the design of this arrangement, the flush ports can be oriented on opposite ends of the gland for easy installation. The re-circulation flush port can be positioned slightly above the bypass flush port or in the highest spot possible in the gland. So positioning the recirculaExternal Flushing Source Drain or Return

Pressure Gauge Flow Meter

API Plan 54. ANSI Plan 7354

Figure 4. Pressurized dual seal external source flush arrangement (Courtesy FlowServe Fluid Sealing Division)

Flow Pressure Meter Gauge

Bypass Flush From Pump Discharge External Flushing Source

API Plan 11. ANSI Plan 7311

Figure 2. Bypass flush arrangement (Courtesy FlowServe Fluid Sealing Division)

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API Plan 32. ANSI Plan 7332

Figure 3. External source flush arrangement (Courtesy FlowServe Fluid Sealing Division) The Pump Handbook Series

API Plan 13. ANSI Plan 7313

Figure 5. Recirculation flush arrangement (Courtesy FlowServe Fluid Sealing Division)

tion port will help vent any vapors that might accumulate in the seal cavity. Seal manufacturers can also drill the flush gland tangentially to increase circulation around the seal. The mechanical seal can provide many years of reliable service if it operates in the environment for which it was designed. The ability to control the environment is important for proper operation and long service. In many instances unnecessary seal failures can be avoided by properly sizing the seal flush. Flush rates calculated from the standard formulas can be off by as much as

45%. Proper flush rates can be determined by measuring the seal cavity pressure in the field or requesting the manufacturer-specified cavity pressure. ■

References 1.FlowServe Fluid Sealing Division, Dura Seal Manual, Ninth Edition 2.Dura Seal Manual, Ninth Edition, EG&G Sealol In-Tandem Engineering Calculation Software, John Crane API 610 Orifice Sizing Chart 3.Graphalloy - Self Lubricating Graphite Metal Alloy, a product of

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Graphite Metallizing Corporation Keith Schindler is a Rotating Equipment Supervisor at Ultramar Diamond Shamrock, Inc. in Commerce City, Colorado. His responsibilities include supervising plant personnel and the repair of all rotating equipment. In his previous position, Keith was a Reliability Engineer at Total Petroleum, Inc.’s Alma, Michigan refinery, and he was a Reliability Engineer with AMOCO oil company in Texas City, Texas. He is a graduate of University of Houston with a degree in Mechanical Engineering and is a registered professional engineer in Michigan

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R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

Winning MTBR Strategies Experts from different industries share their best reliability advice.

Adhering to Specifications and Standards Key for MTBR Improvements at Eastman Chemical By Steve Hrivnak When it comes to pumps, our best practice for MTBR has been paying attention to installation and operation details. A pump is part of a system. That system contains a foundation, a baseplate, grout, piping, a pump and motor, alignment and controls. Ignoring any of the details of these items usually causes short pump life, which manifests itself in short seal life. Properly installed and operated ANSI pumps at my facility experience 42 months MTBR. Our formula for success is: 1. Fabricate or purchase a well-designed pump baseplate. Good designs will install with ease and a machined baseplate with alignment jack screws will make alignments easy! A good design will conform to the following specifications: • The baseplate shall be fitted with one 4” grout fill hole uniformly distributed for every 10 square feet of baseplate surface and/or per subdivided section or raised welded cavity. • Vent holes 1⁄2” in diameter shall be provided for each bulkhead compartment at all corners, high points and perimeter edges. Perimeter vent holes in the baseplate shall be on 18” centers maximum spacing. Vent holes are required on any angle or “C” channel added for stiffness. • Vertical jacking screws 1⁄2” diameter minimum shall be provided around the baseplate perimeter 3” from each anchor bolt location to facilitate vertical alignment of the baseplate. • Machined mounting surfaces for the equipment and driver shall have horizontal jack screw positioning bolts 3⁄8” diameter minimum. • All welding on the baseplate shall be completed before machining the equipment and driver mounting surfaces. • Machined mounting surfaces shall be coplanar to .002”.

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What everyone needs to know about MTBR,” according to consultant Bob Matthews, “is that it is different for each plant, mill and factory.” Pumps and Systems agrees. However, we also believe that techniques used to improve reliability in one industry can have a place in others. In this feature, six soldiers in the battle against low MTBR share their tips with you. Pumps and Systems asked Randall Jones, Reliability Supervisor for AlliedSignal, Bob Amon, Machinery Superintendent for Tosco Refining and Bob Matthews, Rotating Equipment Consultant for GCITexas to describe some of their proven strategies in eight different reliability areas: training and maintenance techniques, tracking and record keeping, pump hydraulics, seals and bearings, installation procedures, couplings and alignment, preventive/predictive maintenance programs and start-up/shutdown procedures. Rounding out this month’s special focus on MTBR are the comments of Steve Hrivnak, Senior Mechanical Engineer at Eastman Chemical Company’s Eastman, TN facility. Steve describes the standards and specifications his organization has followed that have yielded a 42month MTBR for ANSI pumps. And Jody Mefford, Reliability Engineer at Hercules, and Alan Harmon, Reliability Engineer at AlliedSignal, share some of their ideas for first steps and general “big picture” guidelines for increasing MTBR.

Pump Training/Inspection/Maintenance Techniques Jones: We implemented a precision repair program that pays close attention to bearing fits, mechanical seal settings, alignment, etc. Amon: Machinists are provided with written inspection and repair forms to fill out. We have written repair procedures for most equipment and keep them in a readily accessible area by the machine shop. We also formed a dedicated compressor machinist group to handle our 35 major reciprocating pumps, 10 centrifugals and 9 engines. Matthews: Our company has developed pump training programs that fit specific plant needs, and we have found that Competency-Based Training gets the most return. We’ve used disassembly reports to gather information for failure analysis, and we have used assembly reports, procedures and checkoff sheets to support rebuilds. Personnel can see a unit’s MTBR when they

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• Machined mounting surfaces shall extend 2” beyond pump and driver feet on all sides with a 125 micro inch Ra finish. When machining a bent plate as a mounting surface, maximum material removal is not to exceed 1⁄16”. • Provide at least 1⁄8” shim adjustment under driver feet for alignment. • If a spacer coupling is to be used to couple the equipment to its driver, add an additional 1⁄4” axial clearance to the spacer length in the baseplate design. • Anchor bolt holes shall be 1⁄4” larger on the diameter than the anchor bolts. • Tapped equipment bolt holes in the baseplate should be deep enough to enable plenty of thread engagement. On 1⁄2” thick or smaller baseplates, weld 1⁄2” or 3⁄4” thick square plates under or on top of the baseplate to increase the thickness before drilling, tapping and machining. Tapped depth shall be one-and-one-half times the fastener diameter. • All bulkhead cross-bracing on the underside of the baseplate shall have a 2” x 6” opening to enable grout flow from bulkhead to bulkhead. • All corners of baseplate flanges shall have a minimum radius of 1”. All surfaces that will be in contact with the grout shall be rounded to eliminate stress risers. • Angle, channel or stud anchors shall be welded to the bottom surface to act as a shear key in the grout. • Gussets shall be welded inside the driver superstructure to increase stiffness and vibration damping. Most of the above baseplate design points can be found in Process Industry Practice (PIP) RESP002. 2. Grout your baseplates. Baseplates need to be grouted to minimize the vibration that decreases seal life. Regarding the foundation, certain fundamental rules are being overlooked when it comes to freestanding pumps. The foundation mass ratio—the weight of the foundation in comparison to the weight of the machine, driver, baseplate and the liquid or material in the machine—must be of a certain value in order to prevent vibration. For rotary equipment such as pumps, this ratio should be 3 to 1. For reciprocating equipment, it should be 5 to 1. A freestanding pump cannot meet this requirement. You need to start grouting your problem pumps. Note that if you grout a pump to a slab, the slab mass acts as part of the 3 to 1 ratio. Six inches of grout attached to the slab is adequate. For proper grouting guidelines, see the PIP Specification ESS REIE686, chapter 5. 3. Take pipe strain off the pump. Pumps are not

look at the pump shop’s previous workload.

Pump Database Tracking and Record Keeping Jones: Our Reliability Group developed an Access™ database that tracks failures and MTBFs once improvements have been made. We also created a one-page report for recording repair information, and seal vendors are required to issue a one-page failure report for each failed seal sent to them for repair. Amon: Our team installed specific failure codes appropriate to rotating equipment in the accountingbased CMMS. The failure codes are required and must be entered for the work order to be closed. MTBF reports are then easier to compile and are used to justify improvements/upgrades. Matthews: We use the pump history for each unit, and this helps. The downfall is that it adds work for a maintenance team with already overloaded work schedules. Most field technicians don’t have time to chase down a computer and record the information. I talk to some plants that use history records to track MTBR, and they find it can be successful. Again, most MTBR tracking is only for problem areas or isolated equipment.

Modifications/Upgrades to Pump Hydraulics Jones: We resized the minimum flow orifices on two pumps to eliminate flow-induced resonance, and we were able to do it without affecting the process requirements. We also increased pump speed from 1150 to 1800 rpm and removed an impeller on a multi-stage pump to increase the pump flow at the same TDH requirements. Amon: We try to adhere closely to the API 8th edition specifications when repairing earlier edition pumps. Part of this involves keeping a .1000-.1001” fit for sleeve bearing-to-housing and minimizing fits for ball bearings. Our numbers are based on a table that is dependent on shaft size. Matthews: Most pump hydraulic improvements occur during production downtime. We hear of constant engineered improvements on motor refineries, etc. But in small plants and mills with low profit margins, the upgrades are only for a small percentage of the pumps. MTBR is often the reason for most upgrades.

Modifications/Upgrades/Retrofits to Seals and Bearings Jones: Mechanical seals in high vapor pressure applications were converted to tandem seals with the barrier fluid chilled to 50°F. This helps remove heat from the seal faces. We also drilled a small pilot hole in the restrictor bushing in the stuffing box on pumps for these applications. This procedure helps vent the stuffing box during start-ups. Amon: We use a standard retrofit for bearing isolators (Inpro) and “pump pac” pre-loaded bearings on the bottoms of pump areas, or wherever high thrusting applications exist. Matthews: Increasingly, as the emission standards are controlled, older plants are switching to mechanical seals. Newer plants were designed with seals, and those that use or produce hazardous materials have always

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designed to be pipe hangers nor are they designed to pull piping in for final bolting. The following are some typical rules: • Discharge and suction lines shall be straight in the runs adjacent to the pump flanges for a minimum of five pipe diameters or 18”, whichever is greater. (This minimizes the possibility of cavitation and smoothes out the flow entering the impeller.) The pump suction line shall be constructed to provide a spool section for strainer removal. • Discharge valves shall be located directly after the required straight run. Suction valves shall be located directly before the straight run. The suction valve port area shall be equal to or greater than the suction line area. Suction piping shall be sized for 5 ft./sec. (bulk velocity) or less. Hot condensate service, and service for other liquids within 10°F of the boiling point at suction pressure, shall be sized for 2 ft./sec. or less. • Reducers in the horizontal suction lines shall be eccentric and located directly after the suction valve. If the liquid is within 20°F of boiling, place the eccentric reducer at the suction flange of the pump. • Lateral suction piping shall be supported and guided no closer than 1.5 feet (to a maximum of 3 feet) from the flanges. Vertical discharge piping shall be supported from above using spring hangers. Hangers shall be sized so that there are minimal vertical loads on the pump nozzle when it is cold. Piping shall be guided to help prevent transfer of piping moments to pump nozzles. All pump loads shall be below those determined by the pump manufacturer and verified by calculations for both hot and cold service. An exception is that the support design can differ from the above description if acceptable loading is verified by calculations and approved by the appropriate engineer. Special care should be exercised in support design for lines with fast-acting and shockinducing check valves. Installation • Install piping to within five pipe diameters of the pump. Grout the pump baseplate and align the motor using the reverse indicator method to within .002” TIR. Bolt the suction and discharge stub ends and flanges to the pump with gaskets, using four bolts. The spool pieces connected to the pump shall be field-fabricated. Field-fit and tackweld the piping to the stub ends, and then matchmark the flanges. Finally, remove the spool piece and final-weld the piping to the stub ends. The following installation guidelines shall apply and are provided for reference purposes. • If the flanges cannot be aligned by hand (not with hand tools) for the insertion of the bolts, rework the spool piece. • If a set of flange faces are not parallel to each

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used them. With the cost of special stuffing boxes high, we have over-bored the standard box .100” starting at the seal (cartridge) face fit and tapered the last one inch 5°. This in-house machining is very cost-effective. We also use bearing isolators on nearly all pumps, gear boxes, motors, etc. This is the #1 MTBR improvement in plants and mills I see today. On the other hand, I have found that some facilities are converting mechanically sealed pumps back to packing. This is for many different reasons. New fibers carry heat well. Some are lubricated, and others have designs that are proven to pull their weight in today’s industry. Again, not all the new technology may be for you and your plant.

Installation Procedures/Foundations Jones: When we listed it on the one-page repair report, pipe strain was reduced by relieving stress, or by properly supporting or modifying piping. Damaged pump foundations were scheduled to be repaired during outages. We now complete alignment checks on new equipment before the piping and foundation is accepted. Amon: Our standard requires that all grout is epoxy, and any exceptions must be reviewed and approved. We have found that the API 686 latest edition is a good guideline to use. Matthews: I have found that the new grouts and chock systems are excellent and some manufacturers offer books, tapes and other helpful tips for new users. Most vendors will assist with installations and/or training. MTBR is increased when grout and piping are improved. Well-done pipe installations are an art, and most mills have specifications that will work in almost all situations. However, some areas do require expansion joints at pump suction lines. (Discharge lines will also load a pump case or transmit vibration and can require expansion joints.)

Couplings and Alignments Jones: Our team implemented a precision alignment program that uses either reverse indicator, rim and face or laser alignment techniques. Amon: Our facility has standardized on the Thomas Series 71 disc pack, and we require that 90% of alignments be done by laser and kept within .001”. Matthews: My advice is to up the alignment specs. Most modern facilities use laser or reverse dial indicator methods, which are great. But remember that the goal is to get the alignment as close as possible, whichever method you use. I am familiar with the standards of .003” TIR for below 1100 rpm and .002” above that speed. In addition, I think balance really makes the difference. Get the best you can. Good is only good; better than good is best. In my opinion, this is the #2 industry reason for improved MTBR.

Preventive/Predictive Maintenance (Vibration Analysis, etc.) Jones: We use a spike energy system for critical magdrive pumps and routine vibration analysis on all other pumps. Amon: We have a formal PM program that includes

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other to within 1⁄32” across the length of the faces, rework the spool piece. • If the spool piece, which should be five pipe diameters or greater in length, is more than +/1 ⁄8” of true length, rework the spool piece. • The pump shaft deflection shall be checked using the face and rim method during flange bolt tightening. Coupling alignment must remain within .002” TIR. Miscellaneous • Install a strainer on all new pump installations to trap foreign material. The strainer should be removed and cleaned every eight hours until it no longer traps anything. It should then be permanently removed from the piping. A spool piece should be installed to facilitate easy removal of the strainer. • Pressure gauges or taps for pressure gauges shall be installed. (These can be used to troubleshoot pump problems.) See also PIP REIE686, chapter 5, for typical piping practices. 4. Do quality motor alignments. Alignments reduce failures! Maintaining proper alignment is impossible without correctly installed pipes and a solid foundation. Proper coupling alignment for pumps requires a 0.003” offset with a 0.0005” angular tolerance at 1750 rpm, and a 0.0015” offset with a 0.0003” angular tolerance at 3500 rpm. As a minimum, use reverse dial indicator, but preferably use a laser. 5. Keep flow above minimums and below maximums. Stop pump deadheading by using a minimum flow bypass or a power monitor. Vibration and shaft movement is lower at the pump’s Best Efficiency Point (BEP). Excluding the new low flow models, standard ANSI pumps have the following typical minimum flows: at 1750 rpm, minimum flow is about 20% BEP; at 3500 rpm, minimum flow is about 30 to 50% BEP. If you must operate below these levels, consider a re-circulation system. Maximum flows are 110 to 115% BEP. Size pumps to operate within these limitations to reduce vibration and shaft movement. 6. Rebuild pumps back to factory tolerances. Ask your pump manufacturer to supply the tolerance measurements you need to take to ensure that your shaft is not bent, your bearing fits are correct, shaft end play is not a problem, etc. Many pump problems are hidden in worn parts! Doing all six steps is what increased MTBR at our facility!

route-based inspections performed by area machinists. Our vibration program consists of four data collectors. Machinists and electricians collect data and download it. A vibration specialist then analyzes it. We also draw 70 lube oil analysis samples per month on our critical compressors. Matthews: I believe that routes for periodic inspections are a must and have proven very successful. Data collection and control room monitors are necessary. These, however, can be costly and sometimes difficult to justify. In some cases, they are used unnecessarily. With maintenance technicians doing morning walk-throughs in assigned areas to check and adjust equipment, there is ownership; this increases uptime in many operations. It also stretches the MTBR.

Pump Start-Up and Shutdown Procedures Jones: We installed power monitors on mag-drive pumps to prevent deadhead or dry run. Shut off on loading mag-drive pumps is tied to the load cells on the tanks (DCS controls). Amon: We always make sure to warm a pump up 100°F/hr to operating temperature before starting it. Matthews: Training maintenance technicians, operators, planners and supervisors with new installations is, I believe, critical. Most OEMs will assist with the first start-up and any initial training. Most plants have standard start-up procedures that work in a large number of applications. Again, I see more and more in-house training procedures developed everywhere I go. The plant operation must dictate specific start-up and shutdown procedures in dangerous environments.

Other Strategies Jones: We improved the steady bearings on vertical slurry pumps by replacing OEM carbon bearings with heavy bronze bearings that have a grease groove and lip seals on both sides. We also replaced the bowls of our rubber-lined slurry pumps with CDAMCV bowls constructed by a foundry. (These were not available from the OEM.) The new bowls increased pump life by 400%. Amon: We always make a concerted effort to achieve less than a 4 w/n balance specification. Our company holds vendors and outside shops accountable for proper repair standards. We also make sure to provide them with documentation to represent “our” stricter standards, because typical OEM standards are looser than ours. Matthews: With vertical pumps, the use of advanced polymer alloy bearings has a place in today’s search for longer run times. These bearings can take some dry start time, which is needed in many situations. They take impacts without damage to the shaft or bearing, and they have a high resistance to abrasion. Use the tools you can afford to help with reliability and choose them wisely. Don’t waste time and money on sacrificial pumps. "

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Reliability Engineers’ Advice: Make the First Steps the Right Steps Jody Mefford, Hercules I have recently taken a new job where my responsibility is to increase MTBR for the pumps on site. It has been an interesting venture. I have worked for 10 years in pump maintenance and reliability, and this has given me the background I need to help my new employer. One of the first things I did was write an installation specification. It includes sections on concrete, grout, baseplates, alignments, pipe strain and general layout of the pump and system. Most of the information I have included is just good engineering practice, but I have found that some things I thought were common are not. One example is ensuring that concrete is cured before the grout is installed. Even at up to 28 days, the concrete is still giving off moisture and heat that will cause problems with the bonding of the epoxy grout. Our specifications, therefore, include the following: • Let the concrete cure completely. • Fill the baseplate with grout, and don’t leave any voids. • Do a final alignment check before grouting to assure that the equipment will be ”alignable”. • Make sure the baseplate is heavy enough and strong enough to securely hold the equipment mounted on it. I have found that using jack bolts will make alignment much easier. • Don’t neglect the specification for flatness; it’s critical.

In addition to the installation specification, I also have an extensive rebuild procedure that includes, as an example, the following: • All tolerances must be checked before the pump is taken apart and after it is rebuilt. • All pumps are specified back to original manufacturer tolerances. • Once a rebuild has taken place, extensive record keeping is done. Documentation of all parts used and root cause evaluations is done. • Alignment is required before start-up. Another essential issue has been operator training. The manufacturers have been asked to provide numerous short sessions concerning operation of their equipment: How to start, how to stop and some performance keys. We have also asked mechanical seal vendors to provide training on what not to do with a seal, how seals run and why they work like they do. This has helped us eliminate some personnel issues and ward off problems up front.

The general layout of an installation can help or hurt, depending on how it is done. The following guidelines should help:

We have also benefited from a centralized pump repair facility that does all the rebuilds and maintains the documentation and tracking database. In addition, a reliability team performs regular vibration monitoring and infrared analysis of seals and bearings. The team is also consolidating equipment and standardizing on the best and most reliable brands. Mechanical seals, for example, are installed using the acceptable elastomers and seal face combinations for our worst applications. The use of Kalrez® may be overkill in some areas, but it does eliminate confusion at 2:00 AM.

• Check pipe stress and make sure the appropriate supports are installed. • Use at least 10 pipe diameters of straight pipe before the suction to help assure straight flow. • Add suction and discharge pressure gauges so that good analysis of pump performance can be made later. • Assure that the pump is sized for current conditions, not for “possible future expansions.” • Install positive face seals, bullseyes and expansion chambers instead of lip seals and oilers. • Position strainers and valves to avoid flow disturbances.

Alan Harmon, AlliedSignal Polymers We have used a relatively basic technique to improve our MTBR: Root Cause Failure Analysis. On critical and/or burdensome equipment, we identify the root causes of failures and eliminate them. Typical failure mechanisms have been improper alignment, lubricant contamination, incorrect bearings, incorrect lubricant and passage of electric current through bearings. We use vibration and oil analysis to identify bad conditions and support our root causes analyses. Some examples of how we have eliminated root causes are the precision alignment, replacing packing with mechanical seals, sampling oil and installing electrical isolators.

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R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

Reliability Starting with Procurement: Maintenance’s Role in the Purchase Cycle By Timothy S. Stefl, Henkel Corporation ith plant owners continuing to press for increased reliability, many maintenance departments must make aging facilities operate more dependably and at greater capacity than their original design accounted for. Procurement specifications are an important but often overlooked element in the success of reliability initiatives. This article discusses methods of incorporating maintenance requirements into purchasing specifications for both capital and expense projects. With engineering’s focus on cost and schedule and purchasing’s search for low bids that meet minimum operating requirements, people in these departments can tend to disregard future equipment operating and maintenance costs. Maintenance departments should help change the procurement focus to life cycle costing. Pump design has largely been standardized through the use of recognized industry standards such as API 610, Centrifugal Pumps for Refinery Service, ANSI B73.1, Horizontal End Suction Centrifugal Pumps for Chemical Process, and ANSI B73.2, Vertical In-Line Centrifugal Pumps for Chemical Process. ANSI specifications were developed for light and medium centrifugal pump duty with dimensional interchangeability. The API specification was developed for a high quality heavy-duty centrifugal pump. For the majority of process industries in the United States, the ANSI standard has been the ”work-

W

horse” pump specification, in large part due to the high cost of the API 610 standard. ANSI pump standards have required manufacturers to adhere to strict dimensional requirements, making it possible for the end-user to freely interchange pumps of varying manufacturers while incorporating some of the first reliability requirements into their design. External axial impeller adjustments, replaceable casing rings on enclosed impeller designs and shorter impeller overhangs are a few examples of reliability requirements directly attributable to the ANSI standards. However, industry studies indicate that Mean Time Between Failure (MTBF) for the typical ANSI B73.1 process pump is only 13 months and possibly as low as 6 months when installed spares are factored into the equation (Ref. 1). With increasing reliability targets and operating uptime pressures, procuring pumps using the ANSI B73.1 standard may no longer meet the objectives of your plant management team. An upgraded mediumduty pump specification designed to improve mechanical reliability at a modest cost is the UMD 55 Upgraded Medium Duty Pump, referred to as ANSI PLUS by many pump manufacturers. The ANSI PLUS manufacturer designation, however, does not require strict adherence to the entire UMD specification. MTBF for upgraded pumps, depending on the degree of the upgrade, can be as high as 42 The Pump Handbook Series

months (Ref. 2). Shaft runout, balance, bearing design and lubrication requirements are a few of the issues addressed in greater detail by the UMD upgrade. Since upgrading to a UMD type pump will have a negligible effect on the total cost of procurement, it should be used as a starting point for your procurement specification. But since dimensional limits imposed by ANSI B73.1 cannot be achieved using the UMD specification, uninformed buyers often fail to require or allow the procurement of a UMD pump. It is essential that the maintenance engineer responsible for plant equipment operation understand these, as well as any other standards, such as corporate engineering standards, that are involved in the procurement of a pump. The engineer should also actively participate in the procurement process to ensure that plant reliability targets can be achieved. This article uses the UMD specification as the baseline for the procurement process.

API vs. ANSI Although there are no definitive rules for choosing an API specified pump versus an ANSI specified pump, general guidelines do exist (Table 1). Since the majority of API ≥ 350 ft. Head ≥ 300°F Temperature ≥ 150 hp Driver hp Suction pressure ≥ 75 psig > 3600 rpm Speed

ANSI < 350 ft. < 300°F < 150 hp < 75 psig

Table 1. API versus ANSI selection guides

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pumps used in process industries fall within the ANSI guideline, this article focuses on the ANSI pump procurement process. When the pump process specification falls under the API process, the maintenance engineer’s role in the process and review becomes even more critical. An agreed upon API/ANSI division between engineering, maintenance and purchasing should be developed at your facility.

Procurement Cycle The procurement cycle begins when someone identifies the need for a new pump. This can be because of a capital expansion, process modifications or the need to replace a worn out or problem pump. Regardless of need, a process specification must be developed, typically by process engineering (Figure 1a, 1b). Once developed, route the specification for review. Maintenance or maintenance engineering needs to take an active part in this process. They review, among other things, process conditions such as NPSHA, flow, pressure, service and temperature to determine which type of pump best suits the application. Maintenance people need to put to use the lessons learned from the repair histories of pumps in similar applications. From a mechanical point of view, driver types, coupling designs, mechanical seal requirements, etc., should be reviewed to ensure that the specification meets the needs of the organization. This initial review will enable the proper selection of vendors and help you avoid past mistakes that can result in less-than-adequate equipment reliability. Following the review and pump type selection, Requests for Quotes (RFQs) can be prepared and distributed to qualified vendors. All pump standards and specification requirements need to be attached to the RFQ. Vendor selection should be based on the following: 1. Is the vendor in a position to provide extensive experience listings for equipment offered and willing to submit this information without hesitation? 2. Have the vendor’s centrifugal pumps enjoyed a reputation for sound design and infrequent maintenance requirements? 3. Are the marketing personnel thoroughly supported by engineering

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departments? Are both groups willing to provide technical data beyond that which is customarily submitted with routine proposals? Vendors who do not meet the needs of maintenance after the sale should be removed from the qualified vendor list. The CMMS documentation should be forwarded to purchasing when vendor performance becomes unacceptable. After all bids have been received and purchasing has assembled a bid tabulation package, complete a technical review of the bids. It is imperative that maintenance take an active role in this process. Review proposals to ensure that they meet the intent of your standards and specifications. You also need to attempt to make the proposals equivalent. Some clarification may be required and purchasing may need to contact the vendor for additional information. Maintenance and engineering should jointly review written exceptions to the standards and specifications to determine if they will be accepted or denied. All alternate vendor proposals need to be reviewed and evaluated as well. Make any credit/debit adjustments during the technical review. Horsepower consumption, annual maintenance costs, baseplate design and shaft stiffness ratios are examples of items that should be evaluated on a credit/debit basis. Give this information to the vendor with the RFQ documentation. The pump seller’s quotation should be in accordance with all standards provided. Any exceptions or deviations must be noted in a cover letter attached to the seller’s quotation. However, the seller can supplement its basic quotation with alternate quotations, provided that the following two conditions are met: 1. Significant process, price or delivery date advantage is indicated, a superior product is offered at a justifiable higher price or a product of equal quality is offered at a lower price. 2. The alternate quotation states all deviations from the standard and complete engineering details of alternate materials and/or designs are submitted for review. The result of a technical review should be a recommendation to purchase. If any proposal other than the The Pump Handbook Series

lowest is chosen, submit a letter justifying the higher cost proposal, along with the vendor recommendation, to purchasing.

Vendor Data Requirements Proposals The seller shall submit the following as part of the base proposal: 1. The price of delivery for the pump, motor, coupling and baseplate as specified in the standard and attached purchase order 2. Unit pricing for spare parts. The spare parts list shall include all parts required for five years of normal operation and critical start-up spares. 3. Bill of material 4. A parts list with factory cross-reference numbers 5. A completed pump data sheet 6. Preliminary outline drawing showing all principal dimensions, locations of suction and discharge connections, allowable nozzle loads and the direction of rotation when looking at the pump from the driver end 7. A cross-sectional drawing of the pump showing all details of construction. 8. A pump hydraulic performance curve that shows head, capacity, efficiency, rotational speed, NPSHR and brake horsepower. Performance curves for the proposed and maximum impeller sizes shall be included as a minimum. If the unit is steam turbine or variable speed driven, the seller shall submit characteristic curves for each speed required for the normal and maximum conditions specified. 9. 24-hour emergency number for off-hours procedures 10.The seller shall quote the following as separate adders: • Performance test • NPSH test Approval Drawings 1. The seller shall provide and mount on pump body a stainless steel tag identifying seal manufacturer’s name, seal size, type, material code and drawing number. 2. All arrangement drawings or catalog cut sheets shall reference equipment centerlines and elevations. Recommended clearance dimensions required for operating, installing and performing

maintenance (pull space) must be shown. 3. Outline and/or detail drawings shall, as a minimum, show the following information: • Overall dimensions and differential elevations, as required • Nozzle or inlet sizes, ratings, facings, locations and flow directions • Support mounting dimensions and locations • Weight of components and total assembled weight, operating weight and dynamic loading, if applicable • Size (space envelopes) and weight of each separate component and/or subassembly to be furnished by vendor • Purchase order number and the buyer’s identification number • Materials of construction • Leveling instructions, handling instructions and/or tolerances • Location of lifting lugs and/or rigging points 4. After receipt of order, and before any shop fabrications are started, seller shall furnish the following certified drawings: • Outline drawings of complete assembled units. These drawings shall include all principal dimensions, the direction of rotation, the furnished thickness of all flanges and the size and location of every connection to the equipment and base plate. If the rotating assembly of the upper half of the casing exceeds 1000 lbs., the weight of these parts shall be indicated on the drawing. • Mechanical seal cross-section with materials table • Single line seal flush diagram with a table identifying and specifying components, including materials • Drawings of any special auxiliary equipment or piping Final Data Promptly after final inspection, the seller shall furnish: • As-built data sheets and final drawings • Operating and maintenance instructions (IOM manuals), including lubrication requirements • Certified performance curves from the performance test data • Completed customer CMMS equipment forms

Pump Design Specifications The following requirements should be considered for inclusion in any pump standard or specification. Standards and specifications should be evergreen documents. Organizational experience and lessons learned should be systematically added to procurement requirements as necessary. General Pump Requirements 1. Pumps shall be designed for a minimum of 55-months MTBF. 2. Pump design shall emphasize reliability and maintenance simplicity. 3. Pumps shall be sized to approach maximum operating efficiency. Hydraulics 1. Pump and driver, including all couplings, shall be capable of withstanding all start-up and shutdown hydraulic surges and thrusts against closed valves or open pipe connections. 2. Pump shall have continuously rising head to shutoff. 3. Pump head at shutoff shall be at least 110% of the head at specified capacity and 115% of head at BEP. 4. Datasheet shall show minimum continuous stable flow. 5. NPSHR water basis shall be no more than 80% of NPSHA, and never within five feet of NPSHA. 6. Pump designs with mechanical seals shall have a stuffing box pressure greater than the suction pressure to ensure that the temperature and pressure in the stuffing box are adequate to prevent flashing and provide continuous flow though the seal chamber. 7. Correction factors given in the Hydraulic Institute Standards shall be used for sizing pumps handling liquids with a viscosity greater than water. Casing, Nozzles and Pressure Casing Connection 1. Pressure casings shall be selfventing. 2. All suction and discharge nozzles shall be flanged. 3. Pressure casing shall be designed for the maximum discharge pressure, plus allowances for head and speed increases at the pumping temperature. 4. Double volute pumps shall be considered for pumps with discharge nozzle sizes of three inches or largThe Pump Handbook Series

er and required for discharge nozzles six inches or greater. 5. Frame adapters shall be constructed at a minimum of ductile iron. Cast iron is prohibited. Rotating Elements 1. Solid shafts are required, except where conventional packing is used. 2. Seller shall specify shaft deflection as a function of L3/D4. This shaft flexibility factor shall not exceed 65. 3. Impellers shall have a minimum of five vanes. 4. Impeller design shall be neither minimum nor maximum. Design head shall not exceed 90% of the head obtainable with the maximum size impeller at design flow. 5. Rotating elements shall be dynamically balanced to API 610 (ISO 0.665G, 4W/N), maximum residual unbalance. 6. Vibration displacement limit in mils (peak-to-peak) shall not exceed the following limits. • 1200 rpm 1.5 mils • 1800 rpm 0.7 mils • 3600 rpm 0.4 mils Shaft Sealing 1. Unless otherwise specified, inside mounted, balance cartridge type mechanical seals are required and shall conform to the requirements of API 682. 2. Mechanical seals are specified. Bearings and Bearing Housing Design 1. Bearings shall have a minimum rated bearing life of L10 = 50,000 hours. 2. Radial bearings shall be Conrad type with a loose internal clearance fit, AFBMA C3. 3. Axial thrust bearings shall consist of two angular contact thrust bearings mounted back to back; inboard bearing with a 40° conMACHINERY DATA FILES ( ) 1. Installation, Operating & Maintenance Manual (Instructions) ( ) 2. Specification Sheet ( ) 3. Cross-sectional Drawings (if not included in Maintenance Manual), Outline (Dimensional), etc. ( ) 4. Bill of Material (complete parts list) ( ) 5. RPL (Required Parts List) - minimum to be stocked by stored or local vendor ( ) 6. Performance Curve (if applicable) ( ) 7. Mechanical Seal Information (Drawings, etc.) ( ) 8. Design Change Data ( ) 9. Computer Input Form

Figure 2. A machinery data file similar to the one shown here should be the product of accumulation of vendor data during the procurement process.

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Pump Process Data Sheet Unit: __________________________________By: ________________________________Date: ________________________ Appropriation: __________________________Last Rev. By: ________________________Date: ________________________

GENERAL DATA Item Service Number Required PUMPING CONDITIONS Fluid Temp., °F Sp. Gr. @ Temp., °F Sp. Heat @ Temp., °F Vis. @ Temp., °F Max. Vis. @ Min. Temp., °F Vapor Press., psia Solids Corrosive Materials Erosive Materials Chloride Concentration, PPM Particle Concentration, % by wt. Particle Size, Micron Start-up Fluid Min. Flow Rate, gpm @ °F Normal Flow Rate, gpm @ °F Design Flow Rate, gpm @ °F Design Suction Press., psig Min. NPSH Available, ft. Design Discharge Press., psig Design Differential head., ft. BHP @ Design Conditions CONSTRUCTION Pump Type Usage Factor Seal Type Materials - Case Impeller Trim Cooling Water Req’d EMERGENCY CONDITIONS Max. Suct. Press., psig @ Temp °F Max. Suct. Temp., °F @ Press., psig Max. Disch. Press., psig @ Temp °F DRIVER Type Steam Cond. for Steam Turbine Figure 1a. Pump process data specification sheet

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The Pump Handbook Series

Pump Process Data Sheet

GENERAL DATA Item SYSTEM DESCRIPTION SUCTION VESSEL: OPEN TO ATMOSPHERE or CLOSED SYSTEM SUCTION VESSEL PRESSURE, PSIG PUMP LOCATION: BELOW OR ABOVE VESSEL SUCTION VESSEL HAVE LEVEL CONTROL? YES NO YES NO YES NO SUCTION VESSEL HAVE PRESSURE SENSOR? YES NO YES NO YES NO HOW IS SUCTION VESSEL PRESSURE MAINTAINED? Pressure: IF FLUID LEVEL OR TANK PRESSURE DROPS TOO LOW, WILL SYSTEM AUTOMATICALLY STOP THE PUMP? YES NO YES NO YES NO COULD PUMP POSSIBLY RUN DRY? YES NO YES NO YES NO REMARKS:________________________________________________________________________ __________________________________________________________________________________ __________________________________________________________________________________ SITE AND UTILITY DATA LOCATION: INDOOR or OUTDOOR HEATED or UNHEATED PARTIAL SIDES OTHER WINTERIZATION/TROPICALIZATION REQ’D ELECTRICAL AREA CLASS. CL / GR / DIV UNUSUAL CONDITIONS; DUST or FUMES

IN. YES

OUT. NO

IN. YES

OUT. NO

IN. YES

OUT. NO

/

/

/

/

/

/

SITE DATE: ELEVATION_______FT. BAROMETER____________PSIA RANGE OF AMBIENT TEMP. MIN./MAX_________°F RELATIVE HUMIDITY: MIN./MAX.___________________% UTILITY CONDITIONS: STEAM: DRIVERS HEATING MIN.____________ PSIG_______°F_________PSIG______°F MAX.___________ PSIG_______°F_________PSIG______°F ELECTRICITY: DRIVERS HEATING CONTROL SHUTDOWN VOLTAGE _________ _________ _________ ____________ HERTZ _________ _________ _________ ___________ PHASE _________ _________ _________ ___________ INSTRUMENT AIR, PSIG: MAX.______________ MIN._____________ COOLING WATER: WATER SOURCE ________________________________________ TEMP. INLET____________°F MAX. RETURN_______________________________°F PRESS. NORM.__________PSIG DESIGN______________________________________PSIG MIN. RETURN___________PSIG MAX. ALLOW DELTA P__________________________PSI REMARKS_________________________________________________________________________ __________________________________________________________________________________ Figure 1b. Pump process data specification sheet

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tact angle and outboard bearing with a 15° contact angle. 4. Bearing housings shall be fitted with compound labyrinth bearing isolation to eliminate contamination. Conventional elastomeric lip seals are not permitted. Materials 1. The use of cast iron in pressurecasing parts of pumps in not permitted when the fluid pumped is hazardous, flammable or toxic. 2. Gaskets and elastomers shall be compatible with the pumped fluid. 3. ASTM A 193, Grade B7 or approved equal material shall be used for studs and/or bolts in casing joints. Couplings 1. Couplings and coupling-to-shaft junctures shall be rated for 1.5 times driver horsepower, including the motor service factor. 2. Coupling-to-shaft junctures shall be capable of withstanding four times the rated torque without yielding. 3. All coupling register fits shall be perpendicular to the shaft within 0.0001” of face diameter or 0.0005” total indicator runout, whichever is greater. 4. Non-sparking coupling guards of solid aluminum or fiberglass construction shall be provided. The guard material shall be a minimum of 10 gauge (0.135”), firmly bolted to the baseplate and extend to within 0.5” of the pump and motor casings. Guards shall meet the requirements of OSHA and ANSI B15.1. Baseplates 1. Baseplate shall be designed to accept shaft centerline height and support feet of the next larger frame size motor without field welding modification. 2. Baseplate shall be designed with sufficient length to avoid driver and pump suction nozzle from extending over the end. 3. Equipment mounting pads shall be fully machined, flat, parallel and in the same plane within 0.002” per foot of distance between pads. 4. Transverse and longitudinal alignment screws shall be provided on all baseplates without ”C” flange drivers. 5. The baseplate shall have a structural rigidity as specified in the

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$ Labor $ Material $ Total Average direct charges per repair (split: 50/50 labor/materials) 2690 5380 2690 1345 Employee benefits (50% of labor) 1345 --268 Refinery administration & services (9.95% of labor) 268 --3094 3094 --Mechanical O/H (115% of labor) 200 --200 Materials procurement (7.4% materials) $7397 $2890 $10,287 Total Table 2. Total maintenance credits $10,287 per avoided repair Total Number Pumps: 50% in Service:

2754 1377 (1984) 1427 (1983) Average Pump field repairs 100 (1984 ytd.) incl. 40 shop & 60 field repairs Repairs/Month: 90 (1983) incl. 42 shop & 53 field repairs $538k (1984) based on work order tracking Average Pump Maintenance Cost/Month: $651.7 k (1983) Average Pump $5380 (1984) Repair Costs: $6860 (1983) In the past, BTRF has included incremental burden savings in total maintenance credits for reducing pump failures. The credit work-up is shown in Table 2.

Table 3. Pump repair cost calculation for a major refinery

latest edition of ANSI B73.1 for freestanding baseplates, regardless of grouted or freestanding installation. 6. Baseplate shall be fitted with one 6” grout fill hole uniformly distributed for every 9 ft2 of baseplate surface and/or subdivided section. Each grout hole shall have a 1” cylindrical raised lip to facilitate grout filling. 7. Vent holes shall be a minimum of 1 ⁄2” in diameter and be provided for each bulkhead compartment at all corners, high points and perimeter edges.

Conclusion Maintenance organizations have traditionally focused increases of MTBF on better repair techniques and improved operator knowledge. Less-than-adequate pump design by the manufacturer can and does significantly influence MTBF improvements. With industry sources claiming the average cost of a pump failure is anywhere between $2,600 to $10,200, exclusive of production loss, better pump procurement practices must be considered for fundamental improvement to occur. The estimated cost of upgrading a centrifugal pump to ANSI PLUS designs ranges from $1,500 to $4,000 and can be justified by an MTBF improvement of six months to two years. Spending incremental additional dollars on the procurement of pumps will reduce your life cycle cost while freeing up your personnel to work on other problems, all the while increasing your on-stream time. How can you afford not to The Pump Handbook Series

spend time aiding in the procurement of your pumps? “There is hardly anything in the world that some man cannot make a little worse, and sell a little cheaper, and the man who buys on price alone is this man’s lawful prey.” John Ruskin "

References 1. Bloch, H.P. ”Pump Reliability Improvement Methods for Economy”, PRIME Seminar, 1994. 2. Bloch, H.P. ”Practical Machinery Management for Process Plants, Improving Machinery Reliability”, Vol. 1 , Houston, Texas, Gulf Publishing Company, 1982. 3. Garay, P.N. ”Pump Application Desk Book” Fairmont Press Inc., 1993. 4. Hernandez, T. ”How to Order a Pump and Get What You Want”, 10th International Pump Users Symposium, March 1993. Tim Stefl is a mining engineer with 15 years of experience in maintenance and maintenance engineering. He has worked for Peabody Coal Company, Union Carbide and Arco Chemical. Mr. Stefl is currently the Maintenance Engineering Superintendent for Henkel Corporation Chemical Division’s largest North American facility in Cincinnati, Ohio.

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Overhauling 36 Vertical Process Pumps and Motors under Adverse Conditions An “Indiana Jones” of the pump industry ventures out on a tough project. By John H. Burgner, Rotating Equipment Consultant

n American engineering contractor was awarded the task of renovating and debottlenecking a hydrocarbon process plant at a remote overseas site. The rotating equipment portion of the project included 36 identical multi-stage vertical pumps designed for O NPSHr with 60 hp, two-pole motors. The active portion of the job extended over a two year period, and during that time more than 300 pieces of rotating equipment (pumps, turbines, gears and compressors) and 250 electric motors (up to 2000 hp) were overhauled in shops created specifically for this project. As an employee of the American company, I had direct supervision of the rotating equipment shop overhaul and advisory supervision of the mechanical portion of the electric motor repair. Nationals of the host country that were undergoing training supervised by ex-pat advisors in each craft performed the actual shop repair. In addition to the obvious need for timely, cost-effective completion, we established three objectives:

A

1. The finished product in performance, life expectancy and appearance would be equal to that produced by a good, independent pump repair shop in the U.S. 2. All pumps and motors would be repaired so that they would be dimensional and material duplicates of the originals.

3. All pumps would be repaired so that any future maintenance could be performed using OEM parts. (For reasons that will be discussed later, this was not a concern with motors.) In addition to the obvious benefits these objectives would provide the end-user, we gained the advantages of a production operation. Once we agreed upon an approach, we were able to source materials, make drawings, write procedures and train people. All of this was done based on the facility’s 36 units, not just the one in front of us. We did not accomplish each objective completely, but we came close. The only parts that had no significant problems were motor bearings mechanical seals and pump, sleeve type, bowl and column bearings. These were expected areas of concern and adequate preparations were in place. None of our problems would be considered high-tech. The repairs would be considered routine for any competent U.S. pump repair shop. I feel that the interest lies in the methods of solution, the people involved, the location, the scarcity of parts and information, and the apparent degree of success. The problems we encountered were largely related to the age and condition of the equipment and compounded by the jobsite location. The solutions were directly related to the abilities of the host country The Pump Handbook Series

nationals in areas of drafting and machine work, supported by very good ex-pat advisors in machine work and electric motor repair. In an effort to bring organization to the review of what was a chaotic situation, I will list (in approximate chronological order) the problems as discovered and then the solutions. In actuality almost all of the problems, and many of the solutions, occurred simultaneously very early in the project. We had about 18 months to fine-tune and confirm our corrective action.

Impellers and Bowls We knew impellers and bowls were going to be a problem even before we got the first pumps into the shop. The pumps were 3x4x8, 7stage units, which means 252 impellers and bowls. For various complex reasons, we did not have adequate parts, and short of believable death threats, we weren’t going to get them. Please remember that during this time, we were working in an inaccessible location. On the positive side, all 36 pumps and motors were runable. The vibration levels were high, very high on some units, and performance wasn’t all it should have been, but they all ran. By the time the first six or so pumps were in the shop and apart, we had our approach lined out. We did have a few new impellers and bowls available, but we knew that there were none to waste. After measuring

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all the wear ring area critical clearances on the pumps we’d received to date, we felt comfortable proceeding as follows: 1. Bowls and impellers with severe damage were scrapped and replaced. 2. All impeller ring areas were turned to a common diameter (approximately 1⁄8” undersize). These units had not been supplied with impeller rings or bowl rings, which made our job much easier. 3. All bowl ring areas were bored out to a common diameter (approximately 1⁄4” oversize). 4. New bowl rings were fabricated, pressed and lock-screwed in place. When considering this procedure, you must factor in thrust. By reducing the impeller ring area diameter, we slightly reduced the net down thrust. You should also keep in mind that operating hydraulic forces tend to move impeller rings out of position, but they hold bowl rings in position. There were times in the project when we salvaged a bowl with a worn bearing fit by boring and bushing. We could then use a standard OEM bearing. Because we did not initially have easy access to reliable balancing, we balanced all impellers individually on knife-edge rolls using a mandrel. This works very well if the rolls are kept clean and in good condition. Doing all these steps put our bowl and impeller problems behind us for the duration of the project.

Shafts These were our next areas of concern. There were a few new shafts available, but their runouts were totally unacceptable. We straightened them by center-punching the concave area of the bend. This method takes a little practice, but we were able to train the locals to bring runouts of over .010” per 5’ to below .002” per 5’ on 13⁄16” diameter 416 shaft material. The biggest problem was avoiding impeller fit and bearing location when center-punching. With 36 pumps to overhaul, every possible shaft had to be salvaged. This was not easy because nearly all the shafts were damaged in the shaft sleeve O-ring fit area. This area must be right because it is subject to full discharge pressure, is almost impossible to test, and if a failure occurs,

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results in removal and complete dismantling of the pump. A normal repair involves metallization or plating of this area. Metallization was the only procedure available, but I don’t trust it unless I have faith in the metallizer, which I didn’t, and there is a heavy, non-corrosive, ambient liquid involved, which there wasn’t. We were able to salvage shafts where the damage was confined to the O-ring area using the following procedure, which is similar to repair of water well pump shafts: 1. Flip the shaft end for end. 2. Machine the bottom end (formerly the couplings fit) 1⁄4” undersize for a distance of about 6” with a slight radius at step. 3. Fabricate a sleeve slightly oversize on OD, .001” interference to the turned down shaft with .010” or so relief over the center 4” of ID with a slight chamfer to fit shaft radius. 4. Heat the sleeve, install it and pin it in place for extra security. 5. Turn down the oversize sleeve and machine the top end for coupling fit. All of the old shafts required straightening, but employing the above method yielded acceptable runouts. We ran into two other problems. The seal flanges (glands) were carbon steel and badly corroded, especially in the seal seat fit. We fabricated replacements in 18-8 stainless steel; it was an easy fix. Column sections with bearing fits that were worn oversized were repaired by boring and bushing, the same method we had used for the bowls that were worn in a similar manner. As the pump repair was being implemented, the motors were being dismantled, cleaned and inspected. At first glance they looked pretty good; this proved not to be the case. On the positive side, the motors were very well built, heavy cast iron construction and quality throughout. However, the OEM had been bought and was effectively out of business. In addition, not one motor sectional drawing was available. I know this doesn’t sound possible, but a search conducted on three continents failed to uncover one. The Pump Handbook Series

Motor Thrust Bearings The arrangement of these parts was the most obvious area of concern. We were reasonably sure we knew what bearings were required: duplex, 40 degree, angular contact of a size that fit the housing and shaft with a rating suitable for the total down thrust load of pump and motor. (The thrust bearing had been grease lubricated, which I have never felt was suitable for high thrust, two-pole motors.) When the motors came apart, we found that some bearings had been mounted back-toback, some face-to-face and some tandem, facing up. Surprisingly, all of these combinations were in running condition. But the condition of the shafts and housings indicated that numerous failures had occurred. As mentioned, we had no drawings and no usable repair records. If an angular contact bearing is ground correctly it can be mounted in any of the three arrangements. Our measurements indicated that the bearings had been ground to accept face-toface mounting. We also found that there was too much end float of the outer race in the housing. (This needs to be limited.) But there was a circle of about 10 holes, 1⁄4” diameter 3⁄16” deep in the bearing end cap directly over the top outer race, exactly perfect for spring loading the outer race of a face-to-face mounted bearing. There were no springs in place, no pieces of springs and no drawing showing springs, but by installing springs, we put a light, but resilient preload on the upper bearing. This controlled and limited rotor upthrust. Spring loading in this manner is not unheard of, but it is unusual. We put them together and prayed. Events bore us out.

Rigid Spacer Couplings We expected these to be a problem, and we were not disappointed. The seal life of pumps of this design depends on limiting the shaft runout in the seal area. Excessive shaft runout means shortened seal life. Vibration also becomes a problem if coupling precision is not maintained. Our couplings were corroded and had extensive hammer and chisel marks on them. For a part that must be exact in all areas of contact, they were totally unacceptable. Couplings are not “hammer to fit, paint to suit” items. We attempted to

weld up and remachine the fits on the first few units; it took forever to get something we could live with. We found it easier and quicker to scrap the entire coupling and fabricate new ones.

Motor Shafts These shafts presented a problem for the same reason that couplings did. The motor shaft must be exact to avoid extending the runout down through the couplings to the seal area of the pump shaft. Remember, the coupling fit to the pump shaft is always a slip fit; this enables seal replacement. The coupling fit to the motor shaft may be a slip fit or a slight interference fit, but in all designs the motor shaft condition and runout must be as close to exact as shop instruments will measure. Our design called for an interference fit, and some shafts were not acceptable. In order to obtain shaft tolerances we could live with, we replaced the motor shaft whenever a .001” clean up cut would not provide the shaft coupling hub fit condition we needed. (The fact that the bearing fit areas were damaged by bearing failures made this decision a little easier.) Note! Machining the shafts significantly undersize would have required remachining keyways, replacing split rings, and cost us our two secondary goals of interchangability and the ability to use OEM parts. Metallizing was undesirable and plating was unavailable. After the couplings and motor shafts were replaced, the bore runout of the assembled coupling pump hub, when mounted on the assembled motor, was less than .004” TIR.

Motor Rotor Balance This caused us an unexpected problem. When we got the first motor to test, the vibration levels were slightly over .15 in./sec. (filtered to lx rpm). In this situation, rotor unbalance is the expected culprit. We check-balanced the assembled rotor and the readings were good. It has been my experience that a rotor of this size, with a sheet metal fan, can be balanced with the fan in place. However, this was not the case. We dynamically balanced the bare rotor, correcting on each end of the rotor pack, and statically

balanced the fan on our knife edge rolls. Vibration levels dropped below .10 in./sec. We standardized on this procedure for the entire project. After completing the repairs, we started shipping the pumps and motors. Please keep in mind that these vertical process units were only 20% of our shop load. The other 80% were processing through the shop at the same rate. There was a time lag between shipment and start-up that allowed complacency to set in. We were not at all prepared for the next problem.

Motor Lower Bearings These bearings were failing on start-up. None of our previous problems had caused a project delay, and we had handled all of them inhouse. Now we had the very unpleasant prospect of project management involvement. Five lower motor bearings failed within 24 hours of start-up, and we didn’t have a clue as to why. We had somewhat expected thrust bearing problems, but with the lower (radial) bearing on a motor of this design, there is effectively no load. It has been my experience that when Conrad lower bearings in this service fail soon after start-up there is an error in assembly that prevents the lower bearing from moving down in the housing as the rotor expands. This was the first thing we checked. No luck; all was OK there. We checked fits, clearances, lubrication, everything we could think of; and still our no load lower bearing test temperatures were reaching 100°C before we shut down. On initial tests this temperature was not taken. The comment was made that we had to have changed something, so by process of elimination, the lower bearing lip seal was noted. It had been changed to one of higher quality (spring-loaded). The seal was removed from the test motor. The bearing temperature stabilized at 56°C. The increased drag of the spring-loaded lip seal put enough heat into the shaft to expand the bearing in the housing, effectively locking it in place and transferring all axial loading to the radial bearing. We quickly designed and fabricated a non-contacting bronze grease baffle. Problem solved. Or so we The Pump Handbook Series

thought. On the very next motor tested, with a non-contacting grease baffle, we had to shut down when top and bottom bearings reached 95°C. This time when we checked bearing to housing fit, we noticed a small difference. The housing ID was slightly smaller than expected. The bearing we were using had a maximum design OD of 4.5276” and a housing inspection limit of .0005” to .0027” loose. Our bearing and our housing clearance dimensions were within tolerance but on the low end (close to .0005”) and the housing was new (obtained from client stock). We opened the clearance up to .0020” and from then on we were home free. It turns out that in the plant there were motors of the same make, hp and rpm but with a roller, lower bearing. A roller outer race does not have to float in the housing, a ball outer race does. We had received and installed a housing for a roller bearing. In addition, the configuration of the housing tended to retain heat in the bearing area, which again locked the lower bearing in the axial position.

Pump Discharge Heads These heads gave us our final problem. A short way into the project we started getting field complaints about motor-to-pump alignment problems. As part of our assembly checks we would check the motor base face and fit runout as well as the motor shaft and coupling bore runout mentioned previously. As always, pump shaft runout was confirmed and recorded. The clearance of the motor base fit to the pump head was .002” to .006”. We didn’t check the pump discharge head because we hadn’t done anything to it. From our viewpoint they should have been able to obtain an exact alignment. However, to close the only loophole, we started checking discharge head runouts. For alignment purposes there are only three areas of concern. The fit of the column to the head, the fit of the seal flange to the head and the fit of the motor to the head. We turned a fixture on a vertical turret lathe that duplicated the column fit (to head); we then screwed the head in place and checked the radial and axial runouts of the motor and seal flange fits. During years of operation, dis-

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tortion had occurred. We found runouts in the .005” range, which we corrected by welding and machining the radial fits and skim cuts of the mounting surfaces. Final problem solved.

Conclusion The first pumps shipped had about 18 months time in operation when I completed my contract and departed the jobsite. All of them had

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at least three months. We had no performance complaints and the installed, operating, vibration levels on most units were below .1 in./sec. No unit was above .15 in./sec. We did have some seal failures, but nothing that I would consider excessive. To my knowledge no motor bearing failures, other than the initial five, had occurred. Grease lubrication was OK after all. "

The Pump Handbook Series

John H. Burgner is a rotating equipment technician with more than 30 years of practical field experience in both domestic and international operations of centrifugal pumps. He was a speaker at PumpUsers Expo ‘98 in Cincinnati, OH and PumpUsers Expo ‘99 in Nashville, TN.

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Selecting Motors & Drives for Optimum System Reliability Don’t buy a new motor or drive until you ask these three questions. by Jerry Muehlbauer, Marathon Electric hen an AC motor and VFD (variable frequency drive) are to be installed on a centrifugal pump, most of the tough application questions involve selection of the pump, the VFD, or the process controls. Too often, little attention is paid to the motor itself until something goes wrong during start-up. That’s when “finger pointing” begins and the phones start ringing at the office of the motor vendor. Fortunately, with some simple up-front work, most of the potential motor problems can be headed off early in the application process. There are three basic motor questions the specification engineer should ask to help match a standard General Purpose AC induction motor with a VFD and ensure a successful centrifugal pump application.

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Question #1: Will the motor produce enough torque? Start with the Pump The pump’s full load point will normally dictate what size motor and VFD are required. Always begin power calculations at the driven load and work backward to the motor, then to the VFD. If a speed reducer is required between the motor and pump, be aware that they are not 100% efficient. Belt, chain and gear reducers range from 60 to about 90% in efficiency, so their losses must be added-in when figur-

ing input power requirements. Remember also that mechanical speed reducers provide increased torque in direct proportion to their input-to-output rpm reduction; VFDs do not. At best, a VFD can produce the same amount of torque at low speed as at rated speeds (no torque gain produced). This fact is sometimes overlooked when a fixedspeed motor is retrofitted with a VFD and the speed reducer is removed. The result is that the motor may not supply enough torque and it might stall during the initial start-up procedure, if the VFD doesn’t trip out first. Define the Base Load For centrifugal pumps, the base point, where the pump is rated, is normally the highest load point. Base horsepower defines the power needed at the base rating point of the pump. Base speed is the point where the motor produces its rated horsepower (the value stamped on the nameplate) at the listed supply voltage. For a general purpose motor name-plated at 100 hp, 1785 rpm, 60Hz, 460 volts, the base speed is 1785 rpm, base horsepower is 100, base frequency is 60Hz, and base voltage is 460. In Figure 1, base horsepower and base speed are indicated by point A, where the 1785 rpm and 100 hp lines intersect. In this example, the pump requirement matches the motor’s capability and the Ideal Variable The Pump Handbook Series

Torque Load curve intersects with the motor’s base point. List Speeds, Not Frequencies Sometimes a customer specification will list a desired base frequency instead of the pump’s required base speed. This can lead to confusion, since a motor with a 60Hz base frequency may have a base speed of 3450, 1760, 1175, 890 rpm, etc. When this occurs, someone (a base character?) must clarify what the required base speed is so that the correct motor can be supplied. This is normally not a problem if the specification lists the motor’s synchronous rpm (3600, 1800, 1200, 900, etc.), as long as the customer recognizes that the motor will run somewhat slower than synchronous speed at its base point.

Figure 1. Base horsepower and base speed are indicated by point A, 50% speed is point B, 133% load increase is point C.

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No Problem below Base Speed Note in Figure 1 that pump load drops off sharply below base speed, decreasing to only 12% of base power at 50% speed, point B. In an ideal variable torque load, power required varies by the cube of the speed change (895/1785 E3 = 12.6%). Because of this sharp drop-off, the 100 hp TEFC and DP motor’s continuous capability significantly exceeds the load below base speed. This is the case for all energy efficient NEMA (National Electrical Manufacturers Association) design B motors when powering variable torque loads. As a result, we need not be concerned about the motor stalling or overheating below base speed. At very low speeds, when the motor’s cooling air is nearly zero, both TEFC and DP energy efficient motors can still provide at least 25% torque continuously without overheating. As long as the low speed pump load stays at about 25% or less, the motor’s low speed thermal capacity can be essentially ignored. Speed Range Doesn’t Apply Some motor vendors list their NEMA design B motors as being suitable for a 10:1 or 4:1 speed range on variable torque loads. A speed range of 10:1 means that the motor can be operated from base speed down to 1 ⁄10th of base speed and produce the specified torque continuously without exceeding its temperature rating. For a 1785 base rpm motor this would be down to 178.5 rpm. A 1785 rpm, 4:1 motor would be good to 446 rpm, a 2:1 to 892.5 rpm, etc. Speed ranges are pertinent for constant torque loads, where, for example a 10:1 TEFC or ODP motor must produce the same torque at 178 rpm as at 1785 rpm even though its cooling fan is moving at only 1⁄10th speed and a risk of overheating exists. For variable torque loads, however, overheating below base speed is not a problem, unless the VFD is very poorly tuned. As such, energy efficient variable torque motors can operate continuously from base speed down to zero speed (1000:1 speed range) without overheating, making speed range labels unnecessary. However, in some instances, UL (Underwriter’s Laboratory) requires that variable torque explosion-proof motors have their speed range listed on the nameplate.

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Account for Peak Loads In some applications, the pump will run lightly loaded most of the time, but experience sizable overloads periodically. These peak loads must be accounted for during the sizing process, especially if they exceed 105-110% of the VFD’s continuous rating. If the peak loads exceed 110% for more than one minute, a standard variable torque VFD will trip offline, shutting down the process. You could install a more expensive, constant torque VFD to provide 150% overload capacity, but it too would “trip off” if the load remains above 105-110% of the VFD rating for longer than several minutes. This timed-overload-trip feature is designed to protect the VFD’s output semiconductors, not the motor itself. Peak torque requirements exceeding 110% for 60 seconds will normally mandate that the VFD be oversized to avoid trips, so that its continuous capacity matches the load’s peak requirements. An EPACT (the Federal Energy Policy Act) motor, on the other hand, can withstand overloads of 150% for 5 minutes (48-215T frame) to 15 minutes (254T-5013 frame) with no problem. Energy Efficient Motors Handle Peak Loads Better EPACT defines and mandates efficiency levels for most types of general purpose motors up to 200 hp. NEMA MG1 Table 12-10 matches these values and extends the range through 500 hp, but has no legal authority to enforce compliance. The end-user or specification engineer should choose energy-efficient motors that meet or exceed the NEMA MG1 12-10 levels when applying VFDs to pump loads. Not only do they consume less power, but these motors also last longer, can handle higher overloads for greater time periods, withstand over-voltage and under-voltage conditions better and tolerate VFD-generated harmonics better than standard efficiency motors. Since the motor is often one of the least expensive components in a pump system, it makes sense to spend a few extra dollars to increase motor reliability. Although motors powered by VFDs are exempt from EPACT compliance by most interpretations, future policy changes at the federal level could mandate their compliance. The Pump Handbook Series

Figure 2. As frequency is reduced, voltage must be reduced by the same ratio.

Tune the VFD to Match the Motor When it comes to programming the VFD, the base voltage and base frequency must be set to match the motor, regardless of what other scheme may seem to make sense. A 460 volt, 60Hz motor needs the VFD to output 460 volts at 60Hz to operate correctly. Reprogramming the VFD does not change the motor’s need for correct voltage at a given frequency. For example, setting the base frequency of a 460 volt, 60Hz motor to 66Hz (to run the pump 10% faster) will not provide improved motor performance. It will, in fact, be much the same as operating the motor at 418 volts, 60Hz, which will weaken the motor substantially. Add to this the 33% load increase (point C in Figure 1) and a motor stall or VFD trip is just around the corner. A motor must be rewound (or initially over-designed) to operate above base speed on variable torque loads. Select a Variable Torque v/Hz Curve To produce rated torque at normal current requires that the voltage be varied with the frequency according to the motor’s volts-per-hertz ratio (v/Hz). A motor name-plated at 460 volts, 60Hz has a v/Hz ratio of 7.67 (460/60). As frequency is reduced, the voltage must be reduced by the 7.67 ratio to maintain proper performance. This is shown in Figure 2 by the Constant Torque Volts curve. At 50Hz, the voltage is dropped to 380, at 30Hz it’s 230, and at 15Hz the voltage is 115 in order to maintain 7.67 v/Hz. At very low frequencies (below 30Hz on small motors; below 10Hz on larger ones) the voltage must be “boosted” above the normal v/Hz ratio to offset resistance losses in the motor while producing rated torque at rated current. However, too much boost can cause the motor

to be very noisy, “cog” (produce noticeable torque pulsations at the motor shaft), stall during acceleration and sometimes cause an instantaneous over-current (IOC) trip in the VFD. Because variable torque loads don’t require full torque below base frequency, the voltage can safely be reduced below the linear v/Hz ratio and voltage boost can be minimized. Voltage at 30Hz would be 150-180; at 15Hz it could drop off to 75-95 volts. The motor’s ability to produce torque will drop as a result, but this won’t be a problem as long as motor torque capacity exceeds the pump’s load requirement at each frequency. Lowering voltage will reduce motor flux densities, which quiets the motor, reduces torque pulsations and system vibrations, and increases the motor power factor. Most VFDs will provide a variable torque v/Hz curve, similar to the Variable Torque Volts curve of Figure 2. Note that at base speed, full voltage is still applied to the motor. This profile can be set during commissioning of the system to reduce potential vibration and noise. Don’t Exceed Base Speed Just as load drops sharply below base speed, it increases drastically— also by the cube—above base speed. Point C in Figure 1 shows that if speed goes up only 10% (to gain 10% more flow), the power requirement increases to 133%. Properly sized motors and VFDs don’t have enough margin to handle such an overload, especially above base speed. Note that point C exceeds the continuous capability of both the DP and the TEFCs. In fact, it exceeds even the intermittent capability of most motors; a motor stall and VFD overcurrent trip would be inevitable. It is important, therefore, not to exceed base speed by more than 2 or 3%, even on an intermittent basis, without first consulting both the motor and VFD vendors. Both components will often need to be oversized to handle the extra load. A drive can reduce its output voltage below base speed, but VFDs normally run out of voltage at base speed; they cannot increase output voltage above their input level. With a 460 volt input, a VFD can produce 440 to 470 volts at best. In Figure 2, output voltage increases linearly with frequency (Constant Torque

Volts line), showing the proper v/Hz relationship needed for constant torque. As frequency exceeds 60Hz, voltage must continue to increase toward point B if the motor is to produce rated torque at normal current levels. In reality, the VFD runs out of voltage (dashed line to point A) above 60Hz and the motor’s ability to produce torque suffers. It is now “voltage starved.” As a result, load torque must be reduced to keep the motor from stalling. Since the pump’s torque requirement continues to increase above base speed, the motor must be either grossly oversized or redesigned to permit continued operation. Special Base Frequencies Makes Line-Starting Tough If a motor is rewound for a base frequency higher than the VFD’s supply frequency, it cannot be across-the-line-started from the utility mains unless the supply voltage is reduced to match the motor’s v/Hz. For example, if a motor were sized to run the pump at 25% over the 60Hz speed, it would be rewound for 460 volts at 75Hz. When it is line-started at 60Hz, the supply voltage must now be dropped to 368 volts. This makes line-starting more expensive, since either the motor must be overdesigned to handle 460 volts at 60Hz or an extra transformer or specially programmed solid state starter must be added. Keep in mind that the pump will also run 25% slower at 60Hz on utility power than it does on 75Hz VFD power. A 60Hz base frequency enables the bypassing of the VFD when needed. In a bypass system, electromechanical contactors are installed on the input and output side of the VFD. In the event of a VFD or process control fault, the contactors can be switched, bypassing the VFD and connecting the motor directly to utility power. This is often a design specification for municipal pumps and applications that must continue to operate in emergencies. If the drive is to be bypassed for line-starting in an emergency, define the type of starter when you buy so that the correct motor configuration can be provided. Although an inverter duty motor only requires three leads, a part-winding-start motor needs nine and a Y-∆ has six or twelve leads. Extra leads can’t be easily added The Pump Handbook Series

once a motor is in the field. When using a part-winding or Y-∆ connected motor on an inverter, always connect the motor in the “run circuit” configuration, not the “start circuit” configuration. Choose Design B for Greatest Flexibility Centrifugal pump loads are normally a good match for NEMA design B motors. Designs C and D generally sacrifice efficiency and have lower running speeds than design B motors, making them less than desirable for pumps. Design A and E motors have high running speed and high to very high efficiency, but they also have higher starting amps than design B, which can require oversize starters for linestarting. Design B motors are the type most commonly stocked by motor distributors. Because of their suitability for pump loads, compatibility with standard across-the-linestarters, and availability from stock, motor vendors normally select a design B motor as the first choice for VFD-powered pump applications. If you want to use a design A or E, be sure to indicate it clearly at the time of order entry. If you choose a design C or D, get ready to justify it when you receive that phone call from the puzzled motor vendor. No bypass? Save some bucks If the motor will never be bypassed, the motor designer can sometimes optimize its performance by choosing a lower base frequency. For example, a 460 volt, 200 hp, 60Hz, 595 rpm, 5011 frame motor draws 309 amps (due to its 64% power factor). If a 600hp, 1785 rpm, 5011 frame motor is wound for a 460 volt, 20Hz base frequency, it will produce 200 hp at 595 rpm but draw only 213 amps (due to it’s higher efficiency and 90% power factor). Because VFDs are sized by their current capacity, optimization would enable the use of a much smaller drive, which could save the end user a whole bag of money. One catch: The VFD must be programmable to output 460 volts at only 20Hz. Some drives on the market today can’t output their base voltage at such low frequencies. Old-Style VFDs Can Cause Torque to Pulsate Some older VFD designs (6-step and GTO or SCR-based PWM types)

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the mechanical components as the vibration source. Generally, if a vibration’s frequency is 30Hz or less, its source is mechanical in nature, not electrical.

Photo 1. Inverter-duty vertical P base motors with heavy duty drip covers

produce voltage waveforms that cause torque pulsations in the motor, which can then excite the resonant points of an assembled mechanical system. Lowering motor flux significantly reduces these torque pulsations and can quiet the entire system. The first step to take when you want to reduce system resonant and noise is to use a variable torque v/Hz pattern. Optimize the Carrier Frequency Another easy way to reduce motor noise is to increases the VFD’s carrier frequency from 1 or 2kHz to 3 or 3.5kHz. This will produce a smoother current waveform, reduce eddy current losses and help reduce torque pulsations as well. The greatest reduction in noise and vibration occurs when changing from 1kHz to 2kHz, and noticeable improvements are seen when switching from 2 to 3kHz. Only marginal reduction is obtained by carrier frequencies above 3kHz. Some VFDs use a varying, nonadjustable carrier frequency to help minimize noise and vibration. When lowering flux or optimizing the carrier frequency does not eliminate system resonance, the VFD can be programmed to skip the frequencies that cause resonance during acceleration or deceleration. Conduct a coast-down test to determine if the source of vibration is mechanical or electrical in nature. Run the motor and load up to its rated or top speed, then de-energize the VFD and let the motor coast to a stop. Monitor vibration levels during the coast-down. If vibration decreases significantly, you can assume that the source is electrical because the VFD is disconnected and no longer exciting the motor. If there is very little change, look to

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Resonance Is a “Systems” Problem NEMA MG 1-31.40.5.2 states that the system integrator, not the motor manufacturer, is responsible for minimizing noise and vibration. This may seem unfair until the situation is considered from a systems standpoint. For example, the 100 hp motor from Figure 1 has a natural frequency of 5500 rpm. A compatible pump might resonate at 4800 rpm, a mounting base at 5000 rpm, and the connected plumbing may resonate at 3900 rpm. Since the pump and motor would not operate above 1785 rpm, no single component will reach its resonant point and no problems should be expected. However, if all the parts are bolted together, the system’s resonant point can be as low as 1550 rpm: smack-dab (an engineering term) in the middle of the speed range. Not only would the pump vibrate if operated at 1550 rpm, it would shake each time it accelerated or decelerated through 1550 rpm. What the NEMA statement points out is that there is no way, ahead of time, for a motor manufacturer to compensate for the resonance of the countless components that a given motor might drive, so resonance must be corrected at the systems level. Errors Can Cost Lots of Money The customer should always be the one to define the pump’s torque or hp requirement and speed range, not the VFD or motor vendor. The vendors can offer significant technical assistance, but it is the customer who should ultimately “fill in the blanks,” preferably in writing. This helps protect the motor and VFD suppliers from excessive liability in the event of a misapplication. If an undersized motor is selected, it may overheat or even stall. If a larger motor is retrofitted to correct an undersize condition, it may not fit in the same space as the undersized motor, causing installation headaches. Also, although VFDs are advertised by KVA or hp, they are really sized by the current capacity of the output transistors. Larger loads require more current (to produce more torque), which may require a The Pump Handbook Series

larger, more expensive VFD, which may not fit in the available space. In addition, the larger VFD may require more power from the utility and force you to buy larger feeders, transformers and breakers.

Question #2: Will the Motor Overheat on the VFD? Specify an IGBT (Insulated Gate Bipolar Transition) VFD As the VFD’s carrier frequency increases from 1kHz (1985 GTO (Gate Turn Off)-based design VFD) to 3kHz (1993 or newer IGBT-based design) the motor will run 25-50% cooler. A motor design that overheated 10 years ago on a GTO drive may run fine on today’s VFDs. As carrier frequency goes higher than 3kHz, motor temperatures only decrease by another 1-2 %; there’s little benefit to raising carrier frequency above 3kHz (except if audible motor noise is a severe problem). Always call your motor vendor if plans include the use of a GTO or SCR-based VFD. These old designs produce significant 5th, 7th, and 11th harmonics, which dramatically increase rotor losses and can cause severe overheating, especially in motors above 100 hp. The exception to this rule is the new generation of GTO-based medium voltage (23006600 volts) VFDs, which produce much better waveforms and greatly reduce these nasty harmonics. Beware of Larger Motors Larger (100+ hp) general purpose design B motors with die-cast rotors are often designed with rotor bar geometry that compromises motor performance on VFD power. As a result, they tend to run hotter on VFD power than comparable small motors, especially when the VFD’s carrier frequency is below 2kHz. In a recent lab test, a 400 hp ultra-high efficiency TEFC motor was operated at full load on a dynomometer from a VFD with a 1kHz carrier frequency. When transferred from line power to the VFD, the motor’s temperature rose from 64°C to 100°C (the limit is 105°C), an increase of 56%. A less efficient motor would have exceeded its safe temperature and more than likely failed during its first year of operation. It is always best, therefore, to contact the motor vendor before applying a larger

motor on a VFD. Most motor manufacturers have conducted extensive lab tests and created sophisticated computer models to help identify the VFD capabilities of their various product lines. Their charts, like the one shown in Table 1 (for Marathon Electric motors), provide an easy way to identify the limits of standard general purpose motors. Don’t Exceed Base Speed Operation above base speed on a pump load will normally cause motor overheating if the VFD doesn’t trip out first. The motor should have thermostats, thermistors or RTDs (Resistance Temperature Detectors) installed on the windings to detect such over-temperatures and trip off the VFD. This will not only help prevent serious motor damage, but over-temperature detectors can help identify a process or motor problem. If a serious motor problem exists, a lack of thermal protection can cause the motor to overheat and fail catastrophically. It’s often very difficult for the motor shop to diagnose a charred motor and determine what failed first. Many service departments will not honor a warranty claim if the motor windings show evidence of severe overheating. The standard response to an overheated motor is that the customer misapplied it and therefore should not expect to receive compensation for the damage. Many VFDs have solid-state programmable current-vs.-time protection circuits, which can be adjusted to protect the motor as well as thermostats do. Minimize Voltage Boost Voltage boost is required on constant torque applications to aid starting of difficult loads, but variable torque applications need little to no boost. Excessive voltage boost can cause the stator windings to saturate and overheat the motor when it is running below 20-30Hz. Install Energy Efficient Motors A standard efficiency commoditytype motor has two basic requirements: It must produce enough torque to power the load and not start itself on fire. The energy efficient motor must do the same, plus obtain a specific efficiency level. Higher efficiency means lower motor losses, which translates to lower

Product Line Std Eff. ODP Std Eff. TEFC EPACT ODP EPACT TEFC Ultra-High Eff. ODP Ultra-High Eff.TEFC

Maximum 60 Hz Base hp @ Listed rpms 3600 rpm 1800 rpm 1200 rpm 500 hp 450 hp 300 hp 350 hp 350 hp 250 hp 250 hp 250 hp 300 hp 200 hp 200 hp 200 hp 350 hp 350 hp 250 hp 350 hp 350 hp 250 hp

900 rpm 250 hp 200 hp 200 hp 200 hp 200 hp 200 hp

Table 1. Thermal capabilities of design B motors on variable torque loads

operating temperatures. Motors that meet the NEMA MG1 Table 12-10 efficiency levels have more thermal reserve than standard motors, and can handle occasional overloads better without overheating. High Ambient Can Overheat Motors Elevated ambient temperature will often dictate that the motor be over-framed (built in a larger frame size), to help it better dissipate heat. When no information is supplied with an inquiry, the vendor assumes that the motor won’t be operated in an ambient higher than the industry standard of 40°C (104°F). Higher ambient temperature increases the motor’s insulation and bearing temperatures and can lead to premature failure. Once again, since energyefficient motors run cooler to begin with, they can tolerate higher ambient temperatures than their standard efficiency companions. If you suspect that the ambient temperature will be higher than 40°C, be sure to tell the motor vendor upfront, before a frame size is chosen, since a larger motor may not fit in the space initially allocated. Loss of cooling air can cause a motor to overheat, even if it is lightly loaded and the VFD’s load meter indicates that it is “loafing.” The inverterduty vertical P base motors shown in Photo 1 had heavy duty drip covers installed before they were shipped from the factory. They help keep ice and debris from damaging the fan blades and ensure that adequate cooling air is available at base speed. They’re also ideal for the four-man Frigbee Toss at the company picnic. Higher Altitudes Lead to Overheating Altitudes higher than 3300 ft deprive the motor of cooling air and can lead to overheating, especially if the ambient remains at 40°C. NEMA publishes de-rating guidelines for motors operated at high The Pump Handbook Series

altitudes, high ambient or both. As with ambient, if no altitude data is given at the time of order entry, the motor vendor will assume that it is 3300 ft (1000 m) or less. Avoid the Use of Multi-Speed Motors Multi-speed industrial motors have single or dual stator windings and are designed to operate at either of two discreet speeds when line-powered: 3600/1800 rpm, 1800/900 rpm, 1800/1200 rpm, etc. These motors can be designed to provide variable torque, constant torque, or constant horsepower performance when switched from the high speed to the lower speed connection. Multi-speed motors are always a compromise in performance: It is not possible to obtain the same efficiency, power factor and temperature rise as a single speed motor when you put two windings (or one winding connectable for two speeds) in the same size package. Oftentimes multi-speed motors will need to be over-framed to obtain acceptable performance, and they are more likely to overheat when powered by VFDs. They were never intended to be applied on VFDs, but people do it anyway, specifically in retrofit applications where the motor is already in service. To use a single winding variable torque multi-speed motor on a VFD, connect the high speed winding, not the low. For two winding motors, the low speed winding must be disconnected and its leads taped open, not shorted together. There is always detrimental transformer action that occurs between the windings of two winding machines. If the low speed winding leads are shorted together when running on VFD power, excessive voltages may be induced from the low speed to the high speed winding, creating increased insulation stress and possible premature winding failure. Do not use a constant torque or constant horsepower

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Figure 3. The voltage overshoot shown occurs when a fast rise-time VFD pulse encounters the slowing effect of the VFD-to-motor cable.

Figure 5. Waveform overshoots at the motor terminals with 100, 250 and 500 ft. cables

Figure 6. As the VFD’s carrier frequency is increased, more individual voltage pulses are produced per second. Figure 4. Peak voltage related to cable length for a 460 volt IGBT VFD

multi-speed motor on a VFD. That’s just asking for trouble.

Question #3: Will the VFD’s Voltage Spikes Damage the Motor? Corona Energy, Not Peak Voltage, Is The Issue Much has been written regarding VFD-generated voltage spikes, reflected waves and their detrimental impact on the life of motor insulation. The debate is still going on among members of several NEMA Working Groups that are trying to redefine NEMA MG1 parts 30 and 31. These 1993 standards imply that peak voltage and rise times are the main culprits in motor insulation failure. Clarification and standardized test procedures should (we hope) be published in 1999. Until that time, there is no authorized test procedure to determine whether or not a particular motor meets NEMA MG1-30 or 31. Figure 3 shows the voltage overshoot and ringing that occurs at the motor terminals when a fast risetime VFD pulse encounters the impedance of the VFD-to-motor cable. Longer cables produce higher voltage overshoots, due to their increased impedance. Figure 4 graphs the peak voltage

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vs. cable length for a 460 volt IGBT VFD. Note that voltage rises from 1250 volts at 100 ft of cable to only 1350 volts at the 500 ft point. One may consider that an insulation system designed for 1250 volts should be able to tolerate another 100 volts with little trouble, suggesting that cable length might be ignored as a factor in insulation life. From the extensive research conducted by David Hyypio and Neil Roberts at Marathon Electric, however, it is clear that high peak voltage itself is not the main factor in insulation damage. Rather, it is the energy contained in the portion of each spike that exceeds the motor’s corona inception voltage (CIV) level, not the voltage magnitude itself. CIV is the point where the air surrounding a motor’s insulation begins to break down due to voltage stress, causing corona, partial discharges and producing permanent insulation damage. The VFD-to-motor cable length is the most important factor in determining how much corona energy is produced in each output pulse. This destructive energy rises exponentially as a function of cable length, but is largely unaffected by the variations in the voltage rise times present (2000-16000 volts/ microsecond) in today’s IGBT inverters. Long Cables Increase Corona Energy Waveforms of voltage overshoots at the motor terminals with 100, 250 The Pump Handbook Series

and 500 ft. VFD-to-motor cables are shown in Figure 5. The three waveforms are superimposed for contrast. The flat-topped square wave (at the 700 volt level) is the same voltage pulse as measured when it leaves the VFD. The surge impedance of the cable distorts and reshapes this voltage pulse as it travels from the VFD to the motor. Longer cables have higher surge impedance and produce more distortion than shorter ones. The dashed line at 1200 volts is the CIV level for the motor used in this example. The shaded areas indicate the amount of corona energy available to damage the motor insulation. Note that although the peak voltage is only about 100 volts higher at 500 ft than at 100 ft, the destructive corona power is substantially greater. A motor with a 1200 volt CIV level would experience insulation failure much earlier if connected on the 500 ft cable rather than the 100 or 250 ft cable. It is this corona energy, not the ultimate peak voltage, that directly determines insulation life. High Carrier Frequency Makes It Worse As the VFD’s carrier frequency is increased, more individual voltage pulses are produced per second, and for a given application, more destructive corona energy attacks the motor per unit of time (Figure 6). The insulation life of a given motor is reduced when either the cable length or VFD carrier frequency is increased. For example, with a 200 ft cable length, insulation life drops from 90,000 to only 20,000 hours when carrier frequency is increased from 3 to 12kHz. The longest life occurs with short cable lengths and low carrier frequencies, which minimize the corona power produced. Set the Carrier Frequency to 3kHz The main purpose of increasing carrier frequency above 3kHz is to reduce the audible noise that voltage pulses produce in the motor. Carrier frequencies above 3kHz reduce motor temperatures by less than 2 %, and actually decrease VFD efficiency because of the higher switching losses. It is recommended, therefore, that carrier frequency be set at 3kHz except in those applications where audible noise is a critical issue. In those instances, motor cable length should be reduced and voltage rise time increased for best insulation life.

Choose an Insulation That Matches the Installation Needs Tabulated in Table 2 is the recommended maximum cable length vs. carrier frequency for three types of motor insulation systems used in Marathon Electric’s random-wound motors. The insulation systems of other motor manufacturers may not have similar integrity. The main differences between each type listed in Table 2 are the CIV level (higher=better) and resistance to corona damage once partial discharges begin. The better the insulation, the longer the motor lasts once corona starts. Obtaining reasonable insulation life with each type becomes a matter of limiting the corona power, which is most easily done by controlling cable length and carrier frequency. For example, Table 2 shows that with a carrier frequency of 3kHz, a 460 volt standard motor can safely tolerate a 125 ft long VFD-to-motor cable, while the same design motor with inverter grade insulation can handle 875 ft. cables and still obtain the same insulation life. Reducing cable length will increase insulation life for both of these motors. If the planned cable length at an installation cannot be reduced, the specification engineer can call for a motor with a more robust insulation system. Notice that in Table 2 the cable lengths listed for 230 volt motors are significantly longer than the companion 460 volt values. This is because 230/460 volt motors use the same insulation system as 460 ratings, so their CIV level is the same when connected for 230 volts. However, the same cannot be said of all single voltage 230 volt “commodity” motors, where insulation integrity is sometimes reduced to save money. When the VFD input voltage is dropped from 460 to 230 volts, the height of the output pulses is halved. In the Figure 5 example, this would mean that the 500 ft cable length pulse would drop from 1350 volts to about 725 volts while the motor’s CIV stayed at 1200 volts. This would effectively eliminate all corona, greatly extending insulation life. Higher Altitude Requires Shorter Cable Lengths Altitudes over 3300 ft require that VFD-to-motor cable length be reduced below the values listed in

Insulation Grade Standard Insulation Premium Insulation Inverter grade Insulation

Carrier Freq. 3 kHz 6 kHz 9 kHz 12 kHz 3 kHz 6 kHz 9 kHz 12 kHz 3 kHz 6 kHz 9 kHz 12 kHz

230 V 600 ft 600 ft 600 ft 600 ft 1000 ft 1000 ft 1000 ft 1000 ft 1500 ft 1500 ft 1500 ft 1500 ft

Motor Base Voltage 460 V 125 ft 80 ft 65 ft 55 ft 225 ft 135 ft 100 ft 85 ft 875 ft 550 ft 400 ft 325 ft

575 40 25 20 15 60 40 30 25 275 175 125 100

V ft ft ft ft ft ft ft ft ft ft ft ft

Table 2. Maximum cable length for 50,000 hours of insulation life for motors at class F temperature rise

Table 2. An inverter-grade-insulated motor running from a 3kHz drive in Mexico City (~10,000 ft altitude) should have a maximum cable length of only 125 ft instead of the listed 875 ft. The same motor in Denver (~5,000 ft. altitude) could tolerate 250-350 ft of cable without reduction in insulation life. Filters Extend the Maximum Cable Length When the cable length, carrier frequency and type of motor insulation are already fixed in a given application, R-L-C filters can be installed at the VFD output to reduce the corona energy. The inductance of these filters slows the rise time of the VFD pulses and the R-C networks help to trim overshoots. The result is a voltage waveform at the motor terminals similar to that of the older GTO drives (lower voltage stress), yet with the smoother current waveform of the new technology. One filter manufacturer states that a properly designed filter will enable motor cables of 1000 ft for 460 volt standard-insulated motors. The downside is that these filters consume some power, cost money to purchase and install, and take up space in the customer’s VFD area. Simple R-C networks do not slope the waveform and work best when placed at the motor terminals, not the VFD output. Clean, Dry Windings Last Longer Contamination on the motor windings reduces the CIV level dramatically, especially when fast rise time, high frequency voltages are involved. Figure 7 shows the results when graphite and soapy water are sprayed onto the windings of a sample 5 hp motor. As CIV drops, corona power increases dramatically. A drip-proof motor that has successfully operated in a moist, sloppy pump The Pump Handbook Series

pit may fail in short order when transferred from line power to VFD power. This is because contaminants such as oil, salt, acid, alkali, grease, dirt, detergents, disinfectants, carbon black, chlorine and metal dust create conductive paths along the surface of the varnished windings, especially when combined with moisture from the surrounding environment. This facilitates high frequency surface tracking, which can effectively produce short circuits between otherwise insulated portions of the windings. The result is that the winding’s CIV level decreases and partial discharges occur from turn-to-turn, phase-tophase, or from phase-to-ground— damage is inevitable. The best defense against contamination-related failures is to keep contaminants out of the motor in the first place. Select TEFC motors instead of DP designs, which are more susceptible to contamination and moisture build-up. Don’t put DP motors in pump pits or chemical-filled rooms and expect 10 years of insulation life. In retrofit applications, people are willing to spend $25,000 on the VFD and $5,000 on process controls, but they won’t spend an extra $1,500 for a new or rewound, well insulated TEFC motor to make sure that the system lasts longer than three months. They insist on using the 20-year-old dripproof motor that’s sitting inside a wet sump. This doesn’t make sense. For applications where contamination is severe or corrosive gases are present, an explosion-proof motor may be warranted. Standard off-the-shelf explosion-proof motors cannot be applied on VFDs, however. Underwriter’s Laboratory, NEMA and CSA all require explosion-proof motors on VFD power to be specially labeled to show the type

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at base speed. Requesting class B rise on VFD power is even better, but in some cases the vendor will need to over-frame the motor to provide it.

Figure 7. As CIV drops, corona power (in red) increases dramatically.

Figure 8. When a standard magnet wire test sample is heated to class F temperature, its CIV drops considerably.

of load and frequency range. To connect a normal explosion-proof motor to a VFD voids its UL listing, and insurance coverage may be compromised in the event of an accident. Specify space heaters or trickle heaters for motors that remain deenergized for more than a couple of hours at a time. Space heaters keep the motor warm enough to prevent condensation from forming on the windings. Although pure water is non-conductive, it becomes very conductive when combined with dirt or dust on the windings. When contamination and moisture can’t be kept out of the motor, consider installing a 230 volt VFD and motor instead of 460 or 575 volt units. The 230 volt motor’s CIV level will be about 700-1000 volts when wet, which exceeds 85-90% of all the voltage spikes produced by the 230 volt VFD. Little or no corona will occur and motor insulation life will be greatly extended. Keep Motors Cool Higher winding temperatures (whether caused by overloads or elevated ambient) reduce the CIV level. Figure 8 shows that when a standard magnet wire test sample is heated to normal class F temperature (155°C) its CIV drops from 750 volts to less than 580 volts. In an actual motor this would increase the amount of destructive corona power and shorten insulation life. For pump loads, it is best practice, therefore, to specify a motor that operates at class F temperature rise (105°C) on VFD power

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All VFDs Produce Bearing Currents Aside from causing insulation stress, VFDs also stress the motor’s bearings. Stray rotor voltages are generated by both hard and softswitching VFDs, although softswitching designs generate noticeable rotor voltages only at lower fundamental frequencies. These fast rise-time voltage pulses attack the motor bearings at a rate dependent on carrier frequency and can cause permanent damage within only months of start-up. Bearing damage appears to be most severe in constant speed applications, where voltage pulses may sometimes synchronize with bearing speeds. This results in periodic electrical discharges occurring at the same point on the bearing races, blasting grooves into them. These grooves are similar in appearance to the ribs on an old-fashioned washboard; hence it is sometimes referred to as “washboarding” or the “washboard effect.” Photo 2 shows the inside of a washboarded bearing from a 20hp motor. Bearing noise and vibration are greatly increased as a result of washboarding. The washboard pattern becomes most pronounced in belted-load motors, where the rolling elements are pulled to one side by the belt forces. In advanced cases, the bearing races become so rough that the bearings overheat and fail catastrophically. In varying speed applications, the VFD-generated voltage pulses strike the bearings at many different locations, and the resultant pits are evenly distributed over the entire race. This produces more even wear, less vibration, less noise and longer bearing life than that of the constant speed washboarded-bearing motor. Some end-users define increased noise as bearing failure, even if the bearings are still serviceable. Are Pitted Bearings a Problem? It is up to the customer to decide if shortened bearing life is a significant enough problem to require correction, which can be expensive. To completely eliminate bearing damage caused by IGBT VFDs, the motor shop must insulate both bearings, not just one. This is most easiThe Pump Handbook Series

Photo 2. The inside of a washboarded bearing from a 20 hp motor.

ly accomplished by installing insulated bearings. When this is done, the bearings are protected, but the motor shaft remains “hot,” in which case the pump bearings may become damaged. An insulated motor-to-pump coupling is recommended in these cases. If a shaft grounding brush is installed on the motor, the stray rotor voltage energy is reduced by 95% effectively protecting both motor and pump bearings. Unfortunately, these brushes wear down and are subject to failure by contamination. Adding a second brush at the opposite bearing sounds like a good idea, but it can create a path for circulating currents, causing both brushes to wear prematurely.

Summary By answering the above questions the specification engineer will avoid the most common motor application problems. The goal of any pump application should be to utilize standard general purpose motors whenever possible, since they are the least expensive and most widely stocked variety of motor. A definite purpose (VFD-duty by design) motor is justified when the general purpose machine will stall, overheat, fail prematurely due to voltage stress or be installed in a hazardous location. That’s when it’s time to get the motor engineer on the phone. " Jerry Muehlbauer is a Senior Application Engineer in the Inverter duty motor group at Marathon Electric in Wausau, Wisconsin.

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R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

Understanding and Using Shaft-to-Shaft Alignment Measurement Systems Tried and true procedures and advice, plus some tricks of the trade only an expert can provide. By John Piotrowski, Turvac, Inc.

his article explains how to use the three most popular basic shaft alignment measurement methods and how laser alignment systems work. Advantages and disadvantages of each of the methods and of laser systems are detailed.

T

Why Do Alignment? Anyone who has successfully installed or maintained rotating machinery has encountered the critical task of alignment. Done near the completion of the installation, it can affect the life of a drive system and dictate how well that equipment will operate over time. Shaft misalignment has caused a tremendous amount of unnecessary damage to rotating machinery. In spite of all the recent advancements in measurement tooling, aligning equipment continues to confuse and confound even the most talented and intelligent people. A recently completed study at the University of Tennessee shows that even with only small amounts of misalignment present, there is a significant increase in the forces on machinery bearings. Because bearing life is exponentially related to force, rapid equipment degradation occurs right before our eyes. Shaft misalignment disguises itself very well, and the severity of the problem is difficult to detect while the equipment is running. By the time

the symptoms develop, it’s too late; much of the damage has already been done.

Why Is Misalignment Still a Problem? Surprisingly, even the basic alignment measurement methods are not taught in the majority of high schools, vocational schools, junior colleges, or colleges and universities. Lack of knowledge only partially accounts for the poor alignment I and others have seen throughout all of industry. Several other factors are also at work: absence of all the proper tooling required to do the task; insufficient time granted to complete alignment jobs satisfactorily; and, frequently, lack of enthusiasm and ambition in the personnel doing the work. The classic mistake is believing that purchasing an expensive alignment measurement system will automatically solve your alignment problems. However, there can be no substitute for proper training, and it’s not just the trades personnel that need this vital information. Foremen, technicians, engineers, managers, and even production supervisors need to understand why accurate alignment is important, how long it will take to align machinery properly, and why they can’t rely on knowing just one alignment measurement method. Most people don’t even underThe Pump Handbook Series

stand how bad their alignment problems are. Conservatively, more than half of all the equipment operating today exceeds 4 mils/inch of misalignment when running. Figure 1 shows actual field data from machinery measured over one month. Keep in mind that acceptable misalignment is around 1 mil/inch—the first ‘tick’ mark on the Y-axis. Disappointingly, the majority of plant sites cannot produce alignment records for every piece of rotating machinery. Even in facilities where a good preventive/predictive or condition-based maintenance program exists, there is typically 100 times more data collected on vibra-

Figure 1. Actual alignment data taken during one month

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tion, temperature, oil analysis, and motor current than on shaft alignment. Most people who measure vibration and other data have received incorrect technical information about what symptoms are exhibited by misaligned machinery. This frequently results in an incorrect analysis of the problem. These same people are usually not the ones doing the alignment work. So communication gets scrambled between the people identifying the problem, those assigning the work, and those actually doing the work. Mechanics have told me countless times that they were reprimanded for sloppy workmanship because vibration levels on a machine stayed the same or increased after they completed a realignment. Surprisingly to many, this is quite normal.

What Do We Have to Be Concerned With? Shaft alignment is a three-dimensional problem. Not only do you have to align the shafts in the up and down direction (vertical), but you must also align them in the side to side (lateral) and end to end (axial) directions. More equipment is found to be misaligned laterally and axially. The shaft alignment measurement methods discussed here can be used to determine the vertical and lateral positions of machinery shafts. These methods must be done with the equipment off-line and not running—at least that’s what I recommend. However, some rotating machinery drive shafts will move to some other position when operating. This is commonly referred to as “thermal movement” or “hot and cold” alignment. Believe it or not, for certain types of machines you purposely misalign the shafts when the equipment is off-line so that they will move into a collinear axis of rotation when operating under normal conditions. This machinery movement is rarely measured and can cause damage over time. What percentage of rotating machinery in your plant has been measured for this movement?

Before You Start Don’t line up junk. If a drive system has been run under misalignment conditions and has sustained damage to its seals, bearings, cou-

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pling, shafts, and internal components, no amount of re-alignment is going to correct these problems. Becoming conscientious only after damage has been detected will not undo prior negligence. One of the items on your alignment list should be checking and ensuring that your machines don’t have these problems prior to starting an alignment job. Aside from ensuring that you are aligning good equipment, there are several other preliminary steps you should do before you attempt to align machinery. The three most overlooked are: • measuring and correcting excessive amounts of shaft and/or coupling hub runout • finding and correcting excessive “soft foot” problems • checking for and eliminating excessive piping, ductwork, or conduit stress on rotating machinery

the other one was 2.5” in diameter. Most people forget that the majority of machinery runs 24 hours a day, 7 days a week for months or years without stopping. If your automobile ran at 3600 rpm (about 80 mph for most vehicles) without stopping, in one year your vehicle would have traveled 700,800 miles. That’s more than 28 times around the earth, without an oil change. I wonder what the tires would look like if the front end was out of alignment. Using a straightedge during the alignment process is fine for rough alignment purposes. Since some of the measuring devices described later in this article have range limitations, a straightedge and an eyeball are good enough tools to get the shafts close enough to begin using more accurate measurement tools to achieve acceptable alignment tolerances.

This article is principally aimed at showing you a number of shaft alignment methods. However, if you plan on aligning machinery or have been doing alignments and don’t understand or don’t check for the above mentioned items, I guarantee that you will eventually have considerable trouble either aligning equipment or operating the equipment after you’ve completed your alignment job. Refer to the books and technical papers at the end of this article for more information on these frequently overlooked preliminary steps.

Measuring Misalignment with Dial Indicators

Aligning Machinery with a Straightedge Human eyesight is fairly accurate. Unaided, most people can see a 3040 mil (0.030-0.040”) misalignment over a 3-4” separation between two machinery shafts. This isn’t too bad, but this small amount of misalignment can cause several coupling designs to run hot and can cause excessive loads on the shafts and bearings. The result could be premature failure of the equipment. You might ask, “What if the shafts are 20 inches apart, or what if the surfaces you’re measuring on each shaft are not the same distance from their centerlines of rotation?” I wouldn’t recommend it, though I’ve seen someone use a straightedge where one shaft was 2.25” in diameter and The Pump Handbook Series

John Logan invented the dial indicator in 1883 while searching for a more accurate measurement tool to use in his watchmaking business. In 1946, Joseph Christian finally filed a patent describing the use of dial indicators for shaft aligment purposes. Because there is a wide variety of rotating machinery, several different methods had to be invented to adequately align all of the equipment in plants. There simply is no one tool or alignment method that will work everywhere. I have found that less than 10% of the people I have worked with in industry know how to do the three basic alignment methods. To be considered an alignment expert, one must know how to perform several different methods and know which is best suited for the task at hand. Each technique has its own advantages and disadvantages. Make sure to study not only the method but also its limitations. If you can master the three dial indicator methods discussed here, you will probably be able to align 98% of the rotating machinery in existence. I strongly recommend you take a dial indicator to every alignment job. Not only can it be used to measure a misalignment condition, but it also serves as the only reliable tool to

Figure 2. The face and rim shaft alignment measurement method

ensure that another problem does not exist—excessive runout. It is important to understand that shaft alignment is the process of determining the relative positions of the centerlines of rotation of two independently supported shafts. You cannot assume that every point along the length of the shaft is concentric with its centerline of rotation. In other words, you have to be sure that you are not aligning bent shafts or shafts that have coupling hubs that are bored off-center, overbored, or skewed on the shaft.

The Face and Rim Method It is not clear who really invented this method. Robert Voss was granted a U.S. patent in 1949 describing this method, but he probably was not the first person to use this technique for shaft alignment purposes. Figure 2 shows how this method is typically set up across two shafts. You do not have to set up two indicators at the same time (Figure 2). There is nothing wrong with setting up just one indicator, capturing

Figure 3. The reverse indicator shaft alignment measurement method

either the rim or peripheral measurements and then using the same indicator to capture the face measurements. Once the bracket is attached to either shaft and the indicator is placed at the 12 o’clock position, slowly rotate the shafts through 90-degree arcs. Observe the indicator carefully as you traverse through each quadrant (Photos 1-5). Many mistakes are made on these steps, so be careful during this process. When you get to each stopping point (at the 3, 6, and 9 o’clock positions) record your measurements. Advantages of the Face and Rim Method • It is a good technique for situations where one of the machinery shafts cannot be rotated, or it would be difficult to rotate one of the shafts. • The rim (or perimeter) dial indicator shows centerline offset or parallel misalignment and the face indicator indicates if any angular misalignment exists. This enables the user to isolate the two misalignThe Pump Handbook Series

ment conditions so that each problem can be attacked separately. • It is a good method to use when the face readings can be captured on a fairly large diameter (8” or larger). This method begins to approach the accuracy of the reverse indicator technique when the diameters of the face readings are equal to or exceed the span from the bracket location to the point where the rim indicator readings are being captured. Disadvantages of the Face and Rim Method • It is not as accurate as the reverse indicator method if both shafts can be rotated, particularly if the face measurements are being taken on diameters less than 8”. • If the machinery shaft(s) are supported in sliding (plain/sleeve) bearings, it is very easy to axially “float” the shafts toward or away from each other when rotating them. This will result in bad or inaccurate face readings. • Rim and face bracket sag must be measured and compensated for.

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The Reverse Indicator Method

Photo 1. Zero the indicator at the 12 o’clock position.

Photo 2. Rotate both shafts 90 degrees to one side. Observe the movement of the indicator needle carefully as you rotate. Record your measurement and the side you are on (N, S, E, W).

Photo 3. Rotate both shafts 90 degrees again to the bottom. Observe the movement of the indicator as you rotate. Record your measurement.

In this method, you are effectively capturing two sets of rim/peripheral measurements. The reverse indicator method can be used on 60-70% of rotating machinery and is currently the preferred dial indicator method for measuring rotating machinery shafts. It is best suited for use when the distances between measuring points on each shaft range from 3-30”. You do not have to set up two brackets and two indicators at the same time (Figure 3). Set-up one bracket and indicator, capture Figure 4. The shaft to coupling spool alignment measurement method the rim measureA lot of mistakes are made here so be ments from one shaft to another, careful during this process. (Sorry for then remove the bracket from the the repetition, but this is very imporone shaft, turn it around, clamp it to tant!) When you get to each stopping the shaft you just took measurements point at the 3, 6, and 9 o’clock posion, and sweep another set of readtions, record your measurements. ings on the shaft you were just clamped to. Once the bracket is Advantages of the Reverse attached to either shaft and the indiIndicator Method cator is placed at the 12 o’clock posi• Typically, this method is more tion, slowly rotate the shafts through accurate than the face-rim method 90 degree arcs shown in Photos 1-5.

Photo 4. Rotate both shafts another 90 degrees to the other side. Observe the movement of the indicator needle as you rotate. Record your measurement and the side you are on (N, S, E, W).

Photo 5. Rotate back to the 12 o’clock position to ensure that the indicator returns to zero.

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Figure 5. The electromagnetic spectrum The Pump Handbook Series

• If the machinery is supported in sliding type bearings and the shafts are “floating” back or forth axially when rotating the shaft, there is virtually no effect on the accuracy of the measurements being taken. • This method can be used with the flexible coupling in place. (Be careful, though! Refer to the laser experiment later in the article.).

Figure 6. Basic operation of semiconductor diode lasers

because the distance from the mounting point of the bracket to the point where the indicators capture the readings on the shafts is usually greater than the distance at which a face reading can be taken. Also, when capturing a rim reading, the dial indicator is measuring twice the amount that the centerlines are actually offset, making the indicator more sensitive to misalignment.

Figure 7. Basic operation of semiconductor photodiode detector “targets.” The laser beam produces a differential voltage as it changes its position on the target surface.

Disadvantages of the Reverse Indicator Method • Both shafts must be rotated. • It is difficult to visualize the positions of the shafts from the dial indicator readings you collect. • Bracket sag must be measured and compensated for.

The Shaft to Coupling Spool Method Sometimes there is a considerable distance between two pieces of machinery and the two previously described methods won’t work very well. A classic example of this can be found in cooling tower fan drive systems where there is often a separation of 12–20 ft. between the motor and the right-angled gearbox driving the fan assembly. A coupling spool (sometimes referred to as a jackshaft) usually connects the shaft ends together. I use this technique when the shafts are more than 30” apart. Figure 4 shows the typical arrangement for this method. Once again, there is nothing wrong with setting up just one bracket and indicator, capturing the rim measurements from one shaft out to the spool piece, then removing the bracket from that shaft, turning it around, clamping it onto the other shaft, and sweeping another set of readings from this shaft out to a different point along the length of the spool piece. At this point, follow the same procedures listed for the other dial indicator methods. Advantages

of

the

Shaft

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Coupling Spool Method • This is perhaps the most accurate measurement technique where there are extreme distances between shaft ends. • It is a relatively easy method to set up and capture readings. Disadvantages of the Shaft to Coupling Spool Method • Since the coupling spool (jackshaft) must be kept in place, both shafts must be rotated together. • Bracket sag must be measured and compensated for.

Measuring Misalignment with Laser Alignment Systems Figure 5 shows a diagram of the electromagnetic spectrum. The human eye is only capable of detecting a very narrow band of wavelengths across this broad spectral range. In 1954, Professor Charles Townes at Columbia University experimented with having generated radio/microwaves stimulate the emission of energy that was stored in ammonia molecules in a container. His device was named a “maser,” an acronym for Microwave Amplification by Stimulated Emission of Radiation. One of his graduate students, Gordon Gould, continued these experiments and is credited with coining the term “laser” (Light Amplification by Stimulated Emission of Radiation). The lasers used in shaft alignment measurement systems today are low power semiconductor light emitting diodes (AKA LED’s). Figure 6 shows the basic operation of a semiconductor laser. In a sense, lasers are miniature, focusable flashlights. The beam that exits the LED is actually about 1.5 mm x 1.5 mm in diameter, not an infinitely thin strand of light. If you shine it on a wall, you will see a small, elliptical dot of light. To a certain extent, the accuracy of these systems has very little to do with the laser. The most important component of the system is the photodiode target or detector. Figure 7 shows what typical photodiode targets look like. Laser detectors are semiconductor photodiodes capable of detecting electromagnetic radiation (light) from 350 to 1100 nm. When light strikes the surface of the photodiode, an electrical current is

to

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produced. Most manufacturers of laser-detector shaft alignment systems use 10mm x 10mm detectors (approximately 3⁄8in2). A few might use 20mm x 20mm detectors. Some manufacturers of these systems use bi-cell (uni-directional) or quadrant cell (bi-directional) photodiodes to detect the position of the laser beam. When light strikes the center of the detector, output currents from each cell are equal. When the beam moves across the surface of the photodiode, a current imbalance occurs, indicating the offcenter position of the beam center. In effect, the detector replaces the dial indicator since it is capable of measuring distance by how far the beam center moves across the surface of the target. Today there are several different companies manufacturing laser alignment systems. The ways they use the lasers and detectors to determine misalignment between two centerlines of rotation varies some-

Basic Laser-Detector Shaft Alignment Measurement Methods

Figure 8. Three different measurement principles of laser/detector alignment systems

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what. Figure 8 illustrates the three basic methods used.

Dial Indicator Systems: Advantages and Disadvantages Dial indicator based shaft alignment measurement systems have been around since the 1930s, although patents didn’t start appearing until the mid 1940s. These measurement tools and fixtures are rugged and many are Figure 9. Results of a laser alignment system test where 50 mils of shim stock were added under all four feet of a motor after the unit was peradaptable to meet fectly aligned the needs of virtugives him or her a better underally every alignment task. In many standing of the process rather cases, alignment brackets and fixthan trying to rely on a tool that tures are custom made to fit a par“does the thinking for you.” ticular machine or a general purpose bracket can be made from pieces of Advantages Of Dial Indicator angle iron or machined out of steel. Measurement Systems Several companies manufacture • They are relatively inexpensive. general-purpose alignment bracket • Some systems can perform all five systems. The prices for these systems of the dial indicator measurement range from $600 to $6,000. If you are techniques. (There are two more unable to manufacture your own besides the ones shown here.) shaft alignment bracket, it is recom• Accuracies of +/-1 mil mended that you research all of the • A dial indicator can perform available tools before purchasing a many of the preliminary steps (i.e. system. This ensures that you know runout measurement). what their capabilities and their lim• Some systems allow for the couitations are. You’re much more likely, pling to be disconnected when then, to select the right tool for your capturing readings. situation. Even if you have a laser Disadvantages of Dial Indicator system, it is recommended that you Measurement Systems also have a dial indicator system for • Bracket sag must be measured the following reasons: and compensated for. • If your laser system breaks during • User must know how to read a an alignment job, the dial indicadial indicator. tor system can complete the task. • User must know how to graph/ • If your laser system is indicating model the shaft positions from the an alignment solution that doesn’t readings or be able to calculate make sense (e.g. move the outrequired machinery moves. Some board end of a motor 895 mils sidesystems include calculator or ways), use a dial indicator system computer software. to compare the results. Both methods should yield the same result. If Laser Systems: they do not, one of the tools is Advantages and Disadvantages lying. Before you make a silly Laser alignment systems are very move, better find out which one is popular and have been around for telling the truth and fix the other. more than 15 years, so some of the • A dial indicator system forces the initial “bugs” have been worked out. user to understand how shaft The prices for these systems range alignment is accomplished. This The Pump Handbook Series

from $6,000 to $60,000. Again, do your research before buying a system. It’s a large investment and not to be taken lightly. Advantages Of Laser Measurement Systems • Bracket sag does not occur with laser beams, but the user must ensure that the brackets that hold the lasers and detectors are firmly attached to the shafts. • Accuracies of +/-3 microns or better • Some systems enable the coupling to be disconnected when capturing readings. • Most systems include an operator keypad/display module that prompts the user through the measurement steps and calculates the moves for one of the machines. A few systems can actually provide a variety of movement solutions. Disadvantages Of Laser Measurement Systems • Relatively expensive. If you drop it and break it, you cannot use any laser or detector as a replacement. Special equipment is needed to check calibration and accuracy. • Range of measurement somewhat limited, since most manufacturers use 10 mm x 10 mm detectors. Some systems have a range extension feature.

• The system is incapable of measuring runout conditions. (It is possible to align bent shafts and not know it.) • Most systems only determine correction moves for one of the two machines (the “Stationary-Movable” concept). This tends to lock users into one way of thinking. Pumps are always stationary, because they are connected to piping. Motors are always moveable. But think about this. If you are aligning a steam turbine and a pump, which machine is movable and which one is stationary? • Most systems require that the coupling is bolted in place when capturing readings. With moderate to severe misalignment present, incorrect moves will be calculated (Figure 9). • With the great majority of systems, both shafts have to be rotated. • Sometimes it is difficult to capture readings in bright sunlight or well-lit areas. • Accuracy reduced in presence of excessive steam or heat

Conclusions If you’re serious about doing shaft alignment correctly, you should consider this article only a starting point in your educational quest. Its pur-

The Pump Handbook Series

pose is to introduce you to a variety of methods and tooling that can be employed to align rotating machinery shafts. It is not meant as a complete reference guide to the overall scope of machinery alignment and all of the tasks required to accomplish this deceptively simple job. Today, with all of the information and the wide variety of alignment measurement tools available, it is bewildering that misalignment is still a major problem in industry. I am hopeful that this article will start you on your way to proper alignment. "

References 1. Piotrowski, John D., Shaft Alignment Handbook – Second Edition, Marcel Dekker Inc., New York, NY, ISBN# 0-8247-9666-7. 2. Piotrowski, John D., “Why Shaft Misalignment Continues to Befuddle and Undermine Even the Best CBM and Pro-Active Maintenance Programs”, P/PM Technology, Vol. 10, Issue 2, April, 1997. John Piotrowski has been working with rotating machinery for more than 23 years and is considered one of the leading authorities on alignment. He has written more than 30 technical articles and several software programs.

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R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

OEM Pump Materials and Their Relationship to Quality and Reliability Manufacturing processes that exceed ASTM specifications can help improve pump hydraulics. By Colin McCaul, Ingersoll-Rand

umps are an indispensable tool of modern industry. A critical component of many industrial processes, they must operate with a high degree of safety and reliability. The implications of a pump failure often go well beyond simply replacing the pump. Pump failure may shut down a manufacturing process, resulting in lost production and revenue far exceeding the value of the pump. The failure of certain pumps in large electric power generating stations can result in losses exceeding $500,000 per day associated with lost power production. Another aspect of this situation is the desire to reduce maintenance costs and extend the time between planned shutdowns for maintenance or overhaul. Plant operators emphasize the need for more durable equipment that can be operated over a wider range of conditions for longer periods of time. Pump manufacturers, acutely aware of the need

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for ever increasing reliability, respond by putting much effort into design and materials improvements, all in pursuit of this goal. This article will focus on the materials aspect of the problem. Pump manufacturers, by the use of improved materials and processing techniques, can now supply equipment that is more durable and better able to handle difficult pumping applications than ten or 20 years ago. There has been much work as well in what could broadly be called surface modification. This includes a wide variety of coatings and surface treatments, usually designed to produce harder or more wear resistant surfaces. The materials, surface treatments and processing techniques used in the pump industry are often specially tailored for the application. They may differ in important respects from the more generic versions with which many pump users are familThe Pump Handbook Series

iar. Additional technical requirements or special processing techniques are frequently critical to pump performance and reliability. They often represent the critical difference between an OEM component and a replicated part obtained from some other source. The additional materials requirements in the OEM component are not without cost. However, as will be shown by several examples, the added cost is money well spent considering the enhanced reliability it provides and the potential consequences of failure.

ASTM Specifications for Materials In the domestic pump industry, customers frequently specify pump materials using American Society for Testing and Materials (ASTM) specifications. Few pump users have a thorough understanding of these specifications, how they are developed, and their limitations. Pump

desire of producers for standards that are easily met at lowest cost.

Is ASTM Strict Enough?

Photo 1. CN7M impeller experienced intergranular corrosion in seawater application because insufficient cooling rate resulted in precipitation of chrome carbides in grain boundaries.

Photo 2. Impeller has sustained severe intergranular attack.

OEMs, and many other manufacturers as well, often purchase materials meeting their internal specifications. These may be based on ASTM specifications, but they typically include additional technical requirements. This system has evolved because long experience has shown that ASTM material specifications are often not adequate for direct application in pumps. Why are ASTM specifications inadequate? Because they are written by committee. Each committee should consist of members representing producers, users and others. In practice, the committees handling materials for pumps, such as ferrous castings, are either dominated by producers or so structured that producers can dictate the content of those material specifications within the jurisdiction of the committee. The result is that ASTM specifications often have minimal technical requirements, in keeping with the

Those unfamiliar with ASTM material specifications may find it surprising to learn just how loose, or vague, these specifications can be with respect to technical requirements that clearly affect pump component integrity. A particularly good example involves several ASTM specifications that cover austenitic and duplex stainless steel castings, including A351, A743, A744 and A890. These specifications cover stainless steel castings and are frequently specified by pump users for casings, impellers and various other stainless steel cast components. Among the more commonly used pump materials covered by several of these specifications are the austenitic grades CF8 and CF-8M. Both, but particularly the CF-8M grade, are used extensively in the pump industry. These alloys are given a solution heat treatment and water quench to eliminate chrome carbides and produce a microstructure with optimum corrosion resistance. Water quenching from the solution treatment temperature is usually necessary to ensure that the casting will have adequate corrosion resistance. However, water quenching is also a large thermal shock and could occasionally cause thermal cracks in a casting, particularly in a complex geometry such as an impeller. Some foundries prefer a less drastic cooling rate in the belief that it will minimize the possibility of cracking. The heat treatment specified for austenitic stainless steel castings in the ASTM specifications requires that they be “quenched in water or rapidly cooled by other means.” Unpleasant experience has shown that the phrase “rapidly cooled by other means” is vague and can be interpreted to justify any method of cooling, or cooling rate, that one wishes to use. This loophole in the ASTM specification has led to corrosion failures in pumps. The Pump Handbook Series

Failures of ASTM Materials The CN-7M impeller shown in Photo 1 had been used in a seawater application. The leading edge of each vane corroded within the first several months of service. Liquid penetrant examination of the impeller showed several areas where penetrant slowly seeped from the casting surface, indicative of a porous structure. The impeller was found to be sensitized, and the corrosion had been intergranular. The material had been certified to the ASTM A743 specification, and it was determined that the casting had been air cooled. The foundry argued that this produced a rapid cooling rate and met the requirements of the specification. Clearly, in this instance, the corrosion resistance of the material was compromised by the failure to quench the casting in water or by some alternative means that would produce the same cooling rate. The impeller in Photo 2 had been used in acid injection service. Details pertaining to the specific environment are proprietary. However, impellers of the same chemistry, CN7MS composition—similar to alloy 20—had been used previously with good success. Visual examination showed the attack to be intergranular in nature with widespread grain dropping. The microstructure shows a semicontinuous network surrounded by fine precipitated carbides. In this instance the foundry insisted that the impeller had been solution annealed and water quenched. Following an audit of the foundry heat treat practice, it was concluded that intergranular carbide precipitation resulted from an appreciable delay between the opening of the furnace doors and the quenching of the impeller.

Lessons Learned from Failures These failures teach two lessons. The heat treat requirements of the applicable ASTM specifications are inadequate and must be supplement-

387

Photo 3. Ferrite phase in duplex stainless steel casting has been almost entirely replaced by sigma phase due to incorrect heat treatment.

ed to assure castings with optimum corrosion resistance. Also, foundry selection is critical because less technically competent foundries may not have a thorough understanding of specification requirements. The larger pump OEMs procure castings from a select group of audited foundries—those that have demonstrated the capability of producing high quality castings fully meeting the requirements of the material specification. Such attention to quality and detailed technical requirements often differentiates OEM pump castings from those available elsewhere in the market. Insufficient Cooling Rate The problem with an insufficient cooling rate is not limited to austenitic stainless steels. Duplex stainless steel castings, particularly the higher alloyed grades, also require a very rapid cooling rate in order to avoid the formation of sigma phase. This is an undesirable constituent that forms in the microstructure when the cooling rate is too slow. Sigma phase embrittles the material and greatly reduces corrosion resistance. Unfortunately, the ASTM A890 specification, which covers duplex stainless steels, does not specify a quenching method or a minimum cooling rate. This has resulted in the production of castings, particularly from inexperienced foundries that are unfamiliar with this class of materials, in which the ferrite has been entirely

388

replaced by sigma, as shown in Photo 3. This microstructure represents a pump casting sourced from a foundry in the Far East that had no prior experience with duplex stainless steels. The reason for this poor microstructure is a failure to appreciate the importance of very rapid cooling in this class of materials and inadequate requirements in the ASTM specification. This is a more serious problem with large, massive castings, which are more difficult to cool rapidly. Also, the higher alloy grades of duplex stainless steel tend to form sigma phase during cooling more quickly than the lower alloyed grades. Castings with extensive sigma phase, such as in Photo 3, will have very poor ductility and be difficult to machine. They will be less resistant to corrosion as well. ASTM specifications do not adequately address the heat treatment of either austenitic or duplex stainless steel castings. Pump OEMs address this deficiency by using detailed material specifications that have more stringent requirements concerning the method of cooling austenitic and duplex stainless steel castings. By requiring that these castings be water quenched, the OEM is providing a higher quality casting with consistently better corrosion resistance, compared with the supplier that meets only the ASTM specification. ASTM specification A494 covers nickel alloy castings, several of which are frequently specified for use in pumps. Nickel alloy pumps are used in the chemical industry to handle a variety of corrosive chemicals. There has been a long history of inconsistent performance with nickel alloy castings. Some will provide much better corrosion resistance than others of nominally the same grade. One possible reason for this variability can be traced to the manner in which chemical analysis is handled by the A494 specification. This specification, which distinguishes between a master heat and individual melts, has unusual chemical requirements that The Pump Handbook Series

almost certainly account for differing corrosion resistance in castings. A master heat is defined as a single furnace charge of refined alloy that can be poured directly into castings or into remelt alloy for individual melts. The foundry is obligated to perform a chemical analysis of each master heat and report these results to the customer. This means that the customer is given an analysis of an ingot that is subsequently remelted and poured into castings. The final chemical analysis of the casting could, and often does, differ significantly from the master heat analysis. This can occur for several reasons. The melt may pick up carbon or other contaminants from the furnace. Also, the percentages of various elements can change as some are vaporized from the surface of the molten metal. The corrosion resistance of the casting is directly related to its chemistry. So it is apparent why there has been wide variation in the corrosion resistance of castings of a given alloy in a single environment. Bad Chemistry Few customers understand that often the chemistry reported for a casting made to the requirements of ASTM A494 is not that for the actual casting supplied. Rather, it is that for an ingot that was subsequently remelted. This approach implies that castings will always be within chemistry and there will be no rejects for out-of-specification chemistry. The benefit to the foundry is obvious. This same approach is often used with investment castings, and is permitted by some ASTM specifications that cover investment castings. This lax approach to chemistry is a major reason why the corrosion performance of some high alloy castings shows such variation from one heat to another. Recognition of this problem has led to efforts to develop new, more stringent ASTM specifications for high alloy castings. This painfully slow process can be effectively stymied by a small group of material producers.

Photo 5. Fatigue failure of shaft begins with superficial scratches in the keyway that were made when the key was inserted.

Some of the larger pump OEMs use detailed material specifications that require chemical analysis be performed for every heat of castings. This assures the customer that the reported chemistry in fact corresponds with that of the casting. It also results in more uniform and predictable corrosion resistance. There may be some additional cost for the castings made to the more stringent requirements imposed by a pump OEM. However, this expense is insignificant compared to the potential consequences of a single unexpected corrosion-caused pump failure in a critical application. Heat Treatment Another aspect of materials specifications that can affect the performance and reliability of pump components concerns heat treat and other processing steps. Pump shafts are frequently manufactured from stainless steel bar. The martensitic grades of stainless steel, especially type 410, are frequently preferred for their high strength and reasonably good corrosion resistance. These

The Pump Handbook Series

in Figure 1 for 410 stainless steel, which exhibits very large variation in toughness with relatively minor variation in heat treat properties. Charpy impact properties are not a requirement for pump materials that will be operated at ambient or higher temperatures. Nevertheless, consideration of maintenance practices employed during outages reveals that shafting may experience severe loading at much lower temperatures. Charpy testing is a good indicator of material embrittled by heat treatment, a condition that is not easily detected in any other way. Type 410 stainless steel must be tempered at a minimum temperature of 1100°F to avoid potential embrittlement and susceptibility to cracking. Pump shafts that meet only the ASTM A276 specification may have been tempered at a temperature lower than 1100°F. These shafts would fully meet the ASTM specification, but it might have very poor Charpy impact values, and therefore be susceptible to cracking at tiny flaws. The more stringent material specification used by a major pump OEM is based on the ASTM specification but clarifies that tempering be performed at 1100°F minimum, thereby assuring that the material will have good toughness and optimum resistance to fatigue cracking. This clarification regarding heat treatment is crucial to the quality and reliability of a 410 stainless steel pump shaft. Yet it is an item that most customers do not recognize,

RAVENSWOOD CHARPY IMPACT TESTS 120 110 100 90 80

FOOT POUNDS

Photo 4. Boiler feed pump shaft which failed in fatigue after less than three months service.

materials are air hardenable and usually heat treated using a two-step harden and temper operation to obtain toughness and good mechanical properties. Type 410 stainless steel bar is covered by two ASTM specifications. The less stringent of the two is ASTM A276. It lists a Condition H, which is defined as “hardened and tempered at a relatively low temperature” and a Condition T, defined as “hardened and tempered at a relatively high temperature.” There is no additional information clarifying the terms “high temperature” and “low temperature.” So a 410 stainless steel bar can be tempered at any temperature selected by the producer, as long as the material meets the specified mechanical properties. Type 410 pump shafts are used in high pressure multistage pumps in such critical applications as boiler feed service. There have been fatigue failures of shafts in these applications, often destroying the pump in the process. The fractured shaft shown in Photo 4 is typical of this class of failures. Failure occurred after only 2.5 months service in a five stage boiler feed pump operating at 3580 rpm, 230°F and 1100 psi discharge pressure. Metallurgical analysis determined that the shaft met the requirements of the applicable material specification, A276, and had no significant flaws or defects. It was fabricated to drawing requirements and apparently not subject to misuse in service. Failure began in the keyway where there were superficial scratches resulting from insertion of the key (Photo 5). The root cause of the failure was found to be improper heat treatment, which embrittled the material, making it highly susceptible to crack initiation and propagation. This was determined by Charpy impact testing, which showed the material to have absorbed energy of 6 foot pounds, lateral expansion of 2 mils, and zero percent shear at 70°F. These results are very low, but within the wide range possible, as shown

70 60 50 40 30 20 10 0 0

32

70

50

100

TEMPERATURE DEGREES F A

C

D

G

H

Figure 1. Variation in impact characteristics of 410 stainless steel with minor difference in tempering temperature

389

Photo 6. Large crack in superduplex stainless steel forged block, associated with the presence of sigma phase in grain boundaries.

and would be overlooked by some smaller manufacturers of pump shafts and other components, who might not have the in-depth materials expertise required to appreciate this deficiency in the ASTM bar stock specification and its implications for pump performance.

Processing The foregoing examples involve relatively straightforward instances where some deficiency in the ASTM material specification, when recognized and corrected, can provide a superior, more reliable pump component. There are other instances in which the pump OEM uses special processing to improve the product. In many cases, processing know-how is the crucial difference between a reliable component with optimum properties and an inferior component that is likely to have a shorter life and may fail suddenly at the most inopportune time. Forging One example involves duplex stainless steel forgings, which are covered by ASTM A182. As noted, this class of materials requires a very rapid quench from the heat treat tem-

390

perature to avoid the formation of sigma phase. Duplex stainless steels are gradually becoming the preferred material for a number of pumping applications because they have desireable mechanical properties and excellent corrosion resistance. A typical application might be reciprocating pumps that handle corrosive brines in oil recovery. Forgings for this service often have relatively large section thicknesses. Producing these forged blocks without the deleterious sigma phase requires special processing technology. Conventional heat treatment methods, in accordance with the A182 specification, will almost certainly yield inferior microstructure, and the material will be susceptible to either corrosion or cracking. Photo 6 shows a large crack that developed in a large superduplex forged block. Superduplex stainless steels, the highest grade duplex alloys, usually consist of a minimum of 25% chrome, 5% nickel and 3% molybdenum. Nitrogen is also present, and there may be small amounts of copper or tungsten. The superduplex grades are susceptible to a type of spontaneous cracking known as “clinking.” This is caused The Pump Handbook Series

by sigma phase in the microstructure. The photomicrograph in Photo 7 shows the sigma phase at the grain boundaries in this forging. The ASTM A182 specification does not address microstructure, or the role of sigma phase, on material performance. OEMs with special know-how can make large forgings that are free of sigma phase and have an optimized microstructure. Pump users should recognize that material produced to ASTM forging specifications only will probably contain deleterious sigma phase. Therefore, it will not provide required corrosion resistance, and it is likely to crack in service. The consequences of a premature failure extend well beyond the cost of simply replacing the pump. Lost production could easily increase the total cause of the failure by a factor of ten. The special processing techniques developed by the pump OEM are the key difference that makes production of the higher alloy duplex stainless steel forgings possible with the optimum mechanical properties and corrosion resistance. The higher cost associated with special processing is money well spent when one considers the large difference in material quality that results. Special Techniques for Shafts Pump shafts are a component where special processing technology can represent the critical difference between reliable long term service and unexpected problems early in the life of the pump. Many pump shafts are made from stainless

Photo 7. Micrograph showing that fracture path follows grain boundary sigma phase.

steels, with the martensitic stainless steel grade 410 being a common choice. These shafts are made to exacting straightness tolerances. Commercial bar tolerance for runout is 0.125 inches in five feet. A typical runout specification of 0.090 inches in a five-foot length of bar is used to assure that the manufacturing cycle begins with a reasonably straight bar. Nevertheless, maintaining the required straightness tolerances can be difficult, especially for the long thin shafts used in multistage pumps. There is also some risk, with most conventional manufacturing techniques, that the shaft will flex or distort after machining has been completed - possibly after a pump has been placed in service. Distortion of a pump shaft can be a serious problem because it closes the gap between wear rings and other components, leading to galling and possible seizure of the pump. Shaft straightness can change during or after the machining process as a result of several factors. The microstructure of heat treated 410 stainless steel will consist primarily of tempered martensite. However, a small amount of retained austenite may also be present. This can change into martensite during machining. The phase transformation will be accompanied by a change in volume that can cause distortion. Residual stresses, introduced at any of several manufacturing stages, can also be present in bar stock. Stress can be introduced during hot rolling, during heat treatment (in which liquid quenching is often used to obtain specific mechanical properties in heavier cross sections), or during mill straightening, which brings the material into needed tolerances. Thus, the stability of a long thin shaft is often unpredictable. Manufacture of such shafts using conventional techniques is an art, and it may require several iterations of thermal processing, mechanical straightening and finish machining.

Despite best efforts to produce a straight bar free of residual stress, some shafts will not be entirely free of residual stress and will distort during operation, resulting in pump performance problems. Pump OEMs have developed manufacturing techniques that are more likely to produce a straight, stress free bar than other manufacturers who may have less expertise in dealing with this problem. One pump OEM, collaborating with a bar manufacturer, has developed proprietary new heat treat process that produces a remarkably straight 410 bar free of residual stress. The process employs tunnel furnaces with an intermediate water quench. The bar is continuously supported with rollers that rotate and propel it during processing to assure uniform heating and elimination of residual stress. The bar is initially processed in a high temperature austenitizing tunnel where any prior residual stress is removed. The bar is supported along its entire length as it passes through this furnace to assist in removing residual stress. The bar then proceeds through a water quench ring, where it is rapidly quenched to maximum hardness and strength. A final furnace tempers the martensitic stainless steel bar to the desired mechanical properties.

This process, referred to by the trade name Superstraight, can produce long slender pump shafts that are remarkably stable and free of residual stress. These can be manufactured without the intermediate stress relief that would otherwise be necessary between the rough and final machining operations. In some instances, bars manufactured using conventional techniques also required additional straightening and stress relief during final grinding operations. These steps have also been eliminated. The new rotary heat treatment has reduced both setup and manufacturing time while providing a superior final product. Several years experience with this process confirm that Superstraight shafts are stable during machining and after the pump has been placed in service. There have been no reports of distortion in these shafts. This example highlights the significance of processing techniques and their direct relationship to the quality of the pump component. Superior processing techniques often go unnoticed, but they add value to the product. Coatings and Treatments Special coatings and surface treatments can enhance the corrosion and/or wear resistance of pump components without using high

Photo 8. Laser application of 420 stainless steel powder to CA-15 stainless steel impeller hub The Pump Handbook Series

391

alloy castings. Pump OEMs are leading the effort to advance coating technology. Coating processes currently being used include metal spray, diffusion coatings, and several types of plating. One promising example involves the use of a laser to apply metal powder (Photo 8), thereby achieving a metallurgical bond and a coating that is virtually defect free. Laser applied coatings are superior to conventionally applied coatings in several respects. The bond to the substrate is better, and they can be applied in thicknesses up to 1/4 inch. Also, they can often be applied to hardenable base metals without subsequent heat treatment. This represents an important advantage over several other application methods, which require higher heat inputs, and it means finish machined components can be laser treated directly without distortion and the need for remachining. Photo

392

8 shows a coating being applied using a laser. Coatings applied in this manner include 420 stainless steel, stellite and Ultimet, a cobaltbase alloy that combines the wear resistance of stellite with the corrosion resistance of Hastelloy. The use of Ultimet as a laser applied coating is new to the pump industry and offers significantly improved life in applications requiring resistance to both corrosion and abrasion. Laser applied coatings can be used to rebuild worn surfaces and extend the life of the component. For example, type 420 stainless steel can be applied to the hubs of type 410 or other martensitic stainless steel impellers, providing a hardness of Rc 50. Impellers reworked in this way require no subsequent heat treatment. In some instances, this approach will make it possible to eliminate the impeller wear ring. Laser applied metal coatings are an example of leading edge technol-

The Pump Handbook Series

ogy being developed by pump OEMs. This technology is being explored to enhance the quality, reliability and performance of the pumps, in the belief that superior products will have distinct advantages in the marketplace. " Colin McCaul has been the Corporate Metallurgist for Ingersoll Dresser Pump Company for 10 years. A graduate metallurgist and licensed professional engineer with more than 25 years of industry experience, Mr. McCaul is accredited by NACE (the National Association of Corrosion Engineers) as a Corrosion Specialist, and he is a Fellow of the Institute of Corrosion. He has written several papers on corrosion and metallurgical failure analysis and was recently awarded a patent for a cavitation resistant alloy that is being used commercially in the pump industry.

✓

R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

Piping-to-Pump Alignment: Getting It Right! A common sense approach will help eliminate piping errors that lead to pump problems. By Luis Rizo, GE Plastics, and Lev Nelik, Liquiflo Equipment n theory, hardly anyone would disagree that proper piping alignment is crucial to pump operation. In a nutshell, piping should be designed and installed in such a way that no loads are transmitted to the pump. An unstressed pump is a happy pump. Most of us have heard the term “free bolt condition.” The bolts that connect the suction and discharge pipe flanges to a pump should just “drop in,” with no forcing effort applied. If the free-bolt-condition rule is violated, bad things are guaranteed to start happening. A pump structure is flexible, no matter how tough its casing is designed. It does not take much pipe-to-pump misalignment to distort the casing and warp the pump, flanges, feet, seal faces and coupling. The laws of physics are unforgiving. According to Hook’s law, within the elastic range of the material, stress and strain are related:

I

=3 E, where E is Young’s modulus of elasticity (psi), and

3=

L , relative elongation (in/in) L0

E= 7 x 106 for steel. For example, an inch-long bar stretched by just 1/16” will be under significant stress: =7 x 106 x (1⁄16)/1.0=450,0000 psi, i.e. a very small deformation could cause huge stresses. The next time you see a “comealong” employed to move the piping to the pump by “just a little,” remember the large stresses and forces you are transmitting to a pump and take pity on it! Minute distortions transmit huge loads, which, in turn, cause distortions and relative movement between the pump’s internal components. Together, these changes can do a great deal of harm—leaky seals, hot bearings failing prematurely, and broken couplings. Then the discord begins. The user blames the equipment manufacturer, who points at millwrights (who by then are long gone) for improper installation methods, and the vicious circle never ends. In the meantime, the maintenance department spends thousands of dollars in replacement parts and re-installation costs, and the plant loses millions of dollars in lost production. The Pump Handbook Series

Installation Nightmare Why, then, is this simple, common sense rule of not overloading the nozzles violated so often? Actually, no one is intentionally trying to do a poor job. Conflicting goals are usually at the heart of piping problems. The project manager wants the job completed on schedule and within budget. The production manager has a production target to meet. And the design firm wants to complete the job, get closure and start the next contract. Finally, the salesman wants to tell the world he has 10 or 20 pumps running at company “X.” This generates conflicting signals and marching orders. Clearly, something important is missed in the process. If examined closely, it is the rush to “get it done and leave” that puts an installation in trouble. To begin with, a layout of a new plant seldom has the details of the piping in the proximity of the pump. There is always a bigger fish to fry: why worry about a $3000 pump when a multi-million dollar reboiler is being procured? The notion of “taking care of the big stuff” neglects the small but important details. This is the typical crack through which general service pumps often fall.

393

culated by dividing the flow (gpm) by the pipe hydraulic area (in2).

Total cost ($K)

60 50

A=The area of pipe calculated as

40

A= x d2(in2) 4

30 20

So,

V2

10

V1

0 -10

C2

Time

-20

Step 1: Piping Size Considerations Suction Piping There is always a compromise to be made between reducing costs

394

(

0.321 x 4 

)

)

2

1 = 2g

1 2g

The losses are thus inversely proportional to the pipe diameter to the fourth power (h loss ≈ Q2/d4), for a given pump flow. For example, if a 4” pipe is replaced by an 8” pipe, the losses are reduced by 4/84 = 1⁄16, i.e. sixteen times improvement.

Figure 1. Total pump cost, reflecting bad piping, failures and maintenance. 1 = “good” piping practice followed. 2 = “deficient piping practice followed. C1 = C2, initial pump installation cost. Note: Variable expenses such as piping adjustments, gasket replacements, re-torque application to eliminate leaks, maintenance cost and parts are significantly higher for case 2 piping (V2 > V1).

Major piping routes are often laid out and installed by a crew different from the one ultimately in charge of making the final connections to the pump nozzles. If problems surface at that particular time (and they do), they are difficult and expensive to correct, as Figure 1 illustrates. There is little room left to maneuver near the pump. So, instead of cutting portions of pipe and correcting the problem, the piping is instead forced into fitting with a pump. Although these piping errors commonly are made in industrial plants, there is no simple, step-by-step procedure that addresses piping installation and alignment, and certainly not a published U.S. or international standard (such as an API or ANSI). The purpose of this article is to lay groundwork and start a discussion on the subject, which could lead to initiation of such a standard. We are looking for reader feedback, with recommendations and suggestions that will hopefully be compiled and processed as a first edition of a standard.

(

Q2 x d4

C1

2

Q x 0.321 d2 4

V2 = 2g

(smaller pipe size) and providing enough suction pressure, in terms of NPSHa, at the pump inlet (larger pipe size). The smaller the pipe size, the higher the friction—and consequently the greater the losses that take away from available NPSH. Hydraulic losses are proportional to the velocity head V2/2g, where: V=Q x 0.321/A-velocity (ft/sec), cal-

Example: Suppose we have a cavitating/noisy pump that needs to deliver 130 gpm. Its bearings do not last the design life L-10, recommended by the bearing manufacturer. Doing a rough calculation, we find that the system provides the pump with approximately 20 feet of

200 6.56 180

EFF. .8

6.0

160

.7

140

.6

120 T 100 D H 80 (FT)

5.0

6.56

.5 .4

4.0

.3

60 40

.2

20

.1 0

20

40

100 60 80 120 CAPACITY - U.S. G.P.M.

NPSHR=9' N 20 P S H 10 R (FT) 0

140

160

0 180

NPSHA=19' (OK)

6.56

NPSHA=6' (CAVITATION)

Figure 2. Pump performance curve, 1.5x1-6, 3500 rpm (Courtesy of Liquiflo Equipment Company) The Pump Handbook Series

OEM CURVE

1000

FIELD TEST 950

TDH (FT)

900

850

800

750

700 0

25

50

75

100

125

150

175

200

CAPACITY (GPM)

Figure 3. Field test versus OEM curve. Performance decay is caused by radical increase in the discharge pipe size (11⁄2” changes to 4”).

NPSHa, reflecting a mostly static level in a suction tank, but not accounting for the hydraulic losses. Although this sounds like an ample margin, that is not really what is available at the pump—once the pressure drop from the suction vessel to the suction flange is taken into account. Suppose that 14 feet are lost due to friction in a 2-inch line that connects the suction tank to the pump suction flange. This leaves 6 feet of NPSHa for the pump to operate. It is evident from Figure 2 that at 130 gpm the required NPSHr of 9 feet exceeds the actually available NPSHa of 6 feet, so it is no surprise that the pump cavitates. The situation can be remedied by increasing the suction piping size from 2 in. to 4 in. In this case the losses are reduced by 14 x (2”/4”)4 = 1 foot (a big change!), and the available NPSH at the pump inlet is now 20 - 1 = 19 feet. Now we have more than enough required NPSH. This could make the difference between a trouble-free installation and one plagued with problems. Historically, and as a rule of thumb, the suction pipe is selected for flow velocities of 5 ft/sec or less.

Discharge Piping Trying to minimize installation cost by going to a smaller discharge pipe can create many problems, especially if the change in diameter is abrupt (increase or decrease). The turbulence created by an abrupt change in the discharge piping (the “vena contracta” effect) can significantly reduce pump performance. Field testing performed

by one of the authors shows a significant difference in the pump’s performance (Figure 3). As can be seen in this field test, the head at 125 gpm is reduced by about 10%. Modifying the piping size to a more gradual transition restores the pump to original design performance, saving horsepower and making satisfactory performance possible without having to increase the impeller diameter. Since the discharge piping is usually much longer than suction piping, and thus impacts cost more significantly, the rule of thumb and a first approximation is to keep the velocity in the pump discharge pipe under 15 ft/sec, somewhat higher than on the suction side.

Step 2: Bringing Appropriately Supported Pipe to the Pump We can see now that piping issues directly affect pump life and performance. Bringing the pump to the pipe in one operation and expecting a good pump flange or vessel fit is a

Anchors

Pump

Last spools to be made and installed after the pump has been leveled and rough aligned Figure 4. Typical anchors for the pump piping The Pump Handbook Series

395

Figure 5. Rough alignment phase (note that the motor and the pump are not coupled yet, and the baseplate is still sitting free, not grouted.)

Leveling bolt holes

Figure 6. Overhead view of baseplate leveling pads and grout location

difficult, if not impossible, task. When bringing the pipe to the pump, the last spool should always be left until the pump has been leveled in place and rough aligned. At this point the pipe should be securely anchored just before the last spool. This will prevent future growth toward the pump’s flanges. Figure 4 shows typical anchors at the pump suction. The piping layout should not be finalized until certified elevation drawings are received from the engineering group or from the pump vendor. Once the final certified prints are received, final isometrics can be completed and the piping takeoff can be done.

ment is late, it is critical to have certified elevation prints of it. The certified prints that the isometrics require for the piping takeoffs can be made without impacting the construction schedule. If the equipment arrives early, it will be at the site prior to the construction team needing it for installation. In such cases, early preparations must be made for long term storage. It is customary to use oil mist lubrication to keep the equipment in as-shipped condition during the storage. Pressurization of the bearing housing and casing at slightly above atmospheric pressure prevents moisture and contaminants from entering sealed areas and damaging components. Early delivery of equipment to the site has the advantage of allowing for verification of actual measurements. Once the location of the equipment is determined, the baseplate can be put in place, leveled and rough aligned, with the equipment mounted. Rough alignment of the equipment should be done prior to building the grout forms (Figure 5).

Step 4: Grouting Once you are satisfied with rough alignment, remove all the equipment (pump, motor gearbox, etc.) from the baseplate. Level the baseplate to a maximum out-of-standard of 0.025” from end to end in two planes. Use machined pads as the base for the leveling instruments. Inspect the foundation for cleanliness and remove grease and oil with solvent (Figure 6).

Step 3: Delivering and Installing the Pump/Motor/Baseplate On site equipment delivery can be early or late. When the equip-

396

Figure 7. Typical anchor bolt and leveling wedges

The Pump Handbook Series

Allow time for the cleaning substances to evaporate. Form the base using the appropriate techniques to allow for the weight, temperature rise and fluidity of the grout material. Grout the base using epoxy grout. Allow the grout to cure, following the manufacturer’s recommendations. This normally requires 24 hours at 80°F. Remove the forms and clean all sharp residue and edges from the foundation. Figure 7 shows the detail of a typical baseplate, with anchor bolts and grouting. The rough alignment step described above is critical to minimize the changes that will be required to fit the piping to the pump appropriately at a later time. At the last stage, when the final spools are installed, the final alignment can be achieved with only small adjustments. This will minimize potential issues with motor feet/bolts. Unfortunately (motor manufacturers - take heed!), motor hold-down bolts are often too tight and allow for only small adjustments to the motor before it becomes bolt bound. Motor manufacturers could improve this situation significantly if motor feet were slotted, by design, rather than drilled for bolts. Figure 8 shows the tightness of space available to insert the foot hold-down bolt. This illustrates once again why good alignment practice can save the time, expense and aggravation of having to alter motor feet by slotting or reaming.

Step 5: Re-installing the Equipment and Final Alignment Re-install the pump and motor on the baseplate. Rough align the equipment again, using reverse indicator or laser alignment, or similarly accurate techniques (Figure 9). Place an indicator in the horizontal and the vertical planes to assure proper alignment. It should be easy now to fine-tune the motor movements within the allowable alignment target without it becoming bolt bound. This is possible because of the rough alignment completed in Step 4. Note: Never install shims under the pump feet. If the shims are lost or misplaced, then alteration to the piping may be

required to get the pump within the required alignment specification. Normal procedure is to place a combination of shims measuring up to 0.125” under the motor feet. This allows for adjustments that will be required during final alignment.

Step 6: Making the Final Piping Installation Make up the final spool pieces for the suction and discharge spaces. Bring the piping to the pump, but do not connect yet, as shown in Figures 10 and 11.

Step 7: Final Alignment Take final measurements, tack weld the spools in place, and finalize all welding once you are satisfied

Motor hold-down bolt(s)

Washer

Motor foot pad Not enough clearance

Do not connect the piping to the pump yet!

Step 2 - Center the piping using the pump’s suction and discharge nozzles

Figure 10. Illustration of the final connection of the suction piping

with the rough alignment achieved so far. Leave a square and parallel gap between the flange faces. The gap should be wide enough to accommodate the size of the gasket required, plus 1/16” - 1/8”, depending on piping sizing. (This is the only distance over which the piping will be pulled. However, because it is properly anchored before the spool pieces, the relative length is short and stresses are minimized). Final-align the equipment, taking into account hot and cold operating conditions, using two indicators on the pump shaft coupling area (Figure 12). As the piping is tightened into place, the shaft should not move more than 0.002”. Otherwise, modify (or make new) spool pieces until the piping misalignment is fixed.

Suggested Periodic Maintenance

Baseplate Figure 8. Potential bolt-bound situation due to tight clearances between motor bolt, feet and base

Place an Indicator in the horizontal and vertical plane to assure proper alignment

Motor

Figure 9. Rough alignment after grouting The Pump Handbook Series

Several clues are common to piping misalignment. These include leaky mechanical seals, hot running bearings and failures. A quick analysis of failed parts can clearly show the evidence of piping misalignment. To make a final confirmation of the symptoms, unbolt the piping while measuring the movement in the vertical and horizontal planes. Again, piping that moves more than 0.002” (and problematic examples are known to “jump” a lot more than that!) must be modified to correct the situation. See Figure 13 for indicator placement. Place the indicator in horizontal and vertical planes, using the motor and pump coupling halves. Uncouple the pump and motor while watching for indicator movement.

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Start unbolting the flanges and continue watching for movement in the indicators. If the needle jumps more than 0.002”, the piping must be modified.

Summary

Figure 11. Final piping Place indicators as shown on the coupling hobs to start the reverse indicator alignment procedure.

Motor Reverse indicator alignment set up

Figure 12. Overhead view of the motor and pump

Proper piping installation and equipment alignment procedures should significantly improve the life of pumps and other rotating equipment by eliminating and/or significantly reducing pipe stresses and strains. In the long run, it is actually simpler and more economical to do everything right the first time. Rushing through these critical steps can lead to more costly problems down the road. " The authors would like to gather feedback from Pumps and Systems readers, reflecting their input for the next update on this subject. Please email us at [email protected] or [email protected]. Hopefully, such input can lead to a formal “Piping-toPump” standard.

Professional Advancement. 1993 4. “How much NPSHA is Enough?,” Lev Nelik, Pumps and Systems, March, 1995. 5. “Pump Baseplate Installation and Grouting,” Perry Monroe, 5th International PumpUsers Symposium, Houston, Tx. 1988 6. “Centrifugal and Rotary Pumps: Fundamentals with Applications,” Lev Nelik, CRC Press, Boca Raton, FL. 1999. 7. “Machinery Reliability Checklists” Elfer, Inc., Chapter 2 list 2.4, Book I, 1998 (Elfer.com) Luis F. Rizo is a Senior Reliability Engineer and Certified 6 Sigma Black Belt in General Electric’s Plastics Quality Initiative. In his current position at the company’s Selkirk, New York plant, Mr. Rizo is responsible for all facets of equipment design and installation, as well as troubleshooting using root cause failure analysis techniques. He is a graduate of the New Jersey Institute of Technology, with a degree in Mechanical Engineering and a Masters in Industrial Engineering.

References Coupling movement readings

Place an indicator on the flange in two directions to measure the pipe movement.

Figure 13. Piping alignment verification

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1. “Major Process Equipment Maintenance and Repair,” Heinz Bloch and Fred Geitner, Gulf Publishing Company, Volume 4, 1985 2. Hydraulic Institute Standards, page 134, 1983 3. “Machinery Maintenance Cost Reduction Opportunities,: Heinz Bloch and Perry Monroe, The Center for

The Pump Handbook Series

Lev Nelik is a Regional Sales Manager for Liquiflo Equipment Company. Mr. Nelik has more than 22 years of experience working with centrifugal and positive displacement pumps, and as a member of Pumps and Systems’ Editorial Advisory Board, is a frequent contributor to the magazine.

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R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

Alignment Monitoring Report Editor’s Note: This special report was prepared by Pumps and Systems editors with the help of Steven Perry at Ludeca, Inc. Location: Southeastern, U.S.A.; nuclear power station Ludeca, Inc. was asked to supervise the alignment of a boiler feed water pump assembly. The machine assembly consists of two motors, (Siemens 4500 hp, 3600 rpm) mounted in tandem, and coupled to a boiler feed water pump. The pump is a Bingham CD type single stage, double suction, 16,000 gpm, 1400 psi, ≈ 400°F. Following the alignment, Ludeca was also asked to conduct simultaneous on-line positional monitoring between each machine using the Permalign® laser monitoring system. The purpose of the consultation was threefold: • To align a boiler feed water xpump assembly consisting of two motors and a pump according to the approved written procedure. • To monitor the positional changes between the pump and the inboard motor during operation. • To monitor the positional changes between the inboard and outboard motors during operation. Permalign® is a laser sourced measuring system. It senses the movement of a prism across distances of 0 to 30 feet. Its resolution and repeatability are one micron. The maximum error is less than 2% of the displayed

value. Such high resolution, repeatability and accuracy are possible through the use of a linearized photodetector. If the Permalign® monitor is attached to a machine frame, and the Permalign® 90-degree prism to a second machine frame, relative movement between the two can be measured. The primary use for relative measurement is thermal growth compensation. The issue in thermal growth compensation is not how much each machine moves. It is the amount of the difference in their movement: “Does the alignment change?” Thus, relative measurement—the difference in growth of the two machines—is ideal for thermal growth monitoring. For relative measurements to be meaningful, the rigidity of the Permalign® brackets and monitoring fixtures must be assured. Further, the mounting points on the machine must be carefully selected so that the movement of the brackets correlates to the movement of the machine frame, and by extension, shaft centerlines. The usual procedure is to attach

the Permalign® system to a cold machine set, preferably at ambient temperature. The monitors are activated and the machinery started. Upon reaching final operating conditions, alignment changes may be read directly from the PC software display. This data can now be used to deliberately misalign the machine cold so that it will be aligned once reaching operating temperature.

Data Collection and Calculation of Results Permalign® monitoring devices transmit a laser beam that is reflected back to the monitor by a prism (Figure 1). The monitor houses a detector (employing a Cartesian coordinate system) that senses the returned beam to a 1 micron resolution. The monitor itself can sample, collect and average data. Additionally, it performs statistical functions, measures temperature and monitors the quality of the laser beam signal. The system is temperature and vibration stable. The 90-degree prism responds to offset changes in the Y axis and angularity changes with the X axis.

Figure 1. Permalign monitoring devices transmit a laser beam that is reflected back to the monitor by a prism. The Pump Handbook Series

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Therefore, with one monitor and one Procedure According to existing procedure, prism oriented vertically, we can steam was applied to the pump to detect changes in vertical offset and simulate the heated running condihorizontal angularity. With another tion of the pump (about 400°F). At a monitor and prism oriented horizontime designated by the utility, a tally (90 degrees to the vertical monzero-zero alignment was performed itor), we can detect changes in horibetween the pump and the inboard zontal offset and vertical angularity. motor. Upon completion of this, the When the Permalign® system is outboard motor was aligned to the used in conjunction with application software, data are Horizontal Vertical Horizontal collected in a Vertical Offset Angle Angle Offset spreadsheet format. The data collection software Pump/Inb. .14 mil/inch 1.3 mil inboard .05 motor mil/inchusing the same zeroenables the user to set-1.0 themil polling freMtr. zero criteria. Power plant tolerances quency from each monitor and calcufor both alignments called for vertilate changes in the .11 mil/inch .05 mil/inch -1.0 mil Inb. -1.8 mil cal and horizontal gaps to be less aboveMtr./Ot.Mtr. mentioned than or equal to .3 mil per inch. The four degrees of freedom. permissible tolerance for vertical Table 1. Alignment Results

Vertical Offset

Horizontal Offset

Vertical Angle

Horizontal Angle

Pump/Inb. Mtr.

8.0 mils

-10.0 mils

-.41 mil/inch

-0.8 mil/inch

Inb. Mtr./Ot.Mtr.

-1.0 mil

2 mils

.23 mil/inch

.14 mil/inch

Table 2. Alignment Targets

Figure 2: PMP/MTR horizontal angularity

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and horizontal offsets are less than or equal to 2 mils at the point of power transmission (coupling). The final alignment readings taken on each coupling appear in the accompanying table. Note: the values in Table 1 • repeated to less than 1 mil. • pertain to the pump as the stationary machine for the pump/ inboard motor alignment. • pertain to the inboard motor as the stationary machine for the inboard motor/outboard motor alignment. • follow standard Rotalign/Optalign® sign conventions. When the alignment procedure was completed, Permalign® quick-fit brackets were attached to a mounting plate bolted to the pump bearing housing. Clock positions for mounting Permalign® monitors follow the same scheme as for Rotalign/Optalign®. They are defined as looking into the stationary machine from the machine to be moved. A special mounting plate facilitated mounting of Permalign® monitors at the 9 and 6 o’clock positions on the pump. The detector portions of the monitors were oriented toward 3 and 12 o’clock, respectively. Quick-fit brackets to hold the prisms were attached directly to the bearing bracket of the inboard motor. Additionally, Permalign® quick-fit brackets were mounted between the inboard motor and the outboard motor. Another complete Permalign® monitoring set-up was carried out to monitor relative positional changes between the motors. The complete set-up consisted of two complete Permalign® monitoring systems, each to monitor positional changes across individual couplings comprising the machine assembly. The monitors were connected to a PC for the purposes of data collection. The application software was configured to poll data from all monitors every 90 seconds. Following each poll, the software calculated the relative positions of shaft center-

tionary machine for the pump/inboard motor alignment. • pertain to the inboard motor as the stationary machine for the inboard motor/outboard motor alignment. • follow standard Rotalign/Optalign® sign conventions. • are relevant only when steam is applied to the pump to simulate a running condition.

Figure 3: PMP/MTR vertical angularity

lines across the machine train. Results are expressed in terms of angularity and offsets between shaft centerlines. Offsets are displayed at a desired flex plane of a spacershaft. The pump is designated as stationary for the pump/inboard motor test. The inboard motor was designated stationary for the simultaneous test monitoring relative positional changes between motors. Initially, all centerlines were assumed to be collinear. Relative changes between centerlines were monitored as the machine train came on-line. Two data files were collected. File 1VPMP.DAT collected relative positional data between the pump and the inboard motor. File 1VPMM.DAT collected relative positional data between the motors. Both files were initiated on 2/27 between the hours of 1:33 p.m. and 6:58 p.m. It was determined the machine train reached full operating temperature during this time frame. Due to plant requirements, a decision was made to allow the unit to continue operating indefinitely, thus ending the test. A cool-down test was not performed. All Permalign® monitors and brackets were removed from the machine train.

Results The formula for converting Permalign® measurements into alignment targets is: C O L D - H O T = TA R G E T S (Permalign® monitor mounted on stationary machine) Hence, using the final data taken at 6:58 p.m. on 2/27, we must reverse the signs to give alignment targets in Table 2. These values • pertain to the pump as the sta-

Performing the inboard motor/outboard motor alignment subsequent to the pump/inboard motor alignment is recommended. These suggested alignment target values are based on conditions observed during the test cycle. While only small changes in alignment became apparent between the motors, an increase in vertical angularity occurred between the pump and inboard motor. Any modifications in the machine train with regard to the pump, motors, piping, and any other external conditions render all Permalign® data invalid. It is important to remember that Permalign® records relative movement between two centerlines being monitored. In this test, the application software expresses positional changes in terms of inboard and out-

Figure 4: PMP/MTR Vertical offset The Pump Handbook Series

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ors during the test cycle. The alignment between the pump and inboard motor changes significantly from the condition outlined in the utility’s approved alignment procedure. These changes emanate from thermal growth, machine frame distortion due to torque, load, and a variety of other factors. Alignment changes occurring throughout the machine train during operation are enough to violate alignment tolerances outlined by the utility.

Recommendations • Perform a cool-down test with Permalign® across the pump and inboard motor coupling when the unit is taken off-line. • Align the entire machine train to targets. "

Figure 5: PMP/MTR horizontal offset

board motor movement, regardless of the fact that pump growth is the precipitating element for most of the positional change observed during the test cycle.

Conclusions All data collected appeared smooth and consistent with expected machinery behavior (please refer to Figures 2 through 5; pump/inboard motor

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data). It appears that after initial startup changes occur between motors, slight alignment changes manifest themselves over time. These changes were considerably more gradual than alignment changes occurring between the pump and inboard motor. Generally speaking, the motors exhibit moderate angular changes over time. Small to moderate offset changes occurred between the mot-

The Pump Handbook Series

Permalign®, Rotalign® and Optalign® are registered trademarks of Prüftechnik AG, Ismaning, Germany. Steven Perry is an Applications Engineer with Ludeca, Inc. (Miami, FL). He has 18 years of experience with industrial instrumentation and is a graduate of Louisiana State University.

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R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

Construction Impacts on Pumps and Systems Get it right from the start and reduce your maintenance headaches later. By Thomas G. Kahler, Newport News Waterworks Impacts of Construction on Maintenance Better quality in construction of treatment plants, pumping stations, tanks and reservoirs results in less adverse impact on planned maintenance after start-up. Poor quality construction and insufficient testing of materials and equipment during construction will increase maintenance in the first months, and perhaps years, after start-up. This article addresses some of the elements of construction and testing that can have an effect on pumps and process systems. Poor construction will result in increased corrective maintenance beyond anything expected in a new facility. Predictive maintenance in such facilities will require increased analyses of systems and equipment due to potential failures resulting from a lack of comprehensive testing during construction. Preventive maintenance, the staple of a planned program, will take a back seat to corrective maintenance and extensive diagnostics (Figure 1). Having to retrofit, open and inspect and re-test equipment detracts from scheduled maintenance of pumps, motors and associated systems. Another more sophisticated form of maintenance is Reliability Centered Maintenance (RCM) (Figure 2) which involves long range scheduling and usually requires extensive relia-

bility engineering to make program changes. The discovery of misapplied or non-specified materials after startup in plant systems provides new and unwanted replacement projects. The net effect of poor plant construction and inadequate equipment and systems testing is to defer planned maintenance programs in order to accommodate retrofit and repairs of pumping systems after start-up. On the other hand, good quality construction practices, adherence to specifications and a comprehensive testing agenda—system by system— will enhance pump and system start-up. Corrective maintenance is avoided by identification and resolution during construction and testing. In the long term, this approach saves everyone money and time spent in settling disputes over issues that would have been more easily addressed during construction.

Write Specs You Can Live With Ensure that items operationally important to your utility are properly detailed in the construction specifications. The utility staff, plant operators and maintenance supervisors should be involved in the review and development of the specifications. They, after all, will have to deal directly with the material products of those specifications. The Pump Handbook Series

Construction specifications should reflect the level of operation and maintenance desired, as well as the critical nature of the process systems. Generalized specifications will buy materials which are reasonably appropriate, but which are not necessarily what you really want for a specific application. This usually causes problems later in the life cycle of the plant or facility due to material substitution without adequate engineering evaluation. Maintenance contracts, special warranties and specific preferred materials should be addressed in detail in the specifications. Where specific levels of qualifications are desired (of contractors’ craftsmen) the requirements and references should be included in the specs. Performance test requirements are critical to the overall success of any construction project. The personnel performing and witnessing those tests are an integral part of that success and should be qualified in the performance of equipment and systems. Don’t use substitutes or simply “interested” individuals. Your specifications in large part will determine the quality of the plant or facility you construct and with which you must live. Attention to detail in the specs will affect long term maintenance. Getting plant operators and maintenance personnel involved in

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ELEMENTS OF A PLANNED MAINTENANCE PROGRAM •

PREVENTIVE MAINTENANCE

•

PREDICTIVE MAINTENANCE

maintenance plans. The operations and maintenance personnel will own, operate and maintain the equipment, and must be knowledgeable with plant systems long before the facility goes on line. Systems and equipment training related to approved equipment submittals is essential and should start early in construction. If the operations and maintenance personnel are on site when equipment arrives at the construction site and systems start coming together, they will be better prepared through early exposure and can better develop equipment maintenance strategies. Review of equipment submittals and necessary modifications at this level will help the success of the construction project.

Materials and Equipment

•

CORRECTIVE MAINTENANCE

Figure 1. Elements of a planned maintenance program

the development and review of specifications will pay large dividends.

Review of Equipment Submittals As a matter of practice in the water industry, equipment submittals for plant installations are passed from the contractor to the consulting engineer (utility owner’s agent) and then to the owner’s engineer for review and approval. Another group that needs to review and annotate equipment submittals is the maintenance specialists who are responsible for mechanical,

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electrical and instrumentation and control crafts as well as plant operators. These personnel have hands-on experience with plant equipment and can contribute valuable input concerning size, type, configuration and reliability of equipment used in water treatment plants. In addition to citing corrections in equipment specifications and parameters, these technical personnel can plan for school and factory training on the new equipment, develop spare parts listings, evaluate compatibility with existing systems and develop The Pump Handbook Series

One of the most costly factors in plant maintenance and post start-up retrofit is the misapplication of materials. Incorrectly selected materials for machinery applications and process systems produce expensive repairs and retrofitting, to say nothing of downtime and production delays. Attention to detail and substantial experience by designers in the integration of treatment plant systems can preclude improper materials applications. Here again, the owners’ operations and maintenance staff can make significant contributions to the design team if they are brought into planning and design of the facilities early on. During construction it is the responsibility of the project management team to ensure that the contractor, and material and equipment suppliers, provide and install materials called out in the design specifications without unauthorized substitution. This is done through a comprehensive quality assurance/quality control plan. Examples of common misapplications of materials in water treatment plants include the following: • The use of austenitic stainless steel alloys in chlorine solution systems and injection piping, wall sleeves and sensor tubing.

OTHER ELEMENTS OF PLANNED MAINTENANCE

•

RELIABILTY CENTERED MAINTENANCE (RCM)

Figure 2. Elements of a reliability-centered maintenance program

• Mixing 304, 316 and alloy 20 stainless steels in the same piping system or mixing other dissimilar metals in the same piping run or fasteners in mechanical joints. The result is accelerated corrosion. • Direct burial or placement of power, control and signal cables in underground conduits and vaults with improper insulation for the application. • Thrust bearings in pumps and motors with inappropriate load ratings and bearing life requirements. This promotes early overhauls. • Sizing electrical conduits too small for the number of cable bundles applied can cause chafing during pull in and result in eventual multiple groundings within the conduits. • Location, type and height of lighting systems in buildings directly impacts maintenance accessibility and life cycle cost. • Chemical feed systems piping,

valves and fittings should have the proper composition and temperature rating for the product whether plastics, elastomers or metals are used (check manufacturers chemical resistance guides). In addition to misapplication of materials, the designer must do enough research to be reasonably sure the equipment and materials won’t be outdated in the short term and that they are compatible for upgrade and with existing systems.

Setting the Plant Machinery Equipment and machinery must be installed level, plumb and concentric. It is in this area that a millwright’s approach is essential and will assure smooth start-up and operations. If pump foundations are not level and true, without honeycombing and soft spots, you may wonder where those vibrations in your pump are coming from. An improperly cast foundation is usually the last point in The Pump Handbook Series

a vibration analysis to rear it’s ugly head. Natural harmonics between machines and structures do occur, but rarely. It is more normal for the problem to be within the prime mover or driven unit. At a minimum, when major plant machinery is being set, you want these key players present: • project engineer (with detail specifications) • owner’s representatives from operations and maintenance • contractor’s craft supervisor (to provide measurements and verify specs) • manufacturer’s representative (as applicable) This “Tiger Team” concept is very effective and represents everyone’s interest with virtually the same group of people (craft and trade changes excepted) at each installation, checkout and performance test. Those persons assigned to witness the setting of plant machinery and process systems components should be well qualified to address issues raised at this critical point in construction. Personnel present should have structural, mechanical and electrical expertise, as well as knowledge of instrumentation and controls, manufacturing and process operations, depending on the equipment being placed and fitted. In addition to the above “Tiger Team,” the Project Manager should be prepared to provide designers as necessary, on site, to resolve difficult problems related to design. Conference calls do not always reveal the peculiarities of a given situation. How well machinery is set on its foundation, made up to suction and discharge piping, conduits and ductwork will affect frequency of repair and life cycle. All rotating machinery is designed with specific parameters of straightness, parallelism and concentricity, in addition to the prime mover (motor) being within specified alignment tolerances to the rotating equipment. Motors should be thoroughly inspected prior to

405

start-up, especially if they have been held in storage for a long time—or you suspect that this is the case. Within the water industry, in the case of both horizontal and vertical turbine centrifugal pumps, the above mentioned criteria are essential to wear, longevity and performance. The aforementioned factors in general all affect the same component in your plant equipment—the bearings. Although you may be using the proper and recommended lubricants in your equipment, violation of good engineering practices in setting alignment and plumb can cause excessive heat, improper loading and eventual bearing failure. There is no substitute for precision alignment of horizontally seated equipment or plumb and concentricity for vertically mounted equipment, such as vertical turbines. Bearings are probably the most important element to consider in setting machinery, but not the only one. Stresses transmitted to motor frames, pump cases and transmission assemblies will also contribute to increased maintenance frequencies and early failures. Pulling piping into mechanical alignment by using a pump flange to fasten the joint is extremely detrimental. The stresses created travel through the entire machine and prime mover, if close coupled.

The Electrical Suite and I&C The plant electrical suite and instrumentation/control (I&C) package should be selected, configured and installed with a number of important factors in mind. Regardless of the plant’s load requirements, main and secondary switchgear, substations, motor control centers and control panels should be installed with accessibility for maintenance as the primary consideration. Another consideration should be for compatibility in adding second generation equipment during future plant expansions and allowing sufficient room for add-ons. Also, within the electrical systems, allowance should be made in the initial design for spare

406

conduit runs to be left blanked. Twenty percent is usually adequate for future upgrades. Use caution in placement so as not to block future areas for cabinets or equipment. Standby generators should frequently be installed during new construction whether or not they will be used for standby only or for peak shaving. The cost to place generators after initial construction is nearly always higher. Reconfiguration of main switchgear and control systems drives up the cost, as does inflation. Maintenance of generators should be considered well in advance of construction. Will engine and generator maintenance be done in-house or by dealer rep contract? These decisions impacting maintenance should be made early in the project and addressed in your specifications. Electrical terminal strips and lugs can be a real maintenance headache after start-up. Ensure that all strips and lugs in switchgear, sub-stations and starter/controllers are sized large enough to securely hold the number of cables or wires required at each connection. Improper sizing will cause grounds and disconnects due to vibrations. Detailed design consideration should be given to what special tools, test equipment and facilities may be required for the electrical and instrumentation control areas. These requirements should be addressed in the plant specifications, or at least planned for with separate funding. After plant startup is not the time to discover that you need an electronic circuit tracking device to analyze the printed circuit boards from your plant’s programmable logic controller (PLC) system. This type of planning error negatively impacts maintenance and is costly to recover from. In addition, the utility board or Director won’t be impressed with having to purchase expensive electronic test equipment to service what should be a complete, new facility. Energy efficiency is proportionateThe Pump Handbook Series

ly related to maintenance. In new construction the initial costs of high efficiency motors, starters and controllers may be higher; however, those costs are made up in superior equipment components and manufacturing standards. The result is lower frequency of repair and longer life cycles. In most plants a high percentage of the electric utility costs can be attributed to the operation of raw and finished water pump motors. Selecting uninterruptable power supply (UPS) systems can be disappointing. Frequently after you buy it and use it several times you wish you had purchased the upgraded model. Depending upon the load tied to your UPS, (e.g., Central Control Computer, SCADA, PLCs) you can increase backup time from the nominal 20 minutes to as long as one hour for just a few thousand dollars in battery system upgrades. Original outfitting with deep cycle, high efficiency batteries with glass mats and lead-antimony cells is well worth the extra cost. A word of caution here: Exhaust hoods covering battery charging racks should be fitted with an explosion proof fan and ducted to the outside atmosphere, not attached to the building HVAC ductwork or exhausted within the building. Refer to NFPA and NEC standards for details.

Testing the Plant Testing individual components, sub-systems, process systems and finally composite testing of interfacing systems during plant construction is essential to your confidence in accepting that the plant and all equipment is ready to begin the first maintenance cycle. Having an integrated test agenda in the construction schedule is paramount to ensuring that testing will not have to be done in a panic during the substantial completion period, or worse, at the last minute to achieve acceptance. Last minute testing should always be avoided to prevent what can be a grueling and unpleas-

Figure 3. Test memorandum data sheet The Pump Handbook Series

407

DATA SHEET A HORIZONTAL CENTRIFUGAL PUMP NO. ________

Figure 4. Data sheet for horizontal centrifugal pump

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DATA SHEET C

Figure 5. Data sheet detailing vibration analysis summary The Pump Handbook Series

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ant experience that almost never produces the desired results. Everyone is anxious to get through the testing and wrap up the project and thoroughness frequently suffers. A test memorandum (Figures 3 through 5) containing the agenda procedures, criteria and data sheets for testing all plant process systems and equipment should be included in the bid package so potential bidders can review the requirements for testing. This approach affords no surprises for the successful bidder as to what is expected in final performance tests and allows for estimating costs of testing as part of the overall bid. Requiring testing without pre-planning will dramatically increase costs. It’s wise to establish contingency plans in the test agenda for equipment delivery delays and possible failures. Anything can and will happen during operational equipment tests. The established test agenda should be promulgated by the consultant’s test coordinator and conducted with the contractor, manufacturer’s representative, owner’s representative and appropriate project engineers present. Integration of the test agenda into the construction schedule will provide a clear path of equipment testing in a timely manner. However, do not wait until each piece of equipment is set on its foundation to begin testing! Don’t leave all testing of major machinery until substantial completion! Field performance tests should be conducted in a smooth, well orchestrated manner, without rushing and with close attention to details. It is strongly suggested that the owner and manufacturer representatives witness the vibration analysis phase of large rotating machinery such as raw water, finished water and booster pumps. Deviation from specified vibration limits/criteria should not be accepted. If a machine fails one segment of the scheduled test, the entire test process should be re-run.

complete construction and get the new plant on line, an important function is often overlooked—the transfer of spare parts from the general contractor to owner. Failure to adequately plan this event will produce negative effects on maintenance because parts won’t be loaded into your supply system in a timely manner. A spare parts log book (a 3-ring binder) should be established during the equipment submittal reviews. The book lists the spares to be provided by each equipment vendor in accordance with construction specifications. The log book should identify the vendor or manufacturer, the principal equipment or system by nameplate data; the manufacturers’ spare part number, quantity and any other identifying information available for the material. Additional columns should be available for dates, signatures of receipt and inventory accountability. At an agreed point in construction (probably substantial completion), the general contractor’s logistics coordinator and owner’s supply staff should meet to conduct a joint inventory and agree on a spare parts turnover location and procedure. The alternatives to having a plan for spare parts turnover at the end of plant construction may range from your spares being left in their boxes next to the equipment for you to find, or having a tractor-trailer show up at your maintenance facility with one, uninventoried, untagged load of spare parts which could take weeks to sort through and identify. It’s not overkill to have your engineering consultant include a spare parts turnover plan in the plant specifications. This initial effort will help ensure that spare parts are available to perform maintenance in the new plant from day one. If the general contractor is not made aware that he/she is required to conduct an orderly and controlled turnover, there’s no telling how or when you may receive your spare parts.

Equipment Documentation

Certifications and

Visit Manufacturers and

If your plant is equipped with hoists, cranes and other weight-handling equipment, you will want to ensure you have test documentation on file at completion of construction. The signed documentation should indicate compliance with requirements for testing new installations in accordance with federal and state OSHA regulations (29 CFR Part 1910) as well as ANSI B30.2.0 requirements for design and manufacture. In cases where structural steel tanks are constructed as part of your plant or pumping station project, non-destructive test documentation should be available to the owner. Radiographic examination reports of welds should be a part of the project documentation with film interpretation through the consulting engineer or a licensed radiography firm. Owners should retain copies of these inspection reports. Along with the normal motor and pump performance curve documentation, you should also obtain certifications for all pressure vessels installed in your plant. In most states this includes all air compressor receivers of eight cubic feet volume and larger. You can initiate these inspections and certifications through your insurance underwriter’s office. Any elevators installed in your physical plant also require inspection, certification of safe operation and documentation. Also, mezzanine and other suspended floors should be clearly marked as to their load limits per square foot. It’s a good idea not to begin the start-up phase in plant construction without a nameplate database on all equipment that will be included in your maintenance program. Whether you are starting a new maintenance program or adding to an existing one, this database is very important to the continuity of the maintenance plan, and it will be useful later if you upgrade the program.

Spare Parts Turnover Plan In the flurry of activity required to

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Suppliers of High Value Equipment This method of assuring that your utility is getting what it pays for is well worth it. Have your staff visit manufacturers’ plants and service centers prior to or during fabrication of your equipment. Witness motor and pump assembly and performance tests. This is especially helpful in new construction and when you have selected or have been obligated to new types of equipment. The travel costs are a small investment for a utility to make for the resulting personal contacts, first hand impressions and information. Consulting and design engineers do this as a matter of practice, but after initial visits and development of design details for equipment, most of their contact is by phone or correspondence. This belays the personal element and displayed degree of interest by the owner’s staff. If the owner or director of a water utility is about to spend a half million dollars on a series of vertical turbine

pumps, for example, contact with the manufacturers’ designer, production personnel and service facility managers is an effort that will pay off in the future with intimate knowledge of the product and the people. The plant operators and maintenance personnel have a vested interest in knowing what equipment they will be dealing with from the beginning.

tions and review equipment submittals will only serve to improve the facilities. They have a vested interest in new construction as the end users and caretakers. Having a test agenda integrated into the construction schedule is a most important aspect of a successful construction project. It is usually the thing that impacts maintenance more than any other phase of the project. "

Conclusions Depending upon the quality of construction and testing in new plants and facilities, the impact on maintenance can be enhancement or counterproductive. The owner’s operations and maintenance staff can provide positive input if afforded the opportunity. By writing specifications that are meaningful to the success of construction and using a viable test agenda, maintenance in the new or rehabilitated facilities will be most effective. Using the experience and talent of the owner’s operations and maintenance staff to help develop specifica-

The Pump Handbook Series

Thomas G. Kahler is the Facilities and Watershed Maintenance Manager for Newport News Waterworks, Newport News, Virginia. A retired naval officer, he served in the engineering/propulsion and repair disciplines for 23 years. Mr. Kahler has extensive experience in operation, maintenance and condition analysis of nuclear and fossil fuel power plants as well as water treatment facilities. He has presented papers on pump maintenance and plant construction and systems testing at numerous professional conferences.

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R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

Predictive Maintenance Programs: Building a Complete Package People, equipment and technologies must work together for best results. By Kenneth Rankin, Bayer Corporation ith the latest technologies available at more and more reasonable costs, many companies are striving to build complete predictive maintenance programs. As production needs demand less downtime, higher output from machinery, and lower maintenance repair cost, predictive maintenance is becoming a valuable tool for companies. This article will explain some of the technologies used in predictive maintenance. It will also review related considerations such as personnel requirements, cost justifications, commitment from management, in-house expertise versus consultants, and other factors that need to be weighed in making such a program effective.

W

Maintenance Ideologies There are many maintenance strategies being used in industry today. All have a place, and one could argue for or against each. To understand which would be best for your company, let’s examine each and consider the basic concept behind it. And to understand how each maintenance concept affects machinery, let’s look at the failure profiles of a typical machine.

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The first and perhaps oldest used maintenance practice is “run it until it fails.” In years past many plant managers would operate their machines until catastrophic failure and then simply install new equipment. Root cause failure analysis wasn’t performed in most instances. But machinery has evolved over time and become more complex and expensive. Today, equipment cost, lost production due to unexpected downtime, and the need for companies to stay competitive make this— ”run it until it fails”—notion impractical, not to mention unsafe. Preventive maintenance is to perform scheduled maintenance procedures on a piece of equipment as determined by equipment manufacturers or past user experience. Wear is normal in machinery that is in use. As the equipment approaches the wear-out zone, the chance of equipment failure clearly increases. Preventive maintenance can help extend equipment life, but it will not predict failure. Some disadvantages of relying on this approach include the chance of an unexpected failure resulting in production downtime, catastrophic equipment failure, unneeded equipment inspections, unThe Pump Handbook Series

needed oil and lubrication changes resulting in higher waste disposal cost. This will increase man-hours and overtime expenses and will increase maintenance cost. Predictive maintenance could be compared to going to the doctor for your yearly physical and having test results indicate potential problems. Using technologies like vibration analysis, oil analysis, thermography, motor analysis and ultrasonics can help determine the condition of a piece of equipment to the extent that it can be scheduled for service at the best possible time. Determining the cause of failure is extremely difficult after the equipment has destroyed itself. Usually in this situation critical evidence is lost, and preventive measures for the future impossible. By using predictive maintenance technologies, the problem can usually be identified before the failure occurs. This helps in root cause failure analysis and helps to prevent failures of the same type from occurring again. The best results occur when a good preventive maintenance program is combined with predictive maintenance. Such a complete program has many advantages. It can …

without implementing all of these. Due to the magazine space it would take to go into depth on each technology, we will only touch on each.

Vibration Analysis What is vibration? Vibration is the complex, periodic motion of a rotating or reciprocating system in response to one or more forcing functions. The forces that act on a machine internally or externally to cause vibration are often called excitations. Table 1 lists sources of machine excitation and the frequencies associated with each source. Equipment should first be categorized and prioritized so a reasonable monitoring schedule can be deter-

When preventive and predictive maintenance are successfully married, the result is proactive maintenance. The ultimate goal is to get maximum life out of each machine and avoid the “fire fighting” mode when something goes wrong. A proactive maintenance program makes efficient use of all of resources with minimal manpower and parts storage. Root cause failure analysis must be part of any failure situation, even if the failure is noticed before any loss occurs.

The technologies listed below are involved in predictive maintenance. An ideal program would possess all technologies but might not be practical for some companies. With the low cost associated with some of these, it is possible to have a good predictive maintenance program

OPERATING AGE

Figure 1. Causes of start-up and early life failures

mined. Safety concerns, production loss, production downtime, and equipment repair cost are some of the factors that play a role in determining equipment monitoring frequencies. The prioritizing process should include production department input. After determining the monitoring frequency of equipment, a plan can be

PROBABILITY OF FAILURE

WEAR-OUT ZONE

LIFE

OPERATING AGE

¥ EXTENSIVE HISTORIES GIVE EXPECTED LIFESPANS ¥ COMPONENT OVERHAULS AT FIXED INTERVALS ¥ MODERN MACHINERY DOES NOT FIT THIS PATTERN Figure 2. Classic failure profile

PROBABILITY OF FAILURE

Foundation of a Predictive Maintenance Program

EXCESSIVE ROUTINE MAINTENANCE OVERLY INVASIVE MAINTENANCE POOR QUALITY MANUFACTURE INCORRECT INSTALLATION IMPROPER APPLICATION POOR DESIGN

PREVENTIVE MAINTENANCE

Equipment Life Profile The risk of equipment failures is generally highest at start-up and when equipment enters the wear-out zone. Figure 1 shows early life failures. Figure 2 shows failures as equipment starts to wear out. By implementing both preventive and predictive maintenance, equipment life can be noticeably increased (Figure 3).

PROBABILITY OF FAILURE

1. Reduce equipment repair cost 2. Reduce or eliminate production downtime 3. Reduce repair time on equipment 4. Improve MTBF 5. Reduce overtime 6. Improve troubleshooting 7. Improve manpower scheduling 8. Reduce manpower needs 9. Reduce spare parts inventory 10. Eliminate unneeded inspections 11. Increase pm intervals 12. Decrease waste oil disposal cost 13. Lower insurance costs 14. Maintain ISO certification 15. Prevent environmental damage 16. Increase equipment reliability and safety

RANDOM FAILURE

INFANT MORTALITY ZONE

WEAR-OUT ZONE

STARTUP FAILURE

RANDOM/OPERATIONAL FAULTS

WEAR OUT

ACCEPTANCE STANDARD

PREDICTIVE MAINTENANCE PRECISION CORRECTIONS

PREDICTIVE MAINTENANCE

COMMISSIONING

PREVENTIVE MAINTENANCE

ROOT CAUSE

Figure 3. Equipment life increases when preventive and predictive maintenance are followed. The Pump Handbook Series

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established so measurements can be made in an orderly manner. Data is collected at different points on a given piece of equipment or in an equipment train. The analyst first determines what measures will be taken—displacement, velocity or acceleration, for example. The frequency range and application play a critical role in this decision. After determining what measures to monitor, the analyst decides what transducers will be required for the equipment monitoring. Noncontacting relative displacement probes, velocity pickups and accelerometers are all possible options depending on what information is needed. Each transducer has its place, and sometimes equipment monitoring will require the use of more than one transducer. There are various types of test equipment used to capture data and many software programs to help analyze it. The following is a list of instruments commonly used for collecting and analyzing data.

ed as a time waveform or spectrum (depending on your test equipment). The time waveform can be a simple sine wave or a more complex figure. The time domain is viewed as amplitude versus time. Spectral data are expressed as amplitude versus frequency and are derived from the time domain using the FFT (Fast Fourier Transform), which is a calculation method for converting a time wave-form into a series of dis-

crete components of frequency and amplitude (Figure 4). The cost of implementing a program that uses vibration analysis can vary depending on what type of equipment and software is used. The best approach for companies with limited budgets is to start with the basics and slowly upgrade over time. This will give your people time to train and gain experience. Remember that the more informa-

Source

Frequency (Multiples of rpm) Fault Induced

Mass unbalance

1x

Misalignment

1x, 2x, some higher

Bent shaft

1x

Mechanical wear/looseness

Harmonics

Casing and foundation distortion (thermal, mechanical)

1x, some higher

Common Instrument Systems 1. Simple meter 2. Oscilloscope 3. Tracking filter 4. FFT spectrum analyzer 5. Dual channel FFT spectrum analyzer 6. Electronic data logger

Coupling lockup

1x, 2x

Bearings, anti-friction faults

Bearing frequencies

Before buying any vibration hardware, carefully evaluate your program. And before purchasing the first FFT spectrum analyzer or electronic data logger, look at all the equipment options available to you. Some equipment has special features that might fit your program better than others. It would not be a bad idea to talk to analysts at other companies or post questions on the Internet (such as Pumps and Systems’ “Pump-Chat”) to get feedback on a particular brand. Once information is acquired, it can be viewed and analyzed. Measurements are usually present-

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Design Induced Universal joints

2x

Asymmetric shafts and supports

2x

Gears (n teeth)

Nx

Couplings (m jaws)

Mx

Bearings, fluid filled

0.5x, 1x

Blades (m)

Mx

Slider crank mechanisms

Multiples of speed

Impact mechanisms

Multifrequency, depending on waveform

Motor poles and slots

Multiples of poles

Nonlinear supports

Subharmonics and harmonics

Table 1. Sources of machine excitation and the frequencies associated with each source The Pump Handbook Series

2.

3.

4.

5.

Figure 4. Diagram of signal acquisition

tion you have about the equipment, the easier it will be to determine faults. Gathering information on older equipment can be difficult and sometimes impossible. Even if budget restraints keep your plant from starting vibration analysis now, start keeping records of how your machinery performs. Setting up your equipment, alarm parameters and alarm faults is a time consuming job, but it will save you in the long run. Figures 5 and 6 illustrate what typical faults look like using vibration analysis. These are clear and clean data that are very easy to read. Most data are not as simple as this but can be interpreted well after getting some experience and the proper training.

Oil Analysis Oil analysis has been used to evaluate machinery condition for many years. It is the oldest predictive technology. For oil analysis to be effective, it must take into account three vital considerations: oil chemistry, mechanical wear and contamination. Lubricant condition or chemistry deals with the properties of the oil. If oil is breaking down, it cannot lubri-

cate machine parts effectively. Oil contaminants can be solids, liquids or gases. Particulate contaminants can cause excessive wear in rotating machinery while liquid contamination, such as the unwanted presence of water, can adversely affect oil chemistry and promote oxidation in machinery. Many companies pay laboratories to perform tests on oil samples and return the results to the plants. Many larger facilities might receive this service free if they use large amounts of a particular vendor’s oil. With the cost of the equipment that performs these tests on the decline, there are many benefits to doing this in-house. Benefits of onsite testing include immediate data, oil analysis expertise, no cost for additional sampling, complete ownership and integrated reporting. Remember that most mini-labs in the market today can provide only basic tests. More specific tests still might have to be done at a full-scale outside lab. The list below gives specific oil testing equipment one might consider having on plant site. 1. Oil analyzer—Analyzes oil to reveal information about oil contThe Pump Handbook Series

amination, oil chemistry and machinery wear. This is usually done by taking a new sample of the oil you are going to use in your equipment and analyzing it. This sample then becomes your baseline for future tests. Ferrous wear monitor—Measures the amount of magnetic iron present within an oil sample. This will help you detect wear that you normally would not be able to see. The amount of wear is directly proportional to machine damage. Viscometer—These are either mechanical or digital and are used to test viscosity. The viscosity is one of the most essential lubricating properties of a lubricant. Particle counter—Provides information about particle contamination and lubricant breakdown. Microscope—Used to view wear debris and contaminants. By looking at the shape, color and size of the contaminants in the oil, you can determine the amount and type of wear that is occurring in the equipment.

As with all other technologies mentioned here, it is a “must” to have personnel trained. Proper sampling and lab testing procedures will result in accurate data.

Motor Analysis Electrical faults in motors can be determined by mechanical means (vibration analysis) or by measuring electrical current. Collecting and trending current data provides information beyond that obtained in the vibration domain. By collecting flux coil readings and viewing the data in a spectrum, information can be trended and analyzed. Typically, the equipment needed to do this includes vibration analyzer, flux coil and related software to analyze and store data. The data are gathered by placing the flux coil at the end of the motor bell housing and reading the flux field emitted by the motor. This field contains repeatable information which,

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when trended, can show electrical problems as they occur. When the data are viewed as a spectrum, the image resembles a vibration spectrum. Faults can be correlated between the current and flux spectrum, but it is the amount of change seen that will be the same, not the individual amplitudes. Typical faults that can be detected are broken rotor bars,

high-resistance joints, voids in aluminum cast rotors, cracked rotor end rings in squirrel-cage induction motors, stator problems and voltage imbalance. Motor analysis goes along with vibration analysis. The biggest expense for starting this should have been absorbed by your vibration program (analyzer). The software

GEARMESH FAULT ■ HARMONICS AND SIDEBANDS OF GEARMESH DOMINATE

ROLLER BEARING EXAMPLE ■ NOTE HARMONICS OF OUTER RACE FREQUENCY & SIDEBANDS

and flux coil that you need to start this are relatively inexpensive. The database that you will set up on the motors is easy to build and shouldn’t take much time to establish.

Thermography Thermography detects and measures variations in the heat emitted by objects. The process is used to identify electrical problems, friction induced heating, insulation deficiencies and process problems due to heat. It can even find roof leaks. Though this is still the most expensive technology to acquire, it has become available for smaller, more reliable higher resolution cameras at a lower cost. Systems are available so that monitoring routes can be assigned and reports can be generated at the push of a button. The following are examples of faults that can be identified with thermography. 1. Loose electrical contacts or lugs 2. Dirty electrical contacts 3. Insulation deficiencies 4. Heat generated by coupling misalignment 5. Plugged process lines 6. Faulty valves, traps 7. Various machinery problems 8. Roof leaks 9. Deterioration of the walls of a vessel or piece of equipment (such as a boiler) 10.Heating related process problems

Ultrasonics

Figures 5 and 6. What typical faults look like using vibration analysis

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This technology analyzes ultrasonic vibrations to determine equipment faults or leaks in lines. On average, the human ear can detect sounds up to 16 kHz. Some humans can hear sounds up to 21 kHz. Ultrasonic guns can pick up sounds above this range. There are ultrasonic components in practically all forms of friction. In industrial predictive maintenance, sounds above 21 kHz are of most importance. For example, by the time you hear a bearing making noise, it is usually an emergency problem. With an ultrasonic gun,

you can become aware of bearing deterioration in its early stages and make repairs. In fact, ultrasonic inspection and monitoring of bearings is by far the most reliable method for detecting incipient bearing failure. And ultrasonics is the least expensive of the technologies. Initial expense is minimal, and most training is hands-on and gathered by experience. The following list contains faults that can be detected using this noninvasive, nondestructive technique. 1. Pressure or vacuum leak detection—Turbulence creates a broad spectrum of sound called “white noise.” 2. Electrical faults in transformers and switchgear (electric arc, corona, tracking detection)—When voltage escapes in high voltage lines or when it jumps across a gap in an electrical connection, it disturbs the air molecules around it and generates ultrasound. 3. Beginning of bearing wear—Wear and deformation raise the amplitude of ultrasonic frequencies. 4. Brinelling of bearing surfaces— This causes a flattening process as the balls get out of round, and this in turn creates a repetitive ringing and increase in amplitude of monitored frequencies. 5. Flooding of (or lack of) lubricants— The flooding will create a rushing sound that is 8 db over baseline. 6. Faulty steam traps—The turbulence of live steam rushing has a characteristic sound. 7. Problem valves—Fluid leaking also has a sound. 8. Gear difficulties—Faulty gear teeth ring when they come in contact.

First Steps Before getting into the meat of what it takes to start a predictive maintenance program, let’s discuss practices that should be in place to support the program. It would be nice to go out and buy needed vibration equipment and software and then hire an analyst to start a PdM

program and then just sit back and planned for each mechanic before document the cost savings. It is also the beginning of each year. It is unrealistic to think that it works that important to note that with the way! What good would it be to find complex equipment and systems in a piece of equipment with bad bearplace at many facilities today, trainings, only to have it pulled from sering that used to be reserved for vice so that when the repairs are engineers and designers, such as made using bad bearing installation pump selection and design, is a procedures by your mechanics, it must for a qualified mechanic. only lasts a short time after the 3. Documentation—Equipment inrepair? All of the most advanced prespections and failures should be dictive maintenance technologies documented every time. This can being used won’t help when poor be done on a PC or CMMS sysmaintenance practices are in place. tem. When a failure occurs, Below is a list of practices and proattention to all details, no matter grams characteristic of a solid mainhow small, will help to prevent tenance department. another failure. Rotating equip1. Standard maintenance procement files should be kept orgadures—Written maintenance and nized and up to date. Inspection repair procedures on each piece of forms for each piece of equipequipment should include rebuild ment should be easily accessible and inspection strategies, tolerfor the people doing inspections ances, detailed drawings, inspection forms and safety concerns. 2. Training programs—Formal training and refresher training for mechanics performing the repair task can be accomplished inhouse or done by outside consultants. Training on specialized equipment can be written into the contract when it is purchased so that it will be included with the equipment. Many vendors will provide training on equipment that you are currently using at your facility. Vendors who supply parts for your equipment (e.g., bearing vendors) will hold classes on proper installation and storage of their parts. At our facility, we have made time for the mechanics to review manuals and magazines (Pumps and Systems and other periodicals). We also have written procedures for most jobs. Training at and away from the plant is Figure 7. Pump data sheet The Pump Handbook Series

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or making repairs. Digital cameras—a great tool for storing visual data with failure documentation—are relatively inexpensive. 4. Engineering support—Engineering support should be just that—support. Engineers should be involved with equipment failures and work with mechanics and production personnel to correct problems leading to equipment failure. Although some repairs might seem trivial, the engineering group should review them. Maintenance personnel have a good handle on bettering equipment, and the interaction between the mechanics and engineers is essential. 5. MTBF Tracking—Simple programs can be written to track MTBF. This not only helps point out problem equipment, but also will be an excellent tool for reporting equipment improvements and cost savings to management.

3. Oil analysis hardware and software 8k-12k 4. Motor analysis hardware and software 11k 5. Thermography 30k-45k 6. Ultrasound 5k-10k When staffing a rotating equipment department for a PdM program, consider the number of points to be monitored and technologies that will be employed. It is wise to give at least one person (a mechanical engineer or rotating equipment specialist) overall responsibility for the program. This person’s job is to set up

Setting up a PdM Program Budgeting for a predictive maintenance program can be difficult. Money required for hardware, software, personnel and training takes its toll on a department budget. The best approach to setting up a program is planning. Before rushing out and buying equipment and software, department heads should look at what equipment will be monitored and what parameters will be gauged. Basic equipment such as oil analysis, ultrasonics and motor analysis are relatively inexpensive and provide good information. Oil vendors will do general oil analysis at little or no cost. Usually their tests are basic in nature, and the results are not as immediate as tests done in-house but can be an important part of your program. Vibration analysis and thermography are more expensive but provide a much better idea of what is happening in machinery. The list below contains approximate costs for various technologies. 1. Vibration hardware 14k-19k 2. Vibration software 6k-18k

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Figure 8. Motor spec data sheet The Pump Handbook Series

and maintain the program. Deciding manpower needs, equipment monitoring schedules and alarm levels, reviewing and analyzing data, reporting, cost justification and some data collection might also be part of the job description. Trained mechanical personnel can take vibration readings. Trained people in the PM group can do oil sampling and analyzing. Thermography is perfect for members of the electrical group. In each of these areas, these are the personnel who know the equipment and process best. In large plants with lots of equipment, it might be necessary

to have people dedicated full time to PdM. It takes time to gather and analyze data properly. The number of points that can be analyzed depends on the skills of your analyst and the amount of information and data that you have entered into the system. The cost of staffing the program can vary. The engineer or rotating equipment specialist who heads the program is usually comparable to a mid-level engineer. Analysts who gather data and do basic studies on a full time scale seem to make about 5% to 10% more than the maintenance personnel who repair the equipment. Remember that as an analyst’s skills grow, so should his or her pay. Many compa-

nies are starting PdM programs and would gladly pay a higher salary for a trained analyst with some experience. When allocating money for training, consider the cost for the training, travel, food and lodging. Typical training classes run anywhere from $800 to $1600. Vibration and oil analysis and thermography all have different levels of certification. Our experience has shown that two or three training classes a year are a good starting point. The best results come after the training classes are followed with experience. After the analyst has had the training, he or she can apply it on the job. Experience with the proper training definitely makes a better analyst.

Figure 9. Motor vibration spec and repair sheet The Pump Handbook Series

Setting up a PdM program is time consuming. You could buy a data collector and start taking data without any pre-work and do some analyzing, but your results won’t be the best. What you put into your program is what you will get out of it. Our facility has about 500 pieces of rotating equipment that are monitored, and it took about a year to get the database set up to include all pertinent information. The more information that you know about the equipment, the better. Figures 7, 8 and 9 are examples of data sheets that we fill out on every piece of equipment. Developing this information in a database, along with alarm limits and analysis parameter settings, can take some time. When selecting people who will manage or work in your program, make sure that they understand everything involved. They must be willing to learn new skills. Interaction with management, engineering and production are all part of the job. As managers and production operators start to see the benefits and results from the program, the PdM personnel will become more involved in operations. Mechanics who were reluctant to act on your analyst’s data at first will come to trust it as they see equipment run longer with fewer catastrophic failures. The training associated with becoming certified in thermography, oil and motor analysis is relatively easy. Although vibration training requires more math skills, it doesn’t require a Ph.D. Experience at our facility has shown that most people want the training and responsibility that goes with PdM and look at it as a positive career step. Commitment from your analyst is important. Commitment from management to support the PdM is the backbone of the entire program, however. Managers must make sure the needed resources are made available and give the program time to develop. This doesn’t mean you have to spend $150,000

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the first day you start your program to get results. Start with a single technology and build over time. Getting more money for the program will become easier as you document cost savings. 4.

In-house Programs vs. Consultants One could argue for or against inhouse or consultant PdM programs. One thing for sure is that any company that has critical rotating equipment and can’t afford any downtime needs to have one or the other! Before we implemented our own PdM program, we had an outside consultant do all of it. This, along with talking and training with other PdM personnel, has given me some insight on the subject. Pros for an In-house PdM Program 1. Ownership—People who work on in-house programs must take great pride in their work. They are responsible for their program and only their program. With responsibility comes accountability, and you can be assured that the problems at your facility are top priority. 3. Turnover—Many good PdM personnel from in-house programs as well as consultant firms have left due to better job offers. I consider this a plus for in-house programs because managers have some input on the pay and benefits that these people receive and therefore some control on turnover. A plant has no input on what a consultant pays their personnel, so there is always the possibility that you might have new people knocking at your door the next time readings are due. 3. Building equipment and process knowledge—No one knows the equipment and process in the plant better than the people who run and work on it everyday. This is a definite plus when it comes to troubleshooting and root cause failure analysis. Plant employees who work together on equipment

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5.

6.

7.

can gather detailed information so that failure analysis can be as accurate as possible. They know the internal workings of the equipment and process variables better that any off-site personnel. Quick response time for priority problems—In-house programs are quick to respond to problems and diagnose machine conditions. Having trained personnel and PdM equipment on site gives you immediate response to equipment difficulties. On-site personnel can not only troubleshoot immediately, but might also see an anomaly developing elsewhere in the system that is contributing to the problem. Plant personnel can access data, do additional testing and generate reports and recommendations without leaving the plant site for any reason. Budgeting—When allocating money for the department, it is pretty easy to figure hours, equipment and training costs in advance. Besides some occasional overtime, additional testing and troubleshooting should not cost you additional monies. Consultants charge a flat hourly rate or by the measurement point. Any additional test or points are going to cost you more. If the unit has something in particular they want to look at, an on-site analysis can do the job without additional cost. Accountability—With an in-house program you know what you are getting. The program will help meet the plant objectives. Specific measurement points, alarm settings and reading intervals are all easily set up with an in-plant PdM. Consultants can also do this, but it is not as easy for them, and the cost for setting up the initial database can be high. Most consultants take general readings and set alarms for a group of similar equipment. Job opportunities—Many maintenance mechanics will look at the chance of working with PdM The Pump Handbook Series

groups as job advancement. Using the latest technology to diagnose equipment failures is appealing to many mechanics who have topped out in their trade. Pros for Using a Consulting Firm 1. Initial cost—The plant does not have to bear the cost for PdM equipment, personnel and training. Start-up costs can be a big blow to a manager’s budget, and the time it takes to see a return on investment depends how fast personnel can be trained to use it. The consultant also absorbs maintenance and equipment update costs for his PdM equipment. 2.Employee turnover—Personnel losses can be expensive. The cost of training a person in a technology adds up, and when that person leaves the company for another job, the consultant is left with the burden and expense of training a new person. 3.Experience—Consultants work on a wide variety of machinery at many different facilities and therefore are more likely to know of and see faults that are not very common. A consultant will usually work with just one technology and has many different experiences that he or she can call on when troubleshooting. " Kenneth Rankin is a rotating equipment specialist at Bayer Chemical Corporation in Charleston, South Carolina. As a millwright he has spent 10 years in various buildings and divisions of his plant site working on rotating equipment. After becoming certified through the Vibration Institute, Mr. Rankin was promoted to rotating equipment specialist at Bayer’s fibers division where he helped design and implement preventative and predictive maintenance programs. He has also designed and implemented a program for tracking and reporting MTBF on all of the rotating equipment in the building as well as developed detailed repair procedures.

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R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

Best Practice: High Temperature Slurry Applications Using high chrome iron lined slurry pumps increased reliability by more than 3.5 times at this plant. By George C. Schmidt, ALCOA or pumping systems handling abrasive slurries, the major maintenance cost is often the replacement of the wetted parts of the pump that wear out from the abrasive media in the slurry. The most common wear-resistant material used in slurry pumps is the abrasive-resistant high chrome white cast iron—ASTM A-532. There are many slurry pumps on the market that use this material and perform well in moderate services. The most common slurry pumps on the market are suitable for services in the 150250 psig range and limited to operating temperatures below 250°F. There are some processes, however, that handle abrasive slurries at pressures and temperatures substantially above those for which most common slurry pumps are designed. This article presents a case study of the successful use of a specially designed slurry pump for high pressures and temperatures that utilizes the advantages of wear-resistant chrome iron. Design considerations, special operating procedures, spare parts management and wear life improvement results from operating experience are described and an economic analysis is included.

F

Background

tric motor. The pump performance One of the most severe abrasive serfor each operating pump is about vices for slurry pumps is in the high 2,200 gpm at 450 ft. head. temperature digestion process for For many years, the pumps used in making alumina from bauxite ore. In this service were API centrifugal this process, bauxite slurry at approxiunits with cast steel casings. These mately 10-15% solids must be heated pumps had been achieving only to 450°F to facilitate efficient digestion about 675 run hours to complete of the bauxite ore in the caustic liquor. wear out. Severe wear patterns as The heating of the slurry mixture is deep as 1” in the cutwater and sucdone in stages, and the slurry heaters tion areas of the casings made the are usually direct contact heaters casings irreparable. Keep in mind using both flash steam from the that the metal in these areas is about process and live steam. In order to 11⁄4” thick. Photos 1 and 2 illustrate typical wear patterns. move fluid from one heating stage to Because of short life to wear out, another, pumps are needed because high repair costs were being incurred the product must be transferred from and a more reliable pump that would a vessel at lower pressure to the next downstream vessel, which is at a higher pressure. Figure 1 depicts the heating process at the last stage. As shown, there are two pumps operating in parallel with an installed spare. Not shown in the diagram are the valves that enable each of the pumps to be valved out of the circuit so they can be shut down for maintenance. Each pump is driven Figure 1. Pump circuit handling abrasive slurry at high by a 600 hp, 1,750 rpm elec- temperature and pressure. The Pump Handbook Series

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Photo 1. Casing wear at the cutwater area

Photo 2. Casing wear at the suction seal ring area

increase wear life substantially was desired. Wear-resistant coatings had been tried, but a typical coating of .010” thick resulted in only marginal improvement. Other materials such as CD4MCU, CA-15, and CB7CU had also been tried, but none gave sufficient life improvement to justify the extra cost. It was known that the ideal material of construction for this abrasive service is hardened high chrome iron alloy as specified in ASTM A-532. This material is predominantly used in the mineral processing industries for pumps handling abrasive slurries. No one had tried the material in our service is because of its low ductility and the risk that a crack would form and cause a catastrophic failure while the pump was operating under high internal pressure.

The Search for a Better Pump In the early 1980s a search was begun for a pump design suitable for service in the high temperature digestion process that would give the improved wear life advantages of the abrasive cast iron but could also be

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operated safely. Besides the obvious objectives of longer wear life and reduced maintenance costs, some of the design aspects we needed to consider included: • Safety—The design must be such that a catastrophic failure would not result if a crack formed in the hard chrome iron due to thermal shock. • Easy installation—We preferred that it be possible to install the pump on the existing foundation and without major piping rework. • Special design—This was necessary to accommodate the higher operating temperatures without causing internal mechanical stresses that would crack the hard iron liners. • The operating procedures could not be too complex. • The spare parts supply had to be reliable. From a technical literature search, we learned that in the 1970s a research effort was sponsored by the Department of Energy to develop a pump that could be manufactured commercially and was suitable for high temperature, high pressure, abrasive services. The target industrial application at that time was the anticipated growth of the coal gasification industry. Lawrence Pump Company was one of the manufacturers that participated; they subsequently designed and developed a line of abrasive-resistant pumps that could be operated safely for high pressure slurries at elevated operating temperatures. After a review of the bids from three pump suppliers, Lawrence’s pump was selected because their design best fit the specified criteria. Safety The pump is fully lined and uses abrasive-resistant high chrome iron for the high wear parts (casing, impeller and hub disc liner). The wet end of the pump is completely surrounded by a ductile cast steel outer casing that seals the liquid inside the pump. Even in the case of an inadverThe Pump Handbook Series

tent crack in the chrome iron casing, the pressure inside the pump is fully contained by the outer casing. The internal pump parts are specially designed to accommodate thermal expansion during the temperature differentials experienced during start-up and shutdown. Ease of Installation The steel skid base for the pump was designed so it would fit on the existing foundation and anchor bolts. Operating Procedures The only major change in operating procedures was to allow a short warm up period for the pump prior putting it on line with the hot process fluid. Reliable Spare Parts Based on wear life reliability statistics, accurate projections can be made for the supplier to maintain adequate spare parts. See discussion later in this article.

Lawrence Fully Lined Slurry Pump An additional feature designed into the pump is the back pullout design, common in API-design units. This enables removal of the pump’s internals without disconnecting the pump from the piping. In our case, this feature is not utilized because the entire pump is removed and taken to the shop for rebuild. In 1986, three of the existing pumps used in one of the digestion units were replaced with Lawrence 8 x 6 x 22 AL fully lined slurry pumps. In 1990, three more were installed in another unit. In 1999, the last three existing API un-lined pumps in this service will be replaced with Lawrence fully lined pumps.

Results The fully lined pumps using the high chrome iron have demonstrated a life 3.5 times longer than the existing cast steel un-lined pumps. The Lawrence pump has achieved

hrs. Because the spare pump is operated the same amount of time as the others, each pump operates 2/3 of the time. Therefore: 2/3 x 5,760 = 3,840 run hrs per year for each pump 3,840 run hrs/2,300 hrs to wear out = 1.67 pump rebuilds per year for Lawrence pump 3,840 run hrs/675 hrs to wear out = 5.688 pump rebuilds per year for un-lined pump

Figure 2. Lawrence fully lined slurry pump

more than 2,300 run hours to wear out in many cases—a great improvement over the 600-675 hrs for the un-lined pumps. The bottom line results are improved reliability and reduced maintenance costs. Figure 3 shows the wear life comparison of the last 20 wear-out events for the two different pumps.

Operating Considerations Using this pump in high temperature service requires special attention to start-up procedures. To prevent thermal shock of the hard chrome iron internal parts, it is necessary to heat the pump prior to start-up. To facilitate this pre-heating, we initially used live steam that was piped into the suction piping of the pump. Before starting the pump, the steam was slowly bled in and vented out the pump’s vents, which were located at the top and bottom of the pump casing. This enabled a slow warm-up before the pump was opened to the hot process fluid stream. We now accomplish the warm-up procedure by cracking open the suction valve to the pump to enable a small amount of the hot process fluid to preheat the pump. When the valve is opened, some of the hot process fluid flashes to steam and accomplishes the same effect as

having live steam bled into the pump. In the 13 years of operation of these pumps, parts have cracked due to thermal effects only twice. Both times the consequences were minor and the pump was still able to function. The cracked parts were only discovered during routine overhauls.

Wear Life and Maintenance Cost Evaluation To provide economic justification for replacing the existing un-lined pumps with fully lined chrome iron pumps we performed a maintenance cost analysis. The report was based on the comparative wear life between the two pumps and the total rebuild expense. The digestion unit operates 8 months per year, which is 5,760 run

The annual maintenance cost for the chrome iron lined pump was found to be 31% lower than the un-lined pump it replaced. This is true even though the hard chrome iron lined pump has a higher rebuild cost. With this kind of cost benefit, the replacement of the un-lined pumps with chrome iron lined pumps was easily justified.

Spare Parts Partnering The abrasive-resistant chrome iron liners for these pumps are not off-the-shelf parts. For this reason, careful consideration and planning must be given to spare parts management. To assure adequate spare parts availability and timely delivery, the wear-out life of the pumps should be established. For improved accuracy in pump wear life monitoring, actual run hours are preferred over calendar time. On a new pump installation, at least one set of spare parts should be ordered along with

Figure 3. Wear life of Lawrence pumps vs. existing pumps The Pump Handbook Series

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Figure 4. One year timeline projection for spare parts

or shortly after the pump order. This is especially important when the actual life expectancy is not already established. Having a set of the spare pump liners on hand will also cover possible part “infant mortality” during initial start-up. Once we established the average wear life of the pump liners, we created a timeline projection based on run hours to wear out of the pumps. We used the chart to determine when to order parts during the upcoming year (Figure 4). The chart illustrates a one year cycle starting with an assumed starting number of run hours on a group of six pumps starting in the month of February.

Pump Rebuilds and Parts Prediction Timeline In Figure 4, the numbers with the # symbol are those months in which pumps will require rebuilding due to wear out. Pump rebuild activity is accurately projected to occur in the months of January, February, June, July and August. The exact number of sets of parts required can be accurately predicted. In this case, 12 sets of parts are required for the year. Another important observation is that the parts requirements are not uniform throughout the year. A peak demand for parts occurs during January and February and another during July and August. This is very important. If we had ordered parts based on an “average” consumption, say one set per month, there would be severe “out of stock” crises in February, June and July.

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instead of released to the environment if the inner pump liners wear through. • The operating parameters of the pump in high temperature, high pressure services need to be clearly defined in sufficient detail and reviewed with the pump manufacturer to ensure that the pump is suitable for the intended service. • Special consideration needs to be given regarding spare parts supply management. With the use of accurate wear life statistics, accurate parts projections can be made that will facilitate parts stocking and partnering with the pump manufacturer. • These pumps—or pumps of similar design—should be considered as a viable pump in any highly abrasive service where high temperatures and pressures are involved. "

The number of parts that will be needed is reviewed with the pump manufacturer or his representative at least once a year. In our case, an agreement is also reached as to the number of rough un-machined castings the manufacturer will keep in stock at any given time. Having rough castings ready for machining shortens the lead-time for delivery, but it does not entail as high an inventory carrying costs as completely finished parts. This arrangement has worked extremely well and has mutual benefits for both the OEM pump manufacturer and the enduser. The user is assured that no crisis “stock out” situations will occur. The manufacturer knows what volume of parts the user will need for the upcoming year.

George C. Schmidt is a Senior Staff Mechanical Engineer with Aluminum Company of America at the Point Comfort, Texas Operations where he has held various engineering assignments including Maintenance Engineering, Project Engineering and Reliability Engineering. Mr. Schmidt received his BS degree in Mechanical Engineering at McNeese State University, Lake Charles, LA.

Conclusions

Disclaimer

• A properly designed slurry pump using the wear-resistant advantages of hard chrome iron can be successfully and safely operated in high temperature, high pressure services. • A substantial life improvement can be achieved with the chrome iron lined pump. The results of the case study show a life improvement factor of 3.5 over cast steel. • The total maintenance cost of the chrome-lined pump is lower, yielding substantial annual cost savings. • Improved safety is also realized because the liquid contents are still contained in the outer casing

The evaluations and conclusions reached in this case study were based on specific circumstances associated with a specific Alcoa operation and Alcoa makes no representations of suitability for service of the equipment described in this article for other industries or services.

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R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

The Importance of a Bearings Inspection Program Those “little” nicks and dings could lead to costly downtime. Catch them before they do major damage. By Jerry L. Sinclair, Weyerhaeuser Paper Company

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eyerhaeuser’s North Bend, Oregon container board mill was experiencing an 85% increase in spherical rolling bearing failure. Personnel studied the roll environment and increases in machine speeds, and they modified bearing housings. Despite best efforts, the rate of bearing failures continued to increase. As a last resort, operators began inspecting bearings and uncovered serious defects. The inspection program found three major categories for damaged bearings: manufacturing process damage, contamination and assembly damage. This article takes readers through a proper bearings inspection program with tips on what to look for and instruction on how to handle bearings properly.

How the Program Got Started The success story of our mill’s inspection program came about due to an accident. We were taking a bearing out of its box and were given a lesson in gravity. The bearing (weighing 800 lbs.) slid out of the box and landed on the millwright’s foot. While the millwright wasn’t seriously hurt, he did end up with a little free time on his hands. He began logging all the bearings that were removed from a roll. It was during this exercise that he noticed

our bearing usage had tripled in the last year. About this time our maintenance superintendent acquired permission to build a new maintenance building that would include offices for production, engineering and maintenance. This new facility would replace our antique-status shop. For the next few months we worked out of temporary buildings. In February 1990, as we were moving into the new building, discussions about the increase in bearing usage resurfaced, with one main difference: we were now ready to listen. Once the new facility was established, we went back through the bearing book and tabulated the number of bearings that failed from 1989 to 1991. (The bearing book included all large spherical roller bearing failures dating back to 1976. There was also a separate log of all the smaller rolls that were changed out of the paper machine, the position they were in and the bearings that were changed in place.) We were more than a little surprised at the actual figures—the increase in large spherical roller bearing failure was alarming. We presented the problem to our maintenance manager. Convinced by the evidence, he gave us the resources and personnel we needed, and in 1991 we started our in-house bearing inspection program. This set in motion a course of events not only The Pump Handbook Series

for the roll rebuild crew but for the bearing manufacturers (OEM) as well. One journeyman millwright began inspecting bearings full time. From April 1991 to April 1992 we inspected every spherical roller bearing in the storeroom, plus all incoming bearings. During this time we figured that we had inspected every brand of bearing made west of the Mississippi and 60% of the larger bore bearings. Product quality and control quickly became a priority for every manufacturer we worked with—otherwise they couldn’t sell their bearings to us.

Bearing Basics Now is a good time to review some bearing basics. Knowing them is essential to understanding our inspection program, why we looked for what we did, and why what we found concerned us so much. Bearings consist of four components: • Outer ring, which includes 1. outside diameter (OD) 2. outer race (internal portion of the outer ring that is the rolling element contact surface) • Rolling element 1. rollers (spherical, tapered, needle) 2. balls

straight,

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• Cage 1. brass (machined or stamped) 2. steel (stamped) • Inner Ring, which includes 1. the bore (internal diameter (ID) 2. the inner race (external portion of inner ring that is the rolling element contact surface)

Clean Shop Rules Once you understand the parts of the bearing, you can then put together some rules about how to handle them. Bearing life can be greatly extended by adhering to the following: 1. Work with clean tools in clean surroundings. 2. Remove all outside dirt from the housing before exposing the bearings. 3. Handle all parts with clean, dry hands. 4. Treat a used bearing as carefully as a new one.

5. Use clean solvents and flushing oil. 6. Lay out bearings on clean paper. 7. Protect disassembled bear-ings from dirt and moisture. 8. Use clean, lint-free rags if bearings are wiped. 9. Keep bearings wrapped in oilproof paper when not in use. 10. Clean the inside of housings before replacing bearings. 11. Install new bearings as removed from package, without washing. 12. Keep bearing lubricants clean when applying; cover containers when not in use.

Inspection Procedures Once clean shop standards have been developed and applied, you can move on to the actual inspection of the bearings. Before you even get close to the bearing, create a standard bearing inspection form

Figure 1. Bearing inspection form

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(Figure 1). This is the simplest way to ensure that all inspections cover the same items and check for the same defects, regardless of who is doing the inspecting. Ball Bearings Bearing inspection procedures, of course, differ for each type of bearing assembly. First, ball bearings. A shield stops the checking and inspection of many ball bearings. This component is installed to keep out contamination, but can also keep out inspectors. We have managed to destroy a few ball bearings in our attempts to develop an inspection method for them. They have basically the same inner and outer race damage and corrosion that spherical bearings have—it is just much harder to find. The following procedure has been developed after some extended trial and error, and can work for both shielded and nonshielded bearings.

1. Hold the outer race firmly in one hand, placing your other hand in or on the inner race. Hand position on the inner race depends on the size of the bearing. 2. Rotate one of the races while exerting pressure against it. Feel for roughness of any kind. Make sure to rotate the race in both directions! 3. Turn the bearing over and Photo 1. Commonly used inspection tools repeat Step 2. If you don’t detect any rough spots, you should have a good bearing. If a rough spot is detected, you have two options: 1. Refuse to accept the bearing and return it to the vendor. 2. Cut the bearing in half. You can do this by cutting the outer ring, ball retainer or inner ring in half. Once you’ve done this, you can inspect the inner and outer races. Photo 2. Large radius nick on roller

Unfortunately, if you do cut the bearing in half, you have just added the cost of the bearing to your operating cost. However, you now have proof of the damage or defects present in your bearings. Keep these damaged parts safe—you may need them to solve future disputes, not only with the distributor but possibly with the bearing manufacturer. Timken Bearings Those of you who use Timken type bearings may be wondering how the inspection procedure differs. In essence it is the same, in that you are looking for defects. Timken bearings have the same radius nicks as spherical roller bearings. Be careful, the manufacturer may tell you that the roller is designed for full-length surface contact and, therefore, will not be subject to the same problems as spherical elements. In reality, we (as the end-user) install the bearing into a gearbox or a rotating unit because it is a thrust control bearing. We then place the unit in operation. The next

tion involves more than just the actual inspection. There are several different types of “libraries” you must set up. Don’t get intimidated here. These libraries can be as simple as a three-ring binder. The first should contain every bearing manual, engineering bearing reference, failure chart and failure analysis book that vendors and OEMs are able to provide. It should also include blueprints and schematics regarding the rolls and other peripheral equipment. The second library is one you will most likely create yourself— an inspection form library. Some of our commonly used inspection materials are shown in Photo 1. These tools, along with the inspection form, will ensure that each technician follows the exact same procedure when inspecting bearings, and that the same information is recorded before it is irrevocably lost.

Our Findings time we are able to work on it is when it has failed and there is not enough of the bearing left to do a failure analysis. Inspection for defects before installation is absolutely critical with Timken bearings. Other Parts The same careful attention should be paid to shaft journals, housings, brackets and thermal growth. Remember that a large bearing will usually give you a lot of warning before a catastrophic failure—either visual or audible. A small bearing will usually fail without very much warning at all, until smoke and flame appear at the last moments before the machine shuts down.

Setting up a Bearing Inspection Program If you’re going to set up a bearing inspection program, there are several things you should consider. Making sure your bearings are in top condiThe Pump Handbook Series

Once the inspection program was instituted, we made several discoveries about the actual state of our bearings—some very surprising. The following list encompasses the majority of defects we found: 1. Radius nicks (Photos 2, 3 and 4). We originally called these “knife nicks,” as it appeared that something sharp in the production process caused the defect. After visiting four factories and observing their production process, we renamed the defect “radius nicks” because they match the radius of the end of the roller. The sorting machine is the last mechanical operation the roller goes through prior to final assembly. The sorter measures and sorts rollers in increments of only microns and ejects them onto the table. They then roll down and drop into a basket. The edge of one roller striking another already in the

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Photo 3. Radius nick in the center of our load zone

Photo 5. Assembly damage load scrape caused by excessive pressure on the roller

Photo 7. Pit or cavity in the roller

Photo 4. Radius nick on roller

Photo 6. Rust on roller

Photo 8. The shiny lips around the radius nick are a good example of how metal has moved during the time the bearing was in service. This bearing was used for six months.

2.

3.

4.

5.

6.

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basket appeared to be the cause of the damage. Dents. As the basket fills, the above radius nicks turn into dents. The more bearings landing on top of each other, the bigger the dent. Scrapes and scratches (Photo 5). These are another procedurerelated defect. Shipping scatter. The rollers rattling and rotating around between the inner and outer race during shipping cause these blemishes. Rust (Photo 6). Obviously, this is the result of moisture being introduced, either in the form of water or condensation. It can also be caused by a chemical reaction that takes place between two or more different base solvents. Pits (Photo 7). These cavities are created in the metal by the lathing process or during the grinding of the bearing. The pits will appear black if they existed before the hardening process began.

7a. Inner race damage, machined cage. Bearings that are assembled with machine cages have a loading slot (sometimes called a loading gate). All of the rollers are installed using this slot or gate. As the bearing is assembled, the rollers become tighter and the last few have to be forced into the case. Most assembling personnel use a mallet. If the roller is not struck squarely on the end, the edge of the roller dents

the inner race. The number of blows needed to install the roller or rollers clearly contributes to this problem. The more blows, the higher the likelihood of damage. If the components become extremely tight, the machined edge at the radius of the roller will gall the race surface. The pressure will be great enough to gouge into the surface and roll up the moved metal into a ball or series of small balls.

Photo 9. Formula for finding a shrink fit (Courtesy of Weyerhaeuser Paper Company) The Pump Handbook Series

7b. Inner race damage, pressed cages. Bearings that are assembled with pressed cages also have their rollers installed with a mallet or hammer. How much damage is caused when the struck roller collides with the inner race depends on the strength of the installer. Either way, this leaves a very disturbed surface for Table 1. Catastrophic bearing failures resulting in shutdown machine the roller to pass over. 8. Outer race damage. This is usually shipping damage. We affect the operation of the bearing. have received a few parts However, we took pictures of all the where the rollers have rattled defects and filed them along with the against the race hard enough to inspection form. We mailed the form start a gall. and the pictures to the manufacturers 9. Contamination. This comes in for their records as well. Note: always several forms: write down the exact numbers and let• Non-magnetic (brass partiters on the box and on the bearing cles, grindings and filings) itself. Don’t write down just one of • Magnetic (slivers of steel these two sets. The numbers on the from the steel cage roller box do not always match the stamp or socket, usually caused during etched numbers on the part. Don’t get the drilling process) caught with the wrong numbers later! • Sludge from the preservative In 1992 we were removing some of tank the previously inspected bearings to • Metal from a conveyor. In one check if what the manufacturer had case a piece of thin metal said (that running damaged bearings approximately 1⁄2 x 5⁄8 x .0005” was OK) was true. We cleaned and thick was stuck to a roller by inspected each one as if it were new. the preservative oil. We Using the original inspection form, we mailed the part back to the looked at the rollers with documented OEM. They had it analyzed damages. We photographed how the and determined that the metal defects had changed over time. We came from a transition piece found that the metal was becoming between the conveyor rollers. even more damaged during operation. • Dust, dirt and other debris. Most of the radius nicks were in This is usually caused by the line with the roller travel or at an shipping box breaking or angle across the roller face. The damhaving holes punched in it age left “lips” that circled the nick during handling and ship(Photo 8). This supported our theory ping. We have also found that a hydraulic action took place things like sunflower seed each time the nick filled with oil shells and other debris that from the grease and came into conresulting from carelessness. tact with one of the races while passWe documented all of the above on ing through the load zone. In late our inspection form and numbered 1992 we decided that we would no the cages of the rollers that had longer accept damaged bearings, and defects. Manufacturers told us that that running damaged bearings was most of the above defects would not not OK. These included: The Pump Handbook Series

• Any bearing with corrosion, either from rust or a chemical reaction • Bearings with defects on either race • Equipment with damage in either load zone. For all practical purposes, this became the central two thirds of the roller face. To make the process easier, we used black of the paper paint pencils to mark the end of the rollers with damage outside our load zone and yellow to mark damage inside the load zone before returning them to the manufacturer. Each roller was numbered on the cage with a lead or silver pencil. Believe it or not, we received some of those same damaged bearings back from the manufacturer during 1994 and 1995. Our markings were unmistakable. The program progressed well. We had established our library, procedures, forms and marking systems. The roll rebuild crew had been trained, and they now had all the reference materials they could need. We had completed the process of removing bearings on our smaller rolls that registered end thrust against both sides of the inner and outer races. During our search for the cause of this problem, we began looking at other possible problem areas, including: • bearing-to-journal fits • bearing-to-housing fits • housing-to-bracket fits • thermal growth (Our machine has a thermal temperature rise from about ambient to 300°F. Our rolls “grow” nearly 3⁄8”.) Some rolls are engineered so the bearings are held on both ends and the housing slides in the brackets. The bearing housings and brackets were supposed to have a surface finish smooth enough that the housings would slide in the brackets. We found the majority of our housing was roughcast and the corresponding

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brackets were superficially finished. We felt that this was the major contributing factor in the bearing failures in these rolls. We had the bearings analyzed and drew on the expertise of our bearing and lubrication representatives in an effort to determine if we were correct. At the same time, the roll rebuild crew measured the shafts and housings of our rolls. We quickly came to the conclusion that our roll journals were all at the OEM’s minimum recommended dimension. Many were undersized due to the number of bearings that had been removed. In my opinion, one of the most amazing things about machine shops is that if they are given a set of blue prints with the maximum and minimum tolerances for a given journal or shaft dimension, they seem to consistently finish the product at the minimum dimension. Unfortunately, all too often we established new, more rigid tolerances, only to be told by the manufacturer that it is an unreasonable request. This is what happened to us, until we were persistent and found a shop that would machine the equipment to our specifications. The bearing company’s engineering book will usually have a section on the minimum and maximum tolerances for shafts and housings that are to be used with their bearings. Our formula is half way, +.0003. Using this, we now set the dimensions for all our shafts and journals with a predetermined dimension of x amount of inches plus .0003 minus .0005. We figure the thermal growth it takes to install the inner race of the bearing onto the larger diameter shaft and make adjustments to our (now) recommended dimension if necessary.

Conclusion As you can see, the path to success is not an easy one. In the summer of 1992 we finally began to see some positive results from our efforts. We had rejected 98 bearings and there was a 32% reduction in wet end bearing usage. There was

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also a 45% reduction in dryer felt roll bearing usage. Of course the bearing company representatives asserted that the bearing inspection program alone did not account for these reductions, as we had made a considerable number of other changes, like the speed of our machines, and the journals, housings, brackets and grease. They were correct about this in the dryer section. However, on the wet end rolls, bearing inspection was the only change that had been made. May 1993 was the last time we had a catastrophic bearing failure on our paper machine. It was a small bore bearing that cost about $200. It shut our machine down for 16 hours at a cost of $100 per minute. The total loss in downtime was more than $96,000. From May 1993 to the present, our vibration analysis personnel have detected all the bearings that were on the verge of failing prior to failure. All were changed on scheduled downs without any production upsets. The ability to detect a problem, analyze the data and predict a failure time frame for any rotating piece of equipment is one of the most valuable assets a company can have. I would be remiss if I didn’t mention the effect predictive maintenance such as this has had on our bearing successes. A bearing inspection and vibration analysis program can be a company’s silent safety partner. Table 1 shows the reduction in failures that we were able to achieve. We may not have totally eliminated our bearing problems, but for a machine with 225 rolls in operation, we have made a significant improvement. The commitment to a program such as this cannot be taken lightly. It will take at least three years to realize any benefits, plus five years or more to realize the full cost savings. As you can see, there is no quick and easy fix for bearing problems. A company that decides to take on the endeavor of a bearing inspection program must go into it with a firm commitment, and keep a few The Pump Handbook Series

things in mind. Management must ask itself the following questions: 1. Are we ready to commit our maintenance department to this program? 2. Will we support training for maintenance personnel? 3. Will we support our inspection personnel and their requirements in equipment, space, etc.? 4. Can we establish a clean room/ area? 5. Are we prepared to support the necessary documentation program? The accounting department may object to this because of the added paperwork. One suggestion is to order each bearing on a separate purchase order. This will simplify returning defects to the manufacturer. Let’s face the truth. The purpose in producing a product is selling as much of that product as possible. In all honesty, companies like ours only need one manufacturer to address our concerns by correcting their processinflicted damages. If a company were willing to do this, it could become the preferred manufacturer in many industries. Maybe our competitors can keep using bearings from other companies—so we can keep our edge. In today’s marketplace our customers are demanding a zero defect product. In order to accomplish this, we must demand a zero defect product from the companies that supply our equipment and its components. ■ Jerry L. Sinclair is a Millwright at Weyerhaeuser Paper Company’s Bend, Oregon facility. He has more than 39 years of experience in maintenance procedures for pulp and paper processing. In recent years he has developed a bearings inspection and handling program that has been the subject of training sessions at Weyerhaeuser’s Bend, Oregon and Valiant, Oklahoma mills, as well as a feature article in the company’s Maintenance Journal. In addition, Mr. Sinclair presented papers at the Oregon Pulp and Paper Safety Conference in both 1996 and 1997.

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R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

Coupling and Alignment Strategies Choosing the right coupling and keeping your machines aligned can increase reliability and reduce maintenance headaches. By Al Ciotola and Tanveer Khan, Frontline Industries

haft couplings and alignment are critical to many kinds of rotating equipment. The efficiency, smooth operation and mean time between failure (MTBF) of your equipment can greatly depend on the couplings you use and how well you line everything up. Hence, great care should be taken during coupling selection, installation, alignment and maintenance.

S

Coupling Classification Shaft couplings can be broadly classified into two groups: rigid and flexible. Rigid couplings are used in applications where there is absolutely no possibility or room for any misalignment. They are also used in vertical applications where the shaft of the drive has to support the weight of the driven shaft. In this article we will discuss flexible shaft couplings and their alignment strategies, as these are more prone to selection, installation and maintenance errors. Flexible shaft couplings can be divided into two basic groups: elastomeric and non-elastomeric Elastomeric couplings use either rubber or polymer elements to achieve flexibility. These elements can either be in shear or in com-

pression. Tire and rubber sleeve designs are elastomer in shear couplings; jaw and pin and bushing designs are elastomer in compression couplings. Non-elastomeric couplings use metallic elements to obtain flexibility. These can be one of two types: lubricated or non-lubricated. Lubricated designs accommodate misalignment by the sliding action of their components, hence the need for lubrication. The non-lubricated designs accommodate misalignment through flexing. Gear, grid and chain couplings are examples of non-elastomeric, lubricated couplings. Disc and diaphragm couplings are non-elastomeric and non-lubricated.

Coupling Selection Torque The first consideration when selecting any coupling is its torque carrying capacity. The normal torque a coupling has to transmit is a function of horsepower and rpm and is calculated as follows: nominal torque = hp x 63025 (in-lb) rpm A coupling should be able to transmit peak torque that might The Pump Handbook Series

result due to various causes such as start-ups, sudden, cyclical, fluctuating or reversing loads. A service factor is incorporated for this purpose. Almost all coupling manufacturers provide the values for service factors for different types of drive and driven equipment and different service conditions. The design torque of a coupling is determined by multiplying the nominal torque by the service factor: design torque = nominal torque x service factor It is important to note that two service factors must be considered: one for the drive and one for the driven equipment. Sometimes service factors are listed independently for the drive and the driven equipment. Add these together to establish the final service factor for a given application. For example, the service factor for a two cylinder single acting reciprocating pump is 2.0. The service factor for a four cylinder gasoline engine is 0.5. Therefore, the service factor for the coupling that connects these two units is (2.0 + 0.5) = 2.5. An electric motor, since it produces torque that is “non-fluctuating” (constant in time), will have a service factor of zero.

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Choose a coupling that is at the least rated for the design torque. Size The next step is to check whether the selected coupling can accommodate the bigger of the drive or driven shafts. If not, select the next available size that can accommodate this shaft size.

Balancing Requirements Balancing of couplings assumes special importance for high-speed machines. At high rpms, even a slight unbalance can create very high radial loads. These result in vibrations and ultimately can cause damage to the equipment. In recent years, many industries have implemented more stringent requirements for the balancing of couplings. The two most followed standards for the balancing of couplings are A.G.M.A (American Gear Manufacturers Association) 9000C90 and API (American Petroleum Institute) 671. The A.G.M.A balancing limits refer to the maximum potential unbalance as a function of coupling weight and operating speed. Machine sensitivity to coupling unbalance is established and classified as low, average or high, depending on various criteria such as shaft flexibility, bearing loads, machine and foundation rigidity, resonance, coupling length and shaft configuration. The API standard refers to couplings used in refinery applications and relates to the maximum potential unbalance allowable. The potential unbalance is the maximum unbalance recorded after the coupling is disassembled and reassembled several times. This potential unbalance shall not be larger than any of the following limits: U  4W/rpm U  0.008 W U  0.1

Where W=weight of the coupling The value of the unbalance is measured in ounce-inches. Installation Once the selection of a flexible coupling for a specific application has been made, the next step is to install it properly. Before proceeding with the installation, follow these steps: • Verify that the right type of coupling was ordered and that you have all the parts. • Measure the shafts of the drive and the driven equipment, including keyways, keys, tapers and distance between shaft ends. • Measure the coupling hub bores, keyways, tapers, outside diameter and spacer length. • It is always a good idea to assemble the coupling prior to installation to ensure that all components are present and fit together just right. • Verify the proper fit for both hubs, such as taper, interference or clearance fits. Table 1 gives the values for three types of fits for various shaft sizes:

Shaft Dia.

Clearance Fit Class I Class II

∆T = i/α(d-0.002) Where: ∆T = rise in coupling hub temperature from ambient in °F i= interference fit (thousandths of an inch) α = coefficient of thermal expansion d = coupling hub bore diameter in inches For steel couplings, roughly a

Interference Fit

Shaft Dia.

Clearance Fit Class I Class II

Interference Fit

0.500

0.500-0.501

0.500-0.502

0.4990-0.4995

2.375

2.3750-2.3765

2.375-2.377

2.373-2.374

0.625

0.625-0.626

0.625-0.627

0.6240-0.6245

2.500

2.5000-2.5015

2.500-2.502

2.498-2.499

0.750

0.750-0.751

0.750-0.752

0.7490-0.7495

2.625

2.6250-2.6265

2.625-2.627

2.623-2.624

0.875

0.875-0.876

0.875-0.877

0.8740-0.8745

2.750

2.7500-2.7515

2.750-2.752

2.748-2.749

1.000

1.000-1.001

1.000-1.002

0.9990-0.9995

2.875

2.8750-2.8765

2.875-2.877

2.873-2.87

1.125

1.125-1.126

1.125-1.127

1.1240-1.1245

3.000

3.0000-3.0015

3.000-3.002

2.998-2.999

1.250

1.250-1.251

1.250-1.252

1.2490-1.2495

3.250

3.2500-3.2515

3.250-3.253

3.2470-3.2485

1.375

1.375-1.376

1.375-1.377

1.3740-1.3745

3.500

3.5000-3.5015

3.500-3.503

3.4970-3.4985

1.500

1.500-1.501

1.500-1.502

1.4990-1.4995

3.625

3.6250-3.6265

3.625-3.628

3.6220-3.6235

1.625

1.625-1.626

1.625-1.627

1.623-1.624

3.750

3.7500-3.7515

3.750-3.753

3.7470-3.7485

1.750

1.750-1.751

1.750-1.752

1.748-1.749

4.000

4.0000-4.0015

4.000-4.003

3.9970-3.9985

1.875

1.875-1.876

1.875-1.877

1.873-1.874

4.500

4.500-4.502

4.500-4.504

4.4965-4.4980

2.000

2.000-2.001

2.000-2.002

1.998-1.999

5.000

5.000-5.002

5.000-5.004

4.9965-4.998

2.125

2.125-2.1265

2.125-2.127

2.123-2.124

5.500

5.500-5.502

5.500-5.504

5.4960-5.4975

2.250

2.2500-2.2515

2.250-2.252

2.248-2.249

6.000

6.000-6.002

6.000-6.004

5.9960-5.9975

Note: All couplings supplied with Class I clearance fit unless otherwise specified

Table 1. Fit values for various shaft sizes

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• class I clearance fit, which is a snug fit • class II clearance fit, which is an easy slide fit • interference fit A coupling with an interference fit should preferably be installed by heating it in an oil bath or an oven to approximately 200-250°F, and, in some cases, cooling the shaft simultaneously with dry ice. Once the interference fit has been determined by measuring the shaft diameter and coupling hub bores, the maximum temperature increase needed to expand the coupling hub to exceed the shaft diameter by 0.002”, which is required for a slide fit, can be calculated by the following equation:

The Pump Handbook Series

160°F temperature difference is required between hub and shaft for every 0.001”/inch of interference ratio.

Maintenance Considerations ed

Once the right coupling is selectand properly installed, it

becomes the job of the maintenance department to ensure that the longest useful life is achieved for each coupling. The most obvious and perhaps most beneficial maintenance is to periodically check the alignment to make sure it remains within the allowable limits.

Figure 1. Parallel misalignment

Maintenance for Lubricated Couplings Lubricated couplings require a periodic inspection to ensure that the proper amount of grease is present and in good condition. The seals that hold the grease in place also need to be inspected and replaced if found defective. The frequency of the re-lubrication schedule depends on the type of grease used, the rpm and the degree of misalignment. Whenever possible, a “coupling grease” should be used. In most cases, a lubricated coupling manufacturer will provide this specialized grease for their product. However, there are other companies and private individuals that have developed and patented greases uniquely suitable for lubricated couplings. The choice of the proper grease will make a huge difference in the performance of the coupling and save the user a great deal of time and money. Maintenance for Non-Lubricated Couplings Non-lubricated couplings do not need periodic maintenance, as they accommodate misalignment through the flexing of their elastic elements. The elements, though, do need to be inspected for signs of aging, weathering, fatigue and corrosion. In many cases, even when the coupling guard is present (such as expanded wire mesh or guard with an inspection port), a visual inspection can be performed with the help of a strobe light.

Figure 2. Angular misalignment

Coupling and Machinery Alignment When the shaft axes of two or more rotating machines are collinear, i.e., in the same exact line, they are said to be in perfect alignment. Any deviation, either in a parallel or angular direction, or a combination of both, results in misalignment. The deviation could occur in horizontal and/or vertical planes (Figures 1, 2 and 3).

Figure 3. Compound misalignment The Pump Handbook Series

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Any misalignment between the drive and the driven shafts creates undesirable reactionary forces that adversely affect the shafts, the bearings, the seals and the couplings. It shortens the useful life of the equipment and results in poor performance and premature failures. If the machines are in perfect alignment, the only force transmitted through the coupling is a pure torque. The shafts of the drive and the driven machines do not see any additional bending or shear stresses, except for the torsional shear stress due to the transmission of torque. This means that bearings do not have to carry undesirable additional loads. But in actual practice, perfect alignment is rarely achieved. In certain cases, it is not even desirable. Slight misalignment is beneficial in gear or grid coupling designs to maintain the film of lubricant between the points of power transmission during rotation. Alignment should be done such that the machines will be within allowable alignment tolerances under actual operating conditions. Thermal growth and any other potential movement of machines, either in a vertical or horizontal plane, should be taken into consideration. Alignment should always be checked again after the machines have achieved their operating temperatures. Most machinery manufacturers provide the misalignment tolerances required for their equipment. Some plants develop their own misalignment tolerance standards based on their experience and specific needs. If no such data is available, misalignment tolerances published by a leading manufacturer of laser alignment systems can be used for closecoupled, general-purpose machines like centrifugal pumps and motors. (See Table 2.) Why Align Machines? The objective of proper shaft

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alignment is to increase the operating life of the machines. Misalignment results in early and frequent failures of shafts, bearings, seals and couplings. Good alignment will not only increase the life of these parts, it will also reduce power consumption, noise and vibration levels. One can imagine the savings due to a good alignment in terms of increased MTBF, reduced downtime and replacement costs for worn or damaged parts. Symptoms of Misalignment Misalignment shows itself in a number of ways: 1. high rate of failure of shafts, bearings, seals or couplings 2. high level of vibration 3. bearings and couplings running hot 4. excessive leakage of oil through the bearing housing 5. leakage of grease in grease packed couplings 6. broken or loose foundation bolts on the machines 7. broken or badly worn parts in couplings 8. dust generated from rubbing of polymer or rubber parts of elastomeric couplings

What Is Required to Do A Good Alignment? To do a good alignment you will need the proper tools, adequate training and sufficient time. Alignment Tools Laser alignment systems are the most modern, accurate and convenient tools available for the alignment of rotating machines. They are obviously very expensive, but if you have a good number of machines to be aligned in your plant, the amount of time these might save will make them worth considering. A dial indicator with a bracket alignment system is another inexpensive, very popular and effective shaft alignment tool. Besides the actual alignment equipment, the following tools and accessories will go a long way toward achieving a quick and accurate alignment: • precision precut stainless steel shims in various thickness and sizes • jack bolts for horizontal and vertical moves • hydraulic jacks if no jack bolts are available for vertical moves • Chicago bolts to overcome bolt bound conditions Use of hammers and crowbars to move the machines should be avoid-

SPEED (RPM)

PARALLEL MISALIGNMENT

ANGULAR MISALIGNMENT (PER 10” DIA. OF COUPLING)

600

within 0.005” to 0.009”

within 0.010” to 0.015”

900

within 0.003” to 0.006”

within 0.007” to 0.010”

1200

within 0.0025” to 0.004”

within 0.005” to 0.008”

1800

within 0.002” to 0.003”

within 0.003” to 0.005”

3600

within 0.001” to 0.0015”

within 0.002” to 0.003”

7200

within 0.0005” to 0.001”

within 0.001” to 0.002”

Table 2. Misalignment tolerances The Pump Handbook Series

movements of machines relative to each other • check the alignment again after the machines have reached their operating temperatures • do a periodic alignment check • keep records of alignment readings for future reference (the readings will help you in understanding the behavior of the machine over a period of time) Remember that the time and effort spent in doing a good alignment is going to pay you back several times over. ■

Figure 4. Pump and motor alignment (top view)

ed as much as possible. It is very difficult to precisely control the movement of machines. Use jack bolts instead (Figure 4). To get the best results and long, trouble-free service from the coupling, the drive and the driven equipment should be aligned as close to the specifications as possible. Training Training of the personnel responsible for doing the alignment is very important. They have to be well versed in using the alignment tools. They should also be aware of the tolerances for various operating conditions, how to correct soft foot conditions, how to compensate for thermal growth and other movements of machines. It’s also necessary to provide these personnel with all the necessary data, such as how much the machines move relative to each other from their offline position. And most important, everyone concerned with the efficient and profitable running of the plant, including managers, engineers, foreman and mechanics, should be aware of the importance of good alignment. Sufficient Time Shaft alignment is an intricate and precise job. To do a good job, suffi-

cient time should be made available. On an average, it will take two people about four to five hours to align a 50 to 500 hp motor to a pump, provided everything goes just right. Availability and use of the right tools will greatly help in getting alignment right in the shortest possible time and with the greatest accuracy. Only in case of a real emergency or a technical reason like a bolt bound condition or excessive pipe strain preventing a proper alignment should machines be left in a misaligned condition. However, they should be stopped as soon as possible and properly aligned.

Al Ciotola is the President and Owner of Frontline Industries, Inc. He is a graduate of Naples Institute of Technology in Industrial Mechanical Technology and has more than 20 years of experience in the servicing of rotating machinery. Mr. Ciotola owns several patents in the field of couplings and mechanical seals. Tanveer Khan is a graduate in Mechanical Engineering from the University of Bombay. He has 18 years of experience in design and development of mechanical shaft seals and flexible couplings. He is working for Frontline Industries, Inc., as a design engineer.

Alignment Checklist • check the foundation for correct size and condition • check whether the grouting is proper and free of voids • check the base plate for rigidity, leveling, cleanliness and coplanar mounting pads • check for pipe strain (piping should be flexible, well fitted and with adequate support) • clean mounting base, shims and feet of machinery to be moved • rough align the equipment with the help of a straight edge • correct soft foot if any • take into consideration thermal growth and any other future The Pump Handbook Series

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✓

R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

Building A Better Foundation By Scott Bullentini, ITW/Escoweld Epoxy Grout Systems

entrifugal pumps usually instances, the root cause of increases as the cementitious grout lead the list of failure-prone pump/motor failures can be attribbegins to fail. Often the emphasis is equipment in refineries and uted to the poor or improper instalplaced on the symptoms rather than major chemical plants. In lation of the baseplates or soleplates. the cause. Expensive seal designs many instances, this leads to The use of a standardized installation and bearing and coupling configuraunscheduled downtime and costly specification for pumps and drivers tions to “handle” misalignment are repairs. Costs range from industry to is one method of eliminating this employed. The root cause, in many industry. However, one fact is clear: common problem. By improving the instances, is the inadequate installalost production time coupled with pump baseplate installation and tion of the baseplate or soleplate couexpensive repair costs are neither grouting procedures or standards: pled with the use of cementitious desirable nor acceptable in the cost• bearing and seal failures can be grout (instead of epoxy grout) in the driven atmosphere of today’s modreduced original installation. ern industrial production facility. • mechanical alignment can be Why Epoxy Grout? From an end-user’s standpoint, the achieved easier and maintained A pump is part of a system, and challenge becomes how to avoid longer part of that system is the foundation. costly unscheduled downtime, • vibration can be reduced Therefore, it only makes sense to use expensive repairs and low Mean a material that is better suited for Time Between Repairs (MTBR). Many plants employ “generic” dynamic loads. Why is it more World-class or benchmark values installation procedures for baseplate advantageous, then, to use epoxy for machinery reliability and MTBR installation, especially “ancillary” or grout instead of cementitious grout are relative and can vary from comlower horsepower equipment. for baseplate installations? This pany to company, and even from Typically, cementitious grout is used question can be answered, in part, industry to industry. Regardless of to satisfy grouting requirements. by comparing key physical characwhere you are currently, though, These applications will last for a teristics (compressive and tensile there is always room for improveshort time, but the evidence is overstrengths and tensile bond to basement. You have two choices: 1. whelming when a comparison is plates and soleplates) of epoxy grouts Spend money making improperly made between maintenance dollars and cementitious grouts (Table 1). installed equipment right, or 2. spent and cementitious grout failure. The major difference shown in spend money up front with the purThe frequency of pump overhauls Table 1 is the pose of getting it greater tensile right the first time strengths and comand building in Cement-based grouts Epoxy grouts pressive strengths reliability from the Compressive strength 6,000 psi 14,000 psi of epoxy grouts vs. start. Tensile bond to steel 0 2,000 psi cementitious Often overTensile strength 700 psi 2,000 psi grouts. looked but of sinThe objective of gular importance is *Compressive strengths and tensile strengths are expressed as an average of five different grout grouting is to crethe installation and manufacturers. ate a single block grouting of basemonolith, one that plates and solewill act as an effecplates. In many Table 1. Compressive strengths of cement and epoxy-based grouts

C

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The Pump Handbook Series

tive vibration dampener. Properly installed, an epoxy grout will bond to a concrete foundation with a tensile strength greater than that of the concrete foundation. More important, the bond strength of an epoxy to a properly prepared baseplate will be, on average, between 1,600 psi and 2,100 psi. In many instances, the use of cementitious grouts leads to voids because it cannot provide a perfect non-shrink bond to steel. Due to over-watering of cementitious grouts, “settlement shrinkage” occurs (the inability of the cement particles to stay in suspension), which also leads to voids. Voids lead to potential baseplate resonance, which, over time, compromises the installation and in many instances leads to premature seal and bearing failure.

Case History – Epoxy Grout vs. Cementitious Grout The two 15 hp circulating pumps shown in Photo 1 were re-grouted using an epoxy to correct alignment and vibration problems that resulted in premature seal and bearing failure. The baseplates were a simple fabricated steel design and were reused. Cementitious grout was used

Figure 1. Results of the application shown in Photo 1

in the original installation. Over time, the cement grout had come loose from the foundation and was cracking, allowing the baseplate to move and resonate. The pumps were moved and re-grouted one at a time. The results of this application are shown in Figure 1, which illustrates the vibration characteristics of the pumps before and after grouting with the epoxy. The improvement was dramatic.

Standardized Installation Procedures: How to Develop One That Fits Your Needs Obvious to this article’s intent is the implied use of epoxy grouts

Photo 1. These pumps were re-grouted using an epoxy to correct alignment and vibration problems. The Pump Handbook Series

rather than cementitious for grouting of critical equipment. The next step after deciding to use epoxy is to develop a standardized installation procedure that is tailor-fit to your particular industry and pumping applications. Requirements for grouting and machinery baseplate installations are going to be unique to each piece of equipment and each one’s operating environment. Factors such as loading and operating temperatures will play a big role in the decisionmaking process. In general, several key points should provide the framework of a good “operationally consistent” installation specification—one that will ensure a longer life for your equipment. The specification should comprehensively detail all aspects of: • grout selection criteria • baseplate design and preparation and installation • foundation design and preparation • epoxy grout installation requirements • form fabrication • testing requirements • alignment procedures Grout Selection Criteria To adequately complete this area of specification approval, it should be understood that all epoxy grouts are not created equal. Each manufacturer has a unique attribute associated with their product. Therefore, a little research is in order. This first step in determining which grout(s) will best serve your application is to request that several manufacturers perform flow box evaluations at your facility (Photo 2). Several determining factors can be readily identified through this type of testing: • flowability • heat of exotherm (how hot does the grout get after mixing and placement) • effective bearing area (contact area with baseplate) • consistency/working time • compressive strength under plate

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American Society for Testing Materials (ASTM) C 1181 - 91 C 531 - 85 (90) C 579 - 91 C 321 - 88 C 884 - 87 D 307 - 88 D 2471 - 88 D 696 - 91

Test method for compressive creep of chemical-resistant polymer machinery groups Test method for linear shrinkage and coefficient of thermal expansion of chemical resistant mortars, grouts and monolithic surfacings Test method for compressive strength of (Method/B) chemical resistant mortars and monolithic surfacings Test method for bond strength of chemical resistant mortars Test method for thermal compatibility between conrete and an epoxy-resin overlay Test method for tensile strength of chemical resistant mortars, grouts and monolithic surfacings Test method for gel time and peak exothermic temperature or reacting thermosetting resins Test method for coefficient of linear thermal expansion of plastics American Petroleum Institute (API)

API 610, 7th Edition Centrifugal pumps for general refinery service API 686, 1st Edition Recommended practices for machinery installation and design

Photo 2. Flow box evaluations

Once you have determined which grout manufacturers meet your installation needs, the next step is to develop a list of the applicable ASTM (American Society for Testing Materials) tests that can be listed as standards for selection criteria. Some general criteria are listed later; however, each selected grout manufacturer should supply you with specific lab test data to use as your selection criteria (Table 2). The specification section that deals with selection criteria and standards should also incorporate a section detailing post grout pour testing procedures. Most epoxy grout manufacturers can provide testing procedures for your use.

American National Standards Institute (ANSI) Y 14.5 - 73

Dimensioning and tolerancing Steel Structures Painting Council (SSPC)

SSPC SP6

Commercial blast cleaning specification

Table 2. Samples of standards used for criteria when selecting a grout

for all equipment mounting pads to be co-planar level to within 0.002”. A typical fabricated steel baseplate designed for epoxy grouting is shown in Photo 3. The baseplate should be designed with four lifting lugs—one at each corner. This enables the lifting of the

Baseplate and Soleplate Design Criteria Fabricated steel baseplates are replacing older designs, especially those that incorporated cast iron baseplates in the original installation. The reason for the additional cost of a fabricated steel baseplate is simple: it is more versatile. By utilizing jack screws and anchor bolts for adjustment, the baseplate can be “worked” until the machined mounting surfaces (where the pump and driver mount) are level and coplanar to a very tight tolerance. Most installation specifications call

438

Photo 3. Typical fabricated steel baseplate The Pump Handbook Series

complete weight of the equipment and base. It should be fitted with one 6” grout fill hole uniform3 ly distributed for every 9 ft of baseplate. Vent holes no smaller than 3⁄8” should be provided for and located at all corners of any compartmentalized areas and any high points

Photo 4. These pumps were originally installed on cast iron baseplates.

Photo 5. Payback on investment for pumps shown in Photo 4

and perimeter areas where angle iron or a “C” channel is located. Provisions for vertical jacking bolts 5⁄8” diameter minimum should be made, and the jacking bolts should be next to each anchor bolt location. All bulkhead cross bracing on the underside of the baseplate should have a 2”x6” opening to allow grout to flow from each bulkhead to the next. All outside edges of the baseplate flanges (in other words, all outside edges of the baseplate embedded in the grout) should be radiused to a minimum of 1”. This eliminates stress risers from occurring in the grout.

new concrete pads were poured and subsequently prepared for epoxy grouting. The pumps were then reinstalled on fabricated steel baseplates instead of cast iron baseplates. Overall vibration levels for the installation were reduced from 0.7 ips to less than 0.15 ips overall. The initial alignment was easily accomplished; each unit took about 45 minutes (level machine mounting surfaces are co-planar, significantly reducing the amount of time it takes for an alignment job). The total cost, which included relocation, piping modifications, pump overhaul and grouting was approximately $40,000. Payback on the investment was in just over 12 months, as the MTBF increased from .25 to 8 years and still counting (Photo 5).

Foundation Surface Preparation and Design Foundation Design A machinery block foundation supported on soil should have a mass ratio of three times the mass of the machinery for centrifugal applications and five times the mass of the equipment for reciprocating applications. The foundation should be of sufficient width to prevent move-

ment and adequate depth to allow for proper embedding of anchor bolts. A general rule of thumb is that the width of the foundation should be at least 1.5 times the vertical distance from the top of the base to the machine centerline. It is also important to design the foundation with an allowed distance above the finished elevation of the floor slab or grade to prevent damage to machinery from run-off and/or wash-down water. Foundation Surface Preparation The old adage of “you don’t paint over rust” applies when it comes to surface preparation for grouting. Both the foundation and baseplate need to be properly prepared to promote and maximize the excellent bonding properties of the epoxy grout. Freshly poured concrete must be allowed to cure before epoxy grout is applied. Typically, 4,000 psi at 28 days compressive strength is considered adequate strength and cure for new concrete. It is an excellent practice to run an ASTM test for concrete shrinkage to test for moisture content and shrinkage (ASTM C 157 – 80). This test will indicate the end of the chemical reaction of the cement and water that causes

Case History – Baseplate Design Changes Four de-ionized water pumps were originally installed on cast iron baseplates (Photo 4). Maintenance costs over the years had risen to approximately $8,000 per pump per year. The justification for improving the installations was obvious in the failure data gathered by reliability engineers. The pumps were removed,

Photo 6. Hand-held pneumatic chipping hammer The Pump Handbook Series

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best link for the grout (Photos 7a and 7b). Foundation anchor bolts and sleeves should be located to the specified tolerance of the machinery. They should be securely supported to prevent misalignment when the concrete is placed. Be that certain all anchor bolt sleeves are dry and all foreign materials (dust, debris, etc.) have been removed. Also check that all anchor bolt sleeves are filled with a non-binding material such as “spray-in insulation” foam. This will prevent attack from moisture in the period of time that passes prior to grouting. The exposed anchor bolt threads should be wrapped with 2-3 layers of duct tape. This will prevent the grout from bonding to the anchor bolt after the grout has cured and prevent damage to the anchor bolt during foundation surface preparation. These steps are taken to ensure that the anchor bolts are allowed to stretch during final torque and achieve the desired clamping force. Remember that the grout is only acting as a load-bearing and vibration-dampening material; the anchor bolts still must be allowed to stretch and “hold” the machinery to the foundation.

Photos 7a and b. Examples of a good surface profile

the concrete to cure. The general rules of thumb to use if a shrinkage test is not run are: • standard concrete (five bag mix) – 28 days • hi-early (six to eight bag mix) – 7 days Surface preparation can begin after it has been determined that the concrete has cured. Normally, the top 1⁄4” to 1⁄2” of a finished concrete foundation will

440

contain less aggregate and more sand/cement mixture. This portion of the foundation is referred to as laitance and needs to be removed, as it is weaker than the aggregate-filled foundation. This is accomplished by using a hand-held pneumatic chipping hammer with a bushing or chisel bit (Photo 6). A good surface profile will expose a high percentage of the course aggregate. This is the strongest part of the foundation and provides the The Pump Handbook Series

Baseplate Surface Preparation Ultimately, baseplate surface preparation is what can “make or break” a grout job. Just like the foundation, the surface of the underside of the baseplate needs to be prepared with the utmost attention to detail. Typically the OEM will prepare the baseplate for grouting. However, unless otherwise specified, the baseplate will be prepared for cementitious grouting. For purposes of discussion, the baseplate needs to be free of all oils, rust, dirt and foreign materials. It should be sandblasted a minimum SSPC SP-6 profile (Commercial Blast Cleaning) and then coated with an inorganic zinc silicate coat-

ing or a “grout-compatible” two-part epoxy rust inhibitor primer. If the baseplate can be grouted within eight hours after sandblasting, a primer is not needed. However, the baseplate still needs to be sandblasted to an SSPC SP-6 profile. The purpose of priming baseplates is to protect them from damage and/or rust development during shipping and storage. Baseplate Installation Let’s think for a moment of the grout as an adapter that will match the irregular shape of the pump baseplate to the top of the irregular (properly prepared) concrete foundation. The ideal pump baseplate installation is one that is 100% supported by the grout (all leveling devices removed). This can be achieved by using the method shown in Photo 8 (jack screws), which allows the millwright to level the machine mounting surfaces with a high degree of precision. As mentioned earlier, level tolerances of 0.002” co-planar are more easily achieved using this method. After the grout has been poured and allowed to cure, the jack screws can be removed. (The jack screws are waxed or greased for easy removal after grout cure.) Now the baseplate is supported entirely by the epoxy grout, removing problems normally associated with point loading. After the jack screws are removed, the holes are degreased and filled with seam sealant or liquid-only epoxy. To level the baseplate, I recommend using a Starrett 98 machinist level or equivalent and that the baseplate be leveled with the equipment removed. This makes it easier to access the machine mounting surfaces and facilitates a better grout job, because the grout holes are easier to get to. Use an optical level (K&E Model 71-3015 or equivalent) on baseplates in excess of 15’ in length.

Form Fabrication Criteria Grout forms should be constructed from 3⁄4” thick plywood. The plywood should be braced on approximately 12” centers with 2”x4” material. All wood to be in contact with the epoxy must be coated with a minimum of three coats of Johnson Paste Wax. (Of course, remember that epoxy grout has tremendous bonding properties. A good rule of thumb is the person who installs the forms must remove them (Photo 9). All vertical and horizontal corners should have a 45° chamfer strip installed at the final grout level to “break” all sharp edges. The completed forms should be water-tight and sealed with silicone to prevent grout from leaking at the forms during placement. Epoxy Grout Installation Criteria Prior to pump baseplate grouting, a series of checks should be made. It is common to see checklists at the end of installation specifications. The checklist serves as a work audit and verifies that all specified procedures have been followed and met. With this in mind, the following checks should be made:

• baseplate underside is free of oil, dirt, and moisture and is adequately primed and/or prepared • concrete surface is clean, free of oil, dirt and moisture • anchor bolts are protected with duct tape or other non-bonding protection • jack screws are lubricated for removal • forms in contact with grout are properly waxed • grouting materials are in unopened containers, dry and have been stored at temperatures above 50°F and 60°F (The general rule of thumb for equipment and material storage is based on good old-fashioned common sense. All epoxy grout should be stored indoors and kept dry, free of moisture and in the original shipping containers. Storage should be maintained between 50ºF and 80ºF. The materials should be kept between these temperature ranges for a minimum of 48 hours prior to mixing and placement.) • sufficient quantities of grout are on hand at the site to complete the pour (add 25% to the calculated amount needed)

Photo 8. Baseplate installation using jack screws The Pump Handbook Series

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With the temperature-conditioned grout at the pump site and all checklist items completed, it is finally time to mix and pour grout. Epoxy grouts consist of a resin, a converter or catalyst, and aggregate. The liquid resin and converter/catalyst are mixed together first in their shipping containers. Mixing is accomplished with a motorized drill and “jiffy type” paint mixer. Mixing for approximately three minutes allows the resin and converter/catalyst to react and bond. At this point, add the aggregate. The aggregate acts as a filler and a heat sink and affects the flowability of the material and creates some of the physical properties such as compressive strength, flexural strength and modulus of elasticity. Method of Mixing Mixing of the three components is accomplished best via the use of a mortar mixer (Photo 10). The liquids are added first, then the aggregate is added. Mixing is complete when all the aggregate is completely introduced or wetted in the resin system. Do not over-mix epoxy grouts, as this introduces unwanted air and may translate to voids as the air pockets rise to the top.

Photo 9. Grout form

Placement Methods Epoxy grouts should be placed from one end of the baseplate to the other. Starting in the middle should be avoided. By starting on one end, head pressure is generated via the use of “grout-fill funnels” Photos 11a and 11b). As each compartment is filled with grout, the “grout-fill funnels” are replaced with 6” tall standpipes sized to fit the diameter of each grout fill hole (Photo 12). The reason these procedures are useful and successful is that the head pressure created by the “grout-fill funnels” pushes all air out from under the baseplate while maintaining 100% total contact with the grouted item. Never vibrate or strap epoxy grouts; they will flow without mechanical agitation. After the baseplate has been completely filled, leave the full standpipe in place until there is a noticeable increase in heat and the material has stopped flowing from the vent holes in the baseplate. After the standpipes are removed, use the excess grout remaining in the standpipes to mound grouting material in the grout holes to prevent moisture, oil, or chemical accumulations in these areas.

Conclusion

Photo 10. Mortar mixer

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As evidenced from the case histories and information contained in this article, it is necessary to consider a number of factors affecting the final outcome of a pump installation. It should be noted, however, that by utilizing a sound specification for grouting of all rotating The Pump Handbook Series

equipment, one variable in the overall troubleshooting process is eliminated when considering retro-fit applications or questions of equipment reliability. The same holds true for new installations. Epoxy grout is part of the “winning equation” that makes up a sound installation specification for pumps and drivers and any other type of rotating equipment for which proper alignment is crucial. It does truly start at the foundation.

References 1. Monroe, Perry C. The Road to Reliable Pump presented at the Pump Symposium, Houston, Texas 2. American Petroleum Institute. API 610 – 8th Edition – Centrifugal Pumps for Petroleum, Heavy Duty Chemical and Gas Industry Services 3. American Petroleum Institute API 686 – 1st Edition – Recommended Practices for Machinery Installation and Installation Design ■ Scott Bullentini, a graduate of San Francisco State University, is National Sales Manager for ITW/Philadelphia Resins Corporation. He has more than 10 years of experience in the field of industrial construction and management.

Photos 11a and b. Grout-fill funnels

Photo 12. Standpipes are sized to the diameters of the grout fill holes. The Pump Handbook Series

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Successful Submersible Operation Part II: Inspection And Maintenance Upkeep and regular attention keep these workhorses of the wastewater industry running smoothly. By Submersible Wastewater Pump Association

egular inspection and preventive maintenance ensure continued, reliable operation of the entire submersible pumping system. All stations, pumps and operating equipment should be inspected at least once a year—more frequently under severe operating conditions. One of the major advantages of a submersible station is the ability of the service technician to handle most maintenance and service onsite, without entering the wet well. All equipment in the station should be backed by manufacturers’ service manuals. This material should be carefully read, filed and consulted whenever servicing is required.

Safety Precautions

• Be aware of the risk of electrical accidents. • Check the explosion risk before welding or using electric hand tools in or near the station. Never weld or use electrical tools in the wet well after it has been in operation. • Make sure that all lifting equipment, when used, is in good condition and of adequate capacity. • Provide a suitable barrier around the work area—for example, a guard rail. • Make sure that all personnel have a clear path of retreat. • Use safety helmets, safety goggles and protective shoes or boots. • All personnel working with sewage systems must be vaccinated against any diseases that can occur.

CAUTION: Note and read all safety precautions before performing any operation or maintenance procedure. To minimize the risk of accidents in connection with service work, the following rules—as well as all applicable laws, regulations and manufacturers’ recommendations—must be followed: • Be aware of health hazards. Observe strict cleanliness.

Because sewage pumps are designed for use in liquids that can be hazardous to one’s health, make sure that all equipment has been thoroughly cleaned. To prevent injury to the eyes and skin, observe the following rules: 1. Always wear goggles and rubber gloves. 2. Wash and rinse pump thoroughly

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with clean water before starting work. 3. Wash and rinse any components in clean water after disassembly and then dry thoroughly. If you get hazardous chemicals in your eyes, rinse them immediately with running water for 15 minutes, and hold your eyelids apart with your fingers. Contact a doctor immediately. If you get hazardous chemicals on your skin, remove contaminated clothes, wash your skin with soap and water; seek medical attention immediately.

Recommended Inspections CAUTION: Before starting work on any pump, make sure it is isolated from the power supply and cannot be energized. This applies to the control circuit as well. One method is to tag and lock the control panel to let other personnel know that you are working on the station. Keep in mind that some systems have an override switch at the treatment plant or other buildings. Make sure that this switch is also off

Problem 1. Pump will not start.

2.Pump will not start and overload heaters trip.

3. Pump runs but will not shut off.

4. Pump does not deliver proper capacity.

5. Motor stops and then restarts after short period, but overload heaters in starter do not trip.

Possible Cause

Remedy

No power to motor.

Check for blown fuse or open circuit breaker.

Selector switch may be in off position.

Turn to on position.

Control circuit breaker may be tripped.

Reset the circuit breaker.

Overload heater in starter may be tripped.

Push to reset.

Overload heater in starter may be burned out.

Replace the heater.

Unit may be improperly grounded.

Turn power off and check motor leads with megger ohmeter.

Motor windings may not be balanced.

Check resistance of motor windings. If it is three-phase, all phases should show the same reading.

Impeller may be clogged, blocked or damaged.

If no grounds exist and the motor windings check out satisfactorily, remove the pump from the well and check for impeller blockage.

Pump may be air-locked.

Turn pump off for several minutes, then restart.

Lower level switch may be locked in closed position.

Check to be certain the level control is free.

Selector switch may be in the "hand" position.

Switch selector to "auto" position.

Discharge gate valve may be partially clogged.

Open and unclog valve.

Check valve may be partially clogged.

Valve must be cleared—if there is an outside lever, move it up and down.

Discharge line may be clogged.

Use a sewer cleaner or high-pressure hose to clear the obstruction.

Pump may be running in the wrong direction.

Low speed pumps can operate in reverse with little noise or vibration. Check your manual for methods of establishing and correcting rotation.

Discharge head may be too high.

Check total head with a gauge when pump is operating. Compare against original design and precious operating records. If pump has been in service for some time and capacity falls off, remove the pump and check for clogged impeller.

Heat sensors in the motor may trip due to excessive heat.

Impeller may be partially clogged, resulting in the sustained overload, though not high enough to trip the overload heater switch.

Motor may be operating out of liquid due to failed level control.

Check location and operation of level controls.

Pump may be operating on a short cycle.

The wet well may be too small or water may be repeatedly returning to the well due to a leaking check valve. Both must be checked.

Table 1. Troubleshooting checklist for pumps (These are general guidelines only. Consult the specific manufacturer’s manual for detailed instructions.)

and tagged at the other building before you start working on the station. After the pump(s) have been isolated from the power supply and pulled to the top of the station, the appropriate manufacturer’s service manuals should be consulted and the following inspection guidelines adhered to.

Visible Parts on Pump and in Station 1. Check for vandalism or other station damage.

2. Make certain the access cover works properly. Check the holdopen device to ensure that it is engaged. 3. Make certain that the guide rails are completely vertical. 4. Check the condition of the lifting eye, chains, hooks and wire ropes. 5. Make certain that all screws, bolts and nuts are tight. 6. Replace or repair worn or damaged parts.

The Pump Handbook Series

Pump Casing and Impeller 1. If clearance between the impeller skirt and the pump casing or wear ring exceeds manufacturer’s recommendations, it may be necessary to adjust the impeller or replace the wear rings. 2. Wear on the outlet flange from the pump casing usually causes corresponding wear on the discharge connection. 3. Follow the manufacturer’s instructions for disassembly, inspection, and reassembly of the impeller and

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Problem

Possible Cause

Remedy

Service voltage is not on to the panel.

Turn on and check for proper voltage.

Main or control circuit breakers tripped or turned off. Main or control circuit fuses blown.

Turn on or reset and turn on all circuit breakers. Check and replace blown fuses.

Motor heat sensor connections not made properly.

Check motor heat connections and correct to panel.

2. Pumps 1 and 2 will not run in "hand" position. Run lights are on.

Motors not wired properly.

Check motor heat connections and correct to panel.

Voltage starter coils are the wrong ones.

Check and correct starter coils voltage rating to match control circuit voltage.

3. Pump 1 or 2 will not run in "hand" position. Run light not on. One pump operates in "hand" position.

Pump circuit breaker tripped or turned off.

Turn on or reset breaker.

Pump circuit fuse is blown.

Check and replace any blown fuses.

Motor starter overload tripped.

Reset overload after checking motor.

Motor heat sensor circuit open or not properly connected.

Check continuity of motor heat sensor. Correct connections.

Level in wet well not high enough to turn on pumps.

Fill or allow wet well to fill to required levels.

Level float switches may be incorrectly connected or have failed.

Check and connect each float correctly; replace if required.

Air bubbler supply may be off or failed.

Check and ensure air supply is on, bubbler line is working in wet well and has no leaks.

Pressure switches or sensors may not be adjusted or sequenced properly.

Check and adjust pressure switches to correct levels and sequence.

Relay or other control device failed.

Check and replace any control relay, alternator or other device with defective coil or contacts.

No probable panel problem.

None.

Possible system problem, for example, clogged discharge line.

Check and clear check valve or line of obstruction.

Temporary high level condition after power failure or influent surge.

Monitor station operation until "high level" is reduced.

6. Alarm light and/or audible alarm turns on, with one or both pumps not running.

Test hand operation of pump that is not running and refer to Problem 3.

See remedies for Problem 3.

Refer to Problem 4.

See remedies for Problem 4.

7. Circuit breaker tripped for motor power.

Motor not wired properly.

Check and correct connection to panel.

Short in pump cable, wiring or motor.

Disconnect motor and check wiring. Check motor for shorts or grounds.

Size of breaker too small and/or ambient heat problem.

Check and correct breaker size for motor and/or provide ventilation or compensation for ambient heat.

8. Blown fuse for motor power.

Same as Problem 7.

See remedies for Problem 7.

9. Pumps do not alternate. Same as Problem 7.

Alternator relay may be defective.

Check and replace alternator.

Improper sequencing of float switches or pressure sensors.

Check and correct sequence of controls to ensure off, lead, lag sequence.

Pilot light bulb has failed.

Replace lamp with correct voltage replacement.

1. Pumps 1 and 2 will not run in "hand" or "auto" positions.

4. Pumps 1 and 2 will not run in "auto" position.

5. Alarm light and/or audible alarm turns on with both pumps running.

Table 2. Troubleshooting checklist for control panels (These are general guidelines only. Consult the specific manufacturer’s manual for detailed instructions.)

volute. When it is disassembled, check the motor shaft, impeller and volute bore for wear or damage. 4. Follow the manufacturer’s instructions for disassembly, inspection

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and reassembly of the shaft seal. It must be clean and properly seated before reassembly. 5. Always replace worn or damaged parts. The Pump Handbook Series

Electrical Insulation Perform a megger (insulation resistance) test between the pump motor leads and the pump casing. A low— 20 megohms or less—reading indicates

moisture entry into the motor chamber or power cord, or other deterioration of the insulation system. Such problems should be corrected before a major breakdown occurs. Oil Quantity and Condition A. Oil-Filled Motors CAUTION: If there has been any leakage, the motor housing may be under pressure. Hold a rag over the inspection plug to prevent splatter when loosening the plug. Check the oil in both the motor housing and the seal cavity through the oil inspection plugs. The oil level may be low or emulsified (creamlike), which indicates that water has entered the cavity and a leak is present. One possible cause is an inspection plug that is not sufficiently tight. Check the sealing surface of the motor housing, the cable entry and the condition of the shaft seal. Whatever the problem, correct it and make certain that the oil is refilled to the proper level with the motor manufacturer’s recommended oil. B. Non Oil-Filled Motors If there is any liquid in the motor housing, a leak is present and all sealing faces should be checked as listed above under “Oil-Filled Motors.” Cable Entry 1. Make certain that the cable connection is tight. 2. If the cable entry leaks, it may be necessary to replace the cable seal. See the manufacturer’s manual for instructions. 3. When refitting a cable that has been used before, even when the jacket has not been damaged, always cut off a short piece of the cable so that the cable entry seal does not close around it at the same point. 4. If the outer jacket of the cable is damaged, replace the cable. Make sure the cable has no sharp bends and is not pinched.

Controls 1. Check liquid level sensors throughout their entire range of operation. Clean, adjust, replace and repair damaged equipment. Follow the manufacturer’s instructions. 2. The same procedure should be used for checking the balance of the control system. In particular, check signals and the tripping function, and make sure that the relays, lamps, fuses and connections are intact. Replace all inoperative equipment. Piping and Valves Repair all flaws and replace inoperative equipment.

retain this approval label, major motor repair must be performed by an authorized UL or FM repair facility. ■ This article was adapted from Submersible Sewage Pumping Systems (SWPA) Handbook, published by the Submersible Wastewater Pump Association, 1866 Sheridan Road, Suite 210, Highland Park, IL 60035. Phone: (847) 681-1868, Fax: (847) 681-1869, E-Mail: [email protected]. SWPA is a national trade association representing and serving the manufacturers of submersible pumps for municipal and industrial wastewater applications. Founded in 1976, the association’s primary focus is on industry guidelines, education and promotion.

Fault Tracing A voltmeter, ohmmeter, and ammeter—together with the job wiring diagram—are required to test, measure and carry out fault tracing on electrical equipment. Fault tracing must always be performed with the power supply disconnected and locked off, except for those checks which can be performed only with power. Electrical work must be performed by qualified electricians and all applicable local, state and national safety regulations followed. Observe the recommended safety precautions previously mentioned in this article.

Major Servicing Submersible sewage pumps can be serviced in the field at qualified facilities. If the pump is still under warranty, it should be serviced by an authorized shop. Manufacturers’ service manuals provide detailed instructions for the replacement of impellers, stators, seals and bearings. To facilitate field maintenance and service, many manufacturers provide a list of authorized service facilities, recommended spare parts and the maintenance equipment required. If the motor carries an approval label such as Underwriters Laboratories (UL) or Factory Mutual (FM), in order to The Pump Handbook Series

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Using VFD Technology in a Water Distribution System Drive technology matures to conquer pressure and leakage problems while increasing pump life and saving energy. By Satish Kamaraju, Danfoss Drives ariable Frequency Drives (VFDs) on pumps offer many advantages over traditional methods when it comes to constant pressure control. They cost less than water towers, offer precise pressure control, reduce water leakage, increase pump life and save energy. Traditionally, water towers or ground/ underground reservoirs equipped with pumps are used for water storage and maintaining pressure in the distribution systems. Mechanical devices such as control valves are used to regulate pressure. These methods have disadvantages when it comes to operation, flexibility and energy consumption. Water towers can be an expensive choice when it comes to maintenance and system upgrades. Selecting the wrong control valves can significantly reduce pump efficiencies and could lead to routine maintenance problems. With the proper design, VFD technology offers solutions to these problems. A VFD is an electronic controller that adjusts the speed of an electric motor by modulating the power being delivered. VFD and pump combinations maintain a constant pressure in the system by taking 4-20 mA signal from a pressure transmitter and adjusting motor output according to the feedback requirements. VFDs overcome the disadvantages of traditional methods mentioned earlier and offer many benefits—increased pump life, precise pressure control,

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reduced water leakage, easy upgradability and energy savings.

Factors Influencing the Water Distribution System Pressure Fluctuations Maintaining sufficient pressure in the water distribution network is vital for a drinking water supply. Pressure in a water distribution system is affected by numerous factors—water consumption during the day and night, seasonal changes, age of the piping network and topography, to mention a few (Figure 1). A reliable pressure control system is vital for a distribution network.

Figure 1. Daily water demand The Pump Handbook Series

Water Leakage Water leakage (also called NonRevenue Water [NRW]) from a distribution system makes a significant impact on the water management. Water leakage can account anywhere from 10% to over 50% of the total volume treated for supply. Excessive pressures and/or pressure fluctuations in the distribution system can be a major contributor. This is shown well Figure 2, which illustrates water leakage at a given pressure through 1 mm (0.04”), 2 mm (0.08”), 4 mm (0.16”) and 6 mm (0.24”) diameter holes. The plot suggests that reducing pressures whenever possible could lead to significant reduction

constant pressure in the distribution system determined by one factor: the height of the water column. The system requires 2.31 feet of the water column to generate one psi of pressure. Water towers give marginal flexibility in changing the system pressures for leakage control and/or for future needs. Typically, the water level variations (∆h) can only be a fraction of the static head (H), as shown in Figure 3. As an example, a 500,000 gallon capacity spheroid with dimensions of 55’6” diameter and a head range of 37’6” is selected for comparison. Using these dimensions, the required height of the tower to maintain a pressure of 90 psi at high water level (HWL) is approximately 208’ (90 x 2.31). Assuming ∆h can be varied between a low water level (LWL) of 20’0” and 37’6” (HWL), the pressure reduction at LWL is calculated to be 7.6 psi (8%). Figure 2. Water leakage through pipes (Ref. 1)

in water leakage. Reducing pressure during non-peak hours or at night is a good example. Waterhammer Another major factor of the distribution network is the waterhammer or transient internal pressure due to stoppage of flow caused by an abrupt closing of a valve or halting of a pump. This will have a significant impact on pumps, check valves and the piping network.

provide an abundant storage capacity. However, their disadvantages are equally well recognized during renovations and upgrades—the tower has to be taken offline, massive supporting structure for accessibility of the exterior tank installed, OSHA requirements for interior cleaning complied with, to mention a few. Last but not least, the tower demolition charges can be significant. Pressure Control Changing pressure set points is a problem. Water towers provide a

Water Leakage—Difficult to Lower Pressures Reducing water leakage by lowering system pressures in water towers is difficult. The majority of towers are filled to the maximum capacity at night so that they can meet the water demand during the day. Because of this, the water leakage will be higher at night and tend to decrease during the day. An option is to install leak detection systems and implement leak prevention methods throughout the distribution network. This, however, leads to significant O&M costs.

Other Factors Consider the operation and maintenance costs, system upgrades, system closures and the water quality. The following paragraphs compare the traditional methods and how VFDs in combination with pumps help to optimize the distribution system.

Water Towers Be(a)ware of the Costs Water towers have been a good choice for many decades. They have proven very dependable during power outages and

Figure 3. Operation of a water tower The Pump Handbook Series

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Cost Comparison with VFD According a recent survey, the cost comparisons between a water tower of 500,000 gallon capacity and the alternate option, a ground level storage tank and pumps equipped with VFDs of equivalent capacity are: WATER TOWER

STORAGE TANK +VFD

Avg. Installation Costs

$565,000

$325,000

O& M over 60 years

$692,000

$158,000

$1,257,000

$483,000

Overall Costs

Figure 4. Typical layout of a basic storage tank and pump system

Storage tanks with VFDs have installation costs 42% lower than water towers and a significant reduction in O&M costs—77% less than water towers. The cost savings over the life cycle of an average water tower (60 years) is 61%. The survey assumes that VFDs are replaced every ten years and the associated costs are accounted under O&M costs. The survey also includes the installation of a back-up generator for a VFD option during power failures (Ref. 1). Storage Tanks and Pump Systems Choosing pressure controlling devices can be tricky. As an alternative to water towers, underground or above ground storage tanks with pumps are used. While tanks give enough storage capacity, pumps provide pressure to the distribution network (Figure 4). The pressure in the system fluctuates constantly—it decreases as the water consumption increases and increases substantially during the night if proper pressure controlling methods are not used.

Figure 5. Pressure in an over-pressurized system

Pressure Control In a conventional pumping system, one or more pumps are operated at all times based on a predetermined pressure value (Psetpoint) that is selected to meet the peak water demand (Photo 1). At other times, Psetpoint will lead to excess pressures in the distribution system as the water usage reduces overnight. A typical variation of pressure over a 24 hour period is shown in Figure 5. As an example, the resulting water leakage due to these pressure variations through a 6 mm (0.24 inches) diameter hole from a distribution

Figure 6. Leakage corresponding to excess pressures

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Photo 1.

system (Ppeak = 60 psi) is shown in Figure 6 (Ref. 2). An improvement to the conventional pumping system is sending the feedback signal from the pressure transmitter to the pumps. The pump(s) will be turned on and off by a programmable logic controller (PLC) according to the pressure variations. Although this is an improvement, there are some drawbacks with this arrangement. The frequent on of pumps will draw the current abruptly each time the pump(s) starts, subjecting the motor to high inrush current six to ten times the fullload current. The frequent on-off operation will lead to waterhammer problems. This results in high energy costs and reduced equipment life. The following sections discuss VFD technology and the advantages of using it for constant pressure control in a distribution network.

Variable Frequency Drives Principle A VFD has three major components: a rectifier, an intermediate circuit and an

inverter. A rectifier converts the threephase AC voltage (V) from the supply mains (main?) or source to a pulsating DC voltage. The intermediate circuit stabilizes the pulsating DC voltage and sends this on to the inverter. Finally, the inverter converts DC voltage into variable AC voltage with a variable frequency (F). As a result, the motor is supplied with a variable voltage and frequency, which enables infinitely variable speed regulation of the pumps (Figure 7). The effect of varying the frequency is essentially the same as the effect of varying the speed. If the speed of a pump is changed, the flow, head and power will change according to the affinity laws. The laws state that the flow is proportional to the speed, and the head (pressure) is proportional to the square of the speed. The shaft power is proportional to the cube of the speed. Therefore, a reduction of 10% in speed will give 27% energy savings. Figure 8 shows the comparison between a constant speed pump and variable speed pumps in relation to flows, head/pressure and power consumption.

Application — Pressure Control VFDs are added to the pumps, forcing water into the distribution network (Figure 9). The pressure in the lines can be manually controlled by a potentiometer on the drives for applications with minimal pressure fluctuations. For a system with highly variable pressure fluctuations, VFDs are connected to the pressure transmitter to receive 4-20mA signal monitoring the pressure fluctuations. The internal controller (Proportional and Integral Control, PIC) of the VFD monitors the pressure fluctuations and adjusts the motor output to maintain the pressure closer to the set point. As discharge pressure drops, the pump speed is increased to increase the pressure; conversely, when the pressure rises, the PI controller of the VFD decreases pump speed. When properly tuned, this type of control system has the ability to maintain the pressure within 1 psi.

Other Benefits Reduced Water Leakage Dual set points reduce water leakage. The water leakage in a distribution system is proportional to the pressure maintained in the system. Maintaining a constant pressure set point to meet the peak demands will lead to substantial water leakage. From Figure 2, it can be inferred that a second pressure set point (low pressure set point) during the low flow consumption is a viable choice. This will reduce the water leakage substantially, thus reducing nonrevenue water and energy costs. The new pressure control system can be operated on dual pressure set points where a higher set point (Ph) will be maintained during the day (peak

Figure 7. Principle of a VFD The Pump Handbook Series

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demand) and a lower set point (Pl) during the night (Figure 10). As an example, the dual pressure set points are selected for a typical water consumption curve. The pressure Ph of 60 psi will be maintained for 70% for the time and Pl of 40 psi is maintained for 30% of the time (Figure 10a). Reference 2 is extrapolated to calculate the leakage through a 6 mm diameter hole at single and dual pressure set points. A reduction of 22% leakage is estimated from this example. Eliminate Waterhammer The ramp up/down features of a VFD eliminate pressure surges. Waterhammer caused by hydraulic surges can be easily controlled using VFD technology. The properly programmed VFD system will start and stop pumps on a ramped or timed basis. The slow starting and stopping of a pump, coupled with varying the speed to maintain predetermined discharge pressure, can virtually eliminate hydraulic surges in the system, preventing waterhammer.

VFD vs. Control Valves

Figure 8. Constant speed vs. variable speed pump operation

VFDs offer cost savings by more efficient use of power and pump speed. The use of control valves on pumps has been popular as a way of controlling pressure in a distribution system despite their disadvantages (Figure 11). The control valve modulates, controlling pressure while the motor and pump run at 100% speed and develop about 5 to 8 psi head loss from inlet to outlet (Ref. 3). As the valve ages, this head loss increases, causing pump efficiency to decline and leading to frequent valve failures. VFDs offer many advantages over control valves, ranging from pressure control to energy savings. Varying the speed of the pump over the course of the day, to better match actual energy input demands to meet the system demands, can produce significant energy savings. Figures 12 and 13 show the pump curves depicting the distinct comparison of energy savings. The drawbacks of using the control valves for constant pressure are frequent valve failures that could lower pump efficiencies.

Figure 9. A typical layout of a VFD system

Figures 10 and 10a. Dual pressure set points

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the motors is switched on or off, excess energy flows back to the output of the VFD and can damage the unit. For this type of application, look for a manufacturer that has no limit to this type of switching but no additional isolation equipment. Figure 11. Typical application of a control valve

High Starting Currents The power in-rush currents at the time of starting a pump are referred to as Direct On Line (DOL) currents. They typically range from 6 to 10 times its nominal current (Figure 14). Several utility companies penalize pump users by charging under maximum demand charges. The high starting currents can be controlled by “soft starters” to some extent. VFDs eliminate the problem by starting the current at zero and raising it as the load is accelerated with no danger of exceeding full load current. This would cut the electrical charges, significantly eliminating the maxi-mum demand charges and even in some cases qualify pump users for subsidies from utility companies.

are generated by that unit. VFDs with a true displacement power factor of 0.901.00 have added filtration to their equipment to correct the effects of harmonics such as transformer heating and disturbance of sensitive metering equipment. Enclosure Ratings In existing plants, the layout of the control room may limit the application of VFDs. Under such circumstances, VFDs covered by NEMA 12 (typical for water/wastewater plants) enclosures can be installed next to the equipment. Multiple Motor Operation In some circumstances, it may be ideal to run several equal hp motors from one VFD. When one or more of

Software Enhancements There are some beneficial enhancements to VFDs for the water industry that are worth checking into. For instance, your application might require a constant torque start to overcome head pressure. A VFD that has load loss protection will shut down your motor if a coupling fails. Another desirable feature might be an internal PID controller to the VFD that takes two analog feedbacks and evaluates data to produce one of two established set-points. Protective Features Due to the nature of the technology, VFDs generate transients such as Radio Frequency Interference (RFI), voltage rise per time (dV/dt), high peak voltage and harmonics onto the AC line and motor cables. These transients can affect surrounding sensitive equipment

Some Things to Consider When Specifying a VFD For a successful VFD installation and operation, choosing a suitable VFD manufacturer is difficult. The following are the factors to consider before selecting a vendor. Motor Cable Length VFD manufacturers publish a maximum cable length between the VFD and the motor to avoid the risk of stressing the motor and overheating the VFD. Many manufacturers offer output AC reactors to extend their product’s published maximum length. To avoid this additional cost and voltage loss, look for a VFD with long standard (no additional filters needed) maximum cable length specifications. Power Factor The power factor of a VFD is an indicator of the harmonic currents that

Figure 12. Power consumption using a valve control The Pump Handbook Series

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(such as computers and monitoring equipment) and reduce motor life. Check with your supplier for RFI filters as well as built-in reactors that can reduce these effects. Serial Communication Some VFD manufacturers do not include an RS-485/-232 port as standard, so watch out for extra cost. Also, be sure your supplier provides the protocol needed in your application. Well-known

protocols include Profibus DP, DeviceNet and Modbus Plus. Installation and Commissioning Ask your supplier to see a VFD instruction manual. Is it easy to read and understand? Can you easily find critical installation information such as wire gauge and tightening torques? Also, your supplier should be able to show you how to program a working VFD. Text keypads are much easier to

program than numeric ones. Experiment with the keypad to determine how simple or complex the programming is. Ask about programming options to aid commissioning, such as PC software and application-specific macros. Service Ask about the warranty and service policy—sometimes an extended warranty is available. A 24-hour technical assistance number is important as is a local manufacturer’s representative. Find out if the sales representative is capable and willing to technically support the product he or she is selling.

Conclusion VFDs offer many advantages over the traditional methods of pressure control in a distribution network. They provide precise pressure control, reduce water leakage, increase pump life and result in significant energy savings. A careful selection of a VFD supplier should be made for trouble-free operation of the system and to enable you to reap the benefits of the VFD technology to the maximum extent.

References

Figure 13. Power consumption comparison between VFD and control valves

800

Starting Currents

Direct On-Line Star/Delta Soft Starter VFD

700

Power %

600 500 400 300 200 100 0

40 20 30 50 Motor Frequency Hz

60

Figure 14. Direct On Line (DOL) and VFD comparison

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1. Henning Karlby and Inga Sørensen, “Vandforsyning,“ published by Teknisk Forlag, 1st edition, 1998, ISDN 87-571-1964-3 2. Jynson Nguyen, “Water Distribution Systems Today in USA,” paper submitted to the University of Aalborg, Denmark. 3. Philip Lawler, “Water District Saves 40% On Electrical Bill,” Journal, Water & Wastewater International, August 1997. Satish Kamaraju earned his BS in Civil Engineering in India and MS in Environmental Engineering from the University of Southwestern Louisiana, Lafayette, LA. He started his career as a Process Engineer designing wastewater treatment systems and is currently working as Water Applications Expert at Danfoss Drives, Global Water Business Unit, in Rockford IL. He brings 7 years of professional experience in the water/wastewater field.

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R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

Strategies to Reduce Pump Repairs at Petro-Canada Reliability models, tight specs, careful choices and diligence improved reliability dramatically at this refinery. By Ken Noble, Petro-Canada

O

ften when plants look for opportunities to improve pump reliability, attention is focused on the significant few or “bad actors.” This focus does have immediate payback in both money and improved reliability for the targeted equipment, but what about equipment not on the bad actors list? Reliability improvements are not always realized by those machines. Starting in 1993, a reliability program to significantly improve pump reliability was initiated at the Petro-Canada Oakville Refinery. The purpose of this reliability initiative was to identify and implement several specific upgrades that could be done simultaneously when a pump was removed to the shop for a repair.

Reliability Model Figure 1 lists the typical pump components that fail and the common causes of failure. Although the list is not all-inclusive, it does categorize the failure modes that can be applied to those pumps that are typically not on a bad actor list. The figure also highlights that all of the failure modes (with the exception of process upsets) are within the control of maintenance or operations departments. The number of modes that exceed their acceptable operating parameters often reduces the ideal interval between equipment repairs. For

example, all pumps (assuming the hydraulics have been correctly determined) should have a life expectancy of several years. However, in reality, pumps are momentarily run without product or operated with contaminated lubricant or reinstalled with oversize bearing housing bores, etc. Figure 2 shows how the maximum repair interval is reduced by the various off-specification conditions from the time the pump is placed back into service. Based on the common failure modes listed in Figure 1 and the life reduction concept model in Figure 2, several improvement programs were initiated at the Oakville facility. This article reviews these programs as applied to single stage centrifugal process pumps and their respective results, as well as the overall changes to the site reliability.

Understanding the Human Issues Successful programs require communication with all the various parties involved. Trades personnel need to understand the current reliability status, why the programs are being initiated and the goals of said programs. The operations group and senior management must realize the benefits they might gain for the risks they might incur due to the length of time a pump is out of service for repair. Whichever group is responsible for equipment repair costs must know why costs might increase at the start of an improvement program, but will either remain the same or shrink as reliability is increased. Occasional unforeseen problems will arise no matter how well an improvement program is planned—everyone must realize this from the start.

Figure 1. Parallel misalignment The Pump Handbook Series

455

Once a program is initiated, ongoing communicataion and status updates are important. One easy way to provide this

is to publish the trend of equipment mean time between failures (MTBF) and equipment repair costs on a regular

basis. This ensures that people are aware of the tangible benefits. It also helps cement credibility to the program.

Dimensional Checks and Re-Qualification

Figure 2. Life reduction model STEP DESCRIPTION 1. 2. 3. 4. 5. 6. 7. 8. 9. 10. 11. 12. 13. 14. 15. 16. 17. 18 19. 20. 21. 22. 23. 24. 25. 26. 27.

Check impeller for cracks. Check pump for corrosion or broken components Measure shaft runout at impeller end Measure shaft end float. Measure bearing housing flange fit Measure fit between seal chamber bore and gland spigot. Measure runout of cover gasket face. Measure concentricity of shaft and seal chamber bore Measure runout of seal chamber face Check straightness of shaft Measure bearing housing thrust bore Measure bearing housing radial bore Measure fit between impeller bore and shaft. Measure impeller eye wear ring clearances. Measure impeller hub wear ring clearances. Measure throat bushing clearance Measure thrust bearing shaft fit. .Measure radial bearing shaft fits Dye penetrant check bearing housing for cracks Measure coupling fit on shaft Check contact between sleeve & shaft shoulder “hooked sleeves” Incremental balance of pump rotor components. Modify bearing housing covers to accept bearing isolators if available Modify both bearing housing covers for future oil mist connections Drill 1/8” hole at top of seal chamber for vent were possible Re-assemble pump recording all repaired dimensions Check seal setting on non-cartridge seals

Allowable Go/No-Go Value

Repaired Tolerance

None Visual Specified Specified Specified Specified Specified Specified Specified Specified Specified Specified Specified Specified Specified Specified Specified Specified None Allowed Specified

Specified Specified Specified Specified Specified Specified Specified Specified Specified Specified Specified Specified Specified Specified Specified Specified None Allowed Specified

Specified

Specified 4W/N

Specified

Table 1. Single-stage pump repair sheet

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One of the key pump failure modes is “off-specification” fits and runouts. Excessive bearing bore or bearing housing clearances can affect both bearing and seal life; excessive runouts with respect to the seal chamber can adversely affect seal life. On the shop floor, checks were typically performed on an ad hoc basis or only when specifically requested. Tolerances were often left open to interpretation and were not always referenced back to the manufacturer’s or industry-accepted specifications. To remove the subjectivity of the dimensional and runout checks, a 27step repair sheet was developed for single-stage process pumps (Table 1). The checklist required inspection points and “go/no-go” values for specific measurements and runouts. The list also includes the requalification specification, in the event a component exceeds the no-go tolerance. Requalification tolerances and runouts were determined from bearing, seal and pump manufacturer’s manuals. The go/no-go tolerances and runouts were specified to be the maximum original manufacturing tolerance plus 50% of the original manufacturing tolerance spread. Developing the checklist was only half of the solution. The other half consisted of setting aside the required time to requalify the off-spec items. Traditionally, the plant had used the vibration program for monitoring the condition of pump bearings and had gained confidence in the ability of the maintenance department to successfully identify developing problems. Because some of the repairs required the pump to be out of service for up to seven working days, the operations department used a vibration program to validate the condition of the spare pump and monitor it while the designated pump was out for repairs. This combination of developing a consis-

tent set of tolerances and providing sufficient time to perform the full repair had an immediate impact on reliability, as illustrated in Figure 3.

Bearing Isolators Most of the pumps at our site are API 610 units and have original equipment manufacturer’s (OEM) labyrinth oil seals installed at each end of the bearing housing. These OEM seals are a very simple labyrinth design that still allows the ingress of moisture into the oil sump. A third-party bearing isolator with an O-ring separating the stationary and rotating components was selected as the site standard. This isolator has multiple expulsion ports to ensure that the drainage holes aren’t installed incorrectly. Without this feature it would be possible to install the unit oriented the wrong direction. Another option we considered was the use of magnetic bearing isolators. But these were more expensive than the labyrinth isolators and had a specific operating life of less than the targeted pump MTBF of four years.

overall vibration levels. Figure 5 shows the status of the pump vibration levels in 1995. Figure 6 shows the levels in 1998. In 1995, 64% of the vibration readings were equal to or less than 0.10 in/sec. In 1998, the percentage had increased to 77%. By 1998, 90% of all pump vibration levels were equal to or less than 0.15 in/sec, which is the current specified vibration requirement for new pumps purchased using the API 610 pump specification. It should also be noted that approximately 80% of the site’s pump population are 25 to 40 years old and usually operate at 3600 rpm.

epoxy instead of cementitious grout. Grout vendors have studies indicating that epoxy grout has significantly better vibration dampening qualities. In an attempt to quantify the difference in vibration levels, eight new API 7th edition pumps were installed using epoxy grout and eight new pumps were installed using cementitious grout. Figures 7 and 8 show the distribution of the vibration readings taken on the 16 pumps. There was a reduction in vibration levels of approximately 0.03 in/sec using the epoxy grout, but in both cases more than 90% of the readings were below 0.15 in/sec.

Grout During the past several years there have been numerous discussions within the user community about the benefits of installing pump bases using

Dry Couplings As repair intervals increase, problems can arise due to the centrifuging of the coupling lubricant. No

Vibration Levels Vibration levels are often indicative of the overall mechanical condition of a pump. Vibration readings can be used to determine the amount of residual unbalance in the rotor, the condition of the bearings or the accuracy of the alignment between the pump and driver. However, the question is often “What is the optimal vibration level?” Discussions between users and OEMs usually generate disagreement as to what is economically achievable. In the early 1990s, a study of averaged vibration levels vs. MTBF was performed at another Petro-Canada site. The data suggested that the highest MTBF values were achieved when the maximum overall vibration levels were under 0.15 in/sec (Figure 4). Although the Oakville plant did not historically have a large percentage of equipment running above 0.15 in/sec, a balance stand and laser alignment system were purchased to facilitate reduced repair times and decrease

Figure 3. Trend of repairs for 15 services with one or more bearing housings requalified

Figure 4. Snapshot of 1990 MTBF vs. vibration level

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Figure 6. Snapshot of December 1998 pump vibration levels

Figure 5. Snapshot of December 1995 pump vibration levels

coupling lubrication program has been initiated at the site. Instead, all grease couplings are re-lubricated whenever a pump is repaired. Vibration analysis is used to monitor the condition of the coupling between repairs. If a lubricated coupling becomes worn, it is replaced with a non-lubricated coupling. Spacer couplings are replaced with a membrane style dry coupling; nonspacer couplings are replaced with elastomeric couplings.

Handling Vapor in the Seal Chamber Ensuring that the seal chamber is always filled with product is key to obtaining maximum seal life. However, there are numerous sources of seal chamber vapor requiring different types of control. The effects of vapor in the seal chamber usually do not cause an immediate failure. The interaction between the amount of vapor, the time interval that vapor is present, the number of times that vapor is present, and the fluid properties determine the ultimate reduction of seal life. Vapor is always introduced into the seal chamber when a pump is repaired. Figure 9 shows how vapor can become

trapped in a dead-ended seal chamber during the initial pressurization after a repair. Any vapor in the seal chamber will rise to the highest point. When the pump is started, the centrifugal (dynamic) forces will throw the heavier liquid outward, causing the lighter vapor to migrate inward around the mechanical seal (Figure 10). This can cause excessive vibration in seal bellows and sub-optimal lubrication between the seal faces. To enable static venting of seal chamber vapor, 1/8” holes were drilled at the top of the chamber. These provide a leakage path from the seal chamber to the pump volute. Another source of seal chamber vapor is the flush line. The flush line often forms an inverted vapor trap (Figure 11). Even if the seal chamber has an internal vent to enable full flooding during pressurization, additional vapor may be re-introduced from the flush line. A vent valve can be installed on the high point of the flush line to facilitate venting. However, this is a manual process and is not always used. Also, it is sometimes hard to vent because of safety or environmental factors.

Figure 7. Distribution of vibration levels for epoxy and cementitious grouted pumps

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An alternate method is to install a continually rising flush line for all API Plan 11 systems (Figure 12). This prevents vapor from being re-introduced into the seal chamber, since the flush line is automatically maintained in a fully flooded condition because of the self-venting of the arrangement. All of the pumps identified as bad actors that incorporated a Plan 11 flush have had the flush piping reworked to provide this continuous rise. Another source of vapor is from the mechanical seal itself. The rotating faces of the seal produce heat that can, in some cases, cause excessive vapor formation. To minimize the formation of vapor across the faces, most seal manufacturers suggest a 20-25 psi margin between the vapor pressure and the seal chamber pressure. Depending on the impeller design and condition of internal clearances, the required vapor pressure margin may be sufficient. In other instances, a flush must be introduced into the seal chamber to elevate the seal chamber pressure. It is important that this flush be metered, even if only in a rudimentary fashion, to prevent excessive seal chamber pressures and keep fluid from jetting onto the faces.

Figure 8. Cumulative distribution of pump vibration levels for new installations (API 7th Edition pump bases) using epoxy and cementitious grouts

The Pump Handbook Series

To flush tie-in point

P u m p a g e

Trapped Vapour in top portion of Seal Chamber

Continuously rising flush line eliminates trapped vapour

G l a n d

Mech. Seal

Shaft

Figure 9. Trapped vapor during pressurization

At the Oakville Refinery, all API Plan 11 flush configurations incorporate an orifice plate installed vertically to meter the flow. More than one plate may be installed to prevent the orifice plate hole from becoming too small, depending on the differential head developed by the pump. These plates are also installed at least six inches from the gland connection to enable the flow to become re-established before entering the gland. In the late 1980s, most of the seal vendors modified their designs and materials for low specific gravity products to help meet new EPA emissions levels. In many cases, these low emission seals incorporated a new face design and metal-impregnated carbon. Based upon the excellent results achieved using these carbons as a bearing material, and on test seal installations, all API pump seals with carbon faces were changed to the metal-impregnated material. Although the new material is more expensive than standard carbons, it is better suited to operating in sub-optimal conditions during the inevitable process upsets. Along with the changes in face design and material selection, many seal vendors also provide multiple port tangential flush glands on their low emission seals. This flush arrangement helps break up the bubbles that can accumulate at the seal faces. In addition, the flush is more evenly distributed circumferentially around the face, which provides for more uniform face cooling. Providing the fluid is clean, all new seal installations below 350°F incorporate the multiple port flush arrangement.

Figure 10. Effects of centrifugal forces on seal chamber vapor

P u m p a g e

G l a n d Mech. Seal

Oil Mist

Shaft

After a trial test of oil mist at another Petro-Canada site, the first dry sump oil mist system was commissioned for the Oakville Refinery in October 1995. Since then, four additional systems have been installed. Justification for the system was based on the percentage of pump repairs attributed to bearing problems. This data was obtained using the computerized maintenance management system (CMMS). Prior to the installation of the oil mist systems, about 40% of pump repairs were attributed to bearingrelated problems. After installation, the number dropped to just 10%. Vapour trapped in flush line To flush tie-in point

P u m p a g e

G l a n d Mech. Seal

Shaft

Figure 11. Trapped vapor in seal flush line

Bills of Material These documents, created on the plant’s CMMS, are often overlooked. A correct and fully defined bill of material forms a key part of the repair quality control process by ensuring that parts are identified correctly. Parts availability can also be checked before a repair’s scheduled start date, which The Pump Handbook Series

Figure 12. Continuously rising flush line

saves money. Updating bills of material with new components when a program is initially implemented ensures that the parts are available when the pump is removed. The correct parts are also available for subsequent repairs. Other benefits of up-to-date bills of material are inventory reduction and material upgrades. Older plants often implement expansions, which include the purchase of new pumps, incrementally. A review of the component inventory for items such as shafts, wear rings, impellers and throat bushings can indicate that the same part made from different materials might already be stored in the plant warehouse. For example, a roughmachined impeller may be stocked in cast iron, cast steel and 11-13 chrome. Or a wear ring may also be stocked made from cast iron, cast steel and 11-13 chrome. In such cases it is often possible to standardize on a single material (typically the material with the broadest range of application). Although the individual cost of the selected material might be higher, inventory-carrying costs can be significantly reduced. Stock turnover is also increased, and in some cases availability may be increased due to the number of pumps associated with the component. From a reliability perspective, the upgraded material often results in longer component life,

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Figure 13. Annual ratio of repairs for pumps with low emission seals

which in turn reduces repair costs and increases equipment efficiency.

Recycle Back to Engineering Specs The maintenance group initially tested most of these programs as trials on a small group of new pumps or a group of existing pumps with chronic problems. Once the concepts had been proven and accepted it was important to ensure that these changes were incorporated into the standards used for the purchase of new pumps—both at Oakville and other Petro-Canada sites. As with any continuous improvement program, the trick is to make sure the improvements are identified and documented so that the next iteration of improvements can begin.

End Results All the upgrade programs and concepts I’ve described have been validated and implemented at our facility. However, not all pumps have been upgraded to incorporate all the changes. Most units are upgraded only when they are repaired due to worn components. Over the next few years, the remaining upgrades will be completed. Because these are upgrade programs that are only applied when pumps are repaired, the pumps that benefit most (those with short intervals between repairs) have been addressed. Figure 14 shows the relative reduction in pump repairs at the site over the past several years. The programs implemented at the Oakville Refinery were, in essence, a

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Figure 14. Ratio of annual pump repairs relative to 1993

return to basics. Common failure modes were determined for those pumps not identified as “bad actors.” Improvement programs were tested in small pump populations to validate the anticipated results. Once the individual tests were evaluated, all the upgrades/improvements were rolled into an upgrade strategy that was applied to all our pumps (including bad actors) in an ongoing fashion as they were removed from service for normal repairs. This “program approach” allowed for the materials to be prepurchased and updated on the bills of material so they were available for the trades personnel to install when required. These back to basics concepts included: 1. Trades personnel have the “go/no go” values for critical tolerances and fits. 2. If a component requires requalification, repair tolerances are available to the trades personnel. 3. Trades personnel have the required time to perform a complete repair 80% to 90% of the time. 4. Operations personnel know the condition of the spare equipment. 5. Target overall vibration levels to be 0.15 in/sec or lower. 6. Recognize that process upsets occur. Install components that will help mitigate the effects of short-term operation in a distressed mode. 7. Eliminate vapor from the seal chamber at start-up and miniThe Pump Handbook Series

8. 9.

10.

11.

mize the formation of vapor at the seal faces during operation. Minimize contamination of the lubricating oil. Minimize problems associated with couplings by replacing greased couplings with nonlubricated designs. Perform accurate alignments. Ensure equipment CMMS bills of material are sufficiently detailed and accurate. Feed back results to validate the risks that the operations personnel accept by operating without a spare while the pump is being properly repaired.

Identifying improvements that need to be made is easy; obtaining the time to implement them can be considerably more difficult. Additional funding is usually not required for component upgrades (excluding oil mist), since the increased reliability should generate sufficient funds to cover the initial costs. As Figure 14 shows, paying attention to the basics can have a significant impact on reducing the number of pump repairs at a site. ■ Ken Noble has worked in various maintenance and engineering positions with Gulf Oil Canada and Petro-Canada, including having spent 12 years as a rotating equipment specialist. He is currently the Preventive and Predictive Maintenance Supervisor at PetroCanada’s Oakville Refinery. He is a registered Professional Engineer in the Province of Ontario.

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R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

Finding and Solving Vibration Problems Case Histories from the Field By John Piotrowski, Turvac Inc.

remature damage and excessive vibration in rotating machinery can frequently be traced back to improperly designed or installed foundations. These problems can be further compounded when an inadequate amount of contact exists between each piece of rotating machinery in the drive system and the baseplate or soleplate on which the equipment is sitting. This lack of contact is often referred to as “soft foot” and if left uncorrected can distort the machine case and baseplate and cause a considerable amount of frustration during shaft alignment. This article will investigate how problems of this nature caused repetitive failures on several different motor and pump drive systems, illustrate how the problems were identified, show how the problems were corrected and examine what effect the corrections had on the equipment.

P

The Bouncing Baseplate A considerable amount of vibration had existed on a filtered water pump since the initial installation of the system (Photo 1). The operators complained that the unit was very loud when running and seemed to vibrate too much. The bearings had been replaced on the pump and the motor several times with a mean time between failure (MTBF) of about 16 weeks. The

Photo 1. End view of water pump showing anchor bolt arrangement and plastic support plate between channel iron baseplate and concrete foundation

pump’s mechanical seal was also failing prematurely, resulting in production interruptions and unnecessary expenditures to repeatedly remove the unit, disassemble the pump and replace the bearings and seals. Vibration data was captured in both the time and frequency domains on the motor and pump bearings while the unit was running under normal operating conditions. The vibration amplitude levels were markedly higher at the inboard (coupling end) bearings and in the axial direction (Figure 1). The drive system operated at 3600 The Pump Handbook Series

rpm. Notice how the predominant vibration frequency is at the running speed of the machine with amplitudes ranging between 0.22 and 0.5+ in/sec. The vibration amplitude levels at the outboard bearings were around 0.1 to 0.15 in/sec at running speed. During a planned outage, the unit was shut down and safety tagged. The coupling guard was removed and an alignment measurement system was installed between the two shafts. A set of reverse indicator readings was taken to determine the alignment condition. It is good practice to check the alignment of rotating machinery before starting any work on the equipment. Capturing a set of alignment readings with brackets and dial indicators or a laser alignment system doesn’t take much time and will tell you whether shaft misalignment was one of the causes of failure. Additionally, if records are kept on the final alignment condition when the equipment is installed or the last time the unit was serviced, comparisons can be made to the “as found” alignment condition to determine if the equipment has moved out of alignment during operation. Many people are astonished to discover that equipment does not maintain the same alignment condition for extended periods of time and can shift slightly or even considerably during operation. Because it is rare that

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Figure 1. Initial vibration measurements on the inboard bearings of the motor and pump prior to performing any work on the unit

pump foot. As you can see in Figure 2, there were several points at every foot on the motor where there was no contact. This soft foot can cause the motor casing and the baseplate to warp when the motor foot bolts are tightened. About two out of three pieces of rotating machinery have a soft foot condition that is bad enough to warrant correcting. Disappointingly, the vast majority of people who install or maintain rotating machinery are unaware of this problem. It is rarely checked and/or corrected during installation and is frequently the source of damage to internal components. It can usually be traced back to excessive vibration. Warped or distorted baseplates and machine cases cause soft foot problems. Soft foot is sometimes assumed to always be due to warped or distorted baseplates or soleplates. Although there was no soft foot condition under the pump feet of this particular unit, soft foot problems can occur on any type, size or style of machine. The next case history will show how a bad soft foot condition on a pump caused a considerable amount of trouble with another unit. Again, even if someone finds and corrects the soft foot condition, records of the posture must be kept for future reference in the event that the same machine has to be serviced again. The moderate amount of soft foot on the motor feet in this example was corrected and a rough alignment between the motor and pump shafts was performed. An alignment measurement system was installed between the two shafts and a set of reverse indicator readings was taken. Figure 3 shows exaggerated diagrams of the misalignment condition

anyone keeps these records, it is not completely known why machinery alignment changes over time. It is suspected that excessive amounts of piping stress on the pumps causes the movement. It is also believed that settling foundations can cause the movement. In this particular case, no final alignment records were kept on the motor and pump, so it is not known whether the as found measurements were different from where the unit was aligned to during the last rebuild. Work began with inspections of the electric motor—a wrench was used to loosen the motor foot bolts. Surpris-

ingly, we discovered that three out of the four baseplate hold down bolts were loose. The motor foot bolts were removed, and the unit was rotated over onto its side so the motor casing could be inspected for cracks and the amount of contact between the motor feet and the baseplate could be checked. Notice in Photo 2 that the feet are “hollow” and the existing 2-inch shims were only making partial contact between the “U” shaped motor feet and the point of contact on top of the baseplate. The contact points on the baseplate were cleaned and the motor set back upright on the base. A set of feeler gauges were then used to measure the gaps at four points around each motor foot and at four points around each

Photo 2. Close up of “hollow” motor feet and the shims showing the lack of contact

Figure 2. Soft foot map showing the gaps at four points around each bolt hole

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Figure 3. The “as found” and final shaft alignment data

between the motor and pump shafts. The diagrams are called alignment graphs or alignment models. The upper left alignment graph shows the “as found” misalignment condition between the motor and pump shafts in the up and down direction. The suggested shim changes on the motor, which would be required to correct the misalignment condition, are also shown. The lower left alignment graph shows the “as found” misalignment side to side with suggested lateral moves on the motor that would be required for correction. The maximum amount of misalignment at either of the flexing points in the coupling is found and ratioed to the

distance between the flexing points in the coupling to determine the maximum misalignment deviation. This is shown in the bottom left diagram. The graphs on the upper right and lower right show the final shaft alignment positions along with the shim corrections and lateral moves to perfect the alignment. The as found shaft alignment data showed that the pump was operating with a moderate amount of misalignment (3.4 mils/in, refer to Figure 3). The unit was realigned to 0.9 mils/in. Although the final alignment positions are not perfect, the unit is now within acceptable tolerances. As shown in Photo 1, the end bell of the motor was removed and the cooling The Pump Handbook Series

fan was taken off and cleaned. Several of the end bell tie rod nuts were loose. After the soft foot condition was corrected, the fan wheel cleaned, and the shafts re-aligned, the unit was started and the vibration measured again. The before and after vibration data on the motor and the pump shows a slight decrease in vibration levels, but the vertical and axial vibration levels are still higher than desired (Figure 4). The reason for this is the way the baseplate is supported at both ends only on plastic plates (Figure 5). The lower vibration levels at both outboard ends can be attributed to the stiffer assembly at those points than at other points along the length of the baseplate. When the unit is operating, the vibration levels mid-way between the plastic plates (i.e., at the inboard ends of the motor and pump) are higher than the outboard vibration levels (Figure 5). Because the unit is not adequately supported along the entire length of the base, the center of the baseplate is free to move vertically, causing the higher vibration amplitudes. At the next scheduled shutdown period, grouting was installed between baseplate and foundation. Vibration levels were measured again and compared to the as found and the after realignment vibration spectrums (Figure 6). Notice how dramatically the vibration amplitude levels decreased after the unit was grouted. The vibration amplitudes went from 0.5 in/sec to less than 0.02 in/sec. That’s a 25-fold decrease in vibration—a remarkable lowering of vibration on this unit. But it must be understood that the amount of dynamic force in the equipment after alignment and soft foot were corrected, which generated the vibration levels in the middle two spectrums in Figure 6, still existed after the unit was grouted.

The Pump with a Lame Foot A motor and pump drive system experienced several premature bearing failures on a drive motor and premature failures of the bearings and mechanical seal of an ANSI pump. This prompted a thorough investigation of

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Figure 5. Vibration levels at the coupling are higher than the outboard ends due to inadequate support mid way on the baseplate.

Figure 4. Vibration data before and after the shaft re-alignment was performed

Figure 6. Vibration data showing the effect of re-alignment and grouting (The Bouncing Baseplate)

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the drive system. Despite using a laser shaft alignment system to align the machinery, the failures continued to occur. The bearing cap vibration was measured and the results are shown in Figure 7. Vibration data was captured in both the time and frequency domain on the motor and pump bearings while the unit was running under normal operating conditions. After the unit was shut down and safety tagged, the coupling guard was removed and as found shaft alignment measurements were taken to determine if the unit had been subjected to excessive misalignment conditions. The alignment data is shown in Figure 8. Notice that the 1.6 mils of misalignment is not perfect, but it is not severe. In fact, it really wasn’t that bad. After the alignment data was captured, radial runout measurements taken on the motor coupling hub showed 12 mils of TIR (total indicated runout) with 1 mil of shaft

Figure 7. Before and after radial vibration data on the motor (The Pump with a Lame Foot)

The Pump Handbook Series

Figure 9. Motor and pump coupling hub runout

Figure 8. The “as found” and final shaft alignment data Figure 10. Soft foot map showing uneven shim packs found under the motor feet and an excessive gap at the inboard support foot of the pump

Figure 11. Before and after radial vibration data on the pump

“freeplay.” Radial runout measurements taken on the pump coupling hub showed 10 mils of TIR with 1 mil of shaft freeplay (Figure 9). These runout conditions verified that the holes in the coupling hubs had been bored eccentric (off-center) and were the source of some of the vibration. When this was shown to the plant personnel, they became quite perplexed and did not understand why this was not seen by others who had worked on the unit before. The owners of the equipment were unaware that laser shaft alignment systems are incapable of measuring runout conditions. In addition to the excessive runout, there were an uneven number of shims under the motor feet. The 3 mil difference (182 mils vs. 185 mils) between the two outboard feet is not very alarming; the 58 mil difference (182 mils vs. 124 mils) between the two inboard motor feet is inexcusable. When the shims were removed and the motor set flat upon the baseplate, the feeler gauge readings at the four points around each bolt hole indicated that there was no soft foot condition to warrant the difference in the shim packs. In other words, a soft foot condition was built into the motor by the uneven shimming, thus causing the motor frame to warp when the bolts were tightened. Upon examination of the pump feet, an extreme soft foot condition was found under the inboard support foot of the

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Not only was there a remarkable decrease in the vibration levels on the pump, but the vibration levels of the motor dropped significantly after the work was completed (Figures 7, 11 and 12).

Conclusions

Figure 12. Before and after axial vibration data on the motor and the pump

pump (Figure 10). This distorted the pump case so badly that mechanical seal assembly in the stuffing box of the pump was incapable of seating the stationary and rotating seal faces properly. When this was shown to the plant personnel, they again did not understand why others who had worked on the unit before had not seen it. Apparently, each time the bearings or mechanical seal failed, the pump was removed, rebuilt in the maintenance shop, and then installed back on the baseplate the same way it came off. The personnel using the laser alignment system always called the motor the “moveable” machine and never bothered to examine the hold down bolts of the pump, since it was named the “stationary” machine. The world of rotating machinery will be far better off the minute everyone gets this stationary/moveable alignment concept out of their mind. The inboard pump support foot is a separate casting that is bolted to the bearing housing of the pump. The support foot was removed and examination of its underside showed that only 20% of the contact area had been machined flat. The rest was still a rough casting surface, which caused a

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complex, wedge-shaped gap in the mating surfaces. A custom wedgeshaped shim was fabricated to correct the condition. Keep in mind that Ushaped, pre-cut, flat pieces of shim stock cannot correct a wedge-shaped gap. Also be aware that when alignment measurement systems are attached to the shafts and the hold down bolts are loosened, the alignment system is only capable of detecting soft foot; it is not capable of correcting it. After measuring the runout, finding and correcting the soft foot problems and installing an even number of shims under the motor feet, the shaft alignment process was started. One vertical and one sideways move were required to align the unit within acceptable tolerances (Figure 8). It took six hours to measure runout, check the shim packs and correct the soft foot. It took 45 minutes to perform the alignment process with a dial indicator alignment measurement system, just like in the past with the laser alignment system. The pre-alignment checks and corrections are very tedious and time consuming to perform and many people skip them. However, the difference in the performance of the machinery is incredible. The Pump Handbook Series

As evidenced in the above examples, poorly designed support structures where baseplates are not rigidly attached to their foundation with grouting can have excessive soft foot conditions. This can have a profound effect on the vibration levels of your rotating machinery and significantly reduce the operating life by damaging internal components such as mechanical seals, bearings and flexible couplings. The work on the above two units was completed more than three years ago and the machines have been operating continually with no problems. By providing well-designed and installed foundations and correcting any soft foot conditions on every machine in the drive system, vibration levels will remain low and the equipment will operate successfully for a long time with virtually no required maintenance. ■

References 1. Piotrowski, John D., Shaft Alignment Handbook – Second Edition, Marcel Dekker Inc., New York, NY, ISBN# 0-8247-9666-7. John Piotrowski has been working with rotating machinery in industry for more than 23 years. He has written more than 30 technical articles, several software programs and a book. Since founding Turvac Inc. in 1982, Mr. Piotrowski has designed several alignment measurement systems, trained more than 6,500 people in shaft alignment principles and methodology, been a consultant to several major shaft alignment equipment manufacturers, participated in alignment service work in a wide variety of industries with equipment ranging from 1/2 to 25,000 hp, designed training demonstrators for alignment and balancing and invented the Ball-Rod-Tubing Connector system used to measure off-line to running machinery movement.

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R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

Spike Energy™ Measurement New technology combines with tried-and-true engineering practice to aid in vibration analysis of tough applications, including sealless pumps. By Ming Xu, Ph.D., ENTEK IRD International Corporation pike Energy was developed in the late 1970s to detect the signals emitted from defective rolling-element bearings. The term “Spike Energy” was used to describe the very short pulses, i.e., spikes, of vibratory energy generated by the impact of rolling-elements against microscopic cracks and spalls. Spike Energy is a measure of the intensity of energy generated by such repetitive transient mechanical impacts. These impacts or pulses typically occur as a result of surface flaws in rolling-element bearings, gear teeth or other metal-to-metal contacts, such as rotor rub, insufficient bearing lubrication, etc. Spike Energy is also sensitive to other ultrasonic signals including high pressure steam or air flow, turbulence in liquids and control valve noise. Spike Energy measurement utilizes an accelerometer to detect the vibration energy over a pre-determined high frequency range. The mechanical impacts tend to excite the mounted natural frequencies of the accelerometers as well as the natural frequencies of machine components and structures in this high frequency range. These resonant frequencies act as carrier frequencies and the bearing defect frequency modulates with the carriers. The intensity of impact energy is a function of pulse amplitude and repetition rate. The signal induced by such impacts can be measured by accelerometers and processed by a unique filtering and detection circuitry.

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The measured magnitude of the signal is expressed in “gSE” units (acceleration units of Spike Energy). Since its introduction, Spike Energy has been used successfully in many industrial applications and gained acceptance in various industries. Spike Energy measurement can provide early indications of machinery faults and is a very useful tool in vibration analysis. In addition to the traditional Spike Energy overall measurement, Spike Energy spectrum and Spike Energy time waveform were also developed and used in diagnostic analysis in recent years. Although Spike Energy has been used in the field for twenty years, there are still misunderstandings and improper applications of the measurement. For

example, Spike Energy is a high frequency measurement and is inherently different than the conventional vibration parameters, such as displacement, velocity or acceleration. In the conventional vibration measurements, the measured vibration signal is within the linear range of a transducer’s frequency response curve. In Spike Energy measurements, the frequency detection range is beyond the mounted resonant frequencies of most industrial transducers. Consequently, Spike Energy is sensitive to the transducer mounted resonant frequencies as well as mounting methods. It is important to understand such differences and select proper measurement parameters to obtain accurate and consistent measurement results.

Figure 1. Flow chart of Spike Energy signal processing The Pump Handbook Series

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Figure 2. Flow chart of enveloping (or demodulation) processing

Figure 3. Spike Energy peak-to-peak detection

In this article, Spike Energy signal processing, especially its unique peakto-peak detection and decay time constant features, is described in detail. The Spike Energy measurement considerations, including accelerometer mounting and data interpretation, are also discussed. Two Spike Energy case histories of pump applications are presented.

Spike Energy Signal Processing The flow chart of Spike Energy signal processing is shown in Figure 1. The

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vibration signal is measured by an accelerometer and filtered by frequency band pass filters. There are six selectable high pass corner frequencies: 100 Hz, 200 Hz, 500 Hz, 1,000 Hz, 2,000 Hz and 5,000 Hz. The low pass corner frequency is 65,000 Hz, which is the upper limit of the Spike Energy detection range. The purpose of using high pass corner frequencies is to reject lowfrequency vibration signals, such as unbalance, misalignment and looseness. The amplitudes of defect frequencies of bearings and gears could be much lower than those of low frequency The Pump Handbook Series

components. The filtered signal then passes through a peak-to-peak detector, which not only holds the peak-to-peak amplitude but also applies a carefully selected decay time constant. The decay time constant is directly related to the spectrum maximum frequency (Fmax). The output signal from a Spike Energy peak-to-peak detector is a saw-tooth shaped signal. This signal is further processed to calculate Spike Energy overall magnitude and the Spike Energy spectrum. The peak-to-peak detector in a gSE circuit is unique and very sensitive to the defect frequency compared to other envelope detection or demodulation methods. A typical envelope detection processing is shown in Figure 2. In an envelope detection, the vibration signal is first passed through a high pass (or band pass) filter. The filtered signal is full wave (or half wave) rectified. Then, the rectified signal is passed through a low pass filter to separate the modulation (or defect) frequency from the carrier frequency. The lowpass filtering has an averaging effect on the rectified signal and the peaks are smoothed in the demodulated waveform. In contrast, a Spike Energy detection circuit preserves the severity of defects by holding the peak-topeak amplitude of the impulses. It also enhances the fundamental defect frequency and its harmonics by applying the correct decay time constant. Spike Energy peak-to-peak detection is further illustrated in Figure 3. The decay time constant in Spike Energy measurement is a function of measured maximum frequency (Fmax). It is automatically selected by either the instrument (dataPAC™1500) or the host software (EMONITOR® Odyssey or IQ2000™). The decay time constant determines the shape of the signal from the Spike Energy peak-topeak detector. In turn, it affects both the overall magnitude of Spike Energy and the harmonic terms in a Spike Energy spectrum. In order to obtain consistent overall readings, only one fixed decay time constant is used for the gSE overall measurement in both

the instrument and host software. In Spike Energy spectrum measurement, a smaller decay time constant is selected for higher frequency measurements, since the defect impulse occurs more rapidly and the period of impact is more evident when using a shorter decay time constant. The amplitude readings in a Spike Energy spectrum and overall are calibrated to gSE peak values, which are similar to the peak acceleration readings. The measured Spike Energy time waveform and spectrum are shown in Figures 4a and b. The input signal for Figure 4a was from a signal generator with defect frequency at 100 Hz. The Spike Energy time waveform shows a regular pulse train and exponential decay for each pulse. In this particular case, the decay time constant was selected so that the pulse decays to almost zero before the next pulse. The spectrum shows a typical defect frequency and its harmonics. The gSE time waveform and spectrum in Figure 4b were measured from a machine tool spindle. In Figure 4b, the peak-to-peak detection and decay time constant are evident in the real machine gSE waveform. The defect frequency at 71.9 Hz and its harmonics appeared in the gSE spectrum. In applications, the Spike Energy peakto-peak amplitude can be determined from time waveform. For example, the peak-to-peak amplitude for Figure 4b was about 3.5 gSE pk-pk.

Figure 4a. Spike Energy time waveform and spectrum — signal generation data

Measurement Considerations Accelerometers and Mounting Spike Energy is a high frequency measurement and its readings can be affected by accelerometers and mounting conditions. Different accelerometers have different constructions and natural frequencies. The mounted resonant frequencies of industrial accelerometers have a wide range, generally between 10 and 50 kHz, depending upon the transducers’ construction and mounting. For displacement, velocity and accelera-

Figure 4b. Spike Energy time waveform and spectrum — spindle data

tion measurements, the results are usually consistent and repeatable regardless of which transducer is used, provided that the measurement is made within the transducer’s linear frequency range. For Spike Energy measurement, the results very much depend on the transducer’s mounted The Pump Handbook Series

resonant frequency, because most industrial accelerometers’ mounted resonant frequencies are within the Spike Energy frequency detection range. As a result, the Spike Energy readings can be different if different accelerometers are used, unless the accelerometers have exactly the same

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Figure 5. Frequency response of accelerometers (Ref. 1)

frequency response characteristics. This is, of course, rare. Therefore, it is necessary to always use the same accelerometer when collecting Spike Energy data to ensure consistency. For the same reason, it is meaningless to compare the Spike Energy readings of one accelerometer to another and no attempt should be made to do so. Spike Energy measurement also requires more stringent transducer mounting than other vibration parameters. This is because different mounting methods can result in different mounted resonant frequencies. The best mounting method for collecting Spike Energy data is stud mounting. In this case, there is only one interface: accelerometer-to-machine. It enables more transmission of high frequency signal and obtains the most consistent results. Hand-held probes should not be used for Spike Energy measurement because they result in a significant loss of high frequency signal due to their low mounted resonant frequencies. It is a common practice to use magnet holders for quick periodic checks. There are two interfaces: accelerometer-to-magnet and magnetto-machine. In order to minimize the

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loss of high frequency vibration signals in the transmission path, the contact surfaces must be flat, clean, bare and free of rust and paint. The use of a light coating of silicon grease or lubrication oil between the interfaces will improve the transmissibility of high frequency vibration signals. For stud mounting, the threaded holes should be perpendicular to the mounting surface. The length of the stud should be shorter than that of the holes to enable direct contact between the accelerometer and the mounting surface. The cable connector should be sufficiently tightened to the accelerometer to prevent it from rattling and producing erroneous readings. If the accelerometer is studmounted on a moving component of the machine, the extension cable should be attached to the component or the machine to minimize cable movements during the measurement. For magnet mounting, the magnet pole pieces should be free of dents, foreign materials and broken edges. Excellent surface contact from machine surface through any interfaces to the accelerometer is essential to obtain accurate and consistent Spike Energy data. The Pump Handbook Series

Figures 5 and 6 illustrate the mounted resonant frequencies of different accelerometers and mounting methods. The frequency response curves of two accelerometers are shown in Figure 5 (Ref. 1). Curve A and C are Entek IRD Model 943 accelerometer and curve B is Entek IRD Model 970. The accelerometers used for curves A and B are stud mounted on a shaker table and shaken at 1g with frequency sweeping from 0 to 50 kHz. The studmounted resonant frequency of the 943 appears at 22.75 kHz with a large amplitude. The 970 has two smaller resonant frequencies at 31.25 kHz and 32.50 kHz. Curve C is the same accelerometer used for curve A. The only difference is the mounting method. As compared with stud-mounted 943, magnet-mounted 943 has two lower resonant peaks at about 12.50 kHz and 18.00 kHz. The Model 943 also has a resonant frequency at about 32.00 kHz. The acceleration spectra of magnetmounted Model 943 measured from a belt-driven spindle are shown in Figures 6a and b. Figure 6a is a broadband spectrum from 0 to 60 kHz. The high frequency contents were measured up to about 45 kHz. There were almost no measurable frequencies beyond 45 kHz due to the frequency limitation of magnet mounting. The large amplitude at 9.45 kHz is the mounted resonant frequency of Model 943 accelerometer with the magnet holder. The two pole pieces had less than perfect surface contact with the spindle nose due to the curvature, which slightly lowered the mounted resonant frequency as compared to curve C in Figure 5. The zoomed-in spectrum in Figure 6b shows several peaks and haystacks between 30 kHz and 33 kHz. This is the resonant frequency of the Model 943 accelerometer, which was the small peak on curve C in Figure 5 at about 32 kHz. These mounted natural frequencies of accelerometers are excited by the impacts of bearing flaws or other defects. The mounted resonant frequencies act as carrier frequencies for

the defect frequencies. Since the mounted resonant frequencies vary with different accelerometers and mounting methods, the impact induced resonances occur at different frequencies and amplitudes. This will result in different Spike Energy readings.

Applications and Data Interpretation In machinery condition monitoring applications, the most meaningful use of Spike Energy is to trend Spike Energy readings. Because Spike Energy is a high frequency vibration measurement, it is sensitive to machine dynamic characteristics, accelerometer type, mounting conditions and measurement locations. In order to obtain consistent Spike Energy readings, it is important to always use the same accelerometer, the same mounting method and exactly the same measurement location on the machine. Lack of attention to these details may produce large differences in readings. Sealless Pumps Depending upon machine dynamic characteristics, certain machines can be monitored largely based on trend overall Spike Energy magnitudes. One such application is the monitoring of sealless pumps (Refs. 2, 3). There are two types of problems associated with sealless pump operations: process related problems, such as dry running, cavitation, flow change and internal recirculation, and mechanical problems, such as rotor rub, excessive wear of thrust and journal bearings. Conventional vibration measurements, such as velocity and acceleration, have not been very successful in the past in detecting these problems because the internal rotor mass of a sealless pump is relatively small compared to the rest of the pump. The internal processing fluid often creates confusing vibration signals to further complicate the problem. In contrast, Spike Energy is able to detect both mechanical and process related problems. The relationship between patterns of trend overall Spike

Figure 6a. Acceleration spectrum of spindle vibrations measured by magnet mounted Model 943— broadband spectrum with Fmax = 60 kHz

Figure 6b. Acceleration spectrum of spindle vibrations measured by magnet mounted Model 943— zoomed-in spectrum around 30 kHz

Energy magnitudes and sealless pump problems was established via experiments. As a result of Spike Energy monitoring, damage to sealless pumps has been consistently eliminated and the cost of pump overhauls and downtimes reduced substantially. The Pump Handbook Series

Other Parameters In most applications, Spike Energy alone is not sufficient to judge machine condition. Other vibration parameters, such as acceleration, velocity or displacement, should be used in conjunction with Spike Energy

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measurement. When Spike Energy readings are taken, one or more other parameters should also be taken at the same time so that the correlation between them can be established over time. When the trends of Spike Energy increase, it usually indicates that problems with bearings, gears or other components may start to develop. It is not necessary to schedule repair at this time if other vibration parameters are still at an acceptable level. However, it is time to pay attention to the trends of acceleration and velocity. If acceleration readings exceed the allowable vibration limits but velocity readings are still acceptable, vibration spectrum analysis should be performed to confirm the problem and repair should be scheduled at a convenient time. If velocity, acceleration and Spike Energy readings all exceed the allowable levels, the machine is approaching the end of its useful life. Sometimes Spike Energy readings may decrease and, just prior to failure, increase again to excessive values. At this point, shut down the machine to prevent the damage. Using Severity Charts with Caution For conventional vibration overall measurements, a number of general machinery vibration severity charts have been developed. However, it is almost impossible to develop a universal overall level Spike Energy severity chart for general machinery applications. Too many variables are involved: different machine types, operating conditions, accelerometers, mounting methods and ambient conditions. On the other hand, it is possible to develop an overall Spike Energy severity chart based upon empirical data for certain types of machines. The measurement conditions should be specified in the chart to ensure meaningful interpretation. One such chart was developed in the past for ball bearing machine tool spindle applications (Ref. 4). The chart applies to the data taken from ball bearing machine tool spindles operating at idle condition with Entek IRD Model 970 accelerometer and a 65 lb magnet holder. Any discrepancy in the

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measurement conditions can result in misinterpretation. This chart may not be applicable to other machines or measurement conditions. Harmonics and Sidebands Harmonics and sidebands are common phenomena in a vibration spectrum. Harmonics are integer multiples of either the shaft running frequency or certain rotation related frequencies, such as vane pass frequency (number of vanes times shaft rotational speed) or gear mesh frequency (number of gear teeth times shaft rotational speed). Harmonics are produced either by an event that repeats itself several times during one revolution or by a distortion or truncation of a sinusoidal-type signal. The typical harmonics of shaft running speed caused by low-frequency excitations are filtered out by gSE high pass filters. Therefore, harmonics of the shaft rotational speed in a Spike Energy spectrum mean that the shaft rotational speed is modulated with a high frequency carrier, such as a gear mesh frequency. It indicates the problems are associated with the high frequency carrier, for instance a gear riding on a bent shaft. Sidebands are the frequency components equally spaced around a center frequency. In practice, the sidebands are rarely symmetrical to the center frequency due to the non-symmetry of the machine or component. The center frequency is also called the carrier frequency and it may be the gear mesh frequency, multiples of bearing ball pass frequency, resonant frequency of a machine component/structure or mounted resonant frequency of the accelerometer. The sidebands are called modulation frequencies, which are induced by modulation of a signal. There are two basic types of modulations: amplitude and frequency. Amplitude modulation is a variation in amplitude of a constant frequency signal. Frequency modulation is a variation in frequency of a constant amplitude signal. In general, amplitude modulation is associated The Pump Handbook Series

with the change of loading condition and frequency modulation is associated with the speed variation. In rolling-element bearing applications, sidebands are usually one of the bearing defect frequencies and its multiples. The bearing defect frequencies include ball pass frequency-outer race (BPFO), ball pass frequency-inner race (BPFI), ball spin frequency (BSF) and fundamental train frequency (FTF). Bearing defect frequencies are non-synchronous frequencies. Amplitude modulation occurs in rollingelement bearings because the vibration amplitudes vary when the defects on inner race or rolling elements rotate in and out of the bearing load zone. In gear applications, sidebands are either the shaft rotational speed or one of its multiples (n x rpm). Amplitude modulations are present when a gear mesh has an eccentric gear or a gear is riding on a bent or misaligned shaft. A cyclic loading pattern occurs due to the periodic forcing of the teeth into the mesh. A minimum and maximum meshing force occur once per shaft revolution. As the eccentricity increases, the sideband amplitudes will increase. If the gear has a local fault that is associated with individual gear teeth or a small group of teeth, gear vibrations occur when the defect teeth are in mesh. Local gear faults include tooth space error, cracked or broken tooth, tooth surface damage and hunting tooth problems. If there is a local gear fault, the gear angular velocity could change as a function of the rotation. As a result of the speed variation, frequency modulations occur that generate many sideband pairs. In many cases, both amplitude and frequency modulation coexist. For example, frequency modulation may also occur in the case of a gear riding on a bent shaft because the tooth space measured on the pitch circle will vary when the shaft is bent. In practice, sidebands are rarely symmetrical with respect to the carrier frequency due to the non-symmetry of the loading condition and the non-symmetrical design of gear-shaft system. Since the

modulating frequencies are caused by certain bearing, gear or other machine component faults, the Spike Energy spectrum is very useful in diagnosing these machinery faults.

Case Histories Two case histories in pump applications are presented. The first is a centrifugal pump bearing problem and shows the use of both velocity and Spike Energy spectrums to analyze pump bearing defects. The second deals with a sealless pump application and shows the strong correlation between the axial rotor position measurements and Spike Energy measurements. Case History I This case history concerns a centrifugal pump used in a chemical processing plant. The pump unit consists of a vertical motor coupled to a vertical pump, as shown in Photos 1a and b. The running speed of the pump was 3,575 rpm. The vibration was measured on the pump frame close to the bearing. The data was taken by using an Entek IRD Model 9000A accelerometer with magnet mounting. The measured time waveform of acceleration is shown in Figure 7. The impact was obvious and the amplitude was quite large at about 4 g pk. In fact, this pump was very noisy when it was running. The measured Spike Energy overall was also high at 5.11 gSE, indicating high impacts. The pump bearing defect was clearly identifiable in the velocity and Spike Energy spectrums, as shown in Figures 8 and 9. In the velocity spectrum, the higher order harmonics of ball pass frequency – outer race (BPFO) were the main defect frequency peaks in the spectrum. The cage frequency or fundamental train frequency (FTF) was the sideband frequency around some of the high BPFO amplitudes. In the Spike Energy spectrum, harmonics of the FTF and BPFO were evident as defect frequencies. The pump bearing defect was captured by the periodic vibration measurements.

Figure 7. Acceleration time waveform of the pump

Figure 8. Velocity spectrum of the pump

Figure 9. Spike Energy spectrum of the pump The Pump Handbook Series

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Case History II This case history discusses some test results of monitoring a sealless pump using Spike Energy overall measurements (Ref. 3). Figure 10 shows the cutaway of a canned motor sealless pump (Ref. 2). The canned motor pump has one rotating shaft that is an electric motor rotor with an impeller mounted on the shaft. The motor and pump casings are sealed such that there are no shaft penetrations, which is different than conventional mechanical sealed pumps. The stator and motor rotor are protected from the processing fluid by a nonmagnetic containment shell. The rotor is supported by journal bearings. A portion of the pumped fluid is circulated to the motor to provide cooling and bearing lubrication. The size of the canned motor pump used in the test was 3 x 11/2 x 6 and the running speed was 3,450 rpm. The test pump was equipped with the manufacturer’s rotor position monitoring device, which monitors both axial and radial rotor position. Spike Energy overall was measured by an Entek IRD Model 911 accelerometer and the signal

Figure 10. Canned motor design of sealless pump (Ref. 2)

Figure 11. Flow changes (20 gpm increments) (Ref. 3)

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The Pump Handbook Series

Figure 12. Reduced suction pressure (at constant 150 gpm) (Ref. 3)

was processed by an Entek IRD Model 5802 Machine Monitor. The accelerometer was stud mounted on the head of a bolt that attached the pump casing to the main housing of the pump (Ref. 3). During the tests, the pump was subjected to fault conditions to simulate plant operation. The test results are shown in Figures 11 and 12. The testing data indicated a direct correlation between the axial rotor position measurements and the Spike Energy overall readings. Figure 11 shows the changes in measured data as a result of variation in flow over the range of 50 to 190 gpm. When capacities (gpm) were greater than the best efficiency point (BEP) at 150 gpm, the rotor began to move toward the suction flange of the pump and Spike Energy began to move up and increase in width. The trend curves became wider in both rotor position and gSE as the fluid flow and rotor position responded to the changes in pumping conditions. Wider traces of both gSE and axial rotor position indicate abnormal operating conditions and should be avoided in pumps, especially in sealless pumps. The result of the low-suction, pressure-induced cavitation test is shown in Figure 12. The suction pressure was reduced with the pump

capacity held at a constant 150 gpm to observe the effects on both the axial rotor position and Spike Energy. As shown in Figure 12, the axial rotor position moved dramatically toward the suction flange as the suction pressure was reduced. Spike Energy increased overall with a widening trace width. When suction pressure was reestablished at atmospheric normal, the rotor moved back to its normal run position and the Spike Energy also returned to its normal level. Both test cases showed a clear correlation between the changes in axial rotor position and Spike Energy. For sealless pump applications, continuous monitoring of both Spike Energy and rotor position will improve the reliability in determining pump operating conditions as well as the accuracy in pump fault diagnosis.

Conclusion Spike Energy measurement is a valuable tool in machine condition monitoring and fault diagnosis. The Spike Energy circuit has unique peakto-peak detection and decay time constant features in its signal processing. Compared to other bearing defect detection processing, such as enveloping detection, Spike Energy detection not only preserves the severity of The Pump Handbook Series

defects but also enhances the fundamental frequency of the defects. Since most industrial accelerometers’ mounted resonant frequencies are within the Spike Energy frequency detection range, Spike Energy measurements are sensitive to types of accelerometers and mounting conditions. In order to obtain consistent Spike Energy measurement results, the following are important. The best mounting method for collecting Spike Energy data is stud mounting. To ensure consistency, the same accelerometer, mounting method and measurement location should always be used. No attempts should be made to compare and relate Spike Energy readings under different measurement conditions. In most applications, Spike Energy alone is not sufficient to judge machine condition. Other parameters, such as velocity and acceleration, should be used in conjunction with Spike Energy. A Spike Energy severity chart can only be used when the machine type and measurement conditions match those specified in the chart. The most meaningful use of Spike Energy is to trend overall readings. The relationship between certain types of machinery faults and trend Spike Energy patterns can be established over time. Depending upon machine dynamic characteristics, certain types of machines can be monitored largely based on trend Spike Energy overall magnitudes. One successful application in recent years is to monitor sealless pumps using Spike Energy overall measurements. The defect frequency and harmonics in a Spike Energy spectrum are modulation frequencies. These frequencies are related to the high frequency carriers, such as gear mesh frequency, higher order harmonics of bearing defect frequencies, resonant frequencies of machine components and structures, or mounted resonant frequencies of the transducers. In general, the defect frequency in the Spike Energy spectrum does not represent low frequency faults such as unbalance and misalignment. In some cases, the machinery fault is not

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evident in the acceleration or velocity spectrum, but shows clearly in the Spike Energy spectrum. ■

Acknowledgements The author would like to thank Joseph M. Shea of Vibtec, Grant D. Mayers of BASF, Julien Le Bleu, Jr. of Lyondell Chemical, James Lobach of Crane Pumps, Donn Stoutenburg and Aaron Hipwell of Entek IRD for their support and assistance during this study.

References 1. Ming Xu, “Spike Energy( and Its Applications,” Shock and Vibration Digest, Vol. 27, No. 3, pp. 11-17, May/June 1995. 2. Ming Xu and Julien Le Bleu, Jr., “Condition Monitoring of Sealless Pumps Using Spike Energy,” P/PM Technology, Vol. 8, Issue 6, pp. 42-49, December 1995. 3. Julien Le Bleu, Jr. and James Lobach, “Continuous Monitoring of Sealless Pumps - The Next Step,” Proceedings of 15th International Pump Users Symposium, Houston, Texas, pp. 125-131, March 1998. 4. Ford Motor Company Vibration Standards for New/Rehabilitated Machine Tool and Facility Equipment Procurement, Engine Division, General Offices, 3001 Miller Road, Dearborn, Michigan and Transmission and Chassis, Division General Office, 29500 Plymouth Road, Livonia, Michigan, issued on June 26, 1985.

Photo 1a. Vertical pump unit—motor-pump

5. Ming Xu and Joseph M. Shea, “Using Vibration Analysis to Determine Bearing Preload in Machine Tool Spindles,” Proceedings of the 19th Vibration Institute Annual Meeting, Indianapolis, Indiana, pp. 25-36, June 20-22, 1995.

Trademarks Spike Energy, dataPAC and IQ2000 are trademarks of Entek IRD International Corporation. EMONITOR is registered trademark of Entek IRD International Corporation. Dr. Ming Xu is an Advanced Technology Specialist with ENTEK IRD International Corporation. His work involves research and development of machinery condition monitoring techniques. His specialty is in the area of dynamics and vibrations of rotating machinery. Prior to joining IRD, he was a Research Fellow in the Center for Motion Control Research at the University of Pittsburgh. Dr. Xu has published many technical papers on vibration analysis of rotating machinery and shaft alignment. He earned his Ph.D. in Mechanical Engineering from the University of Pittsburgh. He received his BSME and MSME from Beijing Polytechnic University.

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Photo 1b. Vertical pump unit—close view

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R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

In-Plant Perspective: Eli Lilly and Company Case histories from this major pharmaceutical plant illustrate some common sense solutions to chemical-pumping problems. By Gregory E. Henson, Clinton Laboratories, A Division of Eli Lilly and Company linton Laboratories is one of Eli Lilly and Company’s largest processing facilities. Located about twenty miles north of Terre Haute, Indiana, the plant manufactures bulk pharmaceuticals for animal and human consumption. The plant site contains more than 1,200 pumps of various types and sizes that are used in some unusual services. This article shares some of the pump problems and resolutions experienced at Clinton Laboratories.

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Case #1 Two magnetic pumps in parallel moving a water-caustic solution failed just five minutes after start-up for no apparent reason. The obvious cause was a loss of flow; however, this should not have happened because of the continuous feed to both pumps from a level-controlled storage tank. Further investigation found that production personnel had experienced trouble starting the pump(s) earlier that day, and an electrician had been contacted. The electrician found nothing to prevent the pumps from starting, and without communicating his actions, activated a bypass switch that eliminated all permissive from the control circuit. The pump(s) would now run under any and all conditions.

Production wasn’t aware that this was the method used to correct the problem. The electrician was not familiar with the building or the control circuit. The vibration data taken routinely on these pumps showed no increases or reason for concern. As you can see in Figures 1 and 2, the levels were normal. We still didn’t know why the pumps wouldn’t start. In the control logic there is a permissive that if the tank level is not sufficient, the pump(s) will not start. Three days before the failure, the data acquisition system indicated problems maintaining a proper level in the supply tank. The contents are a mixture of caustic and water with a set pH. It appeared that we were not getting caustic into this tank, only water. The pH probes indicated that the tank did not require water, as there was no caustic addition. Again, this should not have prevented the pumps from running out of fluid; the water would have been adequate for cooling the bearings. Further investigation found that a control valve on the caustic header outside the building had closed three days before the failure. The header pressure is normally 5,500 mmHg. Over the three days before the failure, the pressure dropped to 1,300 mmHg. There were no alarms on these valves to draw attention to the The Pump Handbook Series

situation. The low amount of caustic required in the process was satisfied from the residual header pressure, so no alarms were tripped. We had now pretty well resolved how the pumps were operated in an incorrect manner. But how could both pumps fail just five minutes apart? Reviewing the data showed that both pumps were actually in service at the same time. This should not have been possible, because the interlocks only allowed one pump to operate at a time. Our investigation found that when the electrician opened the bypass switch to get one pump operating, he also opened the switch on the other pump. The schematics showed that this did not give any indication into the data acquisition system on operating status. Pumps are started from the control room, but there would have been no operating status visible. Production started one pump, but did not get any indication it was running, so they tried the other pump. When neither pump again indicated having been started, an operator went to the pump area and found both pumps operating. By this time, though, both had actually failed. A detailed explanation of the problem and what exactly the bypass switch does was communicated to all production shifts and the electrician. There was also an information-

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al discussion with the shifts concerning the permissive and what to look for should a similar condition occur in the future. Because not every valve and switch contains alarms, the permissive protects the overall system. This particular case history is a good example of why following through to the root cause of a problem can be invaluable to eliminating reoccurring events.

The last time the pump was pulled, we did a more indepth review. The pump was designed with a suction depth of approximately 5.5 feet. When it was installed in this tank, 12.5 feet short of the bottom, an extension was added to the bottom of the pump. The manufacturer had approved this (Photo 1). The problem was that the head required for the pump to have initial suction had to be

Case #2 A vertical pump that was used to move waste solvents and liquids to an incinerator was notorious for vibration problems (Figure 3). The pump was top-mounted into a tank about 18 feet deep and had lost the bottom bearing next to the impeller a couple of times. The base bolts on top had become loose and the fan guard on the motor had vibrated and fallen off. We continued to make repairs, each time fixing something else, each time feeling we had corrected the problem.

Photo 1. The manufacturer-approved extension (Case #2)

Figures 1 and 2. Normal vibration levels for mag drive pumps (Case #1)

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Figure 3. Vibration problems with a vertical pump (Case #2)

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Figures 4 and 5. Vibration data after re-grouting (Case #3)

above this 5.5 foot depth. It could pump the tank down an additional 6 feet with the extension, but then could not recover. The impeller was still located at the same level as before. The tank was also designed so the liquids in it would at times contain some solids. If kept in suspension, they would not cause any corrosion. But if they settled out, they became very corrosive. For this reason, the pump was usually in service circulating in the tank. With an extra seven feet hanging off the end of the pump, for which it was not originally designed, we had an excessive amount of vibration at various times of operation. One other little item to complicate the situation was that when the pump did lose suction, you had to allow the tank to fill above the originally-designed suction level. That would not have been a problem, had the tank been set correctly. The tank was actually set higher than it was supposed to be by about 3.5–4 feet; the top of the tank was actually higher than the floor drains. This enabled liquid to back up into the first floor drains if the tank level became too high. We decided to purchase an entirely different design and installed a horizontal self-priming pump on top

of the tank with the suction extended to the bottom. Level probes were used to shut the pump on and off, so the level would never get up to where it could back up into the floor drains. This also enabled the pump to operate continually, circulating the contents of the tank so that no solids could settle. The new pump has been in service for more than eight months; no more liquid has backed up into the building and vibration amplitudes are stable at the installed levels. This was a good example where it had been easier to just live with a problem, or in this case several problems, than address the main issue.

with a seal going bad also require documentation for environmental reasons, which increases the costs associated with a seal problem. Evaluating the vibration data and mounting conditions indicated a structural problem that could be improved if the pumps were properly grouted. An epoxy grout was installed under the base of the motor. Following the installation of the new grout, the overall vibration dropped significantly (Figures 4 and 5). This pump has not experienced a seal failure in 18 months. Proper pump installation and grouting can save thousands of dollars over a short period of time.

Case #3

Case #4

The simple act of installing equipment properly can save you money in the long run. A centrifugal pump vibration caused mechanical seal damage on a routine basis, averaging four to five mechanical seals per year. The average cost of the seals was just over $1,000 each, not including lost downtime and labor expenses. The pump used to unload tank trailers operates approximately 1.5 hours per day, usually in the middle of the night. It is critical that the pump run for this entire time without leaking. The leaks associated

Catastrophic failure occurred when the impeller broke in a 50 hp centrifugal pump used for pumping purified water. This clean, warm water contains no solids or contaminants to cause such damage. How does an impeller in this situation break? A mechanic was repairing a leaking joint on the discharge side of the pump when he heard a noise coming from the pump. As he approached the pump, he confirmed that it was coming from the pump housing and immediately shut the pump down from a local control box. As the

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pump coasted to a stop, it began making a great deal of noise. When the pump was disassembled, we found that one of the vanes on the impeller had completely broken away from the hub. The task now was to find out why and if it could be prevented from happening again. Vibration signatures from the previous months indicated that cavitation had been occurring for some time (Figure 6). Close to the actual failure, the signature began showing a great deal of mechanical looseness. The pump’s normal operating range is approximately 325–350 gpm. This pump and flow condition would require an NPSH of approximately 18 feet. As installed and operating, there was a negative calculated NPSH available. Further investigation indicated that the vane broke on the back side of the severe cavitation marks on the impeller. The hydraulic control system was set up for pressure control with a relatively flat line pump curve (less than 5 psig difference from 0–400 gpm). An inspection by the pump supplier showed inadequate piping support on the discharge side that was confirmed with resonance testing. The base was also resonant and filled with grout. The sister pump, next to the troubled unit, revealed the same pattern. First, we rebuilt the pump. During the rebuild, the NPSH issue was addressed by putting a nitrogen blanket of pressure on the tank equal to approximately 90% of the tank rating. When everything was put back together, a support was added to the discharge piping. This reduced the resonance. Some longer term issues also needed to be addressed regarding the hydraulics and controls of the pump. The motor was resized; the 50 hp motor was replaced with a 7.5 hp unit. This changed the operating speed from 3,600 to 1,750 rpm. After this replacement, no additional problems occurred. In this instance, the installation specification was from a previous process that used the pump/system in a totally different service. “A pump is a pump,” is not always true. The entire system and its components need to be reviewed and evaluated before starting up a pump system, especially when using it for something other than what it was originally designed for.

Figure 6. Vibration signature for a 50 hp centrifugal pump used to pump purified water (Case #4)

Figure 7. The coupling for this pump was so worn that the movement looked like an unbalance condition (Case #5).

Case #5 A filtrate pump on one of the centrifuge processes had an abnormal vibration pattern. This centrifugal pump runs continuously during the entire centrifugation process. Without this pump removing the liquids, the process shuts down. The unit is an Ingersoll Rand pump with a 15 hp motor. Vibration monitoring data showed an increase in

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Photo 2. A centrifugal pump used to unload tank trailers (Case #3)

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the 1x operating range. This was a gradual increase over time; eventually the level was unacceptable. We decided to have the machine disassembled and the unbalance condition investigated. Upon removing the coupling cover to start the process, bits of rubber were found lying on the pump base. The rubber grommet coupling had severe wear. We decided at this point to just install a new coupling, check the alignment and test to see if this was indeed the cause of the increased vibration. During the alignment check, we found that the pump was misaligned by more than .040”. This amount should be less than .005”. The pump was aligned using a laser alignment system. The pump was restarted and the amplitudes fell from .168 IPS to only .034 IPS. The coupling in this case had actually been so worn that the misalignment and associated movement looked like an unbalance condition (Figure 7).

Figure 8. The decrease in vibration after a can pump was repaired (Case #6)

Case #6 A different type of pump, sometimes referred to as a can pump, is used for the circulation of glycol through the jacket on a reaction vessel. The pump is completely contained and uses a silicon carbide sleeve bearing. This pump circulates glycol at temperatures from -20 to +120°C. The pump is used to either cool or heat the reaction taking place in the tank. The unit had been monitored for more than a year and showed signs of severe mechanical looseness. Maintenance rebuilt the pump more than once; it failed requalification both times. On the third try, the mechanics contacted the vibration technicians for help understanding if or where the looseness was indicated in the signatures. The mechanics were told to check the shaft-to-bearing clearance and bearing fit. Maintenance’s opinion was that this was a waste of time; the bearings were just off the shelf and brand new. How could they be bad? They found that the shaft sleeve was undersized and enabled excessive clearance in the bearing. This was corrected and the pump requalified without additional problems. It was an excellent example for maintenance; they learned three important lessons. All clearances should be checked on a rebuild, not all new parts are correct, and someone may have made a change that was not documented and might have changed internal clearances. Figure 8 shows the decrease after the corrected repair.

Figure 9. One-year trend on pump bearing (Case #7)

Case #7 This final example is of a 50 hp centrifugal pump that transfers chilled glycol to the various process buildings. This pump contains roller support bearings on each end that are operated in an oil bath. Unlike most roller bearings, which are lubricated periodically by grease, these have continuous lubrication. This instance is one where it is advisable to consider taking more than one type of vibration reading.

Figure 10. HFD levels on pump bearing (Case #7)

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Figure 12. Acceleration signature that shows a raised noise floor at higher frequencies (Case #7)

Figure 11. Acceleration signature of problem pump (Case #7)

The PDM technician uses the overall velocity plot as a first indication for a particular problem. Figure 9 shows the trend on the pump bearing using an overall velocity plot over one year. You can see that there is some increase at the 150-day interval; however, the amplitudes are very low (below 0.08 IPS) and even with the increase, no problem would be suspected. Trends of the HFD levels, which are normally a very good indication of bearing wear or lubrication problems in rolling element bearings, showed no significant increases (Figure 10). There are some increases, but the levels are still well below an unacceptable level or the alert value. The acceleration signal showed something altogether different: increasing, dropping off, and then increasing again (Figure 11). Now something needed to be done. The acceleration signature, Figure 12, shows a raised noise floor at higher frequencies, usually an indication of a problem with a rolling element bearing. When the pump was inspected, the bearing was in pieces. The rollers or balls were so loose that they fell out of the race. About 75% of the rollers or balls had spall marks and the outer race was damaged, apparently from bad installation. A

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section of the outer race was missing, which looked as though it had been broken by an impact, probably from a hammer during installation. Had the technician not reviewed multiple parameters, or had just taken the one type of reading, this machine would have had a catastrophic failure.

Summary The information provided is an attempt to describe various conditions and situations experienced on process pumps. Several of the items discussed were not major problems, but either became costly or could have caused major damage if left unattended. It is hoped that through these examples others might evaluate their own systems and operating conditions to help save time and expenses in daily operations.

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Gregory E. Henson is a Reliability Engineer for Eli Lilly and Company at the Clinton Labs. He is acting Group Leader for the PDM area of the facility, which includes vibration and oil analysis, thermography and motor current analysis. Mr. Henson has 20 years of experience in the rotating equipment field, the past 19 with a public utility. His experiences have dealt with pumps from 1 hp process pumps to 6,000 hp turbine driven pumps. He holds a BSME from Rose-Hulman Institute of Technology and a Level II vibration analyst certification from the Vibration Institute. Mr. Henson has written and presented several papers to the Vibration Institute, EPRI, American Power Conference, PPM Technology and PumpUsers Expo.

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R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

In-Plant Perspective: DuPont What would your day be like if, when you arrived each morning, your pumps were running reliably and you didn’t have any to overhaul? DuPont’s Pumps Running team asked themselves that question and took on the challenge to make it a reality. By Scott White and Tim Holmes, DuPont

any articles have been written about improving pump performance, increasing mechanical seal life, operating pumps correctly and various other details associated with pumps. Everything can usually be reduced to sound, fundamental pump principles for installation, maintenance and operation. These principles have remained constant over the years; only the methods have changed. DuPont’s “Pumps Running” is nothing new as far as pump technology is concerned. It is a new way to organize methods and technology for people to eliminate pump defects. The Johnsonville plant celebrated 40 years of titanium dioxide white pigment production in 1998. Six years prior, in 1992, Johnsonville began implementation of Pumps Running. There are more than 600 pump installations on site, including ANSI centrifugal, progressive cavity, diaphragm, gear, single and multiple stage vertical, sealless and vacuum. Our goal in this article is to define Pumps Running for you, take you through the early years (1992-1995) of the program at Johnsonville, describe the current practices, share some examples of application and outline our plans to recharter the team.

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DuPont Pumps Running — Corporate Definition Pumps Running is a reliability improvement program designed to extend the mean time between failure (MTBF) of pumps to a world-class goal of 24 to 36 months. Pumps Running uses a team that includes people from each of the job functions that can eliminate pump defects: • maintenance • material procurement • operations • technical trades Pumps Running creates proactive behaviors in the manufacturing organization that eliminate defects in pumps from poor workmanship, materials, operations and design, as well as additional defects created by failure events themselves. This method contrasts sharply with a reactive, costtracking approach that causes the organization to focus on cheaper labor or parts.

DuPont Corporate Design Guidelines ANSI pumps are DuPont’s most common, single stage, overhung impeller, chemical pumps. The following Pumps Running design guidelines were initially developed for ANSI The Pump Handbook Series

pumps; however, many of them can be applied to other types of pumps and rotating equipment as well. • DuPont/Pump OEM Joint Product Specification (ANSI B73.1) for original purchases, including sight glass, rear cover options, fasteners, motor alignment jackbolts, coupling guard to ANSI B15.1, etc. • DuPont/Motor OEM Joint Specification and DuPont Engineering Standard for motors • DuPont Engineering Standard for baseplates • mechanical seals from preferred vendor • bearings with C3 clearances from preferred vendor • labyrinth bearing seals from preferred vendor • flexible elastomeric couplings from preferred vendor

Maintenance Specifications 1. Shafts have a maximum runout of 0.001” TIR. No repairs by metallizing. 2. Impeller balance is in accordance with ISO DR1940G6.3. 3. To heat bearings, use induction heaters with a de-magnetizing cycle. Do not heat above 230°F. 4. Mineral oil should be changed every six months, synthetic oil

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5. 6. 7. 8.

9. 10. 11.

every two years. Synthetic oil should be used if temperatures exceed 200°F. Follow the DuPont Engineering Standard for antirust protection during storage. Vibration should be maintained below 0.1 in/sec and checked once per month. NPSH should be higher than manufacturer’s recommendation by 5’ or 35%. Bearing bores should be within 0.0005” TIR, and most other fits for squareness and concentricity should be in the 0.001” - 0.002” range. Shaft and coupling alignment should be within 0.004” TIR parallel and 0.0003 in/in angular. Use a clean area for assembly. Maintain records of Mean Time Between Failure (MTBF).

Sequence of Events for Initiating Pumps Running 1. Invite a Pumps Running consultant to review the program at your site. 2. Visit other sites with existing programs. 3. Obtain management support to initiate program. 4. Select a Pumps Running Team for your site. 5. Attend recommended training. 6. Perform the Pumps Running recommended tasks.

Recommended Tasks — Methodology 1. Eliminate defects introduced through poor workmanship and material deficiencies. 2. Eliminate defects introduced by operations and failures themselves. 3. Eliminate design defects. 4. Focus the entire organization on defect elimination and institutionalizing the process.

Johnsonville Pumps Running (1992-1995) The original Pumps Running team at the Johnsonville plant was chosen from a cross-section of the entire plant. Other DuPont sites implemented Pumps Running one area at a time in order to involve all of the key people in each area. In hindsight, greater success in less time would likely have been achieved at Johnsonville using the area approach. Team members were selected from the following job assignments: • operator trainer • operations mechanical coordinator • maintenance planner/scheduler • maintenance technician • stores/purchasing • first line maintenance supervisor (2) • mechanical reliability engineer Co-author Tim Holmes was chosen as the team leader. All team members attended a one-week training session in

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Houston to get grounded in Pumps Running fundamentals. The mindset that pump failures are inevitable was challenged with questions. “What would your day be like if, when you arrived each morning, your pumps were running reliably and you didn’t have any to overhaul? What else could you do with the manpower and money you have been devoting to pump maintenance and repair?” When the team returned from Houston, it immediately generated a Mean Time Between Failure (MTBF) report for the entire pump population. The report was used to highlight “bad actors” and areas in need of early attention. We also reviewed the maintenance database to ensure that all pumps were set up in the Computerized Maintenance Management System (CMMS) and assigned to an area based on maintenance crew responsibility. Only one type/application of positive displacement pump was excluded from the initial report. No other differentiation by pump type or application was made. A “failure” was defined as a pump that had a repair work order (as opposed to maintenance or service to operations orders) written against the equipment number in the CMMS. The team explained to all operations mechanical coordinators and maintenance planners/schedulers exactly how to enter data into our CMMS so the MTBF would be accurate and representative. Mean time between failure was calculated using the following equation, with a general assumption that across the plant, 50% of the pumps are typically not online. (This lowered our reported MTBFs, because many of our pumps do not have inline spares.) Example

MTBF=

(number of pumps that are typically online) (# months analyzed) (# failures during months analyzed)

Assume an operating unit has 50 pumps, but 10 of them are inline spares that only operate when their sister pumps in the same service are off (failed, taken offline for maintenance, pumps are routinely switched back and forth, etc.). Also assume that during a three-month period, these 50 pumps experienced a total of six failures. MTBF for this example =

(50-10) (3) = 20 months (6)

All pump repair work orders were analyzed every month on a (previous) three-months running basis, and the MTBF for the entire plant and each production unit, operating area and maintenance crew was calculated. The report was issued monthly by e-mail to plant leadership and graphs showing MTBF trends were displayed and updated in each maintenance shop and operational control room. The initial MTBF for the plant’s pump population was approximately six months (Table 1).

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Maintenance Practices The MTBF data quickly exposed individual troubled installations, and from there we were able to identify substandard maintenance practices. One particular pump type was singled out, and a special team from across DuPont’s entire white pigments business was selected to solve the problems. The Pumps Running team chose a few of the bad actors and presented solutions to the respective operating areas, including trimming impellers to eliminate cavitation and replacing baseplates to reduce vibration. Pumps Running’s credibility was established with these early successes.

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thermal growth, realign every time alignment disturbed oil lubrication—keep it clean, sample it, replace it periodically operation hydraulics—operate close to BEP, adequate NPSH predictive maintenance—monthly checks, 0.2 in/sec maximum to prevent seal and bearing damage equipment history, RCFA (root cause failure analysis)

6. Once operation is successfully established, close recirculation line.

Emphasis was placed on how important it is not to run pumps with both suction and discharge valves closed. The consultant recounted a • controlled, catastrophic pump housing failure he was invited to observe. The discharge pressure of a valved-off • pump increased to several hundred psi, and the temperature to several hundred degrees Fahrenheit, within The consultant encouraged standjust five minutes! This raised our site ardizing pump design wherever possiawareness regarding the safety implible, so mechanics become more cations of automatically starting familiar with overhaul procedures and pumps from a remote location using spare parts are kept to a minimum. distributive control systems (DCS). Training The operations training outline The importance of routine operaSome aggressive training plans were reviewed the basic steps in pump starttor rounds also was emphasized. The developed and implemented across the up and shutdown (reverse start-up consultant encouraged the operators to site. A DuPont rotating equipment order listed below): routinely check pumps for leaks, consultant presented several four-hour 1. Prime the pump to remove all air sounds, vibrations, cavitation and seal training sessions on pump basics. Each pockets. pot pressure, and he gave them a basic session was specifically designed for 2. Establish seal/packing water flow. troubleshooting chart to help determechanics or operators, and all were 3. Open recirculation line off mine the root causes of pump required to attend. He also presented discharge back to supply tank (if symptoms. He discussed cavitation, a two-hour seminar on design basics for applicable). what causes it and what operators can pump systems; all mechanical design 4. Start the motor. do to prevent it. Since seal pot systems engineers were required to attend. 5. After a few seconds, open pump with buffer fluids were relatively new The maintenance training outline discharge valve. to the site in the early 90s, he also gave covered the keys to ANSI some basic training on how pump reliability: they are supposed to work. MAY APR AREA PUMPS FEB MAR • overhaul and assembly Incorrect pump operation #1 7 2.6 2.6 1.8 5.3 practices causes as many failures as #2 45 2.2 2.2 2.9 4.1 • storage of parts and assemincorrect maintenance. The #3 105 4.3 4.4 4.9 9.1 bled pumps design training outline #4 25 9.0 5.2 4.5 7.2 • shaft metallurgy and included things such as: UNIT A 182 NA NA 3.9 4.6 dimensions • maintenance and operator #5 116 21.9 17.2 17.2 19.3 • impeller balance access to all equipment #6 62 8.4 16.8 21.0 19.6 • bearing fits and installation (pump, motor, valves, #7 59 7.3 10.2 10.1 10.3 methods couplings, etc.) UNIT B 237 NA NA 15.3 16.2 • cleanliness in assembly • ten diameters of straight area piping on suction side of #8 4 nf nf nf nf • large bore seal chambers pump and no throttling #9 41 19.0 11.4 11.4 15 and cartridge seals valves or strainers directly UNIT C 45 NA NA 12.6 16.5 • non-lubricated couplings upstream • protecting bearings with • good level baseplate per PLANT 464 6.4 6.5 7.1 10.3 labyrinth seals corporate standards • quality control of spare • independently securing Notes: NA = Not Available because we didn’t parts and aligning all piping, and calculate unit MTBF at first • installation and alignment nf = no failures (i.e. small areas) results in infinite installing suction and • baseplates discharge flex joints where MTBF when dividing by zero • laser alignment—eliminate permitted, so pipe strain is pipe strain, account for Table 1. 1996 MTBF in months eliminated • •

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• piping design to facilitate priming and eliminate air pockets • discharge check valve if a waterhammer problem suspected

compromised. In many cases, closer shaft alignment could not be achieved without modification to motor mounts and baseplates, which added even more time to the turnaround.

Alignment Shaft alignment practices varied throughout the plant. We purchased additional precision, laser alignment equipment and conducted training for selected mechanical technicians from each crew. The plant’s basic alignment procedure includes the following steps: 1. Verify lock, tag and try. 2. Disconnect and inspect the coupling. 3. Lift check shaft, then move in and out to identify excessive bearing clearance. 4. Check runout of shafts and coupling hubs. 5. Verify both machines are level and feet are coplanar. 6. Inspect bases of both machines. 7. Install jack bolts for alignment. 8. Eliminate pipe and motor conduit strain. 9. Set the motor at magnetic center (sleeve bearings only). 10. Rough align with a straight edge, making sure the motor isn’t bolt bound. 11. Connect the coupling. 12. Set up a laser alignment system and correct soft foot. 13. Precision align the shafts within tolerances using the laser system, accounting for thermal growth and properly torquing the hold down bolts. 14. Lubricate the bearings and couplings as necessary. 15. Reinstall the coupling guard. 16. Start up the machine, following lock, tag and try procedure. 17. Request vibration analysis check from a predictive maintenance crew. Alignment accuracy definitely improved, but at the expense of increased repair times until the mechanics became proficient. Operations became impatient on some occasions, so the quality of the work was sometimes

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Procurement Pump parts procurement at Johnsonville includes OEM and aftermarket sources. In fact, the Johnsonville plant spearheaded the replica parts movement in DuPont in the late 70s and early 80s. As we began to study pump failures on site and eliminate defects in incoming materials, the Pumps Running team decided to go against the plant’s trend to replicate parts. We found OEM parts were more dimensionally consistent, which helped mechanics achieve Pumps Running specifications for assembly tolerances. To our surprise, we also found that OEM parts were quite cost competitive. The vendor alliance agreements probably contributed to an improved cost position relative to previous years. Another part of the replica part program included two-piece pump shafts. Two-piece shafts were designed for use in slurry applications to minimize cost when packing had to be replaced. Installation of two-piece shafts had become standard practice in almost all pump applications. Early mechanical seal installations failed prematurely partly because of shaft runout and deflection. This fact supported our efforts to use OEM parts because the new OEM standard shafts were redesigned specifically for mechanical seal operation.

Bearings Stores part descriptions were changed to require bearings with C3 clearances from a preferred vendor. That bearing manufacturer had been selected by corporate based upon dimensional consistency, broad bearing product line, worldwide presence in the marketplace and overall price competitiveness. The process Johnsonville went through to convert more than 100 different The Pump Handbook Series

bearing part numbers over to our preferred manufacturer’s product was basically: • Print a list of rolling element bearings set up in CMMS. • Ignore those with low usage levels in previous two years. • Determine preferred vendor’s equivalent bearing part number, wherever possible. • Notify bearing users of proposed conversion to preferred vendor’s bearings. • Process concerns. • Finalize conversion list. • Circulate “Management of Change” document for necessary approvals. • Convert part numbers, manufacturer and vendor to preferred manufacturer and supplier. • Begin ordering new bearings from converged supplier as old inventory is used up. • Provide technical support for a small number of unanticipated problems that surfaced with specific bearings. Team members prepared instructions on modifying existing bearing frames to accept bearing isolators. Our on-site DuPont machine shop performed the modifications as pumps were overhauled.

Standards Several team members were invited to participate in incident investigations associated with pumps. The team influenced site standards for flexible boots on pump suction and discharge flanges, pipe stress/alignment and baseplate selection. These standards were incorporated into site and some corporate project engineering specifications. We learned to be careful with our communications and statements because others tended to take Pumps Running out of context. For instance, a recommendation from an investigation of a failed flexible suction boot on a caustic pump stated that “Pumps Running recommends that flex boots should be eliminated on all chemical pumps.” Of course, that was going to be an expensive recommendation to

After baseline has been established, develop strategy for pump maintenance, training, and purchasing according to this schedule.

Tasks The Pumps Running team will

• Specify area spare requirements—2nd month

• Update existing ANSI pump installation data

• Gather best practice info from engineering standards, OEM,

• Prepare plant maintenance work procedure

etc.—2nd month

• Optimize stores inventory and plant spares • Develop training strategy for operators and mechanics • Develop specific purchase specifications for ANSI pumps.

Purpose

• Issue updated MTBF report—2nd month • Draft training strategy—2nd month • Draft MWP—2nd month • Draft purchasing specification—2nd month

Establish training, improve reliability and reduce cost for more than 200 ANSI pump installations plant wide.

• Review drafts with sponsors and areas—3rd month • Finalize MWP, training, purchase spec—4th month • Present results—5th month

Products Required

• Area implementation—6th month

• ANSI pump maintenance work procedure (MWP) • Updated CMMS (computerized maintenance management system) equipment data, cross reference lists, stores data • Operator and mechanic training plan

Monitor progress and resource areas. Resources will be contacted as needed for help with: • CMMS—item search, data format, update clearances, delete

• Purchase specifications

clearances, order point/order quantity, lead-time, CMMS data extractions

Metrics • Monthly mean time between failure report

• PC—MTBF report, engineering standards

• Stores inventory cost and item reduction

• Reliability—laser alignment, vibration

• Maintenance cost reduction compared to baseline

• Mechanical—seals, pumps, bearings, rigging, couplings

• Conduct RCFA on problem pumps once best practices

• Instrument and electrical—motors, power monitoring • Planning/scheduling—work orders, kits, history

implemented

• Area operations—history, problems, process

Guidelines

• Purchasing—price, vendor contacts, agreements

• A core Pumps Running team will develop plant ANSI pump

• Engineering—pump systems, materials, design parameters,

strategy. Each core team member will lead an area team to implement Pumps Running principles.

process change requests • Training—process change request process, process safety man-

• Monthly team meetings with minutes issued

agement, OSHA

• Team members given time to complete tasks and held accountable

• Accounting—cost data, uptime • Vendors

Timelines

Team Lifetime

Establish base line MTBF and cost data, including inventory, in

Initial charter 3 years

first month to quantify value to plant and business. This includes:

• 1 year to develop plans and implement

• Update CMMS pump data

• 2 years to monitor progress and success

• Organize pump cost data • Organize stores data for pump parts, seals, couplings, etc. • Obtain accurate substores data

Area implementation teams • Structure and representatives chosen by area

Figure 1. Pumps Running team charter

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complete and was not totally true. Once the dust settled, we were able to explain design principles of flexible boots, which led to a more thorough understanding by area operations and engineering personnel. Another example included site engineering not using stilt mounted baseplates as described in the corporate Pumps Running standard. Again, our Pumps Running team was quoted as saying stilt mounted baseplates should not be used at the Johnsonville site. We still don’t know how this happened, but we used this as an opportunity to publish short information bulletins about different Pumps Running principles with specific application information for our site. During this period, plant MTBF reached a high of approximately 16 months. Some areas of the plant consistently met the corporate goal of 24-36 months.

One Area Really Ran with Pumps Running! Co-author Scott White, a first line mechanical supervisor at the time, was asked to implement Pumps Running principles across his area. Various makes and models of mechanical seals were tested in difficult applications. Alignment training of the mechanics was supplemented with pump troubleshooting guidelines and pump assembly/repair specifications. The main emphasis became attention to tolerances, runouts and assembly details, including rebuild sheets for documentation and consistency. Within a year, the pump MTBF in that area went from 12 months to more than 30 months. The principles were successfully adapted to rotary screw, positive displacement blowers that were experiencing catastrophic failures. Improved baseplate and frame stiffness reduced vibration and synthetic oil provided better protection at higher operating temperatures. No catastrophic failures occurred once the installations had been improved.

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Johnsonville Pumps Running 1996-1999 By 1996, several of the original Pumps Running team members had retired or changed job responsibilities. Scott White became the team leader. The focus shifted from individual applications to our management systems and plant practices for pump installations. The team compiled maintenance “best” practices and turned them over to the maintenance training group for use in the mechanic training program. Tim Holmes worked with a summer mechanical engineering student to document system curves and operating points from critical pumps on the plant. The Pumps Running team prepared detailed information sheets for each centrifugal pump installation, with the intent of resolving multiple part listings for particular installations and standardizing pump designs wherever practical. Additional benefits included more rapid conversion to the DuPont/OEM default pump design and the beginning of power end exchanges with the pump OEM. Vendors were asked to provide classroom training sessions and handson toolbox discussions on a variety of topics, including seals, couplings, pumps and alignment. Several of these sessions were videotaped for future use in refresher or new technician training. Our pump OEM provided a mobile training vehicle on multiple occasions for classroom instruction and product information displays. Scott White, now in Area Maintenance Resource, worked with project engineers and mechanics to update installations to full Pumps Running standards. Several of these pumps have been in service since January 1996 without failure. Sealless pump technology was used to solve several seal problems in aggressive chemical applications. After being transferred to a new area of the plant, White helped apply Pumps Running principles to reduce as many as four pump failures per month throughout the area (area MTBF was 612 months). Some operators and The Pump Handbook Series

mechanics in this area had been through the original training, and they began their own personal version of Pumps Running called “The Red Pump Program.” The name came from the red paint used once a pump had been overhauled. Since “red pumps” began to be placed in service in May 1998, none have been removed.

Lost Momentum but Not the Concept In recent years, the plant’s MTBF for pumps has dropped to approximately seven months. Pumps have not detracted from overall plant uptime, nor have they represented a significant cost item when compared to other equipment types. As a result, plant personnel have focused their time and money on other equipment and process reliability challenges, and the Pumps Running philosophy of defect elimination has been at the heart of many of these reliability improvement efforts.

Proposed Charter for “New” Site Pumps Running Team In the near future, our attention will return to pumps (Figure 1). A “new” Pumps Running team is being sponsored by the site Maintenance Leadership Team (MLT). The team will be held accountable for results as detailed in the charter for ANSI pumps. ANSI pumps were chosen based on total number of installations across the site. ■ Tim Holmes is a Mechanical Reliability Engineer at DuPont’s plant in New Johnsonville, Tennessee. He has more than 12 years of experience in maintenance, engineering and was founder and leader of the site’s Pumps Running Team. Holmes led the team from 1991-1996. Scott White is a Maintenance Engineer at DuPont’s New Johnsonville facility. He leads the facility’s Pumps Running Team and serves on the Maintenance Leadership Team. White graduated from the Georgia Institute of Technology in 1989 with a Master’s degree in Mechanical Engineering.

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R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

Building Reliability and Safety into Rotating Equipment By Dennis Weehunt, Weyerhaeuser Corporation

n today’s highly competitive economic environment, companies need to get the maximum useful life out of their equipment with the least possible process downtime. This article introduces the reader to a time-proven maintenance procedure and plan that can maximize the useful life of process pumps, help minimize process downtime and improve safety in the work place. The following procedures have been followed closely at the North Bend Weyerhaeuser paper mill for more than six years with very positive results. There have been significant improvements in equipment life expectancy and production uptime, and product output has set companywide records. Plant safety records also have improved significantly.

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Introduction I became involved in improving the life of process pumps and mechanical seals at the North Bend Weyerhaeuser Paper Mill after attending a three-day course on pumps and mechanical seals. This course opened my eyes to the fact that we had been doing things all wrong at our facility. We were just putting out fires and not looking for the causes of failures or problems with our equipment. We needed to

make some basic procedural changes if we were going to make any improvements to the life expectancy of our equipment.

Implementing Changes To accomplish these fundamental changes, I needed to have a much better understanding of pumps, mechanical seals, pumping systems, alignment procedures and a number of other things related to pump system failures. The first thing I did was sign up for a one-year pump and mechanical seal course offered by the McNally Institute for Pumps and Mechanical Seals. (I was the first person from the West Coast to complete this program.) This course pointed out a number of problems that can limit equipment life. It became clear that the better understanding you have of your equipment and the process it supports, the more you can do to improve the operating life of that equipment. The biggest problem we were facing was that no one wanted to work on the pumps. No one wanted to take ownership or responsibility for these systems. It was assumed at the time that these systems were commercially developed and should last for years and do their job with little or limited maintenance attention. As a result, no one took the time to check the The Pump Handbook Series

important things that make the difference between smoothly operating systems and those destined to be plagued with problems. We were cleaning parts and installing new bearings, but no one was taking the time to determine why bearings or mechanical seals were failing. This had to change if we were going to make progress with equipment and process reliability.

Creating a Procedure Developing a good maintenance and equipment rebuild procedure takes some time and a commitment from everyone involved. Be assured, however, this investment will repay itself many times over. If you and your company are serious about maximizing the life of your process pumps, you need to develop a thorough understanding of proper maintenance and equipment rebuilding procedures, as well as all of the conditions that can affect bearing and seal life. Good pump technicians will want to know much more than just the proper way to remove and install parts. Some of the disciplines to be mastered are: • Understanding how to read a pump curve and how the parameters affect pump performance. Pumps that are operating within their performance envelope will

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have fewer maintenance problems. Process systems that frequently exceed pump parameters, for whatever reason, will tend to experience persistent equipment failures. The process must be carefully examined for anomalies such as stuck or improperly set valves, inoperative sensors, improper chest levels, etc. Knowing the causes of cavitation and how to prevent them. Cavitation can occur on either side of the impeller for different reasons. This is usually a process problem, but the result can be catastrophic for the pump. If the cause of the cavitation is not determined and corrected, you will certainly have persistent maintenance issues to deal with. Recognizing symptoms of pipe strain and how it can affect bearing and seal life. Pipe strain is usually a symptom of more serious underlying system problems. This is another example of where a thorough understanding of the process is as important as knowing the pumping system itself. Knowing the importance of proper pump-to-motor alignment and how to achieve it. Alignment procedures and equipment have come a long way in the past few years. Modern technology has enabled very precise alignment. This new technology and equipment can greatly improve equipment life. Understanding the importance of balancing the impeller and the proper techniques for doing it. You can’t expect a new car tire to last if it is not correctly balanced when installed. It is even more critical that a pump impeller be accurately and properly balanced before putting it into service. Improperly balanced pump impellers lead to premature bearing and seal failures, inefficient pump performance and inconsistent process control. Creating and maintaining good

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pump storage and transport bases. Although this seems like an insignificant part of pump maintenance, it is the single most important precaution you can implement for protecting rotating equipment and the personnel who handle it. Proper handling techniques and the equipment used for storing and transporting rotating assemblies can improve greatly the equipment’s life expectancy and protect personnel from serious injury. Rotating assemblies, typically large and heavy, require handling at numerous locations in a facility. Improper handling can cause bearing and shaft damage and expose personnel to serious injury. • Developing a systematic troubleshooting technique for process anomalies. Because many process pump and seal failures are directly traceable to process problems, it is important that the maintenance operation works closely with process engineers and production personnel to determine the causes of process malfunctions when they occur. Developing a close working relationship between maintenance and process control personnel in a facility will reap large rewards. Proper attention to these items can add to longer bearing and seal life and overall improvements in process pump reliability and process uptime.

Building in Safety and Reliability A very important first step to improving the life of your process pumps is to develop a sensible maintenance S.O.P. (Standard Operating Procedure). Start with inspection forms that are tailored for each brand and size of pump in your plant. Information for these can be obtained from your pump and bearing manufacturers. These inspection The Pump Handbook Series

forms should cover the following: • brand and size of bearings • bearing fit specifications for journal and bore fit • all shaft specifications • impeller size and balance specifications • type of bearing protection • shaft endplay, radial deflection and shaft run-out • location of pump and related components in the facility • date of inspection and initials or signature of inspecting personnel Just as important, and all too often overlooked, is the initial inspection of new pumping equipment before it is put into service. While most maintenance departments completely overlook this step or consider it an unnecessary repetition of the manufacturer’s quality control function, many facilities consider this a sound investment in the future reliability of the equipment and process uptime. The payback in process profitability is usually quite large compared to the time invested. Some pump manufacturers do a better job than others when it comes to assembling their products. It is highly recommended that every unit be checked before putting it into service. Once a facility has established a level of confidence with a manufacturer, inspections can be performed on a sampling basis, but a 100% inspection rate always will provide the highest level of reliability for long term performance. Listed below are examples of some of the flaws we have found on new equipment. In most of these examples it would have been difficult if not impossible to determine the cause of bearing and seal failures after the fact. • rust deposits inside the bearing housing • metal filings from machining, drilling and tapping • deposits of sand • shaft journals under or over size

• bearings that were damaged during transit or installation • lip or labyrinth seals that were damaged and/or installed improperly • impellers that were supposed to be balanced and were found to be poorly balanced or not balanced at all • sunflower seeds and other trash left in bearing housings Remember, these are examples of flaws found on new equipment. This is equipment that had presumably gone through the manufacturer’s quality control process. A typical inspection will take about an hour. The cost to repair the equipment and

Figure 1. Storage and handling base

Photos 1 and 2. Pump storage and handling bases

the loss of product to downtime due to premature failures will cost a facility many times that investment in both time and money. Think of it this way. If you increase the process uptime by the amount of time it took to do the initial inspection, you will have paid for that inspection many times over.

Procedure for Inspecting New Equipment 1. Fabricate a handling and storage base. The handling and storage of heavy process rotating units has always been a potential hazard. How many times have you had a rotating unit roll off of a pallet or slip off of the forks of a lift-truck? The base will provide safety for everyone that comes in contact with the rotating unit as it moves from one shop to the next through the rebuilding process. A pump will be handled many times by any number of people in a typical facility. Steam cleaners, mechanics in the rebuild shop, painters, storeroom personnel and the mechanics that install the unit back into service are examples of different personnel who will come in contact with the equipment. These storage and handling bases will vary from one rotating unit to the next. It is important to custom build the bases for each model and brand of pump in the plant. Figure 1 illustrates a cross-section of a typical storage and handling base. Once the pump is mounted on the storage and handling base it can safely be taken out to be steam cleaned, painted, rebuilt, installed or stored (Photos 1 and 2). 2. Check the general appearance of the rotating unit. Look for damage that might have happened as it was pulled out of service. Check carefully for damage to the shaft and the impeller. 3. Drain all used oil. Note: Oil analysis should be part of a preventive maintenance program and is helpful in determining the causes of The Pump Handbook Series

failures. Be sure to save oil to be tested in a clean container. 4. Completely disassemble the unit and clean all of the parts. Ball bearings should be inspected carefully for contamination and damage. The condition of the bearings will provide useful information about operating conditions in the bearing frame. Lubricant condition and residue should be noted. Oil analysis is often helpful. Bearing damage should be investigated to determine the cause. If the cause is not from normal wear, it should be corrected before the pump is returned to service. 5. Check the shaft, including all of the following areas: • Check bearing fits. If any bearing-to-shaft fits are outside the manufacturer’s specification tolerances, replace the shaft. • Check shaft straightness. Replace the shaft if runout exceeds the manufacturer’s specifications. • Check shaft and sleeve surfaces for grooves, pitting, cuts or wrench marks. Look for things that will damage lip seals or the O-rings in mechanical seals or bearing isolators. Replace the shaft if any imperfections are found. • Check all other dimensions of the shaft against the manufacturer’s specifications. 6. Seal the inside of the bearing housing with a rust preventive paint or epoxy to stop rust from forming in the housing and to prevent any solids from leaching out of the casting pores. 7. Install a magnetic plug in the bottom of the bearing-casing sump. This will catch metal particles and prevent them from getting into the bearings. 8. A sight glass works best to insure that the oil level is correct. The oil level should be maintained at the center of the ball in the lower half of the bearing. Ball bearings are very sensitive to over- and under-

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Figure 2. A strong base for both pump and motor is important for proper alignment.

Photo 3. Motor jack bolt towers. Jack bolts (indicated by arrows) are very important for making fine adjustments.

lubrication. Either condition is detrimental to bearing performance. Excessive heat buildup will result, and this will significantly reduce bearing life if oil levels are not maintained properly. 9. Refer to your bearing supplier’s mounting and dismounting manual for ball bearings. • Use a hot plate heater or an induction heater. Note: A demagnetization mod is abolutely essential when using an induction heater. 10. If the impeller is still good, make sure it is balanced again before it is returned to service.

Maintenance Practices That Cause Seal and Bearing Problems 1. Failure to align the pump and driver properly (Figure 2).

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Figure 3.

• Abnormal loading due to pump and driver misalignment will lead to increased loading and vibration on the bearings. This will cause heat build-up and significantly reduce bearing life (Photo 3). 2. Pipe strain that causes misalignment (Figure 3). • Pipe strain on either the suction or discharge piping can cause misalignment, vibration and excess bearing loading. This inevitably subjects the pumping system to improper forces that result in continuous maintenance problems. • If the pipe flanges don’t meet up to the pump flanges exactly, without forcing them in line, then the piping should be changed. 3. Driving on the coupling with a large hammer (Figure 4). This practice has been a common method of The Pump Handbook Series

installation in the past. This practice is guaranteed to destroy the bearings. • Improper installation of couplings and impellers has been a common practice over the years. Practices like these will only destroy the coupling, impeller and bearings in your equipment. • If the coupling becomes stuck part way on the shaft, stop, hook up a puller and remove the coupling. The first reaction is to pick up a larger hammer or drift to drive it on the rest of the way. Keep in mind that the amount of force applied to the coupling to drive it on the shaft is multiplied dramatically at the ball bearings. The impact will flatten the balls and Figure 4. Driving dent the on the coupling with a large hamraces, mer can destroy resulting the bearings. in significantly reduced bearing life. • Remove the coupling and check the bore of the coupling to be sure that you have the proper fit to the shaft. Make sure the shaft and coupling are free of any burrs or dings that will prevent them from fitting together. • Make sure the coupling is heated to the proper temperature to be installed on the shaft. Most couplings are installed with a shrink fit. The amount of shrink depends on the size of the shaft. 4. Driving on the impeller with a large hammer (Figure 5). This practice has been a common method of installation in the past, and it also will destroy the bearings.

• The impeller should slide on the shaft. If you have to drive it on, then you have damaged the bearings in the bearing housing. Remember, if the impeller is driven on, it will require even more work to remove it for maintenance later. • Check the bore of the impeller and the shaft for any burrs. File and polish as needed. • I have found that most of the problems relating to impellers not fitting the shaft have to do Figure 5. Driving with the fit of the impeller key. Either on the impeller the keyway is not cut deep enough, also can destroy bearings. there are burrs in the keyway or the keyway is not cut straight. 5. Failure to properly balance the impeller. Running an impeller that is out of balance will cause excessive vibration, reducing the life of the bearings in the pump. Impellers should be balanced to operate smoothly at the pump’s operating rpm. Used impellers should be checked for proper balance before they are put back into service. The impeller will normally become out of balance as it wears.

Operation Practices That Cause Frequent Maintenance Problems • Running a pump dry (no product to the pump) will cause overheating and excessive vibration, shortening bearing life. • Deadheading a pump can cause severe shaft deflection as the pump moves off of its best efficiency point (BEP). This translates into excessive heat that will affect both the seals and bearings. • Increased temperature causes materials to expand. Running clearances in bearings might be reduced. • Increased temperatures also can cause elastomer materials in seals to experience “compression set” problems. This will cause leaks, and in some cases, complete failure. • An increase in the bearing case oil temperature is significant because the life of bearing oil is directly related to its temperature. Lubricating oil has a useful life of 30 years at 86°F. Its useful life is cut in half for every 18°F increase in temperature. You can figure the temperature in the bearing is at least 18°F higher than the oil sump temperature. Elevated temperatures carbonize oil by forming a varnish-like film that turns into a hard black coke. These solids will destroy the bearings in time. • Operating off of the BEP (changing the flow rate of the liquid) causes shaft deflection that can cause mechanical seals to fail and overload bearings.

Figure 6. Pump housing and shaft inspection form, Worthington 6 FRBH 142

Figure 7. Required parts form, Worthington 6 FRBH 142

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Figure 8. Pump housing and shaft inspection form, Goulds 3196 MT

Figure 10. Date and location installed form

Figure 9. Required parts form, Goulds 3196 MT

Figure 11. Inspection form for paper machine pumps

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Dennis L. Weehunt has been involved in machinery maintenance and reliability practices for more than 15 years, including his current position as Pump & Mechanical Seal Technician for Weyerhaeuser Corporation’s North Bend, Oregon, paperboard facility. At this site he designed and implemented a comprehensive pump rebuilding and maintenance program that has significantly improved plant operations. Mr. Weehunt received his associate’s degree in Science and Industrial Mechanics from Southwestern College in Coos Bay, Oregon, and continues to speak and publish on pump and mechanical seal maintenance and reliability.

Figure 12. Dial indicator checks recording sheet

Conclusion Global markets are causing process facilities to reevaluate their goals and priorities. Increased economic pressures are forcing companies to improve their financial bottom lines. It is more important than ever to get the most bang for your buck in process equipment. This means optimizing equipment reliability and the life expectancy of equipment components, and improving process efficiencies and process uptime while minimizing maintenance costs. The maintenance and inspection procedures outlined in this article, plus the rotating assembly storage and transport bases described, will go a long way toward meeting these goals. Maintenance personnel need to partner with process control engineers and system operators to establish a comprehensive troubleshooting and diagnostic approach to system problems. Teamwork, sound maintenance procedures and carefully detailed equipment inspection documentation will ultimately force equipment reliability and process efficiencies up while bringing maintenance costs down, resulting in better company stability in the marketplace. ■

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R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

In-Plant Perspective: KoSa The reliability group at this chemical manufacturer found 10 keys to enhancing pump and seal reliability and has agreed to share them with you. By John Fedoronko, KoSa

he KoSa facility at Wilmington, analysis of pumps we were rebuilding proached and solved—also can be used North Carolina, is a petroin the shop. If you are losing bearings, as a model for other types of rotating chemical plant. The site has it is only logical to review lubricants equipment. In 1994, my responsibilimore than 1,000 pumps operatand the operating conditions under ties included the supervision of the ing in many different environments which the equipment was run. During plant shop. During that time, we impleand at a wide range of temperatures. the next one to two years, our field data mented the 10 Keys program in the The Wilmington plant was commisconvinced us to put synthetic oil in following areas: bearings, oil/lubricasioned in 1967 and has continually all ANSI pumps. tion, face plate/rear cover plate design, expanded over the past thirty years. alignment, maintenance repair/analyStandard Duty vs. Extreme Duty Since opening, the site has been owned sis, seal design/auxiliaries, pump Before the bearings were upgraded or operated by several companies, system design/upgrades, equipment in 1995-96, we were finding that two to including Hercules, a Hercules/Fina history and personnel. five bearings per week failed in our Inc. joint venture under the name Bearings ANSI pumps. In a control group of Hercofina, an American Hoechst (now The upgrading of our pump approximately 300 pumps, there were Hoechst Celanese) Fina Inc. joint bearings resulted from the failure six or fewer bearing failures for the venture from 1985 to 1992, and by entire year after the Hoechst Celanese extreme duty bearings subsidiaries since 1992. were utilized. DimenThe plant was acquired by sionally, the bearings were KoSa on December 10, identical; therefore, no 1998. Today, the plant is machining was required. the world's largest producer of dimethyl Bearing Design terephthalate (DMT). Upgrading from 8-ball KoSa's 10 Keys to to 12-ball and standard Achieve Optimum Perduty to heavy/extreme formance/Reliability produty bearings can be one gram is a process that has of the easiest upgrades to evolved since 1994. It is implement on your equipvery important to recogment (Photo 1). The 12nize the process aspect of ball extreme duty bearing this program. We are is currently our site focused on pumps in this article, but the process— Photo 1. Bearing upgrade on our single row ball bearing. The heavier duty bearing standard for ANSI pumps. Our current production the way problems are ap- has 12 balls; the standard bearing only 8.

T

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rates and de-bottlenecking improvements require more from the equipment. For example: A single row standard bearing is rated at 15,300 extended dynamic load (EDL), while the heavy or extreme duty bearing is rated at 20,000 EDL. Shaft vibration is also less with 12 balls than 8, which in turn has direct positive effects on seal life. The normal bearing housing oil temperature will rise, especially with the 3,600 rpm applications, when you switch from an 8-ball to 12-ball design. The benefit of having a heavier duty rotating assembly with a slightly higher running temperature has worked very well for us. One of the first steps in upgrading equipment with high bearing failure rates is to analyze the causes of the failures. Failure may result from an incorrect press fit, oil contamination, loss of lubricant film because of high temperatures, wrong ISO grade, incorrect application, work load exceeding design and other causes. An analysis and clear understanding of where the problem actually lies is the first step to upgrading. Find trends and standardize recommended solutions as much as possible. The additional cost of the heavier duty bearing is only $5–$10 per bearing. The improved performance and reliability far exceed this. Our Experience Confirmed All extreme duty bearings are not created equal. Test for and identify the one that works best in your application. Measure the equipment shafts and journals to be sure all components are within specifications. Spot-check bearings to ensure that they are manufactured according to factory specification. Shaft or journal expansion should be noted if your process requires the use of stainless steel at high temperatures instead of carbon steel or similar high strength alloys. Careful handling and care of bearings during installation is essential to ensuring long bearing life. At our rebuild facility, the bearings are heated to 200–225°F. An induction bearing heater with digital

read-out is used to heat bearings for a hot slip fit onto the pump shaft.

Oil/Lubrication Industrial Grade Oil vs. Synthetics Consider the following key factors when selecting a lubricant for your equipment: 1. What are the OEM's recommendations? 2. What is the running temperature in relation to International Standards Organization weight (ISO WT) in standard oil? 3. What is your site's history with oil-related failure and its effect on mean time between failure (MTBF)? 4. When should synthetic oil be used? Below is a general description of the lubrication approach used at our site. The main question concerning our current lubrication program is whether failures can be identified as oil-related. The OEM's lubricant recommendations are used with new equipment. Once the equipment is in service, check the housing and actual oil temperature (if possible) to be sure the temperature ratings are not exceeded.

Determine if the actual running temperature is in line with ISO viscosity grade recommendations. The synthetic oil ISO running temperature is different than standard industrial grade oil. Synthetics can run at higher temperatures with a lower ISO grade and still provide excellent lubrication. In addition, certain types of synthetics actually decant water from the oil, making it possible to drain most water buildup from the sump drain port. In the first stage of the lubricant upgrade, all pumps with oil temperatures greater than 160°F were switched to synthetic. The oil was changed in several of the pumps while they were still running. After 24 hours of continuous operation, the oil temperature routinely dropped by 15–30 degrees. Our experience has shown that pumps can run in the 220–230°F oil temperature range for months without failing, if synthetic oil is used. However, I do not recommend that you run any pumps with oil temperatures over 190°F. Equipment with oil temperatures greater than 190°F at our facility have been piped with cooling tower water to the oil cooler. Using the oil cooler usually lowers the equipment oil temperature back to a 140–160°F range.

Figure 1. The 10 Keys The Pump Handbook Series

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The lubrication program is computer-based and comprehensive. The type of oil/grease required, with lubrication frequencies, is clearly scheduled. For instance, synthetic pump oils are scheduled to be changed annually. Bearing failures are analyzed to identify the root cause. If the failure trend indicates oil contamination, a more aggressive oil change schedule can be entered into the lube program for that specific piece of equipment.

Faceplate or Rear Cover Plate (Photo 3) For the best performance from a seal, the faceplate and seal must complement each other and the seal auxiliaries. Many older pumps were bought with an old closed face that allowed for packing. At our site, as well as most others, pump packing is almost extinct. Newer pumps feature an open flow modifier that greatly enhances seal lubrication from the pumped product. Mechanical seals have emerged as the industry standard. The old closed face design actually reduces seal life because of air entrapment. The exceptions are flush applications (for example, AP1 11 and AP1 32), which force the air from the seal chamber. The following is a general review of faceplates used at our site.

Photo 2. Notice how large the seal chamber opening is in relation to the shaft diameter. Many of the new high-performance seals require more than one seal chamber.

pump. The second-best option was to bore out the old packing shoulders and, when possible, provide a slight taper. The removal of the packing shoulder enables a complete purging of the air from the seal chamber when the pump is flooded. This is an economical, lowcost way to achieve a more ideal seal condition, which produces longer

pump MTBF. The seal chamber also drains completely. Caution: Boring out faceplates can result in seal cavity erosion on processes with high quantity, abrasive solids—particularly on 3,600 rpm pumps. Of the hundreds of pumps that were modified, only one had an erosion problem using our process.

Closed Faceplates This group of pumps has the old packing shoulder located just inboard of the impeller. This design works well with a clean external seal flush. (The flush lubricates and cools the seal.) The close-fitting shoulder effectively keeps solids away from the seal because of a .8–1.5 gpm flush and a .020–.030" clearance between the shaft and the housing. Straight or Taper Bored With the exception of flush and double seals, all remaining pumps use the pumped liquid to lubricate the seal face. The ideal faceplate is an open flow modifier design. Our process, however, requires stainless steel, steam-jacketed faceplates that cost $4,000–$6,000 per

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Photo 3. Open and closed faceplates. The faceplate on the left has been machined out to enable better lubrication of the seal faces and to eliminate air vapor buildup in the seal chamber. On the right is a closed or conventional design that is standard when using packing. This design is currently used with flush applications such as API Plan 32 and mechanical flush seals. The Pump Handbook Series

Flow Modifier "Speed Bump" Design All new pumps are purchased with flow modifiers. The exception is a pump requiring a flush application. Large Bore/Flow Modifier Some of the new high-performance seals require a large or "big" bore faceplate. Certain series of double gas seals require a big bore box. Repair of Faceplate Faceplates are welded and machined back to OEM dimensions. Repaired units are also hydro-tested. The Pump Rebuilders Dial Ensure that the faceplate to shaft fit (runout) is within rebuild standards. Alignment of Pump to Motor— Vibration Equipment alignment improved when we transitioned from the reverse-dial-indicating method to using laser alignment equipment. Our goal is to install equipment with an excellent alignment of .002" or less on every pump change. The field mechanics have been trained and certified on the laser alignment equipment. Every alignment is completed with a hard copy printout in the equipment history file. Installing equipment by reverse dial indication is an exception to the standard practice. A vibration reading is taken after each installation. The vibration levels indicate the condition of alignment as well as if any soft foot exists. The spectrum analysis benchmarks the condition of other components (e.g., bearings, vane path, gears and more). Our pump installers have successfully completed a seal and pump installation class given by the reliability group. The mechanics are not certified until they install a pump in the field while being observed by one of the trainers. The mechanic must lead and demonstrate the use of required personal protection equipment as well as proper use of an impeller wrench, torque wrench and associated tools. The

Photo 4. The seal from each pump failure is inspected to identify the failure mode. The information is then logged into the company's database.

field installation information checklist also must be completed properly. Quality control is excellent and repeat work for installation errors is minimal. Our vibration analyst notes any unusual vibration history trends. A root cause analysis is carried out in the problem equipment—while it’s running—to prepare for corrective measures on the next downtime if possible.

Maintenance/Repair Analysis All the facility's rotating equipment is rebuilt at one location. The central rebuild system promotes a consistent standard for the repair process. Rotating equipment reliability engineers and specialists are in the same building and have easy access to the shop millwrights. A file is kept in the central shop file room and is maintained by the reliability group. In addition, electronic computer files and databases are kept. Repair orders or notifications are issued for every equipment repair or rebuild. Having orders on each assembly enables us to perform cost/failure analyses using the database for each of the site's cost centers. Check sheets are completed for the equipment in both the shop and the field. This process helps the mechanic The Pump Handbook Series

stay compliant with plant standards. The completed check sheets are then filed in the equipment's history file. Power ends or bearing housings have both unique numerical and barcoded tracking tags riveted to the housing. This unique tracking number is logged on both the shop and field check sheets and enables a historical tracking of the various locations and services in which the equipment has been used. This number is also an excellent identifier for confirming what seal combination a pump has or did have on the previous installation. The field and shop technicians are challenged to analyze and determine what caused the equipment to fail. This analysis process is a culture we encourage and promote. For the most part, this shop and field troubleshooting is informal, but the improvements and upgrades that ultimately result from the process are tremendous. Good communication with your shop and field craftsmen will help in the identification of equipment failure trends (whether by service, area, time or shifts). This early warning can be documented in the equipment history files. The management of change (MOC) process is used on equipment and system changes.

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In addition to the hard copy, an electric reliability database is used for pumps and equipment. This database contains much information about pumps and equipment, such as impeller size, motor, rpm, service and location. There is also a comment section in which information about significant events can be added. Maintenance mechanics are trained, tested and certified on a fieldactual pump installation based on their ability to proficiently execute a pump change according to the established plant standards (Photo 5). Our goal is to have a certified maintenance mechanic on every pump replacement performed on this site.

A thorough review of all seals and elastomers was conducted in the early stage of our pump and seal upgrade program. At the same time, we began reading the seal faces and inspecting elastomers. After several months, failure trends began to emerge. The goal was, and still is, to standardize seals for the site. The goal for the upgrades was to Figure 2. This illustration is from our operator's training manual. The helps our operators visualize where the proper oil level is and identify components graphic how critical it is to reducing bearing failures. that were rated at Seal Design/Auxiliaries least 20% above temperature and were and the whole group of pumps became Pump seal applications also were not adversely affected by our process much more reliable and MTBF rose. being reviewed, along with seal auxilchemicals. After much research and Seal analysis continues on every iaries and faceplate configurations. many meetings with seal representapump change. The more recent failThe seal population was upgraded and tives, we decided on two seal combiures relate to seal leaks resulting from standardized to work with the process nations for the majority of the site's low or no flow conditions that enable conditions. Seal elastomers and comANSI pumps. The seals were stocked the seal faces to run with inadequate ponents that exceeded the process and the pumps were systematically lubrication. environment by at least 20% were upgraded after they failed in the field. The automated process control selected. The failure rates went down; MTBF systems(s) are a large factor in low The following seals are used on began an upward trend. flow or no flow conditions. We are ANSI pumps at our site: The flush seals were matched with using double gas seals with excellent • flush the closed faceplate design and steam results in many of these pumps. The • steam heated seals were matched with the double-wet seal is targeted for use in • heated open or flow modifier designs. high vacuum services. • double-gas Seal auxiliary support systems are Pump System • double-wet also very important to long, reliable Design/Upgrades seal life. A large group of The basic rule of any pump instalpumps, for example, lation is to design the equipment for uses an API Plan 32 consistent flow. Pumps with consisflush. This setup uses a tent flows at their Best Efficiency Point clean cool liquid flushed (BEP) should be very reliable. Unforthrough the seal. The tunately, there are many pumps that liquid cools and lubrido not maintain good flow at BEP. cates the seal. In Many times flows drop below the addition, flushing is minimum requirements, putting enordesigned to keep the mous stress on the seal and bearings. process solids away from Historical data confirms that these the seal. Through seal pumps require change-outs more failure analysis and often than other pumps. research, we concluded There are numerous ways to that the existing flush confirm if the pump is at or near BEP. flowmeter’s capacity The first step is to locate the equipwas too small. The Photo 5. This complete pump assembly is used for training millment's pump curve. Then match the wrights and operators. This particular pump has a double gas seal flowmeters were sized with a nitrogen supply panel. curve with the appropriate impeller for the proper flow rate

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diameter and motor rpm. The curve should indicate gpm available, as well as head and hp requirements. Consult the operator in charge of the pump and find out what the process requires. Next, go to the control room board operator and note flowmeter readouts, if available. If the pump is controlled by an automatic system, inquire about how the logic of the system works. Be on the lookout for level control valves that will close and cause the pump to have a low flow or dead-headed condition. Compare the amp draw of the motor with the motor tag’s amp ratings. Take pressure gauge readings on the pump suction and discharge lines. All of these evaluations will give a good indication of what the pump is doing. Once all of these are gathered, cross-reference the pump curves and determine where the pump actually is in relation to the design curve.

The basic rule of any pump installation is to design the equipment for consistent flow. When you know what a pump is actually doing, the upgrade improvement process becomes much clearer. Control rooms with automated systems, digital control system (DCS) or single loop controllers can be the pump's best friend or worst enemy. The best scenario is for both the equipment and process needs to be met. The key is to communicate the needs of all the varying interests and match it to the control system's capability to give you the best possible solution.

A recirculation line with a controller is a very popular method of keeping the pump flow above minimum. This option can be included on a new installations or added as an upgrade to existing problem pumps. Matching impeller size and pump speed to the actual system will aid the pump in achieving flows at or near BEP. Our challenge for new pump installations is to consider and match process and equipment needs. Long reliable service cannot be expected if the equipment is running at or near maximum design limits.

Equipment History Maintaining accurate equipment history files is the cornerstone of an effective reliability-based operation. Without files and good databases, you cannot clearly identify success or failure trends. Analysis of cost and MTBF of equipment will nearly be impossible to obtain. Our history, including vendor print files, is grouped and centralized for easy access. These files contain original pump specifications and flow curves. Each pump failure is analyzed and logged into the pump database. Seal types, processes, line areas and cost centers are linked to the data. This detailed information alerts us to trend failures, whether they are area-specific or in a site-wide comparison. The reliability group maintains the equipment history files. Hard copies of rebuild check sheets, field installation and laser alignment sheets are kept on each piece of equipment. Each pump has a unique tracking number. This is a numerical tag with a bar code. The number enables us to positively identify spares, regardless of where they are installed on the site. This tracking number is recorded on all hard copy checksheets and has proven to be a very useful tool in our program. If there is ever a question about which type of seal arrangement is on a particular unit, the rebuilder simply refers to the tracking number The Pump Handbook Series

in his logbook and can confirm, with confidence, which seal is in the pump.

Reliability Person/Group Being effective as a reliability person requires the individual to be self-motivated. This individual should have an entrepreneurial outlook about improving the facility's equipment. Concentration should be on areas of greatest impact and return. Also needed is a commitment to continuous improvement of equipment and processes. Continuous improvement is similar to finding weak links in a chain. You fix the weak link and then monitor to find the next; fix, identify, fix, etc. The key to improving MTBF is not in fixing the same link over and over, but upgrading and eliminating the repeat failure mode.

Maintaining accurate equipment history files is the cornerstone of an effective reliabilitybased operation. Another trait is inquisitiveness— looking, probing and asking why— over and over again. There must be a thorough understanding of the equipment and the process before upgrade recommendations can be proposed. Once solutions are found, there is a need to teach; this is an informal process in the field. There are situations when a more formal transfer of knowledge must be made to a group

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of mechanics or operators. The reliability group is very active in training at our site. Thinking "out of the box" is sometimes required to unearth new solutions to old problems. There is always an element of risk to trying new ideas. The risk of new and improved methods or components will be minimal with good research and engineering practices. Question existing practices to ensure that what is being done is really the best way. Some of our major equipment improvements were realized by questioning practices and materials that we had been using for years. Identify, research and eliminate weak links. Many times we have improved equipment with very little capital spending. Staying informed about the newest technology is also very important. Seal technology is progressing at an accelerated rate. Pump conferences and publications are an excellent way to stay up-to-date about new developments. Seal suppliers, on occasion, will give you new design seals for test runs at no cost until the seals have outperformed old designs.

Operators More comprehensive training and certification of maintenance mechanics and process operators also improved the stability and reliability of the general pump population. Experience over the past few years has confirmed that many operators running the equipment have a very limited understanding of what makes the equipment "tick." Let me clarify one fact—these individuals were never given the depth of training needed to fully understand and appreciate the pump's limitations. These same operators, however, are thoroughly trained on how to run the process and what the process limits are. Operators have a tremendous effect on how well and how long our equipment will perform in the field. Through a better understanding of pump and system limitations, the operators will be able to run the

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process within the pump's capabilities. If process conditions are such that damage will occur, the informed operator should take corrective measures to protect equipment and safely restore the process to normal operating conditions. Some cultural changes must occur to promote an atmosphere of ownership and accountability for the equipment reliability, as well as for meeting production goals.

Through a better understanding of pump and system limitations, the operators will be able to run the process within the pump's capabilities. A new centrifugal pump and mechanical seal training program has been developed to increase the depth of understanding on seals and pump limitations. The operators are pretested and post-tested. Employees who take this training must pass exams that are consistent with required training standards. Additionally, area specialists and engineers should take the same training course. This course was developed jointly with the operations and reliability group personnel. The trainers for this course were selected from the operation personnel and spread across the different shifts. Using this approach, there will always The Pump Handbook Series

be trainers (hopefully champions) on the site promoting best practices for running the equipment.

Conclusion Achieving and maintaining an improvement trend in your equipment is an ongoing process. We are fortunate to have a strong, well-rounded reliability group at our site. Our vibration monitoring and analysis program is also a cornerstone in the continuous improvement process. Maintaining accurate and accessible records is another requirement. The equipment data can be entered into spreadsheets allowing easy trending of MTBF, RCFA and cost analysis. ■ John P. Fedoronko has more than 25 years of experience in the petrochemical industry. Previous roles include maintenance planning, area maintenance supervisor (production lines), plant shop supervisor, which included machinist area and rotating equipment rebuilds for the site. In the past five years, Mr. Fedoronko has been intensely involved with the site reliability improvement process, primarily pumps (1200+) and seals. The development, implementation and presentation of pump/seal training classes is currently one of his major responsibilities. The contents of this article reflect the opinions of the author and do not represent the official viewpoint of KoSa or its shareholders.

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R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

In-Plant Perspective: Thorn Creek Basin Sanitary District Strategies for effective control of pump maintenance By George Kracke, Maintenance Superintendent, Thorn Creek Basin Sanitary District

n today’s daily operations of a wastewater treatment plant, it is important to make sure the heart of the system is always working and reliable. The best way to provide this dependability is proper maintenance. Maintenance, that once dirty word, is now becoming the buzzword of the money people. They have come to realize that maintenance really can save dollars. At Thorn Creek Basin Sanitary District, we have been able to enjoy continuous operation somewhat trouble-free. Yes, there are times that a problem arises and causes a glitch, but these occurrences are rare.

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Our Mission Statement The mission of the Thorn Creek Basin Sanitary District is timely and fiscally responsible service to District residents and businesses. The District accepts all reasonable waterborne wastes from residential, business and industrial customers. Our team approach to service enables us to safeguard the environment, enhance the area’s water resources and the inherent ecology, at the lowest possible cost to users. We strive to provide the best customer service within the frame of environmental regulations and laws.

The Facility The Thorn Creek Basin Sanitary District operates under a National Pollutant Discharge Elimination System (NPDES) permit issued by the Illinois Environmental Protection Agency. The plant is rated at a dry weather flow of 16 million gallons per day, but is capable of handling four times that amount for short periods of time. The treatment plant was built in 1933 and continues to be upgraded every so often, as recently as 1996. The facility consists of primary treatment, two stage activated sludge for removal of BOD and ammonia, sand filtration and disinfection by chlorination followed by dechlorination. Gravity thickening, two stage anaerobic digestion, lagoon thickening and storage, and application to grain cropland are used to process solids. The treatment facility is operated by a staff of 35, including maintenance, laboratory, engineering, financial and administrative personnel.

Accomplishments Through good management practices, the District has been able to achieve one of the lowest rates in Illinois. In fact, our rate is 38% below The Pump Handbook Series

the Illinois average for residential users. This is possible because our highly-trained, experienced personnel maintain the level of automation, know current technology and use proper maintenance techniques to minimize cost.

Operation The operation of the plant consists of four off-site lift stations, each with three lift pumps that discharge the effluent into large diameter mains and back to the District plant. There are also gravity mains that feed directly into the plant. The head of the plant processes the effluent through screens to eliminate most large objects that could cause problems. The majority of pumps the District uses are centrifugal vertical lift pumps.

Pumps in Operation We have about every type of pump there is in our operation. There are 120 pumps and they range from just fractional to 125 hp. They include vertical lift, centrifugal, vertical and horizontal, diaphragm, progressive cavity and plunger designs. We have pumps in various sizes, ranging from 1” to 12”. Most are either 4” or 8”.

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Photo 1. This VFD was installed on one of our pumps two years ago. The problem was that the pump would experience two different head pressures, one at normal times (a low head that caused cavitation problems) and a high head on rainy days. Initially, we diverted the flow to a different location, but this caused unbalance. The pumps have run trouble-free since the VFD was installed.

Along with the variety, we also have as many manufacturers of pumps as types. The reason we end up with so many different pumps is that we have to rely on the low bidder and the criteria set by the engineering company designing the project. That, coupled with the 60-plus years we have been in operation, has created a lot of opportunities to test our maintenance skills. The variety of pumps has a definite effect on how we service and stock parts. Because we have so many kinds of pumps, we do not stock a lot of parts. The District has 10 employees in the maintenance department. We do not have expertise in any one pump design or type. Rather, our people have the knowledge and training that comes with time on the job. We do have our people attend schools and seminars to increase their knowledge base. We also take advantage of the training that pump suppliers and manufacturers offer. It is interesting to note that most vendors have good sales engineers or engineers on staff that have the knowledge to help you through an operation or repair situation. I have found their assistance very

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helpful; this is why I recommend that you get to know your suppliers and their abilities. The main technique we use here at the District is the performance of Predictive Maintenance (PM) inspections at regular intervals. This includes both visible and audible inspections. Flow is also monitored continuously by a computer that lets us know not only the current flow but also whether or not the pump output should change. This information is used to help determine how to service the pump. For the most part, we operate on a pump failure mode: a pump is rebuilt when it fails. I must include that if we note visible, audible or flow problems, we immediately take the pump down for repair. Our repairs are about 50% scheduled and 50% failure status. This is not to say this is the most effective way to operate, but it has worked for us. We are fortunate and have the luxury of ”triple redundancy,” that is, we normally run one pump in a station of three, where one pump is able to service the normal load. This is an advantage that many companies don’t have. Most of our failures are seal related. Because we pump solidsladen waste, a lot of debris is captured in the area around the seal and impairs the flushing process. Another cause of failure is the impeller itself. Both age and debris cause unbalance and affect the seals and bearings. Most of the time, we can detect the obvious problems, such as bearing failure and impeller damage, when we do our PM inspections. We do almost all our own pump repairs and rebuilding right on the premises. Over the years, we have acquired personnel that have the skills to make this possible. Recently, we have been taking advantage of current, up-to-date technology. This is often in the form of newer style seals, bearings, lubricants and pump materials. Many of the original pumps installed at our plant are still in service. I guess that says something The Pump Handbook Series

for both the manufacturers and our maintenance department personnel.

Spares and Parts Inventory Having many different pumps has a negative effect on spare parts. Over the years, we have been able to standardize on a seal supplier and have changed many pumps from using packing to employing mechanical seals. We also have a bearing supplier nearby that allows us to get our bearings in about one to two days. On some of the very important pump applications, we do keep a spare shaft and other internal parts on the shelf. There are only about five pumps that fall into this category. We have established a variety of suppliers, e.g., bearings, seals, machine shops, pump rebuild shops, that we can call on as necessary. Again, this is where developing a good vendor base and rapport is important.

Methods Used at TCBSD for Pump Repairs When a pump needs repair, we follow standard lock out and tag out procedures and remove it. The pump is brought back to the maintenance shop and cleaned for disassembly. Care is taken to ensure all parts are removed properly to avoid causing any additional damage. All parts are cleaned for identification and inspection.

Photo 2. This installation solved the problem of transporting sludge of up to 4% solids over a great distance. When the sludge was close to 2% or less, we could use both pumps in parallel. When the higher solids were being pumped, we valved the pumps in series to achieve the needed higher head.

At this point, we try to determine why the component failed, using previous history and recent PM

reports as a starting point. The shaft, impeller and volute are examined for wear; proper repairs are made if any wear is found. These repairs may be in the form of rebuilding or replacement. We try to repair any part within economic reason. When repairs are made to the rotating parts, the entire assembly is sent out for balancing. We found this precaution to be prudent and cost-effective. We have several vendors capable of this work nearby. The pump is reassembled and care is taken to install the bearings, seals and other parts correctly, according to the service manuals that are supplied by the pump manufacturer. Many times the motor is not considered part of the system because it seems to have a very long life span and almost never fails under normal conditions. That thinking could be in error. We examine the motor during the repair and it is evaluated for PM work. We go back to the motor’s history to determine the last time it was serviced. If the service log is seven years or longer for dry locations, or five years for wet, we send the motor to a local motor repair facility. We request that the motor be disassembled, cleaned, dipped and baked with new insulation and bearings. There are some special criteria for motors. We do not repair motors under 5 hp unless they are special duty or a unique design. All motors in this category are replaced. We replace the bearing seals and gaskets during almost all rebuilds. We have found that even though the seal may have failed, it is prudent to replace the bearings and other parts. We normally take this precaution if the system has been in service for longer than one year since its last repair. This is usually economically feasible since the cost of bearings is less than $150. Yes, there is additional labor cost, but we feel it is worth the added effort. Why do our seals fail in less than a year? Most often because debris caused the flushing water to stop

flowing. We have made many inroads to improving the life and reducing the downtime of our pumps. Use of predictive maintenance and other tools and methods has been very helpful. Some of the changes we have made, and the advice I developed from these experiences, follow. • Improve seal water supply source. Make sure that it is always clean and free of obstruction and that the pressure flow through the seal is adequate. Provide visible water wheels to show indication of flow. Inspect the seal water daily. • Where possible, make changes to the impeller size to ensure proper operation. For us, this was a onetime repair that we did because the pump was over-designed for the operation. • Adjust discharge valves so they maintain proper head and keep the pump operating within its curve. We do this if the pump will experience different heads while in service. • Make piping changes to the inlet and outlet to provide laminar flow where possible. This reduces cavitation. When retrofits were made in the past, we installed new pumps or made upgrades at low bid cost. • Make electrical measurements to make sure the system is balanced. One possible cause of system imbalance is corrosion of the starters and contactors by the gases in the plant. We have found over the years that electrical problems can lead to motor failures. We have to take into account that the hydrogen sulfide gases attack the copper wiring and can cause a current unbalance because of resistance. • Make weekly inspections of all pumps and run them. Listen for changes in noise and look for vibration increases. Listen to the operator, because he or she lives with the systems everyday. • Make temperature measurements of the bearing end of pumps and motors at periodic intervals. Look The Pump Handbook Series

for any unreasonable changes. • Ensure that proper lubrication procedures are in place and programmed into the maintenance schedule. This includes making sure the correct lubricants are being used. • Whenever a pump is rebuilt, realign and balance it in place.

Photo 3. These two pumps have basically the same job: recirculating sludge in the digesters. The pumps experienced cavitation when the digester and head were low, causing air to be captured at the impeller. We were able to reduce cavitation significantly by repiping the inlet side of the pump. We no longer have to replace the impeller every two years.

The use of different tools and methods to accomplish the above is important. We make use of infrared thermometers for temperature measurements, handheld vibration meters, amprobes for electrical load inspection and stethoscopes for noise checks. When time comes to do an annual inspection, we contract with a local shop to do the specific test. Once a year, we have all our electrical equipment inspected for temperature faults. This locates the bad connections that can cause motor unbalance. We have a local mechanical analysis shop do our vibration studies and send us a report that shows the comparison to the previous inspection baseline. This has proven to be very valuable and has saved us on several occasions. When we install equipment, we go back through the old records in the data system to determine a baseline for future inspections and repairs. We then adjust the dates for these procedures based on the information we found.

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Common sense, good troubleshooting skills and a dedication to a strong work ethic are the most important attributes of a successful maintenance program. In addition to the methods above, we still rely on good old-fashioned human intervention. The operator who makes rounds each shift notes any odd or different conditions of a pump and uses computers that monitor the operation. Operators note when a difference in flow occurs or a pump is not running.

Conclusion Common sense, good troubleshooting skills and a dedication to a strong work ethic are the most important attributes of a successful maintenance program. Setting up a good control system using software is also very helpful. We use Datastream, but there are a lot of other programs on the market. The ability to go back in time is an important step in making repair decisions. Having confidence in both

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Photo 4. We had a minor noise problem in one of the pumps we had made inlet piping changes to. The culprit turned out to be the eccentric reducer you see here. It is important to make sure that the proper fittings are used and placed in the correct position.

employees and suppliers is also critical. Keep accurate records of all repairs and any other activity that transpires during the process. Training is also very important. We have all the original documents available to reference when working on a pump. Having good general pump manuals is also very helpful. The employees can refer to these for help or to find techniques that will make the repair go smoothly. There are numerous helpful manuals and educational material available for this purpose. One that comes to mind, and that we use here, is the “Pump Handbook Series” published by Pumps & Systems Magazine. Much of information is from real life experiences and includes many good practices. Other reference materials like the Piping Handbook (McGraw Hill, Nayyar), and the Pump Handbook, Second Edition, are also consulted. One thing I find helpful is to work with both operators and maintenance personnel to diagnose the problem and to try and find out why the failure took place. We try to have these employees go into each repair with as much information as we have, so they can see the whole picture. This is not always possible, but can be helpful. I would be remiss if I didn’t include the safety aspect of pump repair. It is very important when working on pumps that all safety issues be addressed to safeguard both the employee and the equipment from The Pump Handbook Series

injury or damage. All employees should be up to speed on safety concerns and how to address each one. Lockout/tagout procedures are the first item on any repair or major inspection task. In pump repairs, as in any situation, you need to “stay on track.” You must continue to update your files and keep track of the latest tools and procedures that become available to help your plant maintenance department. By doing this, you can keep maintenance cost to a minimum. Remember, though, that you cannot hide from the inevitable. A pump will fail and you have to be ready. Be ready with confidence. ■ George Kracke has been the Maintenance Superintendent for the Thorn Creek Basin Sanitary District for eight years. He has more than 27 years of experience in the maintenance field, and has held both Journeyman Electrician and Millwright cards. Mr. Kracke received an Associate’s Degree in both Mechanical and Civil Technology, and he is a member of several associations including ISA, SME and the Association for Facilities Engineering.

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R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

Paper Mill Stock Pump Improvements By Bob Matthews, Owner, Uptime Resource, Inc.

ulp and paper manufacturers are in a constant struggle to reduce the cost per ton of producing paper. Often drastic reductions in the per ton cost are necessary to maintain a competitive edge. In this battle, downtime and maintenance are key issues. One way to keep these costs low is with reliable stock pumps and maintenance techniques. Most stock pumps, no matter who manufactures them, will have some common characteristics that make this article useful. Here are just a few of the many ways to improve these pumps that I have seen or implemented

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Bearings Bearing changes have been written about in many case studies. I remember bearings being the first big improvement we made to our stock pumps when I worked for a major paper mill. Changing the radial ball bearing from the fill slot design, which can enable the ball to get caught by the slot, is one improvement. The use of machined brass cages on the thrust bearings and brass cages on radial bearings also became a mill standard for all stock pumps. The bearing manufacturer’s field-tests show that the brass cages are far more durable than steel or plastic. The feature that impressed

me about them is that they make a lot of noise at the initiation of failure. This alerts the operators and gives them time to set up a spare rotating assembly for change-out. The steel and plastic cages can fail without warning, and these are often catastrophic failures because they seize instantly. The instant lock-up will likely cause the inner race to spin on the shaft. The shaft will be damaged beyond repair, often taking other parts with it.

Which Stock Pump to Use End suction stock pumps with removable suction-side plates are user-friendly. Many of them have a four-bolt, bolt-in arrangement that enables you to install them in any orientation. Removing and rotating these suction-side plates in a timely manner extends their life considerably. The frequency at which you rotate them depends on the amount of wear over time, and that wear is usually isolated more in one area of the suction-side plate. With a little effort, the maintenance team can figure this out. This plate is costly, and rotating can save the cost of the next one or two side plates in an application. The new dual volute casings are very forgiving on suction-side plates, and rotation doesn’t need to be as frequent. The Pump Handbook Series

Upgrading Materials Upgrading brass and cast iron wet end parts to 316 stainless steel can increase the reliability of the entire system. This can be done when the parts are from the same pattern, which will reduce stores inventory and the mill’s taxes. When we started the switch-out to stainless, we did an inventory on the old stores. We had the same parts in stock in three different materials. Most technicians preferred the stainless because chemical additives used today attack the other two materials. The brass and cast iron parts were left sitting on the shelf. The materials department is always looking for ways to reduce inventory, and this is a win-win situation. Some higher upgrades may be necessary depending on the service, and 317 stainless steel and CD4MCu are available from most pump companies.

Dynamic Seals and Seal Boxes Dynamic seals are a great improvement in end suction stock pumps. These seals are essentially the addition of a repeller fitted between the stuffing box and the impeller. The dynamic seal reduces the stuffing box pressure at startup; when shut down, a conventional static seal prevents the fluid from leaking. These are especially

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advantageous when there is no seal water present for most, but not all, applications. There are many new seal boxes, some for packing and some for mechanical seals. Look at the differences and try to make the best decision possible for your application. If you can look at them in operation at the supplier’s shop, trade shows, on videotape, or the web, you can figure out what really works and what doesn’t.

mills have these lines open to the Udrains, but this reduces the seal water pressure, causing stock to flow under

the packing and destroy the sleeve and packing. This stream of water costs most mills about $2,000 per year

Mechanical Seals Here are a couple tips for cartridge mechanical seal applications. Use low profile seals, ones that provide as much space around them as possible. If a bored stuffing box is required, you can bore some of the packed stuffing boxes to save money. Our machine shop would bore 0.050 to 0.100” per side depending how meaty the stuffing box was and what the service was. They started 0.125 to 0.250” inside the box from the gasket face and bored straight through. The last 1” was tapered three to five degrees, again depending on the meat of the stuffing box. This arrangement works well with a seal flush in many applications and can have thread-like grooves machined at the taper in the flow direction. If you are building up or compressing too much stock around a mechanical seal in a packing-style gland add lip seals ahead of the mechanical seal. Some stuffing boxes with short low profile seals will allow one, two or three non-metallic lip seals. Position the lip seal so the seal water flows under the seal and stock cannot flow back. If seal water is a problem, put an inexpensive 110V PD pump there to ensure water quantity and pressure.

Figure 1. Modify a packing style stuffing box for a mechanical seal with little cost.

Packing Packing and packing boxes in stock applications get confusing. There is an inlet and an outlet at the lantern ring port for seal water. The outlet is to check gland pressure. Make sure the lantern ring isn’t plugged. Many

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Photo 1. Stock pump with gland water ports for L5 or 2L3 packing arrangement (This pump is piped 2L3.)

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in water cost alone for just one pump. Multiply that by the number of pumps and you end up with a large sum of money. The fan pump is usually piped this way, and premature failures are a certainty. Many pump stuffing boxes have two sets of seal water ports: one for 2L3 and the other for L5 packing arrangements. The 2L3 is used in diluted applications like white water; the L5 is for stock pumps. When using the L5 arrangement, plug the unused ports smooth inside the stuffing box bore for even take-up on the packing. (This reduced hang up does help.)

Bearing Housing Sealing the bearing housing is necessary where humidity is high and washdowns are common. Good bearing isolators are the first step, and these should keep out all contamination. Plastic isolators did a very poor job in the mill I worked in; the hot areas warped them or they softened and the shaft spun in the rotor O-ring. They did work in ideal conditions, but we did not have the spares for good and bad condition pumps. Brass is a good isolator material, and if, on occasion, contamination is a concern, upgrade to 416 SST or what the industry offers. Pumps should not leak oil, so lubricators are not necessary; this is questionable at some sites. It you use an L type sight glass, put a vent line at the top and

run it to the top of the bearing housing. This enables the bearing housing air and not the outside humid air to balance the oil level. An expansion chamber is a nice enhancement to any closed system. This is simply a can air chamber with a balloon-like bellows that fills and relieves with air as the pump bearing housing expands and contracts with thermal change. Pumps that require lubricators can be fitted with balance line lubricators that don’t ingest moisture.

Rebuilds Better rebuilds are a good goal to set, and we should all try using fewer hammers and more micrometers. Dial indicating the stuffing box face and achieving the required perpendicular alignment to fit in a correctly positioned volute makes alignment easy. When you indicate these faces on horizontal end suction pumps, turn the pump up in the vertical position or shaft sag and bearing fits will sometimes give false readings. This keeps you from having to machine foot bolts or slot motor feet holes. Document everything and keep a detailed equipment history. Good balance with the impeller and coupling, if possible, will add to the equipment life.

Installation Practices During installation, check the pump base. Look for washout, cracks,

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looseness or anything that can cause the pump to shift. Shim specification upgrades are common in most mills. The modern shim makers should guarantee their shim size dimensions. Don’t put more than four to five shims to the stack. All shims are to be free of paint, cracks, dirt and debris, and they should all be the same alpha size. Pump feet and base surfaces are flat and clean to specs. These changes are easy to implement; the hard one is to not use the impact on the pump case. Proper torque will help keep the alignment straight and not squeeze the glue out of the gasket. This article just scratches the surface on things to be done to improve stock pump performance, but it is a great start. If you follow the tips and advice given here, you will be well on the way to better reliability, and, the major goal, lowering that all-important per ton cost. ■ Bob Matthews is the owner of Uptime Resource, a reliability consulting firm in Baytown, Texas. He has more than 30 years of experience in mechanical applications, 10 of which were in hands-on pump repairs. Mr. Matthews spent several years as a reliability specialist at a major paper mill. He is a frequent contributor to Pumps & Systems Magazine and has been a presenter at all PumpUsers Expos. Contact him directly at [email protected] or by phone at 281-421-4918.

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R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

Root Cause Failure Analysis Do you know why your problem solving is ineffective? This new way of thinking can drastically improve your maintenance program. By Dean Gano, President, Apollo Associated Services

he term Root Cause Failure Analysis (RCFA) has been associated with equipment reliability for the past ten or fifteen years and is generally understood to mean find and fix the causes of equipment failure. Many companies have no structured failure analysis. Instead, they operate on the broke-fix system: If something breaks, they fix it, but they don’t ask why it failed. When the problem happens a few more times, they may look for a root cause and fix it differently. These organizations often have measuring systems that record how long it took to repair a failure. Maintenance personnel get very good at repairing the equipment, but the failures keep happening. Organizations that are striving to learn and improve often have some formal root cause analysis program, but their effectiveness is often lacking, and repeat events are not uncommon. Unfortunately, in most businesses today, the majority of failures will be repeated and they will be repeated simply because the employees have never learned effective problem-solving skills. The purpose of this article is to examine the causes of ineffective problem solving and to present a problemsolving methodology that anyone can use on any problem and find effective solutions every time.

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Conventional Wisdom First, let’s examine what conventional wisdom is telling us. Conven-

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tional wisdom would have us find the root cause and remove it, but after 13 years of studying problem solving I have discovered the fallacy of most root cause thinking. It is linear and categorical and thus fails to disclose all the causes needed to find effective solutions. There are many other reasons for our ineffective problemsolving skills, but let’s examine the two basic schemes that have been developed to find root causes.

Figure 1. The Ishikawa Fishbone Diagram

Categorization Methods Over the past 25 years this has been the most popular method, but it is no longer used in many industries because of its history of ineffectiveness. Categorization methodologies provide a predefined hierarchical outline, checklist, or “Cause Tree” from which a root cause is chosen. Through group discussion and consensus, a solution is attached to the defined root cause and implemented. Categorization is a likely choice because it is the “Operating System” of the human mind. We compartmentalize and categorize everything, including ideas. It is only natural, then, The Pump Handbook Series

that we would try to use this process for problem solving. What we fail to understand is that categories are not causes. That is, categorization places causes in predefined boxes, such as People, Procedures and Hardware. Unlike cause and effect relationships, there is no necessary causal relationship between the defined categories. Perhaps one of the most publicized categorical methods is the Ishikawa Fishbone Diagram. This method asks us to guess at possible causes of the event, usually in a brainstorming session with knowledgeable people, place the results of our guessing within predefined categories, like a fish skeleton, and then vote on which causes we think are the root causes. The fishbone diagram is shown in Figure 1. As you can see, causes are listed categorically instead of in time-based causal relationships of “A” caused by “B” and “C”. In addition to this weakness, Ishikawa taught that each cause must be supported by evidence, but like other methods, he did not provide any way to graphically represent evidence, leaving the user to input biases and unsubstantiated wishes. Using the Ishikawa Fishbone Diagram is as random as spear fishing in an icecovered pond. Make several holes in the ice and poke the spear up and down until you hit a fish. If you get nothing from one hole, go poke in another hole (category). Repeat the process until you have several fish and then vote on which one is the best looking.

If we focus on categories instead of cause and effect relationships, we fail to identify the specific causal relationships necessary to understand what happened. By not understanding the causal relationships, solutions are often ineffective and the problem is repeated. A classic example of this can be found in the category of human error. Categorical thinking may set this up as a cause and often dictates punishment as the solution. In the 13 years I have been studying problem solving, I find that placing blame and punishment is used about 20% of the time, and is effective in less than 1%. To stop at a categorical cause such as human error is to ignore the infinite set of specific causes behind the failure. Another fundamental problem with categorization schemes is that the hierarchy of causes is subject to the biases or “realities” of whoever creates the categorization scheme. The notion that any individual or group could possibly create a cause tree that would include all the possible causes of every event-based problem is so arrogant as to be laughable. This will become more obvious when we discuss the cause and effect principle later in this article.

Causal Relationship Methods These methods are based on the understanding that everything that happens has a cause, and by knowing these causes we can control them and hence find effective solutions. While a sound basis, this method is only as effective as the tools provided. All but one of these methodologies fails to understand the four elements of the cause and effect principle. Instead, most of these methods encourage the problem solver to tell a linear story and fill out some forms or charts to document the probable root cause. Without a structured representation of the causal relationships and the evidence to support the causes, individual biases, political power and ignorance prevent effective solutions most of the time. In order to better understand the cause and effect principle, let’s examine it more closely.

The Cause and Effect Principle For at least 5,000 years, mankind has used the notion of causation to express happenings. Unfortunately, we have failed to differentiate the immense power of the Cause and Effect Principle from the simple notion of causation. Causation tells us that everything that happens has a cause; the Cause and Effect Principle provides four basic characteristics that enable us to develop effective problem-solving tools. The four characteristics are as follows: 1. Causes and effects are the same thing. 2. Causes and effects are part of an infinite continuum of causes. 3. Each effect has at least two causes in the form of actions and conditions. 4. An effect exists only if its causes exist at the same point in time and space. Knowing that cause and effect are the same thing only viewed from a different perspective in time helps us understand why people can look at the same situation and see different causal relationships. They are actually perceiving different time segments of the same event. If we treat each perspective as a different piece of a jigsaw puzzle, we can stop the usual arguing and work on putting the different pieces together. Knowing that causes and effects are part of an infinite continuum of causes helps us understand that no matter where we start our investigation, we are always in the middle of a chain of causes. This helps us understand that there is no right place to start. Again, just like the jigsaw puzzle, we can start the problemsolving process anywhere and still end up with a complete picture. This avoids the usual arguments over who is right. Probably the most profound characteristic of the Cause and Effect Principle is that each effect has at least two causes in the form of actions and conditions. This teaches us that every time we ask “Why?” we should find at least two The Pump Handbook Series

causes and for each of these causes we should find at least four more causes, and from each of these four causes we should find at least eight, and on to sixteen, thirty-two, and so on (Figure 2).

Figure 2. The infinite set

With this understanding, we see that there is an infinite set of causes for each effect, limited only by our lack of knowledge. It is this element of the Cause and Effect Principle that is responsible for our pursuit of simpler strategies. We never knew how to deal with an infinite set of causes before now. This is what the Apollo problemsolving method is all about. Cause and effect relationships exist with or without the human mind, but we perceive them relative to time and space. From observation we see that an effect exists only if its causes exist at the same point in time and space. In Figure 3, an open fire exists because conditional causes came together with an action cause at a particular point in time and space. As we can see, three conditional causes: oxygen, oily rags, a match, and one action cause, a match strike, occurred at the same point. If these four causes did not exist at the same time and space, the fire would not exist. For example, if the oily rags where stored in a closed can, or if the match was struck at a different time, a fire could not exist. Understanding this characteristic helps us determine the validity of causal relationship. By understanding these four relationships we can devise some simple tools that will enable us to tap the awesome power of the Cause and Effect Principle.

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Figure 3. Example of cause and effect relationships

Effective Problem-Solving Tools For the past 13 years, I have been working with many different industries to develop a simple set of tools that everyone can use on any event-based problem. This methodology, which is as much a philosophy as it is a set of simple tools, does not require forms or checklists. When fully implemented, it empowers everyone in the organization to be an effective problem solver. It truly changes the way people think about the world around them and improves their ability to effectively communicate what they know. People begin to realize that things do not just happen, and that everything is caused to happen. As a result, a proactive attitude begins to develop, and conditional causes that set people up to fail are removed before they can cause a loss. This methodology, called Apollo Root Cause Analysis (ARCA), provides a four step process based on the cause and effect principle. Step 1. Define the problem by writing the What: Primary Effect (noun verb) When: Relative time of the Primary Effect Where: Location in system, facili-

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ty or component Significance: Why you are working on this problem?

Step 2. Create an Apollo Cause & Effect Chart (Figure 4) For each Primary Effect ask why. Look for causes in Actions and Conditions. Connect causes with “Caused By.” Support causes with evidence or use a “?”

of this later. An added feature unique to this method is the inclusion of evidenced-based causes. When you can’t find evidence, or can’t find the next cause, use a question mark. Evidence should be data we know from using our senses, not supposition. It supports the existence of the various causes. Without sensed evidence, we are subject to our own prejudices and preconceived ideas. It helps to be humble when pursuing causes and evidence. When you don’t know, admit it and use a question mark to signify this on the chart. If there is value in pursuing the unknown, then do so. If no value is perceived, you can make a conscious decision to stop, like we did after “Car Existed,” or leave the uncertainty on the chart, like we did after “Narrow Road.” If evidence is uncertain, use a question mark to express your doubts. Look closely at Figure 5 to see how the question marks have been used. To better understand the four-step process, let’s take a closer look at each step.

Problem Definition Let’s start by defining what a problem is. A problem is the gap between actual and desired. If the goal is to produce a product safely and efficiently and someone is injured in the process, or the cost of production exceeds the sales price, there is a gap

Step 3: Identify effective solutions Challenge the causes and offer solutions. Identify the best solutions—they must: • Prevent recurrence • Be within your control • Meet your goals and objectives

Step 4: Implement the best solutions When completed, an Apollo Cause and Effect Chart may look something like Figure 5. Please note the branched causes connected by the words “Caused By.” Using “Caused By” prevents storytelling by forcing our mind to go from present to past. More on the importance The Pump Handbook Series

Figure 4. Basic chart elements

between actual and desired. Every problem can be defined in terms of a gap. With this fundamental understanding of a problem, every problem should be defined within the context of a goal. Therefore, to define a problem, we must first know what the

goals are. Likewise, we must know what our present state is. Sometimes the goals are not clear. Have you witnessed groups within an organization, such as Maintenance and Operations, which appear to have conflicting goals? Do they end up seeing different problems? Does it lead to miscommunication and conflict? Do different groups within an organization really have different or conflicting goals? It may seem like they do, but if the organization is to be effective, these goals must be aligned. This perception of conflicting goals is the result of not understanding the goals. As an example, consider a football team. The goal of a football team is to win the game. At the highest level everyone’s goal—players, coaches and managers— is to win the game and everyone’s actions are focused on that goal. But let’s look at a different level. What is the goal of the

offense? What is the goal of the defense? At this next level, the offense’s goal is to score points. The defense’s goal is to prevent their opponent from scoring points. The offense and the defense, at this next level, have completely different goals: score points and prevent points from being scored. Different goals at this level, same goal at the higher level. Is there a conflict? No. At the highest level they have the same goal: win the game. Work groups and individuals should always consider their individual, group or team goals within the context of the overall goals. If individual goals are not aligned to the overall goals, sub-optimization can occur at the expense of the organization’s goals. How do we overcome this apparent conflict in goals? By defining the organization’s overall goals within the significance section of the problem definition. Every problem, incident,

opportunity or project should be defined within the context of the organization’s overall goals. A clear problem definition helps to get everyone “on the same page” when an incident occurs. It improves the understanding of what is really important to the organization—a crucial step toward improving the organization’s overall effectiveness and ultimate success in preventing problems from occurring. The four elements of problem definition are listed below. What Asking, “What is the problem?” can generate everything from a common response to a variety of answers from a group. By asking the question, we are beginning to clarify the gap between desired and the actual we are trying to prevent. People will respond from their individual points of view, which may

Figure 5. Example of Apollo Cause and Effect Chart The Pump Handbook Series

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or may not reflect how the overall organization sees the problem. If there are two or more responses, simply separate them with a comma. The “Significance” section will align the different perspectives in an organization by providing the primary effect(s) for the cause and effect chart. There may be more than one problem, and you are going to find it out here. When There are two components to this section. The first part of the “When” is the date and time. Sometimes it might be important to capture an incident in terms of which second the incident occurred in a process step or it may be sufficient to simply state “in the morning”—it depends on the nature of the problem. The second part of this section is the relative timing. It may help to think of this in terms of the contextual factors associated with this incident. You might ask, “Were we using this equipment for the first time? Is this the first time this person has performed this task or used this application?” This may provide important causal reference points once you’re into the analysis. At this point in the problem definition, we don’t know whether they are important or not, so we need to write them down. Where This section should define the physical or process location of the incident or problem. A consistent approach of starting with the higher levels of a system and stepping to the lower levels works well for developing a clear structure that captures the setting. The systems approach helps develop physical locations that accurately reflect the actual layout of a work group or facility. There is also a relative aspect of the “Where” that may help to capture the subtle, but potentially important causes, such as: “In the corner where the light doesn’t shine.” This prompts us to ask during the analysis phase: “Why no light?” Significance The “Significance” section asks the question at the beginning of the analy-

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sis process: “Why are we spending time and resources on this issue?” To take full advantage of the significance, it should reflect the overall goals of the organization. The incident perspective should not be from any one group or individual, like the maintenance group or department manager, but from the stated goals of the company. This helps everyone to see the problem more accurately. It also causes clearer communication regarding priorities and economics. Look at the example below for a difference between a single point of view and one that looks at the overall goals of the organization. The Primary Effect or “What” of this example was stated as a “Pump Loss.”

Significance: (Single point of view) Minor leakage, no reportable spills, $2,500 pump replacement Significance: (Expanded to reflect all company goals) Safety: Near miss of flying pump impeller parts Environmental: Minor Leakage, no Reportable spills Production: 20% reduction in pulp processing line #2, cut back in throughput of 15,000 pounds per hour at $0.30/lbs. for 6 hrs. equals $27,000 Maintenance: Materials = $1,250 pump replacement Labor = $1,250 Frequency: Second time this year, three times in 1998, five similar pumps in unit 3 In the above example the significance has been expanded into five areas that reflect the overall goals of the company. Even though the maintenance group only saw the pump loss as the “problem,” it is important for everyone in the organization to define the problem within the context of the overall company goals. Properly stating the significance generates the primary effect for the cause and effect chart. For example, the primary effect could be changed to “Production Loss,” which is eventually caused by the “Pump Loss.” It is a subtle but important The Pump Handbook Series

distinction required to improve communications, because the focus changes from one of operational or maintenance perspective to one of the company goals—produce product safely. Problem definition is not only the first step, it is critical to effective problem solving because it forces us to understand the problem from all perspectives, thus improving the communications of the organization.

Create an Apollo Cause and Effect Chart For each Primary Effect (the “What” in your problem definition) begin asking why. Use yellow stickies on a board to document your progress. As answers come, connect the causes with the words “Caused By.” This will go a long way to ensuring a good chart, because the words “caused by” force the mind to go from the present to the past, and thus minimize storytelling. When you reach the end of a line of causes, go back to the beginning and start over. Each time you ask why, try to find two or more reasons. Look for condition and action causes. If you have identified an action cause like an operator-pushed button, identify the conditional causes that had to be present at that same point in time and space to allow the effect to occur, like “button existed,” “operator at panel,” etc. After going through the cause chart four or five times, begin to write down the evidence that explains how you know each cause exists. Once you have done this several times it gets very easy and natural, but it may be difficult at first if you are used to thinking and communicating by storytelling. Storytelling is a significant barrier to effective problem solving.

The Problem with Storytelling Storytelling is a linear understanding of an event in a time sequence from past to present, and it significantly violates the cause and effect principle. While stories are our primary form of communication, they conflict with the cause and effect principle in three ways: 1. Stories start in the past, while causal relationships start with the present.

2. Stories are linear, while causal relationships follow the branches of the infinite set. 3. Stories use inference to communicate meaning, while problems are known by sensed causal relationships. Let’s examine a simple little story to see how detrimental these conflicts are: The little crippled boy lost control of the run-down wagon and it took off down the hill on a wild ride until it hit the little blind girl next to the drinking fountain by Mrs. Goodwin. The little boy was in the wagon the whole way, but was not injured. The boy’s mother should never have left him unsupervised. The root cause of the girl’s injury was human error. Stories Start in the Past This story starts in the past at the top of the hill and progresses through time from the past to the present; from the beginning of the ride to the end; from the safe condition to the stated problem of injury. The conflict this creates is that by going from past to present, we do not see the branched causal relationships of actions and conditions. If we could know every cause of this injury example, we would see a diagram of cause and effect relationships similar to Figure 3. That is, we would see a set of ever-expanding causes starting with the injury and proceeding into the past. To express what we know causally in story format, we would first need to express all the causes on the right side of the diagram, i.e., starting from the past. Our language and the rules of storytelling simply do not allow for this. We can not express 16 causes and then tell what they caused and so on. No one would sit still for a story told this way, because stories are about people, places and things as a linear function of time. Stories Are Linear As we look at this simple story (or any story), we find that our language and our mind restrict us to a linear path through time and space. Stories go from A to B to C, without regard for the order of causal relationships. We are told of

the little boy losing control of the wagon as it goes down the hill and strikes the little blind girl. There is no ever-expanding set of branched causes expressed like those in Figure 5. We have the ability to escape this linearity and express branches if we use the words “and” and “or,” but the rules of grammar tell us not to use these connecting words excessively. The best we can accomplish is one or two branches for each sentence. The conflict arises because the cause and effect principle theoretically dictates an infinite set of causes for everything that happens while stories are created and expressed linearly. Stories Use Inference to Communicate Causes Since good stories seem to provide us with a valid perception of what happened, we need to question how this can occur in light of the above two conflicts. The key word here is perception. When we read or hear a story our mind provides most of the information. As we read the words, we are busy creating images in our mind’s eye. These images are created from past experience and assembled into a sequence of events. We don’t necessarily need causal information to create the image; our mind fills in with its own causes. Because stories or the pictures we create do not express the branched causes of the infinite set, we must make up for it somehow, and we do this by inference. We infer causes within the story that are not stated. For example, we read that the little crippled boy lost control of the wagon. Since no cause is stated for how he lost control, we can infer anything our mind will provide, and we do just that if questioned about it. Furthermore, stories imply cause by the use of prepositions such as in, on, with, etc. Prepositions and conjunctions by definition imply a relationship between words, and the relationship is left to the reader. The word “and” is often used to mean “caused.” In this story we read that the boy lost control of the wagon and it took off down the hill, meaning the loss of control caused The Pump Handbook Series

the wagon to take off down the hill. Within this “and,” is the potential for many causal relationships and they are left for the reader to interpret. For me, the “and” between “lost control” and “took-off down the hill” is obviously a broken steering mechanism. You may have inferred that crippled means a paraplegic and this condition was the cause. The next person sees the wagon wheel strike a rock, which causes the wagon to veer sharply, while another person is so shocked by the politically incorrect usage of “crippled” and “blind children,” that she has lost the ability to think about the problem altogether. Because we do not express what is happening causally, each word in the story provides the reader with the opportunity to think they know more about the event than is stated. We interpret the situation from our own biased mind, which is not necessarily what happened or what the storyteller meant. In the end, each one of us thinks we know what happened but we really don’t because stories can not express the full set of causal relationships. Our linear language, and the linear thinking behind it, prevents us from knowing and expressing what really happens in any given situation. And when we get together to discuss our problems, we usually end up arguing and making presumptuous statements like: “It’s obvious why this happened,” or “the solution is clear.” By breaking away from storytelling and creating an Apollo Cause and Effect Chart, we are able to include all possible causes without the usual arguing and politics.

Identify Effective Solutions Once we have an Apollo Cause and Effect Chart, we can look at the causes and propose solutions to remove or control them in a way that will prevent recurrence of the primary effect. This is a two-step process. First the creation stage, where we challenge each cause and offer solutions. Next, we evaluate each proposed solution against the three solution criteria: prevents recurrence, is within our control, and meets our goals and objectives. Looking at

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Figure 5, we can see several opportunities for prevention, such as not parking on the shoulder of the road, or not listening to annoying radio personalities while driving. Each cause provides the potential for effective solutions and the more causes we can put on the chart, the greater the chance of finding a creative solution. Every solution is a function of our goals and objectives. It matters not what someone outside our organization thinks of our solutions, because we are the ones who are going to be responsible for the success or failure. Accepting this responsibility is what makes any solution “right” for our organization. If our team or organization can show that our solutions are supported by clear, evidence-based cause and effect relationships and meet the three solution criteria, then our solutions will be effective. If, along the way to our solution, another stakeholder does not agree with our analysis or solutions, they are enthusiastically asked to share their evidenced-based causes. If they are able to improve our understanding, then we are obliged to add them to the chart and modify our conclusions as necessary. A common occurrence for people who use Apollo Root Cause Analysis is to take the problem analysis with its causes and recommended solution all the way to the president of the corporation and rarely encounter opposition. Why? Because there are no opinionated stories, just evidence-based causes. The chart provides a clear picture for anyone to see how effective the solution will be. Because the causes are clearly stated and supported by evidence, there is no need to test the solutions, as the other methodologies would have us do. During a visit to a manufacturing plant where Apollo Root Cause Analysis has been in use for almost a year, the plant manager confided in me, “It is a little scary letting go, but if I can successfully apply the solution criteria to the corrective actions, then I let it go—regardless of my gut feelings or opinions. I have not been disappointed yet, and this process has truly empowered our workforce. They love it because it really works.”

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Competitive Advantage In the week following a training session on this new approach, a reliability engineer at a prominent chemical company in Europe put these methods to use. After clearly defining the problem and its causes, he and his team fixed an old problem for the last time. After 35 lifetime failures of a main process compressor, their solution completely eliminated the problem. To date, they have realized savings in maintenance costs in excess of $750,000. This company, like many others who have learned to create a common reality, is now using this new approach for solving event-based problems throughout the world and is gaining a competitive advantage by reducing maintenance costs while improving productivity and safety. In today’s global marketplace, everyone can work harder, longer, or faster, but only a few will choose to work smarter. Working smarter means better communications and better problem solving, which translates to a more productive and safer work place. Using the ineffective methods of storytelling and categorization may have been good enough last year, but they won’t take you into the future. Effective problem solving is a definite competitive advantage both from a production standpoint and a safety perspective. We have seen how using the Apollo Cause and Effect Chart can help us communicate better and thus become more effective problem-solvers and decision-makers. By first defining the problem with the What, When, Where and Significance, and then preparing an Apollo Cause and Effect Chart, a common reality is created such that everyone in the organization can provide creative solutions that meet the three solution criteria. As more personnel within an organization learn this simple structured approach to event-based problem solving and effective communications, a questioning attitude begins to permeate the work culture. With this new attitude comes the understanding that things do not just happen, and if we look more closely at causes and effects, we can prevent problems from occurring The Pump Handbook Series

in the first place. With these simple communication tools, employees are able to communicate their ideas like never before. Likewise, the time and money saved not having to resolve repeat events and emerging issues can be used for positive business activities such as increasing productivity or upgrading equipment. You can’t afford not to improve your communications and problem-solving skills. ■

References 1. Churchland, Paul M., 1996, The Engine of Reason, The Seat of the Soul, MIT, Cambridge, MA 2. Covey, Stephen, 1990, The Seven Habits of Highly Effective People, Simon & Schuster Fireside Books, New York, NY 3. Damasio, Antonio, 1994, Descartes’ Error, Grosset/Putman, New York, NY 4. Gano, Dean, 1999 Apollo Root Cause Analysis - A New Way Of Thinking, Apollonian Publications, Yakima, WA 5. Goleman, Daniel, 1995, Emotional Intelligence, Bantam Books, New York, NY 6. National Safety Council, 1983, Accident Investigation: A New Approach 7. Senge, Peter M., 1990, The Fifth Discipline, Currency Doubleday, New York, NY 8. Van Doren, Charles, 1991, A History of Knowledge, Ballantine Books, New York, NY 9. Wilson, Paul, et. al., 1993, Root Cause Analysis, ASQ, Quality Press, Milwaukee, WI Dean Gano is President of Apollo Associated Services, an international business that provides Root Cause Analysis training. Mr. Gano has more than 25 years of experience in process industries, including nuclear, petroleum and chemical, and has been teaching Root Cause Analysis for the past 11 years. He is a member of the American Society of Mechanical Engineers, the American Society of Safety Engineers and the American Society for Quality.

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R E L I A B I L I T Y- D R I V E N PUMP MAINTENANCE HANDBOOK

Continuous Monitoring of Sealless Pumps It’s the next step to reliability. By Julien Le Bleu, Principal Mechanical Engineer, Lyondell Chemical, and James Lobach, Chief Developmental Engineer, Chempump Division of Crane Pumps

ll centrifugal sealless pumps, both canned motor and magnetic drive (Figures 1 and 2), should be monitored to determine mechanical condition. In sealless pumps, the pumped fluid is used as the cooling and lubricating medium for the pump bearings. If only intermittent monitoring is used, the chance of detecting pump damage caused by process changes is very small. Conventional vibration monitoring techniques used with sealed pumps have proven unreliable for detecting problems with sealless pumps. The effectiveness of conventional monitoring techniques is limited by the time interval between measurements, the relative isolation of the inner pump rotor from the outer measuring location and by the pumped fluid. Other factors such as fluid effects and process noises can make interpretation difficult. This article presents the synergistic combination of two relatively new methods of sealless pump monitoring. These methods (monitoring rotor position, vibration analysis) considerably enhance the range and magnitude of mechanical problems that can be identified on this type of pump.

A

Figure 1. Mag-drive pump

Rotor Position Rotor position is a measurement that has been requested by sealless pump users for a long time. As long as

Figure 2. Canned motor pump The Pump Handbook Series

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Figure 3. Pump test loop

Figure 4. Data acquisition block diagram

the bearings are not badly worn, serious damage caused by rotor-tostator contact cannot occur. Rotor position monitoring as a predictive tool is minimal when silicon carbide bearings are used. When bearings made of softer materials are used, such as carbon/graphite, the technique becomes predictive because it enables the user to track wear on the bearings and schedule maintenance before serious damage is done.

Vibration Analysis Overall High Frequency Tracking (OHFT) is a vibration technique used for detecting problems with rolling element bearings. The overall value is as an indicator of pump health or process problems. Experience has demonstrated that a narrow trace, especially at a low value, is desirable. This observation was validated during the course of testing.

Test Conditions The pump was installed in a test loop consisting of instrumentation, a supply tank and associated piping. A schematic of the test loop is shown in Figure 3. The pump was subjected to conditions that attempted to simulate what can be encountered during

Figure 5. Test pump curve

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Figure 6. Flow changes (20 gpm increments)

Figure 7. Large flow decrease The Pump Handbook Series

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plant operation. The pressure on the supply tank could be varied, giving testers the ability to induce or eliminate cavitation in the pump to measure its response with the sensors. Testing was done at a manufacturer’s facility. The pump was equipped with the manufacturer’s rotor position monitoring device, which monitors both axial and radial rotor positions. OHFT was measured using two accelerometers that were connected to a dual channel monitor. It was reasoned that measurements of rotor position, OHFT and power would provide sufficient information to determine the pump’s mechanical condition (Figures 4 and 5). We hoped this combination would also provide advance warning of process conditions that would adversely affect the pump’s health. Measurements were taken for the following pump operating conditions: • Pump capacity range from shutoff to 30% greater than BEP with data taken at 20 gpm intervals in the range • A sudden large increase in pump flow • Air leakage into the suction of pump (injected) • Dry pump operation (part of the “air leakage” test) • Best Efficiency Point (part of first item above, this was considered a base line) • Reduced NPSHA The following information was recorded for the operating conditions listed above: • Motor input power (watts) Labeled (D) (1 volt = 0 kW, 5 volts = 15.5 kW) • OHFT 0-16 gSE on the casing labeled (B) (1 volt = 0 gSE, 5 volts = 16 gSE) 0-5 gSE on the rear bearing housing labeled (C) (1 volt = 0 gSE, 5 volts = 5 gSE) • Rotor position (Axial and Radial) Axial labeled (E) (1 volt = .000” in, 5 volts = .100” in) Radial labeled (F) (1 volt = .000” in, 5 volts = .013” in)

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• Flow labeled (A) (1 volt = 0 gpm, 5 volts = 300 gpm) 3” Brooks MagFlowmeter • Suction pressure labeled (G) (1 volt = 0 psia, 5 volts = 30 psia) Absolute pressure transducer The data in parentheses are the plot scale factors for the data presented in the article. An increase in voltage for the rotor axial position data represents a movement of the rotor toward the suction flange of the pump. The underlined letters in parentheses represent the letters on the graphs for that data set.

Test Results Changes in Flow (Figure 6) The changes in data as a result of variation in flow over the range of 50 to 190 gpm are represented in Figure 6. Note that at capacities greater than the BEP, about 150 gpm, the rotor begins to move axially and OHFT begins to increase in value and width. Neither of these is desirable in sealless pumps. Large Flow Decrease (Figure 7) A large flow decrease from 200 gpm to shut off is represented in Figure 7. OHFT, indicated by (B) and (C) on the chart, is at a high value and a wide trace at 200 gpm, indicating cavitation and pump stress. The axial rotor position also has a wide trace during this part of the test, indicating the rotor was “hunting” to find its hydraulic balance. When the flow is decreased from 200 to 150 gpm, the rotor moves to a normal and more balanced axial location and OHFT decreases from 10 gSE to approximately 4 gSE. Large Flow Increase (Figure 8) Large flow increases are illustrated in Figure 8. At very low flows, the OHFT value is high and the rotor axial position is near center. At BEP, 150 gpm, OHFT is at a minimum level and rotor axial position is slightly toward the motor, relative to its position at BEP. When flow is increased to levels significantly higher than BEP, the onset of cavitation is indicated by an The Pump Handbook Series

increased width of signal in both the OHFT and axial rotor position data. Rotor axial position oscillates at the high flow level, and this “hunting” of the rotor is indicated by the wide data trace (E). The rapidly changing pressure balance on the impeller during cavitation causes this oscillation. The OHFT signal also becomes less stable and “nosier” when the pump is cavitating. Suction Side Restriction (Figure 9) The suction valve was closed at a constant flow to measure the response in terms of the measured parameters. The results are illustrated in Figure 9. The width of the trace representing axial rotor position and OHFT begin to increase. Both of these indicate instability. One is in the rotor position and the other is in the pumped fluid. This could represent a suction strainer plugging in the suction line if there were one or a valve that was not opened fully. It also can represent a fluid that has become too hot and is flashing in the suction of the pump. Failure of sealless pumps due to inadequate suction conditions or other cavitation-inducing operation can be minimized through the connection of the monitors to appropriate operator alarms. Air Injection and Dry Run (Figure 10) The air injection and resulting dry run test in Figure 10 is a graphic representation of the effects on axial rotor position when air leaks into the suction. OHFT immediately begins to decrease in amplitude. The radial position of the rotor changes because the loss of fluid in the radial bearings and around the rotor reduces the radial support stiffness. OHFT levels are very low because of the loss of coupling fluid and transmissibility between the pump casing and the rotor. The watt meter shows no load on the pump and the flow has fallen to zero. This test does show that a “dry run,” as with a tank pump out, does not result in an instantaneous

Figure 8. Large flow increase

Figure 9. Suction side restriction The Pump Handbook Series

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catastrophic failure. The effects of dry running and severe cavitation are cumulative in our experience. This is especially true of mag-drive pumps. Low Suction Pressure Induced Cavitation (Figure 11) While factory NPSH testing defines the onset of cavitation as a 3% loss of head, cavitation effects are sometimes seen well before a measurable head loss occurs. Continuous monitoring of OHFT and rotor axial position represents a practical method to measure the actual onset of cavitation through the direct measurement of pump response to hydraulic conditions. A test was conducted where the suction pressure was reduced with the pump capacity held at a constant 150 gpm to observe the effects on the monitored parameters. Figure 11 is the graphic representation of the test results.

Results Table 1 represents all of the data that was recorded during the testing of the pump for each measured parameter and test condition. The table can be used with whatever combination of sensors that exists in the user’s plant and will be helpful in the interpretation of indications. For example, if OHFT is added to a system that already has a power monitor, useful information can be obtained that should minimize maintenance costs and catastrophic failures.

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Overall High Frequency Tracking (OHFT) During pump testing, several new items were noted regarding OHFT. • The mounting of the sensor was found not to be as critical as originally suspected. That is, if the sensor is mounted solidly to the pump casing, the orientation as to axial or radial did not significantly

Table 1. “Truth” table of results

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change the magnitude of the overall readings. Lower baseline OHFT readings will result when the accelerometer is mounted on the upstream side of the cutwater. Higher baseline readings resulted when the accelerometer was mounted downstream of the cutwater. Presumably, the increased baseline noise was caused by possible turbulence or hydraulic noise associated with liquid passing by the cutwater. A rule of thumb when using OHFT is that less is better. The quieter the pump is, the better and more trouble-free it will operate. The exception to this rule is dry running. Wider traces of both OHFT and axial rotor position are indicative of operating conditions to avoid in pumps, especially sealless pumps. A time interval of one second or less should be used as a sample rate for

Figure 10. Air injection

Figure 11. Reduced suction pressure (at constant 150 gpm) The Pump Handbook Series

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capturing OHFT data. This will capture all of the fast-changing operating and mechanical data that can take place within the pump. • Mounting the sensor closer to the source of the stimuli is better, because the signal is stronger. Usually that means mounting on the pump casing. • Conditions that raised the OHFT level caused the rotor to move significantly in the axial direction.

Rotor Position Monitoring The rotor is monitored through a series of wound coils located outside the primary containment and protected from the process fluid by the stator liner. Electrical signals received from the coils are used to continuously monitor the actual running position of the rotor. The device detects any change in rotor position in the axial and radial directions simultaneously. By comparing the instrument’s output to the original factory test baseline of a new pump, the condition of the internal radial and axial bearings can be determined. After the initial calibration, radial bearing wear is determined by a change in output in the radial direction greater than the baseline data for new bearings. The amount of wear is proportional to the change in signal. It should be noted that normal operation of a sealless pump does not promote wear of the radial and axial bearing surfaces. The process upset conditions leading to lack of lubrication and rapid heat rise are the main causes leading to the wear of these surfaces. Continuous monitoring enables users to trend these damaging events to predict and improve the mean time between maintenance intervals.

Conclusions There are many benefits to monitoring sealless pumps continuously. Since the trend in process plants is to use distributed control systems, much of the plant equipment is being remotely

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operated. Presently, the board operator in a control room has little or no feedback on the operating condition of most of the pumps. These parameters are not instrumented with trip or alarm limits based on pump health. It is possible for the pump to be operating in an off-design condition or have mechanical damage and the operator not know it. The feedback presently comes in the form of failed equipment and expensive repairs. With continuous monitoring, it is possible to get immediate feedback on the condition of the pump and on the process conditions. Cavitation, dry running and extreme operating situations that will result in pump failure can be immediately detected. Armed with this information, the operator can make decisions to improve operating conditions that will prolong equipment life and maintain product quality. For new pump installations, use the latest technological advances being offered by sealless pump manufacturers. This includes rotor position monitoring on canned motor pumps and heat detection on the containment shell in the center of the magnet area of mag-drive pumps. It is best to utilize both radial and axial monitoring of rotor position when available. If you want to retrofit sealless pumps with OHFT monitoring systems, follow the guidelines below. • When metal-to-metal contact is detected, the pump should be stopped and scheduled for maintenance. • When OHFT is added to a pump, the condition of the pump should be known and a baseline representing that condition recorded. OHFT values should be fairly low, in the 10-40% of full-scale range. • A baseline set of readings should be taken with the pump operating at its normal pump curve capacity. This should be done even if the discharge valve has to be throttled to achieve this with the size of the impeller that is being used. • When an increase in OHFT is detected, the process should be The Pump Handbook Series

varied, if possible, to eliminate the “noisy” condition. This will help to determine if the increase is due to mechanical or process conditions. • If the OHFT levels are reduced to a relatively low value by adjusting process conditions to normal pump design values, the pump most likely does not have a mechanical problem but is probably not being operated on its curve. To prove this, allow the process to settle for a short while. Start the standby pump, if one exists, and look at its OHFT readings. If they are substantially lower than those of the recently running pump, leave the spare pump in service and put the other in standby mode. Schedule the recently stopped pump for maintenance. • If switching pumps cannot reduce the OHFT noise, it may be an indication that the pump is incorrectly sized for the process and will be a maintenance problem. The pump application should be investigated for proper sizing and adequate suction conditions. The greatest savings will come from detecting the conditions that will cause a problem in a pump early enough to eliminate them and thus prevent a failure. If early detection of off-process operation is not possible, the next best maintenance practice is to detect a problem at its inception and schedule the pump for maintenance at a point when the problem will result in minimal maintenance costs, business interruption and no leakage. ■

Nomenclature gSE - Dimensionless unit used in detecting problems with rolling element bearings.

Acknowledgments The authors would like to thank Crane, Chempump division, and Entek/IRD for the use of their equipment and facilities in conducting these tests. ARCO/Lyondell Chemicals should be thanked for their help and support in allowing this article to be written.

References 1. Shea, J. M. and Taylor J. K., 1990 “Using Spike Energy for Fault Analysis and Machine-Condition Monitoring”, IRD Mechanalysis, Inc. 2. Le Bleu, J., Jr., 1994 “Monitoring Sealless Pumps for Metal-to-Metal Contact of the Internal Rotor”, Proceedings of the 11th International Pump Users Symposium, Turbomachinery Laboratory, Texas A&M University, College Station, Texas. 3. Le Bleu, J., Jr. and Xu, Dr. M., 1995 “A New Approach for Monitoring Sealless Pumps”, Proceedings of the 19th Vibration Institute Meeting, Indianapolis, Indiana.

Julien Le Bleu is the principal engineer for rotating equipment for ARCO chemicals in Lake Charles, Louisiana. He has more than 25 years experience in the field of rotating equipment, and is responsible for all rotating equipment in the Lake Charles facility. He has authored several articles and has lectured at the Pump Symposium previously. Mr. Le Bleu is presently a member of the advisory board for the Texas A&M Pump Symposium. He received his bachelor’s degree from the University of Florida (1974). James Lobach is a Chief Developmental Engineer with the Chempump Division of Crane Pumps and Systems. He has had extensive experience in the design and application of high speed rotating machinery. For the past five years, he has been closely involved with canned motor pump design and innovations, including low specific speed

The Pump Handbook Series

pumping, pump hydraulics and performance, and monitoring equipment. He has provided field service engineering in the chemical and petrochemical industries for the past 15 years. Mr. Lobach received a B.S. degree in Mechanical Engineering from the University of Colorado (1969). He is a registered Professional Engineer in the state of Colorado. This article has been reproduced with permission of the Turbomachinery Laboratory from the Proceedings of the 15th annual Pump Users Symposium, Turbomachinery Laboratory, Texas A&M University, College Station, TX, pp. 125-132. Copyright 1998.

525

Increased Reliability at a Small Refinery How, despite a limited maintenance budget, they got it done By Eddie Mechelay, Maintenance Specialist, VECO Rocky Mountain Inc., Denver, Colorado

I

ncreasing equipment reliability at any location provides many challenges. This article chronicles the approach taken by Total Petroleum in Commerce City, CO, beginning with the hiring of a new maintenance supervisor and continuing through the implementation of a full reliability program. During the 31/2 year time frame, the plant MTBF increased from 2.3 years to 4.5 years, while at the same time, overall maintenance costs were kept flat. In the late 80’s and early 90’s, the Total Petroleum Plant was experiencing a general increase in reliability due to the cooperation of maintenance and operation management and skilled craftsmen. The pump shop was staffed by four pump mechanics and a capable supervisor. During this time frame, the plant began the process of some root cause analysis, but still relied heavily on the input and direction of vendors for recommendations and solutions. Yet, while commitment to a quality product produced by the shop was evident, diffi-

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culty remained with the overall knowledge base at the plant. Despite such limitations, significant strides still were made in productivity and reliability. Over time, though, improvements began to taper off and refinery management went looking for a solution.

specific goals that needed to be achieved. It was these goals that set the path forward.

Setting Goals

When moving forward and implementing change, it is important to set goals to benchmark progress. The first Have A Plan goals set at the refinery were task Refinery management reviewed other specific and not the typical benchsuccessful plants and determined a need marks of MTBF, MTBR and cost. for a rotating equipment supervisor, Instead, these first goals were to: preferably an engineer, with supervisory experience. Management took the time • Reduce electric motor repair cost to develop an in-house test to ensure that • “Clean up” turbo-machinery spare the successful applicant would have the parts proper technical knowledge to achieve • Implement a lubrication program results immediately. The refinery’s plant • Implement process safety managemanager also was particularly concerned ment repair procedures with non-technical aspect of the candi- • Update the plant vibration monitordates who were interviewed. The ability ing program to coach and mentor mechanics, train and educate operators, and foster an Motor Repair Total Petroleum’s Commerce City environment for change were critical site had a population of roughly 400 factors in reviewing all applicants. After hiring the new supervisor, the motors and was averaging three failures maintenance manager provided five a month with an annual cost of $30,000. The Pump Handbook Series

Although this figure may seem low, it was the inability to determine the root cause of failure that led to the effort to improve. The decision was made to visit the motor repair shop and review repair procedures with the shop supervisor. Refinery management also wished to determine the level of knowledge at the shop and its acceptance in converting electric motors from grease lubrication to oil mist lubrication. During the shop visit, it was learned that the motor mechanics were not measuring the bearing fits and that the bearings were typically cold pressed on the shaft. This technique can damage the bearing and will likely cold roll the shaft, thus creating a poor fit for the bearing. Yet, when the idea of an induction bearing heater was discussed, shop personnel were non-committal, tending to believe that such an intervention was not necessary. In addition, while reviewing the techniques for oil mist on the electric motors, the shop superintendent suggested that no oil mist could lubri-

cate as well as a good grease. Clearly, this particular motor shop was not interested in sharing knowledge or working as a partner in reliability. Contrast this visit with the shop that was awarded future motor repairs. The second shop had performed repairs for the refinery in years past, but typically did not perform them in a timely manner (as defined by the refinery management at the time). During the kickoff meeting with the second shop, its management committed to turning motors around in three days for most failures. They also agreed to upgrade the existing shop induction heater to one that limited bearing temperature to 230°F. When asked about the oil mist conversions, their response was, “Please get us the information and we will be happy to work with you on your new application.” With this new partnership in place, motor repairs were reduced to one per month, at a yearly cost of only $14,000. Additionally, in 1994 and 1995 not one motor experienced a catastrophic The Pump Handbook Series

failure since each failure was caught before it became catastrophic, either by routine observation or through vibration monitoring.

Turbomachinery Spare Parts One-fourth of the plant’s total spare parts inventory value was used to support two assets at the plant: the Reformer Compressor/Turbine Set and the Blower for the Cat Cracker Unit. Historically the plant had hired outside consultants to manage the overhauls and recommend spare parts. Unfortunately, these individuals typically had only a day or two to review the parts on the shelf. Paying someone $150/hr to review spare parts on the shelf did not appear to be good stewardship. The only technique for correctly reviewing parts for these critical pieces of machinery is to obtain equipment specific parts list from the Original Equipment Manufacturer (OEM) and review the items individually, taking care to look for exact part numbers.

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When moving forward and implementing change, it is important to set goals to benchmark progress. The first goals set at the refinery were task specific and not the typical benchmarks of MTBF, MTBR and cost. During such a review, it was determined that numerous duplicates were on the shelf, yet many long lead items were not in stock. By obtaining credit for some duplicates, deleting some unnecessary items and planning expenditures, the parts for these critical assets were brought in line. It should be noted that practically every description and part number in the plant CMMS system had to be updated to insure that correct parts were purchased in the future. Retagging the parts and placing them on the shelf only would have corrected the situation until the next overhaul. It was during the

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description updates that a technique was to be removed from this work unless developed to mark those items in invena real plant emergency occurred; tory that were “repairable,” thus prevent- • Partnered with a local vendor to ing the purchasing department from assist when unique situations arose buying new without first contacting the and additional support was needed. rotating equipment department to determine if those items could be Using this process, the refinery was repaired economically. able to reduce the number of oils to two: an ISO 32 for the turbo-machinLubrication Program ery and an ISO 68 for the plant pump “What do you mean lubrication population. The grease chosen was a program? We lubricate our equipment. polyurea grease, because it was used …Well, no, we don’t get all of them, in the sealed and shielded bearings for but we get most of them… What SKF and MRC that the plant had grease? We use the grease in the selected to stock in inventory. Subsewarehouse… How often do we check quently, not only was there a signifithe oil? We check it when it gets low. cant reduction in bearing failures, but . . I don’t know the required viscosity by routinely inspecting equipment the for that pump. . . ” mechanics began to identify other Although the plant had a group of problems before they became larger, excellent mechanics who truly wanted including, bad drain valves, excessive their equipment to run more reliably, vibration, seal leaks, etc. they were lacking in their understanding of lubrication techniques and properties. Therefore, the mechanics and reliability engineer teamed together to identify every bearing in the refinery and determine the proper lubrication required for each point. This effort can be best demonstrated with the following: • Determined the location of every lubrication point—bearing housings, grease fittings, oil mist systems, etc.; • Used manufacturers’ recommendations to determine the proper viscosity of oil and grease properties; • Evaluated these recommendations and compromised to reduce the number of unique lubricants for the plant; • Educated personnel outside of the rotating equipment department on the reasons for using different Process Safety Management grease for rotating equipment; • Published a plant-wide listing of Mechanical Integrity Procedures Writing procedures for maintaining each lubrication point accessible via or repairing equipment is a difficult the LAN; • Divided the plant into smaller more task. This difficulty stems as much from manageable areas and greased or the culture of the maintenance business as much as it does from the technical inspected oil reservoirs; • Made this work a priority and did component of the work. Before jumping not allow the mechanic or operator into the process, the refinery mechan-

Paying someone

$150/hr to review

spare parts on the

shelf did not appear to be good

stewardship.

The Pump Handbook Series

ics held a procedure-writing work session to determine the obstacles and opportunities related to the task. There was no resistance to writing these procedures, since OSHA CFR 19.10 required that they be written. The consensus of the meeting was to write procedures that allowed the mechanics to perform their jobs better than they had in the past. This was accomplished by identifying approximately 45 unique procedures, then writing up the most common one first—a single stage overhung centrifugal pump—thus building a template to use with all of them. The benefits of this process became evident when manuals that had not been available had to be obtained from local vendors. In addition, datasheets were developed for many of the procedures, which allowed a “go/no go” system to be implemented as a part of the repair process. In other words, if certain fits were out of tolerance, the equipment was kept out of service until it could be repaired to the proper specifications. A year after this process had been implemented, the mechanics were surprised at the number of times they referred back to the procedures that they had written. This was especially true of pumps that were one-of-a-kind and not serviced frequently.

Vibration Program The supplying vendor had included alarm limits and alarm banding with the vibration program installed at the refinery. However, without specific, detailed input of experienced refinery personnel, the effectiveness of the program was minimized. Typically, the mechanics were not closely involved with identifying the cause of vibration—they just removed the pump, installed new parts and then replaced the pump. It should be noted that the operations personnel at the plant were extremely supportive of the vibration program and routinely switched all spare pumps on a weekly basis, including services with common spares. (See Inset.) An additional

overhaul or repair. If sufficient time was not available to perform the alignment during a particular day shift, the equipment was kept as a standby and immediately aligned the following day. The emphasis on alignment led to additional workload as pipe strain, soft foot and modifications for equipment being bolt bound were corrected. The mechanics determined that new motors installed in existing applications often became bolt bound. Consequently, it became standard practice to drill out the holes in the motor feet prior to field installation. After a year of mastering the reverse indicator method, funds became available for a laser alignment system. This system allowed the mechanics to expand their precision alignment to cooling towers and large reciprocating machinery that did not lend itself to the reverse method. However, to insure they understood the mechanics and process of machinery alignment, new Alignment mechanics entering the pump shop Although it was not detailed in the were required to learn the reverse original goals set by the refinery’s method before being trained on the Maintenance Manager, machine align- laser system. ment appeared to be the largest contributing factor to the increased Preventative Maintenance Even with a proactive maintenance reliability at the refinery. Prior to 1994, the refinery essentially had no align- crew, the refinery had never implement program. Single indicator and rim mented an effective preventative and face methods were the chosen maintenance system for rotating equiptechniques, and often times no align- ment. The entire system consisted of ment was performed. Interestingly, the two PMs, one to check the bolts on the typical excuse of “insufficient time” was coupling of a vertical pump (that had not a factor in the lack of alignment. failed three times in one year), and a The major issue was lack of knowledge PM to inspect packing on critical valve in the refinery. about up-to-date techniques. Utilizing the experience of the The refinery invested in a quality reverse indicator alignment kit capable mechanics and performing a basic risk of servicing all but a handful of equip- assessment of the plant, the PMs were ment on the site. A training program was increased from a task list of two to over implemented to teach mechanics the 60 and then scaled back to a total of 50 process of reverse indicator alignment, PMs. This basic risk assessment resulted in including the pitfalls of: pipe strain, soft foot, bolt-bound equipment and lack of numerous assets from plant support jack bolts. In addition, acceptance systems being added to the PM schedstandards for alignment tolerance and ule. Assets such as the plant air compresestimation of thermal growth were incor- sors, the discharge water air blower and porated into the training. Also during this flare pump were all included as a result time, all equipment was aligned after an of this review. Items that were expenbenefit of this switching was an increase in operator confidence in the spare equipment Since the original installation, new equipment had been installed without having been added to the vibration routes. The vibration analysis/trending software was DOS-based and was quickly becoming obsolete. The software was upgraded and the entire plant database upgraded to include all equipment deemed important enough for vibration readings. On some of the smaller horsepower equipment, the determination was made not to remove it from the route, but to significantly reduce the number of points taken per item. Over the course of the upgrade, the mechanics and supervisor became a team to identify alarm levels and nomenclature. It was during this time that the mechanics were taught basics of vibration such as how to identify looseness and misalignment.

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sive to repair, such as Sundyne™ pumps, were included. These pumps were overhauled on a schedule of two years. Adding these pumps to the preventative maintenance list included adding significant inventory to the plant parts system. The results are indisputable: zero pump/gearbox failures in six years. The scaling back of PMs was achieved by effective use of predictive maintenance tools such as oil analysis and vibration analysis, as well as overcommitting to assets the mechanics “felt” needed preventative maintenance when, actually, periodic inspection was all that was required. In addition to the total number of tasks, the frequency of the PMs was continually modified as more information was learned about the equipment. Communication is the key element when making these changes to a PM system. It is imperative that operations and maintenance discuss the reasoning for adding/deleting PMs, as well as for changing the frequencies. If operators are not kept informed on issues found during PMs, their buy-in on preparing equipment for work is not as great.

Next Steps At this point, the majority of the “low hanging fruit” had been identified and corrected. Over the next years, the progress slowed. However, due to increased reliability of the equipment, staffing in the shop was reduced from four to three mechanics after one contributor retired. It was during this period that the shop began to manufacture 95% of the shafts and wear rings for inventory. No longer did the purchasing department expedite long lead items for inventory. Mechanics were scheduled on a routine basis to fabricate parts from drawings the mechanics had documented. This effort actually made it easier to control shaft modifications that were made for specialized seals. The refinery did not need to insure that vendors had the required information for modified shafts, it only needed to track these modifications in-house.

530

One mechanic was assigned to review the plan’s bearing inventory. During this review he determined that approximately 33% of the inventory was obsolete. Through his efforts, a complete overhaul of stock descriptions and stock levels was made to more closely mirror the needs of the equipment.

Although it was not detailed in the original goals set by the refinery’s Maintenance Manager, machine alignment appeared to be the largest contributing factor to the increased reliability at the refinery.

mechanics the skills, it provided them with reference material to successfully repair and maintain the equipment. Much of the mechanics’ training on bearings, couplings and lubrication was furnished by vendors. Often times, mechanics or supervisors with a particular strength would train the others. Examples of in-house expertise included areas such as lathe training, alignment and unique positive displacement pumps. The key to the success of this informal program was to identify the needs that were affecting the reliability of the plant and then to create an action plan that provided the proper education.

Measuring Performance To achieve maximum benefit from a reliability effort, key indicators to measure performance need to be tracked, so the managers and mechanics at the refinery discussed items that needed attention. It was important to select only a few issues to begin with so that incoming information would not be overwhelming. In order to obtain the best information, the plant printed out all rotating equipment work orders for the previous five years and manually entered the data into a spreadsheet to track MTBF. In addition to this measurement, the refinery tracked measurements such as: motor repair cost, seal failures/cost, horsepower uptime, and inventory value. The results of these graphs were posted on a monthly basis in the shop and were referred to on a continual basis to drive performance.

Conclusion

Mechanic Develoment Although a formal skills training program was never fully developed, significant and relevant training was given to the mechanics. Maintenance of the previously mentioned Sundyne™ pumps required sending the mechanics to a three-day course at the OEM factory. This comprehensive program not only taught the The Pump Handbook Series

The reliability success at Total Petroleum in the 1990’s is a story of commitment and communication. The principles used to increase the reliability of the equipment at this site are available in numerous formats to everyone in the pumping industry. Management’s strong desire for the program to succeed allowed those in the pump shop to “think outside the box” and implement programs and solutions that had been off-limits in the past. The teamwork between operations

and maintenance also was an integral component of success. Determining specific actionable problem areas and developing goals to improve as was done at Total Petroleum is vital to the success of any reliability effort. Tracking the performance and sharing the successes will continue the process of improvement.

experienced greater than a 50% turnover, the basic reliability principles were not compromised. The original commitment and foundation agreed upon by management has never wavered, even though the plant has had four plant managers. ■

The author would like to thank Keith Footnote Schindler with Flowserve and Dick From 1993 through 1999, the MTBF Brown with Ultramar Diamond of the plant continued to increase, even Shamrock for their contribution and with constant personnel change in the guidance in preparing this material. pump shop. The original four mechanics were reduced to two and then Eddie Mechelay has been working in increased to three when one mechanic the petrochemical and refining induswas trained and promoted from within try for the past 13 years. A graduate of the maintenance department. Texas A&M University with a B.S. in When the original supervisor left the Mechanical Engineering, he is a member company, the shop replaced him with of the Pumps & Systems User Advisory another engineer who continued to Board and a frequent contributor foster an attitude of communication and to the magazine. Contact him at “thinking creatively.” After an additional [email protected] eighteen months, the shop promoted a mechanic to the supervisor position and This article was excerpted from the another internally trained shift opera- official conference proceedings of tor to the mechanic position. PumpUsers Expo 2000. In summary, although the shop

The Pump Handbook Series

The principles used to increase the reliability of the equipment at this site are available in numerous formats to everyone in the pumping industry.

531

Water Contamination of Equipment-Lubricating Oil Is it a misunderstood killer? By Brad Rake, Vice President of Strategic Development, Trico Mfg. Corp.

Figure 1. Water absorption as a function of relative humidity (Courtesy of Noria Corporation)

Field Perception: Is an Ounce of Prevention Good Enough? Perceptions of what good lubrication practices are, and even how effective they are, can vary significantly. Field surveys have shown that the amount of water in oil thought to be safe by the majority of people is significantly higher than the amount actually recommended by equipment and

532

component manufacturers. To complicate matters, safe levels can vary considerably by oil type, additive package and operating temperature— making general guidelines ineffective. One type of oil can be used safely at 750 parts per million (PPM), while another may be damaging at 200 PPM at the same operating temperature (Figure 1). Although many plants have implemented maintenance programs The Pump Handbook Series

that have resulted in major improvements, even more can be done. Most preventative and predictive maintenance programs make regular oil changes a requirement. With positive results such as longer mean time between required maintenance, the prevalent attitude regarding noncritical equipment maintenance is, “just change the oil regularly.” Although more diligent oil changes often result in significant improvements, many times they do not. Consider the following statistics: • New oil tested “straight from the bottle” has contained as much as 1,500 PPM of water. • 60% of all oil changes/samples taken occur after damaging levels of water already exist. • 25% of all oil changes/samples taken occur before damaging levels of water exist. While changing oil routinely does not guarantee optimum lubrication performance, it has become widely accepted as one of the most effective means of decreasing costs associated with equipment lubrication. This is despite the fact that much of the cost incurred could be avoided completely by changing oil “smarter,” not just more often. Many of the hidden costs result

from “routine” oil sampling and changing procedures, which are often unnecessary efforts (Figure 2). Even more challenging, though, is the problem of how to keep skilled tradesmen from engaging in these types of mundane, time-consuming procedures that clearly yield insufficient results.

gas. Maintaining the dew point then becomes critical. Seal oil moisture levels of less than 50 PPM may be necessary, depending on the oil type, to maintain hydrogen dew points below 15°F, as required. Such low levels are extremely difficult to detect without frequent sampling and testing.

How Much is Too Much?

How Moisture Contaminates Oil

Just how much of a Knowing how to prevent problem water contamination water from getting into the in lubricating oil can be is oil is one of the most effecdependent on many factors, tive ways to increase oil and Figure 2. Percentage of time the lubricant is changed vs. when it needs to including type of equipment, be changed equipment life. Don’t how the equipment is used overlook the basics of oil and lubrication maintenance methods. between the two elements until the storage and handling. Common avenues Even very low levels of moisture in oil viscosity increases to a point where it for water ingression include contamican damage steel. For example, in acts as a solid. This enables the ball to nated funnels, oilcans, loose plugs on concentrations well below the saturation roll cleanly and effectively against the equipment and vented applications in level, water accelerates fatigue spalling inner and outer races. Water contami- humid operating conditions. Standard in steel rolling element bearings. One nation can limit the lubricant’s ability “globe oilers” have been used successstudy found that by reducing water to form a “solid” and reduce the film fully for many years to maintain proper concentrations well below saturation, strength to the point where surfaces levels of oil in wet sump applications bearing life increased by a factor of five. begin to contact one another (Figure 4). such as horizontal process pumps. This can result in significant savings The amount of moisture allowed in However, these oilers are vented to the where bearings are a predominant the lubricating oil of a typical hydro- atmosphere, enabling humidity to be maintenance concern (Figure 3). gen cooled turbine/generator as speci- absorbed by the lubricant in the oiler Very small concentrations of water in fied by ASTM D95 is 2,000 PPM. If the sump, then transferred into the equipoil accelerate degradation by oxidation. lubricating oil mixes with the seal oil, ment sump. To prevent this type of Combined with increased temperatures, it can come in contact with hydrogen ingression, closed system or internally the effective life of the oil is greatly reduced. For every 18°F rise in oil temperature, the oxidation rate doubles, leading to a change in viscosity, development of sludge and possible additive depletion. Hydrogen embrittlement weakens the steel surfaces of bearings and gears. Dissolved water penetrates into surface cracks of rolling element bearings and reduces the fatigue life. Additionally, rolling element bearings rely on elastohydrodynamic (EHD) oil film thickness to protect the various surfaces of the bearing. EHD is dependent upon operating conditions such as surface speed, load, lubricant viscosity and the pressure-viscosity relationship. As the pressure between surfaces increases, the lubricant is “squeezed” Figure 3. Life extension factor (Courtesy of Noria Corporation) The Pump Handbook Series

533

Figure 4. Water contamination limits a lubricant’s ability to work effectively by reducing its film strength.

vented oilers are recommended for use in humid operating environments. Pressure differential between the equipment housing and surrounding atmosphere is a leading cause of moisture ingression. Several conditions can produce this atmospheric exchange as pressure is equalized, including equipment operation where housing temperature fluctuations occur during frequent on/off running conditions, process fluid temperature changes, outdoor use and air flow over the equipment. When moisture is introduced into the housing, the oil absorbs it at a variable rate depending on temperature, type of oil and lubricant agitation. Other major causes of contamination are the methods by which oil is stored, handled and transferred. Bulk storage is a common source of ingression, with pressure differential again being a leading contributor. Storing barrels where sunlight can reach them can cause a pressure increase as the oil temperature rises in the sun and a

Photo 1. On-line monitoring device capable of testing for Saturated Relative Humidity (SRH) (Courtesy Trico Mfg. Corp.)

pressure decrease as the temperature drops when the barrel is in the shade. This pressure increase/decrease is enough to draw humid air in as the pressure inside the barrel decreases, and the oil will absorb the moisture. Additionally, if the barrel is stored vertically, any water that collects on top of the barrel can be drawn in through the bung and contaminate the oil.

Measuring Water Contamination As mentioned previously, safe levels of water PPM in oil vary. The widely accepted method of measuring water is to take a sample and perform a Karl Fischer titration test. Newly introduced on-line monitoring devices measure water contamination by percentage, or Saturated Relative Humidity (SRH) (Photo 1). This measurement is taken below the saturation point of the oil, before phase separation will occur. Just as in the air we breathe, this saturation point varies by temperature—hence the

“relative” aspect of SRH. How many PPM of water in oil should be considered safe is relative to temperature. For example, 500 PPM at 150°F may be safe, or below the saturation point, but the same 500 PPM at 75°F will be above the saturation point. Phase separation will occur and the water will condense. This is very similar to the dew point of the atmosphere— when the temperature and dew point converge to within five degrees, moisture begins to condense—wherein fog or even rain results. Although there are many ways of detecting water in oil, most require that a sample be taken from the piece of equipment. Research has shown that more than 85% of all sampling or oil changes were done either unnecessarily or after damage had already occurred. As indicated in Figure 5, it is virtually impossible to “see” when it is time to either change or sample the oil. Additionally, most on-line water sensing available today begins between 300 and 600 PPM, where an already potentially damaging condition exists. As shown in Figure 6, the common measurement of moisture contamination has been in PPM, when the most universal critical property is the level of SRH. Factors such as whether the oil is mineral or synthetic, and what type of additive package is present, can change oil’s ability to absorb moisture.

On-line Monitoring

Figure 5. Water in oil only becomes visible when it’s already at dangerous levels.

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The Pump Handbook Series

Although condition monitoring has been a significant component of preventative maintenance programs for many years, the focus of this effort

has been directed toward critical equipment such as turbines and larger pumps. Concentrating on eliminating “surprise” equipment failures and extending MTBR has resulted in appreciable savings in many industrial facilities. Although it had been cost-effective to monitor “critical” equipment, it was difficult to justify the expense of installing monitoring devices on non-critical equipment such as process pumps, compressors and blowers—even though the overall maintenance cost was higher on “noncritical” equipment. For example, in one independent study, more than 70% of the maintenance budget of a chemical processing plant was spent on non-critical equipment maintenance, including oil changes, oil sampling and scheduled repairs. It was not statistically evident what percentage of the cost was associated with lubrication, but all of the professionals involved in the study agreed that lubrication was a factor in many component and pump failures.

Cost Justification It is now possible for individual companies to justify the cost of proactive monitoring and maintenance of non-critical equipment. After entering plant and equipment information into an easy-to-use program, savings projection calculations such as Return on Investment (ROI), Net Present Value (NPV) and Internal Rate of Return (IRR) are provided. Easy-to-read summaries and explanations, as shown in Figure 7, make it possible for company personnel to determine which products and maintenance procedures make financial sense. It is estimated that the average centrifugal pump mean time between planned maintenance is less than 18 months, and the average cost to repair a failed process pump approximately is $3,500, not including lost production. Over a typical 20-year installation life, pump repair costs can exceed $45,000. Although it is difficult to isolate the costs specific to water in oil, as seal technology continues to improve,

Figure 6. The most common measurement, PPM, doesn’t always tell the complete story.

bearing failures will become the predominant cause for maintenance—and water is a leading contributor to premature bearing failure. Bearings for process pumps are designed for a minimum life of two years, with the theoretical life estimated at greater than five years in the vast majority of applications. Imagine what might happen if simply keeping the oil dry five years becomes a reality! With the current improvements to bearing housing isolators (seals), and the trend toward non-vented, or closed, housings, particulate contamination is significantly reduced, and water is becoming the number one cause of premature bearing, gear and seal failure.

Summary Until recently, water contamination of equipment-lubricating oil was often underestimated regarding equipment failure. For many years, visual checks were considered the norm. Waiting for the color of the oil to change, or for water to collect in sump bottles installed at the bottom of equipment sumps was considered prudent—and effective. As component manufacturers, including those of bearings and seals, began to campaign for dryer oil, a new generation of maintenance standards and procedures was born. Although component and equipment life increased, the new challenge became that of how to determine when maintenance such as oil changes was required. Visual checks were no longer effective—water could not be seen. Even worse, no technology was available as a cost-effective alternative—until recently. Low cost monitoring devices and easy-to-install water contamination prevention products are available— and have proven effective. With facilities such as oil refineries, chemical processing plants, and pulp and paper facilities operating with fewer maintenance personnel, there is an increased emphasis on equipment efficiency and reliability. Guesswork, predictive maintenance and historical trend analysis are becoming outdated methods of ensuring effective lubrication. What’s taking their place is the type of equipment and devices that provide immediate, concise and accurate indication of lubricant condition—all of which will become part of the “prescription for healthy equipment.” ■ Brad Rake is the Vice President of Strategic Development at Trico Mfg. Corp. He received his degree from the University of Kansas and has more than 21 years of experience in the research and development of industrial equipment. Mr. Rake is a member of the Society of Tribologists and Lubrication Engineers (STLE), and serves as chairman of the Bearing Housing Technical Committee of the Hydraulic Institute (HI). He holds one patent in lubrication technology, with four patents pending. Trico specializes in the engineering and manufacturing of industrial lubrication equipment, including monitoring equipment, dispensers and high performance fluids. Mr. Rake can be reached at [email protected].

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EXECUTIVE SUMMARY Date: Tuesday, November 28, 2000

Lubrication Problem: MOISTURE Equipment Identification: Pump 29 Product Description: •Passive moisture removal system with changing color indicator •Viewport of oil level •Cartridge replacement

Product Solution: Air Dryer

Financial Report Total Annual Savings: equipment rebuilds, increase production, oil tests & changes 1st Year Expenses: for purchase, installation and maintenance of Air Dryer ROI (Return on Investment) DPP (Discounted Payback Period) NPV (Net Present Value) for 5 years at 14% rate of return IRR (Internal Rate of Return) for 5 years

$22,148.00 $420.50 6.94 days 7.63 days $75,613.08 4000%

Monetary figures in US Dollars *

USER INPUT

EQUIPMENT DATA

FINANCIAL DATA

GOALS

Annual prorated rebuild cost = $2800

3-inch bearing shaft

Calculate for 5 years IRR & NPV

Reduce moisture from 1000 ppm to 180 ppm

Equipment failure rebuild time = 6 hrs.

.25 gallon oil sump

10% discount rate of money

Minimal Rate of Return = 14%

Mean Time Between Failure (MTBF) = 2 yrs.

14% Rate of Return

Production loss avoidance = $10,000 per hr. 6 lab tests per year 6 oil changes per year

OUTPUT RESULTS

1 gallon new oil = $35

Bearing life extension factor =3

Disposal 1 gallon used oil = $20 Oiler = $30 per hour Millwright = $75 per hour Electrician = $60 per hour

Figure 7. This executive summary details the financial benefits of proactive monitoring and maintenance of non-critical equipment.

536

The Pump Handbook Series

SUMMARY REPORT cont. Product: Watchdog Air Dryer Equipment Identification: Pump 29 ROI (Return on Investment) Calculations The following savings projections are based on: 1. Life extension factor for reducing moisture 2. Increasing the MTBF 3. Production loss avoidance 4. Changing oil when 70% SRH is indicated. ·

·

·

Total Annual Savings = (Rebuild Savings ($1,867) + Increased Production ($20,000) + Oil Change Savings per year ($116) + Lab Analysis savings ($150) + Oil disposal savings ($15) ) = ($1,867+$20,000+$116+ $150+$15) = $22,148 First Year Expenses = (product price ($41) + installation time (0.5 hr) * millwright rate ($75/hr) + routine maintenance time (6 hr) * oiler rate per hour ($30/hr) +6*replacement cart price ($27) ) = ($41+0.5*$75+ 6*$30+6*($27) ) = $421 ROI(Return on Investment) = ( First Year Expenses ($421) / Total annual savings ($22,148) )*365 = ($421/$22,148)*365 = 6.9 Days

NPV (Net Present Value) Calculations The following NPV Calculations are based on: 1. 14% return rate 2. Payback period of 5 years · ·

Relevant interest factor = (100%-(100%+ acceptable internal rate of return (14%) )- number of years into the future to project (5 yr) )/( acceptable internal rate of return (14%) ) = (100%-(100%+14%)-5)/(14%) = 3.433 NPV = ( Total annual savings ($22148) )*( Relevant interest factor (3.433) )-( First Year Expenses ($421) ) = ($22,148)*(3.433)-($421) = $75,613

IRR (Internal Rate of Return) Calculations The following IRR Calculations are based on: 1. Payback period of 5 years · ·

Interest factor = Initial Expenses ($421) / Total annual savings ($22,148) = $421/$22,148 = 0.019 IRR = 4000%

DPP (Discounted Payback Period) Calculations The following DPP Calculations are based on: 1. 10% cost of capital · · ·

Discount factor for cost of capital = (100%/(100%+ discount rate (10%) )year cumulative cash flow becomes positive(1) ) = (100%/(100%+10%)1) = 0.909 Discount Cash Flow = Total annual savings ($22,148) * Discount factor (0.909) = $22,148*0.909 = $20,133 DPP = Initial Expenses ($421) / Discount cash flow ($20,133) )*365 = ($421/$20,133)*365 = 7.6 Days

List Prices are in US $ Figure 7. Continued

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Pro-Active Maintenance for Pumps A good program takes time, but it’s time well spent. John Piotrowski, President, Turvac Inc.

K

eeping pumps operating successfully for long periods of time requires careful pump design selection, proper installation, careful operation, the ability to observe changes in performance over time, and in the event of a failure, the capacity to thoroughly investigate the cause of the failure and take measures to prevent the problem from re-occurring. Pumps that are properly sized and dynamically balanced, that sit on stable foundations with good shaft alignment and with proper lubrication, that operators start, run, and stop carefully, and that maintenance personnel observe for the appearance of unhealthy trends which could begin acting on and causing damage to, usually never experience a catastrophic failure. While true with a large percentage of pumping systems, it is definitely not true with all of them. Pumps frequently are asked to operate far off their best efficiency point, or are perched atop unstable baseplates, or are run under moderate to severe misalignment conditions, or, having been lubricated at the factory, are not given another drop of lubricant until the bearings seize and vibrate to the point where bolts come loose. When the unit finally stops pumping, new parts are thrown on the machine and the deterioration process starts all over again, with no conjecture as to why the failure occurred.

538

Recently, a supervisor at a pharmaceutical company who had been trained in root cause failure analysis suggested that whenever a failure occurs on a piece of machinery, to find the primary source of the problem, the situation should be treated like a police crime scene in which corrective action is not initiated until all the evidence has been gathered by the crime lab. That’s because, until the real criminal is apprehended and banished, the crime is likely to occur again and again. How many times have you and your organization so thoroughly investigated a failure that you found the primary cause(s)?

The Four Maintenance Philosophies If you work long enough in industry, you may get to observe all the different “styles” of maintenance. How a maintenance organization operates usually can be categorized in one of four different ways: • Breakdown or Run-to-Failure Maintenance • Preventive or Time-Based Maintenance • Predictive or Condition-Based Maintenance • Pro-Active or Prevention Maintenance Briefly, here’s what each one is about. The Pump Handbook Series

Breakdown or Run-to-Failure Maintenance (Figure 1) This maintenance philosophy allows machinery to run to failure, providing for the repair or replacement of damaged equipment only when obvious problems occur. Studies have shown that the costs to operate in this fashion are about $18 per horsepower per year. The advantages of this approach are that it works well if equipment shutdowns don’t affect production and if labor and material costs don’t matter. Where would this apply? The disadvantages are that the maintenance department perpetually operates in an unplanned, “crisis management” mode, with unexpected production interruptions and the need to keep a high inventory of spare parts in order to react quickly. Without a doubt, this is the most inefficient way to maintain a facility. Futile attempts are made to reduce costs by purchasing “cheap” parts and hiring “cheap” labor, further aggravating the problem. Personnel often are overworked and understaffed, arriving at work each day to be confronted with a long list of unfinished work and half a dozen new “emergency” jobs that occurred overnight. It is not uncommon to send someone out to work on an emergency job first thing in the morning, and by 10 o’clock, half way through the

Breakdown or Run-to-Failure Maintenance cost : $18/hp/yr

basic philosophy • allow machinery to run to failure • repair or replace damaged equipment when obvious problems occur

advantages • works well if equipment shutdowns don’t affect production and if labor and material costs don’t matter

disadvantages • unplanned/”crisis management” maintenance activities • unexpected production interruptions • must have high parts inventory to react quickly • most inefficient maintenance operation (up to 30% higher maintenance costs required to operate in this fashion) Copyright © Turvac Inc. 1996 All Rights Reserved

Figure 1. Breakdown or Run-to-Failure Maintenance Philosophy

job, stop their progress and send them on a new “higher priority” emergency job. Despite the wonders of modern life in the new millennium, many such operations still exist.

Preventive or Time-Based Maintenance (Figure 2) This philosophy entails the scheduling of maintenance activities at predetermined time intervals, where damaged equipment is repaired or replaced before obvious problems occur. When it is done correctly, studies have shown the costs of operating in this fashion to be about $13 per horsepower per year. The advantages of this approach are that it works well for equipment that does not run continuously, and with personnel who have enough knowledge, skills and time to perform the preventive maintenance work.

The disadvantages are that the scheduled maintenance may be done too early or too late, and that reduced production could result due to potentially unnecessary maintenance. In many cases, there also is a possibility of diminished performance through incorrect repair methods—perfectly good machines disassembled, good parts removed and discarded, and new parts improperly installed. For some, squirting grease into bearings every month is their idea of a preventive maintenance program.

Predictive or Condition-Based Maintenance (Figure 3) This philosophy consists of scheduling maintenance activities only if and when mechanical or operational conditions warrant—by periodically monitoring the machinery for excessive vibration, The Pump Handbook Series

temperature and/or lubrication degradation, or by observing any other unhealthy trends that occur over time. When the condition gets to a predetermined unacceptable level, the equipment is shut down to repair or replace damaged components so as to prevent a more costly failure from occurring. In other words, “Don’t fix what is not broke.” Studies have shown that when it is done correctly, the costs to operate in this fashion are about $9 per horsepower per year. Advantages of this approach are that it works very well if personnel have adequate knowledge, skills and time to perform the predictive maintenance work, and that it allows equipment repairs to be scheduled in an orderly fashion. It also provides some lead-time to purchase materials for the necessary repairs, reducing the need for a high parts inventory. Since maintenance work is

539

Preventive or Time-Based Maintenance cost : $13/hp/yr

basic philosophy • schedule maintenance activities at predetermined time intervals • repair or replace damaged equipment before obvious problems occur

advantages • works well for equipment that does not run continuously and personnel have enough knowledge, skill and time to perform the preventive maintenance work

disadvantages • scheduled maintenance may be done too early or too late • reduced production due to potentially unnecessary maintenance • possibility of diminished performance through incorrect repair methods Copyright © Turvac Inc. 1996 All Rights Reserved Figure 2. Preventive or Time-Based Maintenance Philosophy

the pump being subjected to run under a slight (0.1 to 2 mils/inch), moderate (2.1 to 10 mils/inch) or severe (10+ mils/inch) misalignment condition? Since a good proactive maintenance program requires that you keep records of alignment on all the rotating machinery in your facility, compare the as found alignment to the last final alignment on the unit. Has the alignment shifted? If so, by how much and what caused the shift? (It’s an easy question to ask, but usually quite difficult to answer.) 5. Now begin to disassemble the flexible coupling. Were all the bolts tight? Are any parts missing? If the coupling is a mechanically flexible type (gear, metal ribbon, chain, Ujoint, etc.), is there any lubricant still in the coupling? If so, does it look like fresh grease or oil, or is it discolored? If it is grease, did the grease centrifugally separate into oil and its base? (For example, is there a build-up of thick brown or

540

gray sludge or powder in the couand replace them, since you probapling?) If possible, scrape off some bly changed their metallurgical lubricant for analysis, then wipe all characteristics or thermally warped of the grease away with rags and them. In some cases, coupling hubs solvent, if necessary. Inspect the are so tight or have “rusted” to the coupling for excessive wear. If you shaft so that they must be cut off aren’t sure what excessive wear carefully. If it appears that the couwould look like on that particular pling hub was previously hit with a coupling, get a new coupling and hammer, it is possible that the shaft have it there for visual comparison. is bent. (Check item #8 below.) If the coupling is an elastomeric 6. Visually inspect the pump for any type, is the rubber hard and no obvious problems such as loose longer pliable? Does the elastomerfoot bolts, loose pump casing ic element appear to be worn or is bolts, cracked casing, low lubriit cracked? How long has the elascant level, loose shim packs or tomer been in service? Are the set missing shims, leaking mechaniscrews still in place, and are they cal seal, leaking oil seals or disloose? If the coupling is excessively coloration in the shaft. worn, it will probably need to be 7. Determine if there is an excessive replaced. Use an appropriate puller amount of shaft “freeplay”. This is to remove the coupling hubs—do fairly easy to do and can tell you if not beat them off with a hammer. there are potential bearing probIf the coupling hubs had an interlems in the pump or driver. Attach ference, fit and heat is required to a fixture on the driver shaft, span remove them. Try not to heat the across the coupling (engaged or hub above 275°F. If you have to get disengaged) and place a dial indithem “cherry red” in order to cator on the top of the pump shaft remove them, throw them away (or coupling hub) and zero the The Pump Handbook Series

only performed when it is needed, there is likely to be an increase in production capacity. The disadvantage is that maintenance work may actually increase if personnel improperly assess the level of degradation in the equipment. In addition, to adequately observe the unhealthy trends in vibration, temperature and/or lubrication, this approach requires the procuring of equipment to monitor these parameters, as well as the training of in-house personnel. The alternative is to outsource this work to a knowledgeable contractor to perform predictive/ condition-based duties. If an organization previously had been operating in the breakdown/run-to-failure mode and/or the preventive maintenance style, its production and maintenance management would need to conform to this new philosophy. This could be problematic if the maintenance department were not allowed to purchase the necessary equipment, were unable to provide adequate training in the new techniques, were not given the time to collect the data, or were not permitted to shut down the machinery when problems were identified.

Pro-Active or Prevention Maintenance (Figure 4) This philosophy utilizes all of the previously discussed predictive/preventive maintenance techniques, in concert with root cause failure analysis. This not only detects and pinpoints precise problems that occur, but ensures that advanced installation and repair techniques are performed, including potential equipment redesign or modification, thus helping to avoid problems or keep them from occurring. According to studies, when it is done correctly, operating in this fashion costs about $6 per horsepower per year. One advantage to this approach is that it works extremely well if personnel have the knowledge, skills and time to perform all of the required activities. As with the predictive-based program, equipment repairs can be scheduled in an orderly fashion, but additional improvement efforts also can be undertaken to reduce

or eliminate potential problems from repeatedly occurring. Furthermore, it allows lead-time to purchase materials for necessary repairs, thus reducing the need for a high parts inventory. Since maintenance work is performed only when it is needed, and extra efforts are put forth to thoroughly investigate the cause of the failure and determine ways to improve machinery reliability, there can be a substantial increase in production capacity. The disadvantages are that this method requires employees to be extremely knowledgeable in preventive, predictive and prevention/pro-active maintenance practices themselves, or outsourcing to a knowledgeable contractor that can work closely with the maintenance personnel in the root cause failure analysis phase, then assist in repairs or design modifications. This also requires procuring equipment and properly training personnel to perform these duties. Again, if an organization previously had been running in the breakdown/run-tofailure mode and/or the preventive maintenance style, the production and maintenance management would need to conform to this new philosophy. This could lead to problems if the maintenance department were not allowed to purchase necessary equipment, provide adequate training in the new techniques, were not given the time to collect the data, were not permitted to shut down the machinery when problems were identified, or were not given the time and resources to conduct the failure analysis, nor to modify components or procedures to increase reliability.

What to Do Before Rebuilding or Replacing a Pump Effective problem identification and problem avoidance requires a rigorous investigation process. When a pump failure occurs, it is very tempting to remove the pump, replace the defective parts (or the entire pump), install the new or rebuilt unit and get the unit back on-line as quickly as possible. However, if several critical checks are not made during the removal and disassembly The Pump Handbook Series

process, important clues as to the cause of the problem may be overlooked. To assist you in identifying the source of a failure, here is a recommended checklist to use when any pump is removed from service. In fact, it may not be a bad idea to perform many of these checks on an annual basis. During the disassembly process, the following things should be checked: 1. Was the coupling guard rubbing against the shaft or the coupling? 2. For mechanically flexible couplings (e.g. gear, metal ribbon, chain), is there grease or oil on the inside of the coupling guard and on the baseplate? If so, did it come from the coupling, the bearings on the motor or the pump, or from someplace else? 3. Once the coupling guard is removed, does the flexible coupling show any obvious signs of distress? Don’t take it apart yet— just visually inspect it, slowly rotating the shaft by hand. For example, does it appear to have been running hot, or are any of the coupling bolts loose? If it is an elastomeric type coupling, does the rubber appear to be cracked or is there rubber powder on the baseplate? If it’s a flexible disk type coupling, are the disk packs cracked, or do they show signs of cyclic fatigue? Is there an excessive amount of “backlash” across the coupling? Do the coupled shafts seem to turn easily or at least consistently through the entire 360 degrees of rotation; are they very difficult to rotate; or do they rotate smoothly for part of the rotation and then seem to bind for the rest of the rotation? 4. Prior to disassembling the coupling, capture a set of shaft alignment measurements (Figure 5). It really does not matter what type of alignment method or tool is used to capture these measurements. What is the amount of misalignment in mils per inch? Was

541

shifts more than that amount, you should seriously consider providing adequate piping supports on the suction and/or discharge piping to reduce or eliminate the stresses. If the movement is severe (i.e. 20+

Misalignment Tolerance Guide

1.5

ep

c

1.0

0.5

angle degrees

re a

0.10

li g

nm

en

0.08

tn

ac

maximum deviation at either point of power transmission

mils per inch 2.0

el

c

542

observing the indicator as you loosen each bolt. If the pump shaft does not shift vertically or laterally more than 5 mils, there probably is not an excessive amount of piping stress on the pump. If the pump

ta b

ec e

ssa

0.06

ry

le

0.04

ex

indicator. Lift the shaft from underneath and observe the dial indicator (Figure 6). If the pump shaft is supported in rolling element bearings, you should not see any more than 1 mil of movement. (Of course, if you use too much force when lifting the shaft, it is quite possible to elastically flex the shaft, giving you a false reading of the looseness of the assembly.) If the pump is supported in sliding type bearings, the amount of shaft movement should be within the radial bearing clearance range. 8. Check for excessive shaft or coupling hub runout. This can be done by placing a dial indicator on the shaft or coupling hub surface, holding the indicator in place with a fixture or magnetic base, then slowly rotating the shaft and observing the needle of the indicator. The rule of thumb for excessive radial runout is that there should be no more than 4 mils of Total Indicated Runout (aka TIR) for machinery running up to 1800 rpm; 2 mils of TIR for machinery running from 1800 to 3600 rpm; and less than 2 mils of TIR for equipment running over 3600 rpm. If the runout exceeds these levels, a series of radial runout checks at different points on the shaft and coupling hub should be performed to determine if the coupling hub hole has been bored off center, bored at an angle, or overbored, or if the shaft is bent. Remember, it is possible to have excellent shaft alignment and terrible runout, as well as to have good runout and terrible alignment. 9. Determine if there is an excessive amount of piping stress on the pump. There are several ways to do this. One way is to attach a fixture on the driver shaft, span across the coupling (engaged or disengaged), place a dial indicator on the top and one side of the pump shaft, and zero the indicators. Loosen the pump base bolts, one at a time,

2

le n 4

0.02

t 6

8

10

12

14

16

18

20

22

24

speed (RPM x 1000) Figure 5. Misalignment Tolerance Guidelines

Shaft ‘Lift’ Check Place the indicators on top of the shaft or coupling hub and hold the dial indicators steady.

1

1 2 3

2 3 4 5 4

1

1

2 3

2 3 4 5 4

lift upward on each shaft and note the dial indicator readings Figure 6. Shaft ‘Lift’ Check Procedure The Pump Handbook Series

26

Predictive or Condition-Based Maintenance cost : $9/hp/yr

basic philosophy • schedule maintenance activities when mechanical or operational conditions warrant • repair or replace damaged equipment before obvious problems occur

advantages • works very well if personnel have enough knowledge, skill and time to perform the predictive maintenance work • repairs to equipment can be scheduled in an orderly fashion • allows some lead-time to purchase materials for the necessary repairs • increases production capacity

disadvantages • improper condition assessment may actually increase maintenance work • requires procurement of equipment and training of in-house personnel or outsourcing this work to a knowledgeable contractor to perform duties • production and maintenance management must conform to this new philosophy Copyright © Turvac Inc. 1996 All Rights Reserved Figure 3. Predictive or Condition-Based Maintenance Philosophy

Pro-Active or Prevention Maintenance cost : $6/hp/yr

basic philosophy • utilizes predictive/preventive maintenance techniques with root cause failure analysis to detect and pinpoint the precise problems, combined with advanced installation and repair techniques, including potential equipment redesign or modification to avoid or eliminate problems from occurring

advantages • works extremely well if personnel have enough knowledge, skill and time to perform all of the required activities • repairs to equipment can be scheduled in an orderly fashion and improvements can be made to reduce or eliminate potential problems from occurring • allows some lead-time to purchase materials for the necessary repairs • substantially increases production capacity

disadvantages • requires extremely knowledgeable employees in preventive, predictive and prevention/ pro-active maintenance practices • requires procurement of equipment and training of in-house personnel or outsourcing this work to a knowledgeable contractor to perform duties • production and maintenance management must conform to this new philosophy Copyright © Turvac Inc. 1996 All Rights Reserved Figure 4. Pro-Active or Prevention Maintenance Philosophy The Pump Handbook Series

543

mils either direction), you may have to cut and re-fit the piping. Sorry! 10. Disconnect the piping and check for excessive “soft foot” problems. Soft foot conditions can be detected fairly easily, using magnetic bases and dial indicators placed near each tightened foot bolt, and successively loosening each bolt to see if the foot lifts up or drops away. If more than 2 mils of movement is observed at any foot, further investigation is warranted. The amount of lift (or drop) seen by the dial indicator is only an indication that a problem exists, and is not necessarily an indication of how the soft foot should be corrected. Soft foot checks also can be made with almost any alignment measurement system, by setting up the tooling on the shafts, zeroing the instruments in the 12 o’clock position and loosening up the foot bolts, one at a time, noting any changes that occur as each bolt is loosened. Again, the amount of lift (or drop) seen by the alignment measurement system is only an indication that a problem exists, and is not necessarily an indication of how the soft foot should be corrected. Since many pumps are driven by motors, personnel who install and align these systems frequently refer to the motor as the

544

movable machine in the alignment process. Therefore, soft foot problems are often corrected on the motors. However, in such cases, even though the pump is referred to as the stationary machine, it would be incorrect to assume that it does not have a soft foot problem. Be aware that soft foot on pumps can be as severe as on any other piece of rotating machinery. This checklist of ten “pre-removal” steps can provide valuable information on the source of the problem with a pump you are about to overhaul or replace. A pro-active/prevention-based maintenance program requires that you thoroughly investigate each failure to determine the root cause of the problem. It also requires you to take steps to prevent the problem from occurring again. Sometimes, none of these ten steps indicate a problem. However, one or more of these checks often can lead you to the source of the malady and provide you with the means to prevent the failure from occurring over and over. These steps may take some time to carry out, but it always will be time well spent. ■

including the Basic Shaft Alignment Workbook, the Shaft Alignment Handbook (ISBN 0-8247-9666-7, now in its second edition), and has written several software programs on alignment and balancing of rotating machinery. He also has contributed to the 4th Edition of the Shock and Vibration Handbook (ISBN 0-07-026920-3, Chapter 39. Part II: “Vibration Induced by Shaft Misalignment”). Since founding Turvac Inc. in 1982, Piotrowski has, among other things, designed several alignment measurement systems, trained over 6,500 people in shaft alignment principles and methodology, been a consultant to several major shaft alignment equipment manufacturers, participated in alignment service work in a wide variety of industries, designed training demonstrators for alignment and balancing, and invented the BallRod-Tubing Connector system used to measure off-line to running machinery movement. Contact him via e-mail at [email protected] .

(Editor’s Note: This article has been adapted from the Official Proceedings of the PumpUsers Expo 2000 Technical Conference. A highly popular instructor (and a very energetic and engaging speaker), Mr. Piotrowski is once again John Piotrowski has been working with scheduled to present on this topic rotating machinery for over 26 years, and during PumpUsers Expo 2001 in Baltiis considered to be one of the leading author- more Maryland, September 17-19.) ities on its alignment. He has authored over thirty technical articles on the subject,

The Pump Handbook Series

PS0901pg26

C M Y K

Bearing Protection Devices and Equipment Reliability: Part I – Constant-Level Lubricators

Pressure to improve: the risk of failure outweighs any disadvantages.

Heinz P. Bloch, P.E, Consulting Engineer, Process Machinery Consulting

To this day, important lubricant-application and bearing-protection issues are overlooked or misunderstood in the typical process-plant environment. Solid explanations are not widely disseminated and industry sacrifices potential equipment reliability For decades, constant-level lubricators have been the standard lubricant-supply system for huge numbers of bearing housings. They have been used on millions of centrifugal pumps, gear units, and electric motors with oil-lubricated bearings. However, certain types of constant-level lubricating devices require an understanding of how they function, and can be misapplied. When lubricators are misapplied, bearings could be deprived of lubrication and may fail. More recently developed rotating labyrinth seals (commonly called “bearing protectors” and “bearing isolators”) can now be found in many OEM and retrofit applications. They are an improvement over the old stationary labyrinth, but they come in different versions, and some may not perform as expected. There are pressure-relief poppet valves being marketed for the bearing housings of centrifugal pumps and other machinery. It can be shown that they are rarely, if ever, of any merit. Finally, a new generation of on-line moisture monitoring devices and desiccant-containing absorber cans are now offered by competent manufacturers. They are intended for mounting at the top of the vapor space of pump bearing housings but are not always cost-justified. Part I of this article examines when and why some constant-level lubricators are often involved in equipment failures. Part II will cover: The Pump Handbook Series

• Why some bearing-housing seals, or bearing protectors/isolators may not fulfill their intended function, • When bearing housing pressure-relief poppets serve no useful purpose, and • Where it makes economic sense to apply on-line moisture monitoring and/or moisture reduction devices.

Understanding Different ConstantLevel Lubricators As the name implies, constant-level lubricators are designed to supply lubricant, as needed, to maintain a constant level within the bearing housing (see sidebar, “The Basics”). The design relies on simple physics, and the regulated flow from filler bottle to bearing is quite reliable. Problems can arise, however, when reverse flow to the lubricator occurs. To begin with, the user community needs to be aware of the two distinctly different types of constant-level oil lubricators in use on the majority of small machines. The most widely used “opento-atmosphere” (OTA) type is shown in Figure 1. It does not incorporate a balance line connection to the bearing housing and, if used on bearing housings with restricted venting provisions, may not adequately protect against catastrophic bearing failure. On the OTA-type shown here, the desired oil level is determined by the height setting of a wing nut, threaded into a centrally located rod. The oil level at “x” is open to the atmosphere so higher-thanatmospheric pressures in the bearing housing will cause level “x” to rise, often to the point of overflowing (although the risk of an overflow condition can be minimized with extended-height surge chambers). 545

PS0901pg27

This phenomenon is easily explained by two phenomena. First, any increase in housing pressure (∆p) will be equilibrated in the lubricator (as per the static-fluid form of the Bernoulli equation). Some lubricant will be forced up into the filler bottle, to compress the space at the top. Oil also will be forced up into the OTA surge chamber, until the liquid head there is sufficient to balance ∆p. Secondly, oil is essentially incompressible, and the entire volume of lubricant displaced from the bearing housing must go somewhere. Even a fractional psi pressure increase, pushing the large-area lubricant surface in the bearing housing down by only a small distance (∆x) would force the fluid a much larger distance (∆X) up into the narrow surge chamber. Which of these processes predominate will vary with system geometry but without a properly sized surge volume, the rising fluid could overflow the rim (or reach an air-vent passage present in some models) and be lost from the system. Figure 2 shows a different constant-level lubricator. An internal O-ring or similar sealing means, prevents the ambient

atmosphere from reaching the oil level at “x.” This closed-loop, pressure-equalized lubricator (PEL) incorporates a balance tube, or pressure-equalizing line to prevent pressure differentials between the bearing housing and the lubricator.

What Makes Housing Pressures Rise?

Note that the location of the wing nut in Figure 1 or slanted tube in Figure 2, determines the oil level in the bearing housing. Either the bottle height in Fig. 1, or the bottle-and-tube height in Fig.2, is usually set with the machine not operating. At that time, temperature equilibrium exists, i.e. the machine components, lubricating oil and surrounding environment are probably all at or near the same temperature. Most bearing-housing vents, and especially small vents, offer a restriction that might allow a small amount of pressure build-up in bearing housings. It is also reasonable to expect that some oil will get flung into the close-clearance region, typically fitted with lip seals or labyrinth seals, where the shaft penetrates the bearing housing. The oil Constant-Level Lubricators: has film strength, which makes it cling to The Basics surfaces. This oil now At rest, the lubricator very much tends to bridge the gap resembles a simple, mercury between the rotating barometer. As long as the neck of shaft and the surroundthe filler bottle is immersed in the pool of lubricant, preventing air ing stationary compointrusion, the column of liquid in the nents. In that event, the bottle is supported by the atmostrapped air above the phere. The pressure in the air/vapor oil level will constitute space at the top of the bottle, plus a closed volume. the head of fluid, exactly balances As this volume of air ambient air pressure (p). is warmed by sun exposure or by frictional If depletion of lubricant causes the heat generated in the level (x) to drop below the bottlebearings, its pressure neck, it allows air to enter. As air bubbles up into the bottle, it will increase in accordisplaces lubricant, which pours dance with the perfect down into the pool below. gas law: (P2) = (P1)(T2/T1) When the pool surface rises Using Rankine sufficiently to again seal the bottle (absolute) temperafrom air, lubricant flow stops. The tures and absolute vapor space and new, lower, liquid pressures, it is easy to height have re-equilibrated with see how relatively ambient pressure. minor temperature

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increases may cause pressures to rise by amounts that cannot be ignored. Granted, as pressures go up beyond the rupture strength of oil films, the sealing oil film will be temporarily interrupted. The bearing housing seal will “burp,” or open up for a fraction of a second. Nonetheless, rising bearing-housing pressure may lead to one of two undesirable events: • During a pressure rise, the oil level will go down in the bearing housing and will rise in the lower surge chamber supporting the lubricator bottle (Bernoulli’s law). If the housing level drops sufficiently, the bottom of the bearing may be deprived of oil, or slinger ring immersion may no longer be sufficient for satisfactory lubrication. • If oil overflows the surge chamber during a pressure rise, it is lost from the system, while the pressure in the lubricator adjusts to the new housing pressure. Oil from the bottle refills the housing when the system re-equilibrates with each “burp.” As the cycle repeats itself, more and more oil will be lost. These problems are caused by system features that allow for pressure differentials between the housing interior and the environment. There is evidence that installing bearing isolators and other narrow-gap housing closures increases the likelihood of potential problems with lubricators that had served perfectly well as long as not-so-close fitting labyrinth seals were being used. Fortunately, the pressure-balanced constant-level lubricator in Figure 2 does not introduce the same risk. As mentioned earlier, the oil levels “x” at locations inside the bearing housing and inside the lubricator are always exposed to identical pressures. The problem is solved, and another step towards increased equipment reliability has been implemented. Only PEL’s and not OTA-type lubricators should be used in reliability-minded plants. By providing a piping connection between the lubricator and bearing housing, lubricator port “x” and the bearing housing operate at the same pressure and the risk of oil leakage or unintended lowering of oil levels due to pressure buildup

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in bearing housings is eliminated. This approach provides protection from airborne contaminants, as well. Disadvantages? The pressure-balanced constant-level lubricator probably costs a few pennies more, but the resulting reduction in the risk of failure far outweighs any disadvantages, both perceived and real. ■ Heinz P. Bloch, P.E., is a consulting engineer and ASME Fellow with offices in Montgomery, TX. He advises industry world-wide on reliability improvement and maintenance cost-reduction issues, and

continues to teach in-plant courses on all six continents. Before retiring from Exxon in 1986 after over two decades of service,his career included an assignment as Exxon Chemical’s Regional Machinery Specialist for the U.S., as well as machinery-oriented staff and line positions with Exxon affiliates in the USA, Italy, Spain, England, The Netherlands and Japan. He is the author or co-author of thirteen texts and over 200 technical papers and articles. He also functions as the Reliability/Equipment editor of Hydrocarbon Processing Magazine and chairs

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the yearly International Process Plant Reliability Conference. Email him at [email protected]. Editor’s Note: This excerpt is adapted from “Constant Level Lubricators and Other Bearing Protection Devices: Smart Retrofits will Improve Equipment Reliability,” by Heinz P. Bloch, P.E. Part II will explore the use of other bearing-protection devices, including bearing protectors, pressure-relief poppets, moisture monitoring devices and lubricant desiccants.

∆X ∆x x

∆p

Figure 1. Unbalanced constant level lubricator (Adapted from: Trico Mfg. Corp., Pewaukee, WI; www.tricomfgco.com)

∆p

∆p

x

∆p

Figure 2. Balanced constant level lubricator (Adapted from: Trico Mfg. Corp., Pewaukee, WI; www.tricomfgco.com) The Pump Handbook Series

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Bearing Protection Devices and Equipment Reliability: Part II – What is Really Justified? An ounce of prevention is worth a pound of cure.

Rotating labyrinth seals (commonly called “bearing protectors” and “bearing isolators”) can be found in many OEM and retrofit applications. Bearing isolators are an improvement over the older stationary labyrinths. They come in different versions, however, and some may not perform as expected. Pressure-relief poppet valves are marketed for the bearing housings of centrifugal pumps and other machinery. It can be shown that they are rarely, if ever, of any merit. Finally, a new generation of on-line moisture-monitoring devices and desiccant-containing absorber cans are offered by competent manufacturers. They are intended for mounting at the top of the vapor space of pump bearing housings but are not always cost-justified.

Rotating-Labyrinth Bearing-Housing Seals

Heinz P. Bloch, P.E, Consulting Engineer, Process Machinery Consulting

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It has long been recognized that as many as 91% of all the rolling element bearings installed in the world’s machinery fall short of reaching the manufacturer’s calculated L-10 life. L-10 is defined as the number of operating hours at which 10% of an identical bearing population will either have failed, or will exhibit visible or measurable damage. Simple subtraction reveals that 90% of bearings should be expected to survive to L-10 life. Research and follow-up analysis have established lubricant contamination as the predominant cause of failure. Much airborne contamination finds its way into bearing housings through openings where shafts protrude through The Pump Handbook Series

bearing housings, or at vents and breathers on the lubricated assembly. Bearing housings undergo temperature shifts between day and night, and in operating/idle cycles. With increasing temperatures, the vapors above the liquid oil will expand and, with decreasing temperatures, they will contract. In a closed volume, increasing temperatures cause pressures to go up, while decreasing temperatures cause pressures to decrease. In an effort to reduce the pressurerelated, in-and-out flow of contaminated air, conventional labyrinths are often replaced by rotating-labyrinth seals (which we will describe as protector/isolators, for short. A typical version is shown in Fig. 3). These devices are designed with inherent clearances, so an air gap separates the rotating and stationary elements. Except when the clearance is bridged by an oil film, this gap is large enough to allow the exchange of air with its ever-present contaminants, water vapor and airborne dust. It has been pointed out that bearing protector/isolators work best when the housing vent is plugged. To quote one prominent manufacturer (Ref.1): “If the housing vent is left open, the slight vacuum created by the contaminant expulsion elements will induce the flow of airborne dust, dirt, vapors and everything available in the immediate environment through the bearing enclosure not unlike an oilbath vacuum cleaner. This action is constant and the amount of induced debris build-up can be significant.”

“To this day, some important lubricant-application and bearing-protection issues are overlooked or misunderstood in the typical process-plant environment.” In Part I of this series concerning those issues, the author compared the types of available constant-level lubricators, and found the pressure-equalized (PEL) designs superior to those that were “open-to-atmosphere” (OTA). In this conclusion, the noted lubrication expert reviews various contamination-prevention and pressurerelief alternatives, with focus on cost and reliability-based justifications.

Bearing isolators fitted with dynamic Orings (Fig. 4) try to close the gap through which airborne contaminants can enter the bearing housing. The expectation is for the O-ring to effectively seal off the gap at standstill (Ref. 2). The designer/manufacturer hopes that centrifugal force, acting on the rotating O-ring during operation, will cause the ring to lift off sufficiently to avoid the scraping and galling wear modes noted on circumferentially contacting dynamic Orings (such modes are why O-ring manufacturers do not recommend high-cycle, dynamic, circumferential sealing applications). Bearing protector/isolators equipped with dynamic “vapor blocking” O-rings are likely to outperform isolators that do not incorporate a dynamic O-ring, in that they do prevent air intrusion on shutdown. Current production of these components exceeds 175,000 per year (Ref.2). Many practicing engineers, however, have voiced concerns with mistaken claims that these devices provide “hermetic” sealing. Reliability professionals correctly reason that if the O-ring does lift off, there still will be a gap through which contaminated air moves. Conversely, if there is no gap, there will be wear. Thus, contrary to written claims dating back about a decade, even bearing protector/isolators designed with dynamic O-rings will not achieve hermetic sealing during operation.

Whenever there exists a thin film of clean oil between either spring-activated or magnetactivated seal faces, long seal life and hermetic containment of the lubricating fluids result. Tens of thousands of the magnetic face seals shown in Fig. 5 have been used in aircraft task pumps, as aircraft generator seals, and on vertical stabilizer units (Ref.3). They can tolerate rubbing velocities as high as 86 m/s (17,000 fpm) and temperatures to 200 °C (392°F), and have frequently served to everyone’s satisfaction for over 50,000 operating hours. These seals have performed equally well in such industrial applications as gun drills, gearboxes and pump housings. They use a single Al-Ni-Co magnet annulus to attract the opposing seal face. A rather similar seal (Fig. 6) incorporates a series of strong rare-earth magnetic rods to

Figure.3: Rotating labyrinth seal (bearing protector/isolator) (Source: Inpro/Seal Company, Milan, Illinois; www.inpro-seal.com)

Hermetic Sealing is the Preferred Solution Hermetically sealing the bearing housing implies that nothing enters and nothing escapes. Only face-to-face sealing devices meet this definition. In view of the generally limited axial space between bearing housings and fluid casings of centrifugal pumps and other machinery, narrow-width, magnetically closing face seals have been developed in recent decades.

Figure.4: Bearing protector/isolator with dynamic O-ring (Source: Inpro/Seal Company, Milan, Illinois; www.inpro-seal.com) The Pump Handbook Series

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attract the opposing face. Superior face- vaporized oil is led off or collected at the atmospheric to unseat the steel ball (although material combinations achieve coefficients bottom-center location of the bearing PVC balls would relieve at lower pressures). of friction that, even without lubrication housing (Fig. 8, and Ref.6). If equipment is fitted with labyrinth seals present, rival those of PTFE. The stationor bearing protector/isolators (either with or ary seal face of the product in Fig. 6 (also Pressure-Relief Poppet Valves without O-rings), slightly negative pressures shown in insert, Fig. 8) has a diamond-like are Rarely Needed may actually exist inside the bearing housing hardness (RC90, Ref.4). Thousands of Pressure-relief poppets, similar to the one due to pumping action brought on by the machines in different industries have been shown in Fig.9, have been offered as retro- contaminant expulsion elements (Ref. 1). fitted with this particular type of seal, and fit items for bearing housings of centrifu- Bearing protector/isolators with designs that yearly production exceeds 60,000 (with gal pumps and other machinery. The weight avoid such pumping action have a liberal primary applications in centrifugal pumps of the ball exerts a downward force on the separation between rotating and stationary in hydrocarbon processing and related seat. Whenever the product of pressure elements, and will provide pressure equalindustries). inside the bearing housing multiplied by ization between housing interior and the Figure 7 shows a magnetic seal whose the exposed area (= upward force) exceeds surrounding, external atmosphere. Consiscontacting faces are pushed together by the the downward force on the seat, the ball will tent pressure build-up can occur only in truly repelling action of rod magnets of like polar- unseat itself and will allow excess internal hermetically sealed bearing housings,(i.e., ity. Although repulsion magnetic seals pressure to escape to atmosphere. with magnetic seals). However, lightly loaded embody certain advantages over the pulling A steel ball of 3/8 inch (9.5 mm) diame- seals with face orientations per Fig. 8 would configurations they are more expensive to ter weighs 0.00646 lb (2.9 gm). Assuming “burp” before the poppet would release. Only make and take up more axial space. Very an exposed area of 0.11 sq. in., it would take heavily loaded magnetic face seals would few of these seals have been sold since their a pressure increase of 0.059 psi [or benefit from poppet relief valves or expanintroduction in 1992, and only a few dozen 1.6 in. (40 mm) of H2O column] over sion chambers (described below). Experiwere produced in 2000. ence shows that the seal of Magnetic seals obtain lubriFigs. 6 and 7 will not need either cation and cooling from the everdevice. present oil fog that surrounds oil-lubricated bearings. Properly Expansion Chambers: designed, using appropriate face Use where Needed materials and applying suitable Expansion chambers (Fig. 10) are designed to absorb the expanselection criteria, they represent sion of gases or fluids in closed the ideal choice of bearing(hermetically sealed) systems. housing seal to prevent both A rolling diaphragm provides a egress of lube oil and ingress of variable volume that, when atmospheric contaminants. correctly dimensioned, will Should the temperaturereduce pressure buildup. On dependent pressure inside a non-vented bearing housing rise Figure.5: Magnetic seal used in aerospace applications (Source: some seal models, this will extend seal life or prevent “burping.” above the magnetic closing force, Magnetic Seal Corp., Warren, RI; [email protected]) the seal would release pressure Moisture Monitors and by opening and immediately reDesiccant Cans closing. There is ample evidence that Magnetic seals of the type free water in lubricating oils shown in Figs. 5 and 6 will seriously curtails bearing life. perform flawlessly, either with Water vapor, from in-leaking pressure-balanced constantmoist air, will condense once level lubricators or non-vented saturation limits are exceeded. bearing housings, and they are Since the 1990’s, competent ideal for closed, environmenmanufacturers have offered ontally acceptable oil-mist lubriline monitoring devices, capable cation systems. In a closed, of annunciating maximum safe oil-mist application, the oil mist levels of relative humidity is introduced in the space between magnetic seal and Figure.6: Magnetic seal applied in process equipment (Source: Isomag (Ref. 7). Also, add-on air dryers are now available. bearing (Ref.5). Excess liquid or Corporation, Baton Rouge, LA; [email protected]) 550

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of an air dryer requiring replacement upon color change and, thus, future maintenance. Smart companies design away maintenance, not add to it. On the overwhelming majority of pumps and similar process plant machinery, it makes far more sense to invest in failure avoidance (i.e. preventing the moisture intrusion), rather than investing in either moisture removal or moisture annunciation. Hermetic sealing is usually the right strategy, and renders measures to remove moisture unnecessary. There are, however, applications where moisture monitoring and/or desiccant-based moisture removal makes technical and Figure.7: Inpro “RMS 700” (three-piece) repulsion magnetic seal (Source: Inpro/Seal economic sense. Machine components that Company, Milan, Illinois; www.inpro-seal.com) both rotate and undergo axial movement, or gearbox installations where hermetic sealing Desiccant technology works. It has been is not practical, are where moisture monitoraround for hundreds of years and packets of ing and removal often are easily justified. a chemically suitable desiccating substance are found in the packing containers of Conclusions are Supported cameras, sunglasses, kitchen appliances, etc. by Basic Physics At issue is the technical and cost justification • A reliability-

Fig.8: Magnetic seal hermetically sealing-in oil mist applied to modern centrifugal pump bearing (Source: Isomag Corporation, Baton Rouge, LA; [email protected])

Fig.9: Pressure relief poppet valve

• Bearing protector/isolators with expulsion vanes have been known to create small pressure differences that promote the outward leakage of oil. • Magnetic bearing-housing seals are a cost-effective means of precluding the alternating, in-and-out movement of airborne contaminants. • Magnetic seals are the only practical hermetic bearing-housing closure. Hermetic sealing optimally extends the life of lubricants and bearings. Precluding lubricant contamination also makes the use of more expensive, but superior, synthesized hydrocarbon lubricants economically attractive. • Magnetic seals render constant-level lubricators obsolete. Constant-level lubricators are no longer needed in hermetically sealed bearing housings. • Poppet relief valves are very rarely needed. Their usefulness is a function of seal closing forces and must be deter-

Fig.10: Bearing housing expansion chamber (Source: Trico Mfg. Corp.; Pewaukee, WI; www.tricomfgco.com)

minded process plant will give serious mined on a case-by-case basis. consideration to upgrading from the • Moisture monitoring and removal are of “traditional” non-balanced constanteconomic value in equipment whose geometry and component features level lubricator. Pressure-balanced preclude the hermetic exclusion of water configurations are clearly becoming the from the bearing housing. norm among Best-Of-Class perform• Expansion chambers are low-cost, ers. readily justified add-ons where undue • Bearing protector/isolators perform pressure rises may either jeopardize the better than lip seals and/or labyrinth life of hermetic sealing devices, or allow seals. • Bearing protector/isolators with outward leakage of lubricant. dynamic O-rings perform better than There are compelling reasons, then, to those without the O-ring. adhere to the old dictum: “An ounce of The Pump Handbook Series

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prevention is worth a pound of cure.” We have the means and the knowledge to reduce the risk of bearing failure. Superior constant-level lubricators prevent pressuredriven oil loss. Hermetically sealing the bearing housing is feasible, and prevents the intrusion of water and airborne dust. Expansion chambers protect certain hermetic seals against pressure-induced decreases in anticipated life. Desiccantbased means of moisture removal and costly moisture monitoring devices will not be needed once the well-known water entry passageways have been closed (in a proven manner, which duplicates the sealing action of literally hundreds of millions of mechanical face seals). However, moisture monitoring and removal are important reliability-improvement measures in equipment where moisture intrusion cannot be prevented by economic means. ■

and ASME Fellow with offices in Montgomery, Texas. He advises industry world-wide on reliability improvement and maintenance costreduction issues, and continues to teach inplant courses on all six continents. Before retiring from Exxon in 1986 after over two decades of service, Mr. Bloch’s professional career included an assignment as Exxon Chemical’s Regional Machinery Specialist for the United States, as well as machineryoriented staff and line positions with Exxon affiliates in the USA, Italy, Spain, England, The Netherlands and Japan. He is the author or co-author of thirteen texts and over 200 technical papers and articles. In his spare time, he functions as the Reliability/Equipment editor of Hydrocarbon Processing Magazine and chairs the yearly International Process Plant Reliability Conference. Email him at [email protected].

Retrofits will Improve Equipment Reliability,” by Heinz P. Bloch, P.E.)

References:

Ref. 1: Orlowski, D.C., Gibit Gambits, Volume 1, No. 71, June 13, 1989 Ref. 2: Inpro/Seal Corporation, P.O. Box 260, Milan, IL 61264 Ref. 3: Magnetic Seal Corporation, 365 Market Street, Warren, RI 02885 Ref. 4: Isomag Corporation, 11871 Dunlay Ave, Baton Rouge, LA 70809 Ref. 5: Bloch, Heinz P. and Abdus Shamim; “Oil Mist Lubrication: Practical Applications,” The Fairmont Press, Lilburn, GA, 1998 Ref. 6: Bloch, Heinz P.; “Practical Lubrication for Industrial Facilities,” The Fairmont Press, Lilburn, GA, 2000 Ref. 7: Rake, Brad; “Water Contamina(Editor’s Note: This excerpt is adapted tion of Equipment-Lubricating from “Constant Level Lubricators and Oil,” Pumps and Systems, January Heinz P. Bloch, P.E., is a consulting engineer Other Bearing Protection Devices: Smart 2001, pp. 28-35

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Maintenance Outsourcing – Critical Issues

An increasingly popular option, but do your homework first

There are a number of issues facing organizations that are considering maintenance outsourcing as an improvement initiative. First, a critical examination of economic factors, relating to the firm and to the market in which it operates, is required to determine whether this sometimes vital business function should be outsourced at all. There is then the consideration as to which maintenance activities, and whether other related activities, should be included within the scope of the outsourcing arrangement. Finally, the unique nature of the relationship between a client organization and its maintenance-services provider must be reflected in the contract itself, and in the process by which it is tendered and administered. All of these issues should be addressed before a contract is let.

To Outsource or Not to Outsource – A Strategic Decision Will it Better the Organization?

Sandy Dunn, Managing Director, Assetivity Pty Ltd.

Conventional wisdom regarding the outsourcing decision states that organizations should outsource “non-core” business activities. The difficulty with this approach, however, is that it provides no guidance for deciding which activities are “non-core”. In many organizations, the discussion about what is a “core” business function and what is “non-core” becomes highly subjective and, in the end, one person’s opinion prevails over another’s. A better approach, and the one that Assetivity Pty Ltd. typically adopts in advising clients about outsourcing, is to look at the decision in terms of the matrix shown in Figure 1. Consider the outsourcing decision along two dimensions. The Pump Handbook Series

The first, Strategic vs Non-Strategic, concerns how important the activity is to the organization in achieving longterm, strategic competitive advantage in its chosen marketplace. In terms of maintenance, this varies widely, depending on the industry in which an organization competes and its chosen strategy for competing in that industry. A contract mining organization, where low production costs largely drive competitive advantage (and in which maintenance and asset ownership costs typically comprise 55-60% of total costs), maintenance clearly is of strategic competitive importance to the firm. Outsourcing maintenance in this environment would, in effect, be handing over control of this potential source of competitive advantage to an external party. Maintenance to a hospital, though, may be of less strategic importance and, therefore, could potentially be a candidate function for outsourcing. The second dimension, Competitive vs Non-Competitive, refers to how competitively the function currently is being performed within the marketplace. This relates primarily to the cost of the service, but also could include service elements such as response time. Putting the two elements together gives four possible outcomes. 1. Those functions that are of strategic importance to the firm, and which are currently being performed competitively, require no further action; the status quo should be retained. 2. Those functions that are of strategic importance to the firm, but which currently are not being performed competitively, should not (in the long run) be outsourced. A better longterm option is to re-engineer these 553

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functions to ensure that they are performed at a competitive cost. In the transition to competitive performance, an interim, tactical decision to outsource is possible. As a source of potential competitive advantage, however, the function should be retained in-house over time. 3. Those functions that are not of strategic importance to the firm, and which are not currently being performed competitively within the market, should be outsourced. There is little value in investing to improve these functions. 4. The final combination, those functions that are not of strategic importance to the firm, but which are being performed competitively within the external marketplace, is more interesting. A number of options exist for this type of activity, including a. selling the function as a going concern, b. extending the function to provide services to external customers, c. outsourcing the function, or d. raising the profile of the function, to turn it into a source of strategic competitive advantage. The preferred option depends largely on the function being considered.

to one exclusively), awareness of this possible outcome prior to establishing an outsourcing strategy is vital to the outsourcing organization. Otherwise, the organization may find itself “locked in” to a sole provider.

What to Outsource – The Scope of the Initiative How Much Maintenance to Outsource

An alternative approach is to outsource all of the shown activities except the analysis and work-identification steps. In this approach, the contractor is permitted to plan and schedule the work, and to decide how and when work is to be done, but the outsourcing organization retains control over what is to be done. A third approach is to outsource all facets of maintenance, thus giving control over the development of equipmentmaintenance strategies (i.e., Preventive and Predictive Maintenance programs) to the contractor. In this instance, the contract must be structured around the achievement of desired outcomes in terms of equipment performance, with the contractor being given latitude to achieve them.

An important consideration in making the maintenance outsourcing decision is what aspects of maintenance to outsource. If we consider the maintenancemanagement process as consisting of the six major steps shown in Figure 2, then a number of options exist. First, organizations may choose simply to outsource the work-execution step, while retaining the remaining steps in-house. This often is done, on a limited basis, when employing contractors to supplement an inhouse work force during times of high workload (during major shutdowns, for Does a Competitive Outsourcing Market Exist? example). This is A second consideration for outsourc- the minimalist ing, related to the above model, is to approach to outFigure 1. Decision matrix determine whether a competitive market sourcing. for supply of the outsourced services actually exists. In particular, when dealing with highly specialized mainteWork Work nance services (such as specialized Identification Planning turbine maintenance) or with maintenance occurring in remote areas (such as at remote mine sites), the selection of an outsourced maintenance-service provider may create large barriers to entry for other potential maintenance-service History providers wishing to enter into the Analysis Recording market. While these barriers may be overcome by adopting an appropriate outsourcing strategy (such as letting work to two or more contractors, rather than Figure 2. A world-class Maintenance Management System 554

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Work Scheduling

Work Execution

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There are advantages and disadvantages to each option. The most appropriate approach depends on the client’s particular situation.

Maintenance-Related Functions No organizational function exists in isolation from the others. Looking at how maintenance fits into the wider asset management strategy of an organization (as illustrated in Figure 3) also raises challenges. How the maintenance contractors will interface with the production operators, and the relative responsibilities and duties of each party, is one such challenge. Many organizations today are adopting Total Productive Maintenance principles, which encourage production operators to take a higher level of responsibility for equipment performance and to perform many minor maintenance tasks. There is also a growing realization that the manner in which equipment is operated can have a huge bearing on maintenance costs, and on the maintenance activities required for performance targets to be met. A high level of teamwork between the maintenance contractors and the production operators is, therefore, vital to the successful completion of the contract. This can lead to the view that an alternative, and possibly better, approach to the outsourcing of maintenance is to include plant operation in the scope of the contract. Hence, we see the letting of Operations and Maintenance contracts, particularly in the Power Generation industry. There is also a growing realization that maintenance is limited in achieving higher

equipment performance by the fundamental design of the equipment being maintained. The best that maintenance can achieve is the inherent reliability and performance of the equipment that is built in by design. There is, therefore, a school of thought that says that the best way to overcome this limitation, in an outsourcing environment, is to give the contractor responsibility for the design of the equipment. This can be limited to ongoing equipment modifications, or include responsibility for the initial design of the equipment (as in a BOOM, or a Build, Own, Operate and Maintain contract, which is gaining favor in many infrastructure projects).

Needless to say, the decision to outsource

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key aspects. Of particular importance will be the explicit consideration of risk at various key points in the contracting process, and the identification of appropriate strategies for managing those risks. These could take the form of either shaping or hedging actions. Shaping actions are those actions undertaken to minimize the likelihood of the risk factor occurring. Hedging actions are those actions undertaken to minimize the impact of the risk factor, should it occur. In addition, the evaluation criteria for the selection of an appropriate maintenance contractor are likely to be quite different from those for a major capital project. Significant work probably will be required to develop appropriate criteria, and to ensure that sufficient information is obtained from tenderers to be able to make an informed decision.

Establishing an Appropriate Specification of Requirements

The specification of requirements during the tendering process will need to be carefully considered. In particuas maintenance, is not lar, for contracts involving large-scale one that should be taken outsourcing of most maintenance functions, ensure that the requirements lightly… specification is outcome-based rather than input-based. The specification will need to detail what is to be achieved from the contract, not how it is to be achieved, Mechanics of Partnership or what inputs will be required for its achievement. Establishing an Appropriate In Assetivity Pty Ltd.’s experience, Tendering Process ensuring that all the required outcomes The tendering process for a major are specified is a major undertaking. outsourcing contract is likely to be differ- Agreement on how the achievement of all ent than for major capital works in a few of these outcomes will be measured is

any major function, such

Figure 3. Components of Asset Management The Pump Handbook Series

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also, potentially, a huge undertaking. For example, in one recent outsourcing contract, a desired outcome was the achievement of long-term plant integrity. Deciding how to measure that was a difficult process.

Establishing an Appropriate Contract-Payment Structure

processes, and equip people with the skills to perform the required duties. Many potentially successful outsourcing contracts fail, simply because the client does not manage the contracts effectively.

Establishing an Appropriate Structure for the Contract Document

There are a number of alternative contract payment structures. These include: • Fixed or firm • Variable price • Price ceiling incentive • Cost plus incentive fee • Cost plus award fee • Cost plus fixed fee • Cost plus margin

Most standard contracts in place at most organizations are not well structured for large outsourcing agreements. Many Standard Terms and Conditions are inappropriate for large, long-term servicerelated contracts, particularly those that are of a partnering or gain-sharing nature. Usually, it is best to combine Special Conditions of Contract with revised Standard Conditions of Contract to Each of these price structures represents develop a new contract structure that is a different level of risk-sharing between the appropriate for the particular contract contractor and the outsourcing organiza- being let. tion, and a number of considerations will need to be made in determining the most Managing the Transition to the appropriate payment structure. These Outsourced Arrangement include: There are many issues to be addressed • The extent to which objective assess- by the outsourcing organization in the ment of contract performance is possi- transition to the new arrangements. ble Among these are matters such as: • The ease with which realistic targets • Staff – Which personnel will be can be set for contractor performance retained by the organization? Will any • The administrative effort involved with be employed by the contractor? Which each payment option will be let go? • The degree of certainty with which the • Drawings – Who has responsibility for desired contract outcomes can be ensuring that drawings are kept up to specified date? Who will be the custodian of site drawings? Transition arrangements may be put in • Computer systems – Will the contracplace to gradually transfer the payment tor have access to the client’s Computstructure from one method to another erized Maintenance Management over time, as a greater degree of certainty System (CMMS) or maintain its own over the requirements of the contract and computerized maintenance records? more accurate knowledge of target levels Who is responsible for ensuring that all of performance are established. CMMS data are accurate? • Materials Management – Will the Establishing an Appropriate contractor provide materials, or will the Contract-Administration Process client? and Structure • Workshop facilities and tools – Who owns and maintains these? Before the contract is let, the client will need to have decided on the appropriate contract-administration process, and Agreeing to Contract Termination on the roles and responsibilities of the Arrangements staff in managing the contract. ManageAnother critical issue that needs to ment must establish the structures and be addressed before the contract is let, is 556

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how any decision to terminate the contract would be managed. In particular, agreement needs to be reached regarding the duties and obligations of the outgoing contractor in handing over to an incoming contractor (or to the client organization, should they decide to bring maintenance back in-house).

Summary Needless to say, the decision to outsource any major function, such as maintenance, is not one that should be taken lightly. Careful consideration of all major issues is vital, if the transition to contracted maintenance is to be smooth and satisfactory to both parties. ■

Alexander (Sandy) Dunn is the Managing Director of Assetivity Pty, Ltd., a consulting organization assisting capital intensive industries improve asset performance through effective Maintenance and Operations Management. Dunn is also the webmaster of the Plant Maintenance Resource Center (www.plant-maintenance.com). His 20 years of experience include positions as National Manager of Wishaw Engineering Services, and Specialist Maintenance and Mining Consultant with PricewaterhouseCoopers. A frequent contributor at the WA Maintenance Conference and the Asian Maintenance in Mining Conference, Dunn delivered “Condition Monitoring in the 21st Century,” and “Maintenance and the Internet” to the 2000 International Maintenance Management Conference. He holds a Bachelor of Engineering degree (Hons) from the University of Melbourne, Australia, and a Master of Business Administration from Curtin University of Technology, Australia. Email him at [email protected].

(Editor’s Note: This article was adapted from a paper of the same title, published on-line at the Plant Maintenance Resource Center (www.plant-maintenance.com) and is published here with permission of the author.)

PS1101PG022

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Pump Reliability – Hydraulic Selection to Minimize the Unscheduled Maintenance Portion of Life-Cycle Cost

Pick the pump that prevents potential problems.

Allan R. Budris, P.E., Director of Product Development, Eugene P. Sabini, Director of Technology, and R. Barry Erickson,Ph.D., Vice President of Technology, Industrial Pump Group, ITT Industries

Recently, significant attention has been focused on the total cost of owning a pump, over the entire operational life cycle. Major components of the cost of ownership are initial cost, installation cost, operating cost and maintenance cost. In process plants, unscheduled maintenance often represents the most significant cost component of pump ownership. Although numerous papers have been presented on the subject of pump reliability, that literature primarily addresses mechanical means of improving reliability. The result of this attention to mechanical issues has been a marked increase in the “Mean Time Between Repair” (MTBR) for process plants. This has been achieved largely through improved installation practices and increased attention to operating procedures. Efforts such as these will continue to yield improvements in MTBR, but will be limited in potential. A more holistic approach also would give attention to the best hydraulic fit to optimize reliability. The authors believe that four basic hydraulic factors can have a significant effect on pump reliability. They are Operating Speed (relative to the rated maximum), Percent of Best Efficiency Flow (Flow Ratio), Suction Energy and NPSH Margin Ratio. These last two factors have further been combined here into an NPSH Margin Reliability Factor (NPSH-RF), which is shown to be reasonably effective in predicting the reliability of High Suction Energy pumps. Laboratory tests on 3 API end-suction pumps, and maintenance data collected on 119 operational ANSI and split-case pumps, were used to validate the chosen The Pump Handbook Series

hydraulic-selection reliability indicators. Operating Speed, Flow Ratio and the new NPSH Margin Reliability Factor generated reasonable trend-lines that can be used to predict pump reliability. Hydraulic selection, based on these factors, can reduce the maintenance component of life-cycle cost. The laboratory reliability factors presented here (from ref. 1) are based on correlation of Bloch and Geitner’s reliability factors(2) with laboratory pump bearing-frame oil temperature, and vanepass vibration tests on 3 API (end-suction) pumps, plus published mechanical-seal face and abrasive-wear rates. The fieldtest reliability factors presented are derived from curve fits (trend lines) of actual MTBR data on 71 ANSI and 48 split-case pumps in two process plants. There was much scatter of the field data, due to the fact that the records were not cleansed of failures caused by factors other than hydraulic selection (such as human error, difficult to handle liquids, system interactions, or the mechanical design of the pumps). The duty cycles (operating times) varied between pumps, especially those in standby service. Also, the pumps were not always operated at the conditions of service analyzed. Despite the resulting large scatter in the data, definite trend lines could be and were developed, on the strength of the large number of pumps evaluated.

Operating Speed Operating speed affects reliability through rubbing contact (such as at seal faces), reduced bearing life through increased cycling, lubricant degradation and reduced viscosity due to increased 557

PS1101PG023

temperature, and wetted-component wear due to abrasives in the pumpage. High operating speed also increases the energy level of the pump, which can lead to cavitation damage. Figure 1 compares the API-610 pump laboratory reliability-predictor test results with the reliability trend line from MTBR data on the 119 process pumps, as a function of the ratio of the operating speed to the maximum rated pump speed (Speed Ratio). The reliability factor for the field-test data was the calculated ratio of each actual MTBR (defined here as the duration of the study period divided by the number of repairs) to the maximum MTBR in the pump population. It was assumed that pumps with zero repairs in the 48-month study period, and for which no meaningful MTBR could be calculated, would have required repair at a time increment following the study period. Conservatively, these pumps were assigned an MTBR of 72 months (and a Reliability Factor of 1.00). Both curves show a marked increase in reliability with reduced speed.

Percent Best Efficiency Flow Rate (Flow Ratio) The Flow Ratio affects reliability through the turbulence that is created in the casing and impeller as the pump is operated away from the best efficiency point (bep) flow rate. Resulting hydraulic loads, which are transmitted to the shaft and bearings, increase and become unsteady with increased divergence from bep. The severity of these unsteady loads also can reduce mechanical-seal life. Operation at reduced flow rates that put the pump into its recirculation mode can lead to cavitation damage in High Suction Energy pumps. (Refer to ANSI/HI 9.6.3, ref. 5, for more guidance on the allowable operating region for centrifugal and vertical pumps.) The field-data to laboratory reliability comparison, plotted against the Flow Ratio (actual flow as a percentage of bep flow) are presented in figure 2. The field data are, however, only based on the 48 splitcase pumps, since no definitive trend line could be established from the ANSI data. 558

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Figure 1. Speed Reliability Factor

Also, for trending purposes, the 1.00 Field End Suction Pump: Reliability Factor is based on a MTBR of De = 0.9 x Suction Nozzle Size 52 months. Correlation between the field Split Case/Radial Inlet Pumps: and laboratory data is good in the normal De = 0.75 x Suction Nozzle Size operating range, with the maximum reliability occurring around 90 percent of the Budris and Mayleben(3) also have best-efficiency flow rate. proposed distinct gating values for defining High and Very High Suction Energy Suction Energy pumps. The defining suction-energy Suction Energy is another term for the levels for end-suction and radial-suction liquid momentum in the suction eye of (also know as split-case or doublea pump impeller, which means that it is suction) configurations are based on the a function of the mass and velocity of the analysis of hundreds of pumps from liquid in the inlet. Suction Energy, as origi- several manufacturers. nally approximated by Budris and Start of High Suction Energy: Mayleben(3), is defined as follows: End Suction Pumps: Suction Energy (S.E.) = De x N x S x s.g. S.E. = 160 x 106 (Equation 1) Split Case/Radial Inlet Pumps: Where: S.E. = 120 x 106 De = Impeller Eye Diameter (inches) N = Pump Speed (RPM) S = Suction Specific Speed (RPM x Start of Very High Suction Energy: (GPM).5 )/ (NPSHR).75 End Suction Pumps: [NPSHR in feet] S.E. = 240 x 106 s.g. = Specific Gravity of Liquid pumped Split Case/Radial Inlet Pumps: Since the suction-energy numbers are S.E. = 180 x 106 quite large, the last six digits are normally dropped (i.e., S.E. x E6). It should be The above definitions of Suction noted that, if not readily available, the Energy (Equation 1), and of High and Impeller Eye Diameter can be approxi- Very High gating values are consistent mated as follows: with values presented in ANSI/HI 9.6.1(4). The Pump Handbook Series

PS1101PG024

Pumps with values of suction energy below these values are considered to have low suction energy. Generally speaking, Low Suction Energy pumps are not prone to noise, vibration or damage from cavitation. However, there could be detrimental effects on mechanical seals from the air or vapors that may be liberated from the liquid during the formation of the cavitation bubbles under low NPSH Margin conditions (NPSH Margin Ratio below 1.1 – 1.3). Figure 3 is based strictly on the field data for 77 ANSI and split-case pumps,

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with 42 Low Suction Energy pump failures below 48 months deleted. It is unlikely that these failures were caused by factors related to Suction Energy, mainly cavitation. Here also, a 1.00 Reliability Factor equates to no failures in 48 months, or an MTBR rate of 72 months. The trend is unquestionable, with higher suction-energy pumps requiring the most frequent repairs.

from the application, divided by the NPSH Required (NPSHR) by the pump. By Hydraulic Institute definition, the NPSHR of a pump is the NPSH that will cause the total head to be reduced by 3%, due to flow blockage from cavitation vapor in the impeller vanes. NPSHR is by no means the point at which cavitation starts. That level is referred to as incipient cavitation. It can take an NPSHA of from 2 to 20 times NPSHR to fully NPSH Margin suppress cavitation within a pump, NPSH Margin Ratio is defined as the depending on pump design and Flow NPSH Available (NPSHA) to the pump Ratio (percent bep). The higher incipient-cavitation margins are normally associated with high suction energy, high specific speed, pumps with large impeller inlet areas or reduced-flow operation in the region of suction recirculation(4). This means that a high percentage of pumps are operating with some degree of cavitation. It is the amount of energy associated with the collapse of the cavitation bubbles that determines the degree of noise, vibration or damage from cavitation, if any.

In process plants, unscheduled maintenance often

Figure 2. Percent Flow Factor

represents the most significant cost component of pump ownership. Figure 4 shows the affect of the NPSH Margin Ratio on pump reliability, based on the 77 field pumps. Again, the Low Suction Energy failures (below 48 months) were deleted, because it is unlikely that these failures were caused by factors related to Suction Energy. Based on these data, the NPSH Margin Ratio does have a definite influence on pump reliability. This is especially true for High

Figure 3. Suction Energy Factor (without S.E. Failures) The Pump Handbook Series

559

PS1101PG025

and Very High Suction Energy pumps, the 77 ANSI and split-case pumps (again, due to the fact that some cavitation usually without the 42 Low Suction Energy pump exists below a Margin Ratio of 4.0. failures below 48 months), as shown in figure 6. Although not perfect, the agreeNPSH Margin Reliability Factor ment is quite good. It must be rememThe NPSH Margin Reliability Factor bered that the NPSH-RF only applies to (Fig. 5) was developed to quantify the High Suction Energy and Very High relationship of both NPSH Margin and Suction Energy pumps. Suction Energy to pump reliability. The NPSH Margin Reliability Factors are based Conclusions on the fact that, for applications above the The Speed, Flow Ratio, Suction Energy gating suction-energy values (lower limit and NPSH Margin reliability propositions of High Suction Energy), greater suction energy makes it more important to suppress the residual cavitation, cavitation that exists above the NPSHR, to prevent damage. This reliability factor is only applicable within the allowable operating flow region, above the start of suction recirculation (see ref. 5). Much higher NPSH Margin values are required in the region of suction recirculation for High and Very High Suction Energy pump applications.

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and methodologies were confirmed by actual field experience. The Mean Time Between Repair and Life Cycle Cost of most centrifugal pumps can be improved if slower pump speeds are used, and if pumps are selected to operate in their preferred operating range (70% - 120% of bep flow rate, ref. 5). One way to reduce pump speed is to select larger, slower pumps. Another is to select a Variable Frequency Drive (VFD) over a control valve for

The authors believe that four basic hydraulic factors

Figure 4. NPSH Margin Factor (without Low S.E. Failures)

can have a significant effect on pump reliability. The diagonal lines (in Fig. 5) are lines of constant relative Suction Energy (x 106). Therefore, (for example) the line marked “180/240” (Double-Suction pump Suction-Energy level / End-Suction pump Suction-Energy level) represents the start of Very High Suction Energy. Pumps of this suction-energy level require a minimum NPSH Margin Ratio of 2.5 for maximum reliability. To validate the NPSH Margin Reliability Factors in figure 5, NPSH-RF values were plotted against the field reliability of 560

Figure 5. NPSH Margin Reliability Factor The Pump Handbook Series

PS1101PG026

Figure 6. NPSH Margin Reliability Factor (w/0 Low S.E. Failures)

process-control applications, which normally will result in lower pump speeds and in operation closer to the preferred operating region. Further, the Mean Time Between Repair of High and Very High Suction Energy pumps can be increased by keeping the NPSH Margin Ratio above the values recommended in Figure 5, and/or by reducing the Suction Energy. The easiest way to lower the Suction Energy and to increase the NPSH Margin of a pump application is by lowering the speed of the pump. Allan R. Budris, P.E., is Director of Product Development for the Industrial Pump Group of ITT Industries, in Seneca Falls, NY. Previously, he was the Director of Engineering for ITT A-C Pump, Manager of Engineering for ITT Marlow in NJ and he has held various positions with Worthington Corp./Dresser Industries Incorporated. Mr. Budris is Chairman of the Hydraulic Institute Pump Piping Work Group. He has authored several journal publications, is a member of ASME, and has been awarded a number of patents. Mr. Budris received his M.S. degree in Mechan-

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R. Barry Erickson, Ph.D., is currently Vice President of Technology with the Industrial Pump Group of ITT Industries, in Seneca Falls, NY. His prior industrial experience includes 17 years in engineering at Allis Chalmers in Cincinnati and 8 years in product development at Duriron (now Flowserve) in Dayton, OH. Dr. Erickson has been active on many ASME and Hydraulic Institute technical committees, and is presently Chairman of the Centrifugal Section of the Hydraulic Institute. He holds two patents, and has authored numerous technical papers and standards. Erickson received his Ph.D. from the University of Cincinnati. During his tenure there he taught thermal and fluid science courses in the Mechanical Engineering Department and, after graduation, he continued as an Adjunct Assistant Professor for 20 years.

ical Engineering from Newark College of References Engineering and has attended courses at 1. Erickson, R. B., Sabini E. P. and Rochester Institute of Technology, AmeriStavale, A. E., October 2000, can Management Association, Dresser “Hydraulic Selection to Minimize Management Development, and the Goldratt the Unscheduled Maintenance Institute. He is a registered Professional Portion of Life Cycle Cost”, Pump Engineer in the State of New Jersey. E-mail Users International Forum 2000, him at [email protected]. Karlsruhe, Germany. 2. Bloch, H.P. and Geitner, F. K., 1994, Eugene P. (Gene) Sabini is the Director of “An Introduction to Machinery Technology for the Industrial Pump Group Reliability Assessment”, Gulf of ITT Industries, in Seneca Falls, NY, and is Publishing Company, Houston, TX. responsible for applied research and hydraulic 3. Budris, A. R. and Mayleben, P. A., design of all new products and field re-rates. 1998, “Effects of Entrained Air, Mr. Sabini was previously Manager of Energy NPSH Margin, and Suction Piping Engineering Design with Goulds Pumps. Mr. on Cavitation in Centrifugal Sabini has 33 years of experience in the Pumps”, International Pump Users pumping industry including design and develSymposium proceedings, Texas opment of many centrifugal pumps for the A&M University, Houston, TX. chemical, API, power utilities, and municipal 4. ANSI/HI 9.6.1 – 2000, “Centrifugal industries. He spent 25 years with Worthingand Vertical Pumps for NPSH ton Pump, designing, engineering, and testing Margin”, Hydraulic Institute, custom centrifugal pumps from both a Parsippany, NJ. mechanical and hydraulic standpoint. Mr. 5. ANSI/HI 9.6.3 – 2000, “Centrifugal Sabini has written numerous papers and holds and Vertical Pumps for Allowable four patents. He received a BSME and M.S. Operating Region”, Hydraulic degree from Stevens Institute of Technology. Institute, Parsippany, NJ.

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561

? PUMP AND SYSTEM TROUBLESHOOTING HANDBOOK Table of Contents Inlet Piping Configuration and Motor Efficiency.......................................................1 Cavitation and NPSH in Centrifugal Pumps..............................................................3 Pump Suction Lift, NPSH and Priming .....................................................................7 How Much NPSH is Enough?....................................................................................11 Variable Speed Drives: Advantages and Pitfalls .......................................................13 Estimating Maximum Head in Single and Multi-Stage Pump Systems..................16 Elements of Minimum Flow .....................................................................................18 NPSH Required for Reciprocating PD Pump ...........................................................22 Pump Suction Conditions .........................................................................................24 Where’s the Prime?....................................................................................................27 Pump System Design – Part 1...................................................................................29 Pump System Design – Part 2...................................................................................32 Pump System Design – Part 3...................................................................................35 Pumping System Piping ............................................................................................37 Have You Broken Any Shafts Lately?........................................................................39 Pumps in Parallel – When One is Not Enough........................................................41 Pumps in Series – For More Pressure.......................................................................43 Vertical Pumps with Integral Thrust Bearings.........................................................45 Why doesn't the pump pump? .................................................................................48 Pumping Downhill ....................................................................................................50 Selecting a Pump- The Right Start Leads to the Right Finish .................................52 Selecting a Pump- What Type Should It Be?............................................................53 Selecting a Pump–Suction Pressure..........................................................................54 Troubleshooting Pump Performance Degradation...................................................55 Vibration Amplification.............................................................................................58 Selecting a Pump? Define Efficiency First!...............................................................60 Selecting a Pump: Will it Operate Where You Want it to? ..................................... 62 Performance Curve vs. System Curve .......................................................................64 Calculating Shaft Deflection......................................................................................67 Pumps in Parallel with Variable Speed Drive ..........................................................69

? PUMP AND SYSTEM TROUBLESHOOTING HANDBOOK Table of Contents Sizing Pumps for Complex Systems: Part 1..............................................................71 Sizing Pumps for Complex Systems: Part 2..............................................................72 Sizing Pumps for Complex Systems: Part 3..............................................................74 Pumping with Air......................................................................................................75 Troubleshooting Centrifugal Pumps.........................................................................77 Water Hammer - Containing the Surge ....................................................................80 Reducing Noise in API Process Pumps.....................................................................82 Pump and Turbine Rerate Problems........................................................................85 Looking for Hidden Dangers .....................................................................................93 The Correction and Prevention of Low NPSH Conditions .............................95 Torsional Vibration Linked to Water Pumping System Failure ......................98 Modeling Pump Intake Noise.......................................................................100 Motor Bearing Failures in Cooling Tower Water Pump................................103 Cavitation in a Nutshell ...............................................................................106 Hydraulic Pressure Reactions in Pump Piping Systems ..............................109 Developing Meaningful Pump Failure Data ................................................112 The Mystery of Cooling Tower Pump Noise ................................................117 Controlling Surge and Pulsation Problems ..................................................120 Avoiding Gremlins and Alligators at Pump Start-up ....................................126 Experience with Replacement of Boiler Feed Pumps for Reliability Enhancement..130 Troubleshooting ANSI Pumps Using Limited Information ..........................134 Why Doesn’t Your Self-Priming Pump Work? ..............................................137 In-Plant Perspective: Eastman Chemical Company ....................................140 Root Cause Failure Analysis ........................................................................148 Coupling Alignment .....................................................................................................155

All materials © 2002 Pumps & Systems, LLC. No part of this publication may be reproduced without the written consent of the publisher. The publisher does not warrant, either expressly or by implication, the factual accuracy of the articles or descriptions herein, nor does the publisher warrant the accuracy of any views or opinions offered by the authors of said articles or descriptions.

?

PUMP AND SYSTEM TROUBLESHOOTING HANDBOOK

Inlet Piping Configuration and Motor Efficiency BY ROSS C. MACKAY AND JOHN T. DAVIS The article entitled ”Pump Suction Conditions“ in the May 1993 issue of Pumps and Systems stated that, ”The suction of a pump should be fitted with an eccentric reducer positioned with the flat side uppermost.“ However, some papers seem to indicate that there may be an exception to this rule when the pump suction is coming from an elevated tank. Can you provide a clarification of this point and an explanation of the reasons for the positioning of a reducer in the cases of drawing from a sump, an elevated tank, and a side suction condition? David W. Lawhon, Rotating Equipment Engineer & Jeffrey A. Robinette, Project Engineer Basic Chemicals Group Occidental Chemical Corporation, Ingleside, TX

Q:

The physical location of a pump relative to its supply source does not change the rules of inlet piping design, and the reason for their existence. When designing the layout of suction piping, it‘s not necessary to be influenced by whether or not it‘s approaching the pump from the side, the top or the bottom. The only concern over piping configuration is the need to deliver the liquid to the eye of the impeller in a smooth parallel flow of uniform velocity. An inverted eccentric reducer and a concentric reducer produce the same problem in a horizontal line (Figure A). The flow through an elbow in a horizontal and a vertical line, creates a similar uneven flow pattern (Figure B). Some consideration has been given to the possibility that an acceptable flow pattern could be achieved by leading a vertical line, down through the elbow, into a horizontal eccentric reducer in the inverted position. However, when we consider the

A:

combination of these, it is evident that it doesn’t really help (Figure C). If it is absolutely essential that an elbow and a reducer be adjacent to each other, close to the pump suction, consider combining the two into a reducing, wide sweep elbow (Figure D). Severe space restriction is the only sound reason for not providing suitable inlet piping for centrifugal pumps and, if you’re in this position, you may be faced with some choices. At that point, you may wish to reconsider the five rules previously outlined: 1. provide sufficient NPSH 2. minimize friction loss 3. no elbows on the suction flange 4. eliminate vapor from the suction line 5. ensure correct piping alignment

FIGURE 1 Fig. A

Air Pocket

Air Pocket Fig. B

Air Pocket Fig. C

Any compromise of these rules will reduce pump reliability, while conformance to them will ensure that inlet piping design will not be a factor in any future pump problems. Fig. D When selecting an electric motor for a centrifugal pump, is there an advantage to purchasing a motor with an efficiency of 96% in lieu of a motor with an efficiency of 87% if the pump operating efficiency is only 75%?

Q:

Albert M. Diaz, Production Engineer Advisor, Unocal, Energy Resources Division, Midland, TX

A:

The efficiency of the electric motor driving a pump or any other load will directly affect the efficien-

The Pump Handbook Series

cy of the pump/ motor system. The financial advantage gained by selecting a motor higher in efficiency depends on numerous factors, including the cost of electricity, the hours of operation, the load imposed on the motor, and the differential initial cost of the motors being compared. In a motor retrofit, analysis of

1

the operating speed of the system should be considered to avoid having the higher efficient motor consume more power, producing additional and possibly unnecessary process output. Simplistically speaking, motor efficiency multiplied by pump efficiency determines the system efficiency. There are other losses to consider, but given the motor and pump efficiencies in your question, the system efficiencies would be 72% versus 65.3%. This system efficiency differential might require considerable additional expenditure. A

2

common rating used on a horizontal injection pump is 300 hp Two Pole. Given • 7200 hours per year operation • 6 cents per KW-hr power cost • 300 hp motor • operated at full load continu ous Calculated energy costs per year: motor (96.2% efficient) $100,500 motor (91.0% efficient)$106,243

The Pump Handbook Series

The savings is $5743 per year! This is probably a worst case analysis: running all year, expensive energy, large efficiency differential—but obviously a significant amount of energy costs are involved and it is common for the extra cost of a higher efficient motor to be justified by energy savings. Analysis that includes cash flow, tax rates, power factors, and other issues can be performed for a more complete review. ■ Answer to first question provided by Ross C. Mackay, second answer by John T. Davis.

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PUMP AND SYSTEM TROUBLESHOOTING HANDBOOK

Cavitation and NPSH in Centrifugal Pumps BY PAUL T. LAHR

C

FIGURE 1 985

800

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22.5"

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BU

30

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60 50 40

ABSOLUTE PRESSURE–LBS. PER SQ. IN.

Flow is reduced as the liquid is displaced by vapor, and mechanical imbalance occurs as the impeller passages are partially filled with lighter vapors. This results in vibration and shaft deflection, eventually resulting in bearing failures, packing or seal leakage, and shaft breakage. In multi-stage pumps this can cause loss of thrust balance and thrust bearing failures.

300

E

FID

L NE M SU HA RO EN ET LUO G O IF E DR TR E IN HY RO EN OR LO YL HL P H C O OC PR ON

200

80

400

E

AN

H ET

AR

C

BUBBLE FORMATION PHASE

1. Mechanical damage occurs as the imploding bubbles remove segments of impeller material. 2. Noise and vibration result from the implosion. Noise that sounds like gravel being pumped is often the user’s first warning of cavitation.

E

ID

OX

N BO

300

100

600 500

E

ID OX

29" 29.1" 29.2" 29.3"

.40 .30

29.4"

.20

29.5" 29.6"

NET POSITIVE SUCTION HEAD

29.7" 29.72"

.10

When designing a pumping 60 30 0 30 60 90 120 150 180 system and selecting a pump, one TEMPERATURE–F must thoroughly evaluate net positive suction head (NPSH) to pre- Vapor pressures of various liquids related to temperature. vent cavitation. A proper analysis

The Pump Handbook Series

210

240

3

VACUUM–INCHES OF MERCURY

600 500 400

GAUGE PRESSURE–LBS. PER SQ. IN.

-60° to 240°F 1000

CAVITATION EFFECTS

BUBBLE COLLAPSE PHASE

friction in the suction pipe is a common negative component of NPSHA, the value of NPSHA will always decrease with flow. NPSHA must be calculated to a stated reference elevation, such as the foundation on which the pump is to be mounted. NPSHR is always referenced to the pump impeller center line.

involves both the net positive suction heads available in the system (NPSHA) and the net positive suction head required by the pump (NPSHR). NPSHA is the measurement or calculation of the absolute pressure above the vapor pressure at the pump suction flange. Figure 2 illustrates methods of calculating NPSHA for various suction systems. Since

C

avitation is the formation and collapse of vapor bubbles in a liquid. Bubble formation occurs at a point where the pressure is less than the vapor pressure, and bubble collapse or implosion occurs at a point where the pressure is increased to the vapor pressure. Figure 1 shows vapor pressure temperature characteristics. This phenomenon can also occur with ship propellers and in other hydraulic systems such as bypass orifices and throttle valves—situations where an increase in velocity with resulting decrease in pressure can reduce pressure below the liquid vapor pressure.

4

FIGURE 2 4b SUCTION SUPPLY OPEN TO ATMOSPHERE -with Suction Head

4a SUCTION SUPPLY OPEN TO ATMOSPHERE -with Suction Lift

CL

PB

Ls LH PB

NPSHA=PB + LH – (VP + ht)

NPSHA=PB – (VP + Ls + ht)

CL

4d CLOSED SUCTION SUPPLY -with Suction Head

4c CLOSED SUCTION SUPPLY -with Suction Lift

p

CL

LH

Ls

NPSHA=p + LH – (VP + ht)

NPSHA=p – (Ls + VP + ht)

CL

p

Calculation of system net positive suction head available (NPSHA) for typical suction conditions. PB = barometric pressure in feet absolute, VP = vapor pressure of the liquid at maximum pumping temperature in feet absolute, p = pressure on surface of liquid in closed suction tank in feet absolute, Ls = maximum suction lift in feet, LH = minimum static suction head in feet, hf = friction loss in feet in suction pipe at required capacity.

FIGURE 3 ENTRANCE LOSS

FRICTION

TURBULANCE FRICTION ENTRANCE LOSS AT VANE TIPS

INCREASE PRESSURE DUE TO IMPELLER

E D

B C

POINT OF LOWEST PRESSURE WHERE VAPORIZATION STARTS

A

INCREASE PRESSURE

It is a measure of the pressure drop as the liquid travels from the pump suction flange along the inlet to the pump impeller. This loss is due primarily to friction and turbulence. Turbulence loss is extremely high at low flow and then decreases with flow to the best efficiency point. Friction loss increases with increased flow. As a result, the internal pump losses will be high at low flow, dropping at generally 20–30% of the best efficiency flow, then increasing with flow. The complex subject of turbulence and NPSHR at low flow is best left to another discussion. Figure 3 shows the pressure profile across a typical pump at a fixed flow condition. The pressure decrease from point B to point D is the NPSHR for the pump at the stated flow. The pump manufacturer determines the actual NPSHR for each pump over its complete operating range by a series of tests. The detailed test procedure is described in the Hydraulic Institute Test Standard 1988 Centrifugal Pumps 1.6. Industry has agreed on a 3% head reduction at constant flow as the standard value to establish NPSHR. Figure 4 shows typical results of a series of NPSHR tests. The pump system designer must understand that the published NPSHR data established above are based on a 3% head reduction. Under these conditions the pump is cavitating. At the normal operating point the NPSHA must exceed the NPSHR by a sufficient margin to eliminate the 3% head drop and the resulting cavitation. The NPSHA margin required will vary with pump design and other factors, and the exact margin cannot be precisely predicted. For most applications the NPSHA will exceed the NPSHR by a significant amount, and the NPSH margin is not a consideration. For those applications where the

A

B

C

D

E

POINTS ALONG LIQUID PATH RELATIVE PRESSURE IN THE ENTRANCE SECTION OF A PUMP

The pressure profile across a typical pump at a fixed flow condition.

The Pump Handbook Series

The system designer FIGURE 4 should also calculate the system suction specific speed by substituting design flow rate and the Q1 system designer’s NPSHA. Q2 The pump speed N is generally determined by the 100% CAP Q3 head or pressure required in the system. For a lowQ4 3% maintenance pump system, designers and most user specifications require, or prefer, Ss values below 10,000 to 12,000. However, as indicated NPSH above, the pump Ss is dictated to a great extent by the system conditions, Typical results of a four-point net posidesign flow, head, and the tive suction head required (NPSHR) test based on a 3% head drop. NPSHA. Figures 5 and 6 are plots of Ss versus flow in tem requiring a 3,500 rpm pump gpm for various NPSHA or NPSHR with 20 feet of NPSHA, the maxiat 3,500 and 1,750 rpm. Similar plots mum flow must be limited to can be made for other common 1,000 gpm if the maximum Ss is to pump speeds. be maintained at 12,000. Various Using curves from Figure 5 and options are available, such as Figure 6 allows the system designer reducing the head to allow 1,750 to design the system Ss, i.e., for a sys-

SUCTION SPECIFIC SPEED The concept of suction specific speed (Ss) must be considered by the pump designer, pump application engineer, and the system designer to ensure a cavitation-free pump with high reliability and the ability to operate over a wide flow range. N x Q0.5 Ss = —————— (NPSHR)0.75 where N = pump rpm Q = flow rate in gpm at the best efficiency point NPSHR = NPSHR at Q with the maximum impeller diameter

NPSHR

TOTAL HEAD

NPSH A is close to the NPSH R (2–3 feet), users should consult the pump manufacturer and the two should agree on a suitable NPSH margin. In these deliberations, factors such as liquid characteristic, minimum and normal NPSH A, and normal operating flow must be considered.

FIGURE 5 1 9 8 7 6 5 4

S1 Suction specific speed

3

2

V=2

HS

1 9 8

3

4

7

5

6

6 7

5

8

4

9

V=2

HS

10

4

2

V=1 14

HS

16

18

28 32 36

20

50 55

5

V=4

HS

40

60

65

3

Solution for S=N

2

Q Hsv0.75

for N=3,500 rpm 1 1

2

3

4

5

6

7

8

9 1

2

3

4

5

6

7

8

9 1

2

3

4

5

6

7

8

9 1

Q, Capacity, gpm

A plot of suction specific speed (Ss) versus flow in gallons per minute (gpm) for various NPSHA or NPSHR at 3,500 rpm. (Single suction pumps. For double suction use 1/2 capacity). Hsv=NPSHR at BEP with maximum impeller diameter.

The Pump Handbook Series

5

FIGURE 6 1 9 8 7 6 5 4

S1 Suction specific speed

3

2

1 9 8 7

=1

HSV

6

2

5 3

4

4

14

7 8 9

HSV

HSV

6

2

=45

=12

5

3

16

18

28

=24

HSV

20

32

50

40

36

Solution for

10

S=N

Q Hsv0.75

for N=1,750 rpm 1 1

2

3

4

5

6

7

8

9 1

2

3

4

5

6

7

8

9 1

2

3

4

5

6

7

8

9 1

Q, Capacity, gpm

A plot of suction specific speed (Ss) versus flow in gallons per minute (gpm) for various NPSHA or NPSHR at 1,750 rpm. (Single suction pumps. For double suction use 1/2 capacity.) HSV=NPSHR at BEP with the maximum impeller diameter.

FIGURE 7

2 3

NPSH - FEET

HEAD

1

4

C

GPM

A

B

A typical plot of the suction and discharge systems. Curve 1 = pump head capacity performance, curve 2 = total system curve, curve 3 = suction system curve NPSHA, and curve 4 = pump NPSHR.

6

rpm (Figure 7). This would allow flows to 4,000 gpm with 20 feet of NPSHA. It is important for the pump user to understand how critical the system design requirements are to the selection of a reliable, trouble-free pump. Matching the system and pump characteristics is a must. Frequently, more attention is paid to the discharge side. Yet it is well known that most pump performance problems are caused by problems on the suction side. Figure 7 is a typical plot of the suction and discharge systems. It is important that points A, B, and C be well established and understood. A is the normal operating point. B is the maximum flow for

The Pump Handbook Series

cavitation-free operation. C is the minimum stable flow, which is dictated by the suction specific speed. As a general rule, the higher the suction specific speed, the higher the minimum stable flow capacity will be. If a pump is always operated at its best efficiency point, a high value of Ss will not create problems. However, if the pump is to be operated at reduced flow, then the Ss value must be given careful consideration. ■

REFERENCES 1. Goulds Pump Manual. 2. Durco Pump Engineering Manual. 3. Hydraulic Institute Test Standards—1988 Centrifugal Pumps 1.6. Paul T. Lahr is the owner of Pump Technology, a consulting firm. He serves on the Pumps and Systems Editorial Advisory Board.

?

PUMP AND SYSTEM TROUBLESHOOTING HANDBOOK

Pump Suction Lift, NPSH and Priming BY STEPHEN D. ABLE INTRODUCTION

FIGURE 1

In cases where suction performance is evaluated in terms of suction lift, actual pump performance may not meet the user’s expectation. Suction lift is generally sufficient when handling nonvolatile liquids (with a specific gravity near 1.0) from open supply tanks at ambient temperatures. If the supply is pressurized, the fluid is volatile, or has a specific gravity other than 1.0; the use of net positive suction head (NPSH) would be indicated. Several examples are presented which illustrate the need to evaluate pump suction performance in terms of NPSH. Confusion can arise concerning the meaning and use of “priming” verses suction lift, and the meaning of “dry” and “wet” suction lift.

Atmospheric Pressure (14.7 psia) + 5 psig + 19.7 psia

Entrance

5 feet

NPSHA According to Hydraulic Institute Standards (Ref. 1), net positive suction head available (NPSHA) is the total suction pressure, including allowance for acceleration head available from the system at the pump suction connection, minus the vapor pressure of the liquid at the pumping temperature. NPSHA for a reciprocating pump is normally expressed in pounds per square inch (psi) or feet. If the pressures and temperatures are provided at some distance from the suction of the pump, frictional head losses must be calculated and subtracted. NPSHA = (P – Pv – Pf) x

2.31 sp. gr.-H

where P = absolute pressure on the surface of the liquid (psia, i.e., atmospheric pressure) Pv = vapor pressure of the liquid (psi) Pf = pressure loss due to inlet piping friction sp gr = specific gravity of the liquid Hs = height of the pumps suction inlet above the tank level

NPSHR Net positive suction head required is the total head that the

Gate Valve

4 inch Pipe (10 feet Total)

90° Elbow

Suction piping schematic for Example 4, pumping a viscous fluid pump requires to operate properly without cavitation (local boiling of the fluid), a reduction in pressure-flow performance, excessive vibration or noise. NPSHA must always be greater than NPSHR.

SUCTION LIFT Hydraulic Institute defines “total suction lift” as the pressure below atmospheric at the inlet port of the pump plus the suction system frictional losses. Thus, total suction lift is equivalent to the height of the pump suction above the tank level plus the piping system frictional losses. Many pump companies use a closed loop test arrangement with water to establish the NPSHR of their pumps as well as the head, flow and efficiency. The Pump Handbook Series

PRIMING AND DRY SUCTION LIFT Suction lift (also NPSH) and priming are two separate issues. Priming has to do with how well the pump can draw a vacuum and create a suction in the inlet. In the case of positive displacement pumps, the “dry lift” capability of a pump indicates how well the check valves work when sealing against a vacuum. If a pump has perfect valve sealing, it could lift a column of water 33.96 feet at sea level; that is, 14.7 psia x 2.31/1.0 x sp gr.

SPECIFIC GRAVITY AND VAPOR PRESSURE EFFECTS ON SUCTION LIFT If the liquid has a specific gravity of 2, the pump could only lift the liquid a maximum of 16.98 feet (that

7

is, 33.96/2). If the liquid has a significant vapor pressure, it would have the effect of lowering the differential pressure across the liquid column. In other words, at sea level there could no longer be a total of 14.7 psi available to lift the liquid column. If, for example, the liquid has a vapor pressure of 1 psi and a specific gravity of 2, a pump with ideal valving could lift the liquid only 15.82 feet; that is, 14.7 psia – 1 psia x 2.31/2 sp gr. Reduced atmospheric pressure has a similar effect.

FIGURE 2

PIPING LOSSES If the pressures and temperatures are provided at some distance from the suction of the pump, frictional head losses must be calculated and subtracted. To calculate or look up piping loss information through piping of various sizes and viscosities see Cameron Hydraulic Data or Thermodynamic Properties of Steam (Ref. 2 and 3). These references also provide information on losses through fittings, elbows, etc.

EXAMPLE 1: LIFTING COLD WATER An application requires a pump to lift 20 feet at the flow rate specified. The factory test data indicate that the pump has an NPSHR of 12 feet at that flow rate. Determine the NPSHA of the application and static height that the pump could lift the fluid for this application. Note: The vapor pressure of water at 70°F is equal to 0.3631 psi (Ref. 3). Assume that frictional losses are negligible. NPSHA = [(14.7 – 0.36 – 0) x 2.31] / (1.0 - 20) = 13.1 feet Static height from the NPSH equation: Hs = [(14.7 – 0.36 – 0) x 2.31]/1.0 – 12 (NPSHR) = 21.1 feet Therefore, the pump selected can be used in this application because the NPSHA is greater than the NPSHR. The resulting calculation for the static height indicates this as well.

8

Friction factors for commercial pipe (Cameron Hydraulic Data)

EXAMPLE 2: LIFTING A HIGH SPECIFIC GRAVITY FLUID If the fluid pumped in Example 1 is changed to sulfuric acid (sp gr = 1.83), the suction performance would change for this application. Determine the static height, Hs, that the pump would lift. The Pump Handbook Series

Note: At 330°C, sulfuric acid has a vapor pressure of 1 atmosphere (Ref. 4). Because water exhibits this pressure at 100°C, assume that the vapor pressure for sulfuric acid at room temperature is negligible. Hs = [(14.7 – 0 – 0) x 2.31]/1.83 – 12 (NPSHR) = 6.6 feet

that can be lifted by the pump. Note also that atmospheric pressure drops about 1 psi for 1000 feet of elevation. Therefore, elevation can affect pump suction performance as well as line losses.

FIGURE 3

EXAMPLE 4: PUMPING A VISCOUS FLUID For viscous materials that are at or near room temperature and have a supply tank at least a few feet away, use an NPSH calculation rather than only suction lift. Suction lift does not take into account fluid viscosity and the resulting frictional losses in the suction system. For this example: • tank is pressurized to 5 psig •

atmospheric pressure equals 14.7 psi

•

pumped fluid is at room temperature with a viscosity of 100,000 cP

•

10 feet of 4-inch pipe on suction

•

one 4-inch elbow

•

one 4-inch open gate valve

•

specific gravity equals 1.0

•

pump suction is 5 feet below minimum tank level

•

maximum flow rate during stroke is 2 gpm

Since it is generally difficult to find tabular data for friction loss in piping, valves, and fittings at higher viscosities the pipe friction loss can be calculated as follows: Hf = f x L x V2/D/2/g Note: This equation is for nonturbulent (laminar) flow. If Re > 2000 flow is considered turbulent you must use a “Moody Diagram” (Figure 2). where Resistance of valves and fittings to flow of fluids in equivalent length of pipe (Hydraulic Institute Standards) Therefore, a pump that can lift 21.1 feet on a cold water test can only lift 6.6 feet of cold sulfuric acid.

EXAMPLE 3: LIFTING COLD SULFURIC ACID This example is for the same pump as in Examples 1 and 2, but

with 1 psig of frictional line pressure loss (Figure 1). Hs = [(14.7 – 0 – 1) x 2.31]/1.83 – 12 (NPSHR) = 5.3 feet The effect of the line losses further reduces the static height of fluid The Pump Handbook Series

f = friction factor = 64/Re Re = Reynold’s number = 3162 x Q/d/k [Ref. 4] Q = flow, gpm during the stroke of the plunger d = diameter of pipe, inches D = diameter of pipe, feet L = pipe length, feet

9

k = kinematic viscosity, centistokes = v, absolute viscosity (centipoise) x specific gravity V = velocity of fluid, during the stroke of the plunger = 0.408 x Q /d2 [Ref. 4] g = gravitational constant, 32.2 ft/s2 So: Re = 3162 x 2/4/100000 = 0.0158 V = 0.4085 x 2/42 = 0.05 ft/s f = 64/Re = 64/0.0158 = 4050 Hf = 4050 x 10 x (0.05)2/(4/12)/2/32.2 = 4.7 ft = 4.7/2.31 = 2.04 psi Avoid using the head loss formula (hf = k x V2 / 2g) because it is not corrected for viscosity effects. The friction losses at the tank exit, valves, and fittings are given by looking up the equivalent length of straight pipe for the restrictions (Figure 3): 1.

L (4-inch elbow) = 10.1 ft of pipe

2.

L (open gate valve) = 2.3 ft of pipe

3.

L (tank exit is the ordinary exit) = 6 ft of pipe

The equivalent length of straight pipe for the three restrictions is equal to 18.4 ft of 4-inch pipe (10.1 + 2.3 + 6). The head loss through the three restrictions can be found by knowing

10

the loss in 10 feet of straight pipe. From the earlier calculation, there is a 4.7 foot head loss in 10 feet of straight pipe. Therefore, the loss in the three restrictions is: hf multiplied by the three restrictions = 4.7 ft x 18.4 ft of pipe 10 ft of pipe = 8.7 ft (or 3.7 psi) The total frictional losses are:

pump will perform better than predicted, since most high viscosity fluids are shear-thinning). Depending on pump type, pump operation and efficiency can be altered when handling viscous fluids. ■

REFERENCES 1.

Hydraulic Institute, 1983, Hydraulic Institute Standards for Centrifugal, Rotary, and Reciprocating Pumps, Fourteenth Edition.

2.

Heald, C.C. (1988), Cameron Hydraulic Data, Seventeenth Edition.

3.

Keenan, J.H. and Keyes, F.G., (1964), Thermodynamic Proper-ties of Steam, Thirty Sixth Printing, page 28.

4.

The Crane Company (1988), “Flow of Fluids through Valves, Fittings and Pipe,” Technical Paper No. 410, page 32.

5.

Weast, R.C. (1969), Handbook of Chemistry and Physics, 50th Edition, page D144.

hf = 4.7 + 8.7 = 13.4 ft (or 5.8 psi) The NPSHA is equal to: Absolute tank pressure: 5 + 14.7 = 19.7 psia (or 45.5 ft) vapor pressure of the fluid at the pump suction: assume it is zero piping system losses: 5.8 psi (13.4 ft) height of fluid above pump suction : 5 ft NPSH A = [ (14.7 + 5) – 0 – 5.8] x (2.31 / 1.0) + 5 = 37.1 ft In this example the NPSHA is greater than in the other examples due to the addition of tank pressure and the maintenance of a minimum tank level above the pump. Fluids with a high viscosity often undergo changes in viscosity when set in motion, which can significantly affect the prediction. (Normally, the

The Pump Handbook Series

Stephen D. Able is a Senior Engineering Consultant with ARO Fluid Products Division, Ingersoll-Rand in Bryan, OH.

?

PUMP AND SYSTEM TROUBLESHOOTING HANDBOOK

How Much NPSH is Enough? BY JOHN SIDELKO Knowing how important NPSH is to the satisfactory operation of our pumps, my colleagues always recommend that we provide more than the amount cited by the pump vendor. How much more NPSH is enough? The Net Positive Suction Head (NPSH) required, quoted by the pump vendor, is based on Hydraulic Institute Standards. It is a measure of the total energy in the fluid on the suction side of the pump. This value of low suction energy at any given flow and speed causes a 3% reduction in the normal differential head created by the pump. The overall performance is adversely affected because the pump is moving liquid away faster than it is allowed to enter. This problem occurs more frequently when handling liquids close to their boiling point. Vibration (generally not associated with any mechanical frequency in the machine), shortened bearing and seal life, and possibly even impeller and case damage will happen repeatedly. Other criteria subject to evaluation and justification during the purchasing process such as total installed cost, efficiency, and maintenance, are overshadowed by the absolute necessity for enough NPSH. In low energy pumping systems (low flow, low speed, and low horsepower), even though something is wrong within a pump operating at its value of NPSH required, the amount of energy involved is too low to cause damage. Hydrocarbons, due to their chemistry, also don’t dissipate much energy during cavitation. It is common in these situations to supply two feet more than the maximum NPSH required by the pump. Critical services that are the heart of the process, and/or high energy machines (high flow, high speed, or high horsepower) are the ones that should compel pump users to add a margin of safety

Q:

A:

so that system NPSH available always exceeds the pump NPSH required. Figure 1 will help determine the amount of cushion necessary to achieve reliable pump operation. Many pump manufacturers generate diagrams like Figure 1, based on factory tests, to accurately determine the Hydraulic Institute definition of NPSH required over the complete range of speed and flow for each model they sell. Figure 1 clearly depicts the rate of pump performance degradation due to decreasing values of NPSH. It is this rate of degradation and the amount of energy being added to the fluid that determines if a pump will perform satisfactorily under marginal amounts of NPSH available. Figure 1 represents additional information that users should request of pump vendors because it is not commonly published data. The illustration should represent the pump’s behavior at the most difficult operating point—the maximum rated flow and speed for a specific installation.

Sufficient NPSHA is a safe operating region for any centrifugal pump.

Two situations exist: either new equipment is under evaluation to complete a specific process requirement, or the pumps and system are already in place.

NEW EQUIPMENT

Many articles have been written and commercial software has been developed to help calculate system NPSH A. When NPSH available is only zero to ten feet greater than NPSH required, the pump vendor should be asked to provide a graph similar to Figure 1 for the pump operating at its maximum rated speed and flow. Try to select a pump that interacts with NPSH available in the system to ensure 99% differential head or better. If not, contact the pump vendor to see if any internal hardware changes can be made to improve the suction performance. Often larger inlet eye diameters are used to lower NPSH required. High energy Piping alterations on the suction side of the machines machine also affect should be NPSH available. The pump user must deteroperated mine this value and make the necessary with enough improvements. Larger system NPSH suction lines, shorter distance, fewer bends and available to restrictions, or higher tank level maintain 99% minimum help raise NPSH available in the system. differential

•

Insufficient NPSHA (below the Hydraulic Institute definition of NPSH required) will cause unreliable pump operation.

•

Marginal NPSHA is the transition zone where pumps may or may not work well. Low energy pumps will be fine. Critical services or high energy machines should be operated with head enough system NPSH available to maintain 99% differential head or more. Depending on the inlet geometry of the pump, sometimes only several feet of increased NPSH may be needed to push the pump into the area of reliable operation. The Pump Handbook Series

•

or more.

EXISTING EQUIPMENT

Although it is not as easy to solicit cooperation after the equipment sale, manufacturing tolerances are accurate enough by today’s standards that a typical representation of Figure 1 can be provided by the pump vendor. Instrumentation should be employed whenever possible in the field to

11

FIGURE 1

Increasing NPSH available versus percent of pump differential head.

measure suction pressure, temperature, flow, fluid characteristics, and barometric pressure. This information can be applied to the following equation: NPSHA = PB – Vp ± GR + HV where: PB = barometric pressure converted to feet Vp = fluid vapor pressure converted to feet GR = gauge reading at pump suction converted to feet HV = velocity head = V2/2g V = fluid velocity at pump suction flange in feet/sec

(32.174 ft/sec2) Plotting the NPSH measured in the system on Figure 1 will quickly assess the health of your installation. As noted above, some hardware adjustment within the machine may help. Sometimes radical changes must be made, such as adding an inducer to the pump or changing manufacturers. System changes to lower the pump, raise the tank, reduce process temperature, increase line size, shorten the suction run, or minimize elbows and restrictions may easily prove to be more expensive. Misunderstanding NPSH can be a costly error. There simply must be enough NPSH available for any

g = gravitational constant

12

The Pump Handbook Series

pump installation to operate reliably. ■ John E. Sidelko is Product Development Manager for A.R. Wilfley and Sons, Inc. He serves on the Pumps and Systems Editorial Advisory Board.

?

PUMP AND SYSTEM TROUBLESHOOTING HANDBOOK

Variable Speed Drives: Advantages and Pitfalls BY DAVID WILLIAM SPITZER

A

variable speed drive is a device that changes the speed of a driven load, typically by using electrical, mechanical, or hydraulic means (Figure 1). Considering the wide range of overlap in their application, it should be no surprise that the use of variable speed drives often falls between the cracks of the traditional technical disciplines.

A VARIABLE SPEED DRIVE: 1.

may be an electrical device specified and installed by an electrical engineer, such as an electrical switch gear

2.

may be a mechanical device specified and installed by a mechanical engineer, such as a gear reducer or hydraulic coupling

3.

may be a final control element, such as a replacement for a control valve that can affect the control of the process, an area important to the instrumentation and control engineer

4.

may affect the electrical power distribution system by reducing electrical demand

5.

may affect the operation and proper functioning of the mechanical equipment by operating slower, and potentially reducing maintenance

6.

may affect the operation of the process and be the key component required to improve the process because desired operation cannot be achieved with a control valve

Knowledgeable generalists who appropriately rely on specialists (such as electrical, mechanical, process, instrumentation, and process control engineers) are in the best position to apply variable speed drive technology - and they are most likely to come up with viable, technically correct solutions. With the increased trend toward specialization of job functions, however, people with broad knowl-

of flow as compared to a control edge and appropriate skills are bevalve installation, in that it virtually coming more difficult to find. eliminates valve-related hysteresis. The primary advantage of variSuch an installation may significantly able speed drive technology is the economy gained by reducing electrical energy consumption; howFIGURE 1 A & B ever, other advantages that A. VSD application with variable motor speed are difficult to quantify may Variable be achieved in Speed most applicaDrive tions. These Motor and include reducEquipment tions in mainteSpeed Varies nance requirements as a result Mechanical Motor of fewer rotaEquipment tions, reduced forces on internal components and seals, and B. VSD application with constant motor speed improved overConstant Equipment all performance Motor Speed as a result of Speed Varies tighter control of Variable the process. For Mechanical Motor Speed example, a variEquipment Drive able speed drive installation can improve control

FIGURE 2 Full speed operating point Full speed pump curve

Pump curve at reduced speed

System Pressure

Increasing shaft horsepower Flow

The Pump Handbook Series

Reduced speed operating point

13

improve product quality and process efficiency.

FIGURE 3 A. Typical full-speed pump flow control

CONSTANT TORQUE MECHANICAL LOADS Constant torque loads, such as those found in positive displacement equipment, can be analyzed in a relatively straightforward manner because they require nearly constant torque input as speed is varied. As such, shaft horsepower requirements, equipment capacity, and electrical energy consumption vary directly with speed. For example, if a piece of equipment must produce 75% of its capacity (at rated speed), it will operate at 75% of rated speed, require approximately full torque, and consume approximately 75% of its electrical requirements at rated speed.

CENTRIFUGAL MECHANICAL LOADS Centrifugal pumps and fans follow the affinity laws, which are nonlinear, adding complexity to the economic analysis of any given application. Graphical representations of variable speed drive operation clearly show a reduction in shaft horsepower (which will reduce electrical energy consumption) while providing the required hydraulic energy output (Figure 2). Economic analysis is beyond the scope of this article (see reference). However, it is important to recognize that the brake horsepower of centrifugal loads varies with the cube of the speed. So, reducing speed by only 10% (to 90% of rated speed) will reduce the brake horsepower by 27%: [100(1 – 0.93) ]

FIC

FIC

14

Flow

Flow

the output from one piece of equipment feeds multiple users because more than one flow cannot be independently controlled by only one final control element. In addition, care should be taken when attempting to manipulate the operating speeds of equipment working in parallel to ensure that each piece of equipment supplies an appropriate part of the total load. Whereas many variable speed drive applications may replace a control valve or damper, the best applications improve the process as well. In one such application, piping leaks that were caused by full-speed operation because the pump produced too much head were minimized by open-

ing the bypass valve to the feed tank, causing the pump to ride out on its curve, effectively reducing discharge pressure (Figure 4A). While this approach ”worked,“ it also increased operating costs significantly because the increased recirculating flow (to the tank) produced no benefit to the process. A variable speed drive was installed on the centrifugal pump to control the header pressure, allowing the bypass valve to be closed (Figure 4B). This effectively eliminated the bypass flow and reduced energy consumption by 73%. In addition, tight control of the discharge pressure was achieved under varying operating conditions that resulted in more uniform flow to process equipment. It

FIGURE 4 A & B A. Full-speed operation Throttled to reduce supply pressure Supply to Plant

Return

P

Q B. Variable-speed operation Closed

TYPICAL APPLICATIONS Variable speed drives typically alter equipment speed to provide the rotational energy input necessary to supply the hydraulic energy output to the process. In most situations equipment speed is varied to control level, pressure, or flow (Figure 3). With the exception of header pressure control, variable speed drives are generally not applicable when

B. Typical variable-speed pump flow control

Supply to Plant

Return PIC

P Q The Pump Handbook Series

should be noted that variable speed drive applications that replace control valves typically reduce electrical energy consumption by 30–70%. However, these electrical energy savings may be small compared to process-related savings in applications where the variable speed drive improves the process.

POTENTIAL PITFALLS The potential pitfalls of applying variable speed drives to rotating equipment include the following: 1. Constant torque variable frequency drives must be able to provide full motor current at all speeds. To avoid premature failure, the electrical equipment operating these loads should be specified, designed, and rated for constant-torque operation. 2. In larger applications, electrical equipment should be suitably designed to be integrated into the electrical system to avoid problems such as electrical noise, ringing, and harmonic distortion. 3. The motor must be capable of dissipating the heat it generates at all speeds, including low speeds, when the cooling fan also operates slowly. Solutions to this problem include appropriately limiting the motor operating speed range, upgrading the motor winding insulation, oversizing the motor to provide more heat dissipation, utilizing a totally enclosed non-ventilated (TENV) motor, or installing auxiliary motor cooling to dissipate heat. This problem is especially prevalent in constant torque applications where full-load current flows, even though the effectiveness of the motor cooling fan is reduced or nonexistent. 4. In Division 1 hazardous locations, the National Electric Code

5.

6.

7.

(NEC) requires that the motor be approved for the hazard. While full-speed motors are approved for full-speed operation at maximum load (and maximum fan cooling), variable speed drive motors must be approved over the entire operating speed range. In centrifugal applications, the worst case is usually at full speed, a condition for which the motor was designed. However, in constant torque applications, maximum heat dissipation is required at all speeds. Approved motors for each type of variable speed drive load are becoming available to meet these requirements, utilizing appropriate heat dissipation techniques described in item 3 (as required). The mechanical equipment must be capable of operating at reduced speeds. For example, a liquid-ring vacuum pump is a questionable candidate for variable speed drive operation because it fails to produce a vacuum below approximately 80% of its rated speed. The mechanical equipment must operate without damage at reduced speeds. Some equipment can be damaged by inefficiencies and slip at slower speeds that cause overheating. Equipment can also be damaged when its lubrication system becomes inadequate at reduced speeds. In most installations variable speed drives cannot securely stop the process fluid—that is, provide tight shutoff. When tight shutoff is required for process reasons, a control valve can be installed in addition to the variable speed drive. In certain applications, a check valve may be required to prevent reverse flow conditions that can compromise safety or upset the process.

The Pump Handbook Series

8.

9.

Consideration must be given to the fact that a mechanical variable speed drive is an additional piece of equipment with moving parts subject to the wear and tear, heat and maintenance needs of industrial machinery. Hydraulic variable speed drives involve two energy transformations (rotating to hydraulic and hydraulic to rotating) that may add to electrical energy costs. In addition, leaks and other maintenance headaches will sometimes occur, particularly in high-pressure situations.

SUMMARY This discussion, by no means complete (Ref. 1), should give readers a basic understanding of some of the advantages and pitfalls associated with a technology that can be the key to significant process improvements and energy savings. While there are a number of pitfalls, some of which are outlined, the reader should not be intimidated because the great majority of problems can be solved by utilizing the knowledge and skill of specialists. An individual familiar with the mechanical nature of the equipment is in a unique position to initiate and support the application of variable speed drive technology when equipment speed can be safely and effectively reduced. ■

REFERENCE D.W. Spitzer. Variable Speed Drives—Principles and Applications for Energy Cost Savings, 2nd Edition Revised, Instrument Society of America, 1990. David William Spitzer, P.E., is Manager of Utility and Instrumentation Engineering for Nepera, Inc. in Harriman, NY.

15

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PUMP AND SYSTEM TROUBLESHOOTING HANDBOOK

Estimating Maximum Head in Single – and Multi-Stage Pump Systems BY JAMES NETZEL

The maximum head or discharge pressure of a centrifugal pump can be easily estimated if the impeller diameter, number of impellers used, and rpm of the driver are known.

Q: A:

How can you estimate the maximum (shutoff) head that a centrifugal pump can deliver?

The maximum pressure a centrifugal pump delivers should be known in order to ensure that a piping system is adequately designed. Any pump that operates at a high flow rate could deliver significantly more pressure at zero (0) gpm flow, such as when the discharge valve is closed, than it delivers at operating flow. The maximum head or discharge pressure of a centrifugal pump, which usually occurs at shutoff con-

FIGURE 1

17 16 14

Head in Feet x 1000

13 12 11 10 9 8 7 6 5 4 3 2 1 1 2 3 4 5 6 7 8 9 10 11 12 13 14 15

RPM x 1000 Rotations per minute (rpm) vs. head in feet to estimate maximum head

16

The Pump Handbook Series

ditions (0 gpm), can be easily estimated if the impeller diameter, number of impellers used, and rpm of the driver (electric motor, gas engine, turbine, etc.) are known. Let’s say we have a singlestage pump with a 10-in. diameter impeller and an 1,800 rpm driver. To determine the head in feet, simply take the impeller diameter in inches and square it. Our 10-in. impeller at 1,800 rpm would yield 102, or 100 ft of head. An 8-in. impeller would yield 82, or 64 ft of head, while a 12-in. impeller would yield 122, or 144 ft of head. Now let’s assume that our 10-in. diameter impeller is driven by a 3,600 rpm motor. We first determine the head at 1,800 rpm, but then multiply this value by a factor of four. The basic rule is that every time the rpm changes by a factor of two, the head changes by a factor of four. The head at 3,600 rpm for our 10-in. impeller is therefore 102 x 4, or 400 ft of head. Our 8-in. impeller at 3600 rpm would give us 82 x 4, or 256 ft of head, and our 12-in. impeller would give us 122 x 4, or 576 ft of head. For multiple stages (more than one impeller), simply multiply the final head for one impeller by the total number of impellers in the pump. For a pump with three 10-in. impellers and a speed of 3,600 rpm, we get (102 x 4) x 3 = 400 x 3 = 1,200 ft. of head. Now what happens if we reduce the speed below 1,800 rpm? The same rule still applies: a change in speed by a factor of two changes the head by a factor of four. Therefore, a 10-in. diameter impeller spinning at 900 rpm delivers only one fourth the head it would at 1,800 rpm: 102/4 = 25 ft. Plotting several head-versusrpm points on a curve will allow the user to estimate the maximum

head at any given speed. Let’s say we have a turbine-driven pump that injects water into the ground to raise the subterranean oil reserves to the surface for processing. The vendor tells you that the maximum head is classified, but you have been requested to resolve system problems that you believe are pressure related. The vendor tells you that the pump has four 8-in. diameter impellers and is driven by the turbine at 13,000 rpm. You would estimate the maximum head as follows: Step 1 Determine the head at 1,800 rpm: 82 x 4 stages = 256 ft Step 2 Multiply the head at 1,800 rpm by four to get the head at 3,600 rpm: 256 x 4 = 1,024 ft Step 3 Multiply the head at 3,600 rpm by 4 to get the head at 7,200 rpm: 1,024 x 4 = 4,096 ft Step 4 Multiply the head at 7,200 rpm by 4 to get the head at 14,400 rpm: 4,096 x 4 = 16,384 ft Step 5 Plot the rpm-versus-head points to obtain the curve shown in Figure 1. As you can see, the estimated head at 13,000 rpm is 12,500 ft. To convert head in feet to psi, simply divide the head by 2.31 to get 5,411 psi. Ray W. Rhoe, PE, has a BSCE from The Citadel and 15 years’ experience with pumps, testing, and hydraulic design.

elastically under contact pressure. This deformation creates larger film areas and very thin films. Such lubrication systems are normally used to control wear in rolling element bearings. In seals A mechanical seal is where the viscosity of designed to the fluid sealed increasoperate in many es with increasing prestypes of fluids. sure, elastohydrodynaThe product A mic lubrication occurs. sealed becomes the lubriBoundary lubricacant for the seal faces. mechanical tion is important for Many times the fluid seal faces that are movbeing sealed is a poor seal is ing very slowly under lubricant or contains abraheavy load. Here, sives that must be taken hydrodynamic and elasinto account in the seal designed tohydrodynamic lubridesign. The design of the cant pressures are seal faces, materials of to operate insufficient to separate construction, and seal the seal faces. The slidlubrication play an imporin many ing surfaces are protecttant role in successful ed by the tribological operation. Achieving a types of properties of the materihigh level of reliability als of construction. An and service life is a clasfluids. example of a seal operatsic problem in the field of ing within this lubricatribology, the study of tion system is a dryfriction, wear, and lubrirunning agitator seal. cation. Mixed-film lubrication, a comThe lubrication system for two bination of all the previous syssliding seal faces can be classified as tems discussed, occurs in all follows: 1) hydrodynamic, 2) elastocontact seals. Here the fluid film hydrodynamic, 3) boundary, and 4) becomes very thin and is a combimixed film. nation of both the liquid and the Hydrodynamic conditions exist gas phases of the fluid sealed. when the fluid film completely sepaAsperities from one surface may rates the seal faces. Direct surface penetrate the lubricating film and contact between seal faces does not contact the opposite surface. The take place, so there is no wear, and seal face load is then supported heat generation from friction is zero. partially by the fluid film and parThe only heat generation occurs tially by solid contact. If the generfrom shearing of the fluid film, ated head at the seal faces is not which is extremely small. A hydroremoved, surface wear and damdynamic seal may rely on design feaage can occur. For applications tures such as balance factors, surface where the seal face load is too waviness, or spiral grooves to sepahigh or the fluid viscosity is too rate the seal faces. The Society of low, designs of seal faces can be Tribologists and Lubrication changed through balance and face Engineers (STLE) guideline in geometry to improve seal perfor“Meeting Emissions Regulations mance. ■ with Mechanical Seals” lists hydrodynamic seals as a technology to James Netzel is Chief Engineer control emissions. at John Crane Inc. He serves on the Elastohydrodynamic lubrication Editorial Advisory Board for Pumps (EHD) is also found in sliding surand Systems. faces, but more often this involves rolling surfaces separated by an oil film. Here the moving surfaces form an interface region that deforms

Q: A:

What different types of seal lubrication exist?

The Pump Handbook Series

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PUMP AND SYSTEM TROUBLESHOOTING HANDBOOK

Elements of Minimum Flow BY TERRY M. WOLD

M

inimum flow can be determined by examining each of the factors that affect it. There are five elements that can be quantified and evaluated: 1.

Temperature rise (minimum thermal flow)

2.

Minimum stable flow

3.

Thrust capacity

4.

NPSH requirements

5.

Recirculation

mechanical handbooks. What is the maximum allowable temperature rise? Pump manufacturers usually limit it to 15°F. However, this can be disastrous in certain situations. A comparison of the vapor pressure to the lowest expected suction pressure plus NPSH required (NPSHR) by the pump must be made. The temperature where the vapor pressure equals the suction pressure plus the NPSHR is the maximum

allowable temperature. The difference between the allowable temperature and the temperature at the pump inlet is the maximum allowable temperature rise. Knowing ∆T and Cp, the minimum flow can be determined by finding the corresponding head and efficiency. When calculating the maximum allowable temperature rise, look at the pump geometry. For instance, examine the vertical can

FIGURE 1

The highest flow calculated using these parameters is considered the minimum flow.

TEMPERATURE RISE Temperature rise comes from energy imparted to the liquid through hydraulic and mechanical losses within the pump. These losses are converted to heat, which can be assumed to be entirely absorbed by the liquid pumped. Based on this assumption, temperature rise ∆T in °F is expressed as:

H

1

778 x Cp

η–1

SUCTION Low Pressure Lower Temperature

DISCHARGE High Pressure Higher Temperature

∆T = ————— x ——————

where H = total head in feet Cp = specific heat of the liquid, Btu/lb x °F η = pump efficiency in decimal form 778 ft–lbs = energy to raise the temperature of one pound of water 1°F To calculate this, the specific heat and allowable temperature rise must be known. The specific heat for water is 1.0, and other specific heats can be as low as 0.5. The specific heats for a number of liquids can be found in many chemical and

18

A high-pressure vertical pump. Asterisks (*) denote where lowtemperature fluid is exposed to higher temperatures. Flashing and vaporization can occur here. Temperature increases as fluid travels from A towards B. The Pump Handbook Series

pump in Figure 1. Although pressure increases as the fluid is pumped upward through the stages, consider the pump inlet. The fluid at the inlet (low pressure, low temperature) is exposed to the temperature of the fluid in the discharge riser in the head (higher pressure, higher temperature). This means that the vapor pressure of the fluid at the pump inlet must be high enough to accommodate the total temperature rise through all the stages. If this condition is discovered during the pump design phase, a thermal barrier can be designed to reduce the temperature that the inlet fluid is exposed to. Some books, such as the Pump Handbook (Ref. 5), contain a typical chart based on water (Cp = 1.0) that can be used with the manufacturer’s performance curve to determine temperature rise. If the maximum allowable temperature rise exceeds the previously determined allowable temperature rise, a heat shield can be designed and fitted to the pump during the design stage. This requirement must be recognized during the design stage, because once the pump is built, options for retrofitting the pump with a heat shield are greatly reduced.

MINIMUM STABLE FLOW Minimum stable flow can be defined as the flow corresponding to the head that equals shutoff head. In other words, outside the ”droop“ in the head capacity curve. In general, pumps with a specific speed less than 1,000 that are designed for optimum efficiency have a drooping curve. Getting rid of this ”hump“ requires an impeller redesign; however, note that there will be a loss of efficiency and an increase in NPSHR. What’s wrong with a drooping head/capacity curve? A drooping curve has corresponding heads for two different flows. The pump reacts to the system requirements, and there are two flows where the pump can meet the system requirements. As a result, it ”hunts“ or ”shuttles“ between

these two flows. This can damage the pump and other equipment, but it will happen only under certain circumstances: 1.

The liquid pumped must be uninhibited at both the suction and discharge vessels.

2.

One element in the system must be able to store and return energy, i.e., a water column or trapped gas.

3.

Something must upset the system to make it start hunting, i.e., starting another pump in parallel or throttling a valve.

FIGURE 2 A

Note: All of these must B be present at the same time to cause the pump to hunt. Minimum flow based on the shape of the performance curve is not so much a function of the pump as it is a function of the system where the pump is placed. A pump in a system where the above Recirculation zones are always on the criteria are present should pressure side of the vane. A shows disnot have a drooping curve charge recirculation (the front shroud in the zone of operation. has been left out for clarity). B shows Because pumps with a inlet recirculation. drooping head/capacity curve have higher efficiency bearings can be sized to handle and a lower operating cost, it would the thrust. Thrust can be balanced seem prudent to investigate the instalby the use of balanced and unballation of a minimum flow bypass. anced stages or adding a balance THRUST LOADING drum, if necessary. These techAxial thrust in a vertical turbine niques for thrust balancing are pump increases rapidly as flows are used when high thrust motors are reduced and head increased. Based on not available. It is worth noting the limitations of the driver bearings, that balanced stages incorporate flow must be maintained at a value wear rings and balance holes to where thrust developed by the pump achieve lower thrust; therefore, a does not impair bearing life. Find out slight reduction in pump efficienwhat your bearing life is and ask the cy can be expected, and energy pump manufacturer to give specific costs become a factor. thrust values based on actual tests. NPSH REQUIREMENTS If a problem exists that cannot be How many pumps have been handled by the driver bearings, conoversized because of NPSH availtact the pump manufacturer. There able (NPSHA)? It seems the easiest are many designs available today for solution to an NPSH problem is to vertical pumps (both single and mulgo to the next size pump with a tistage) with integral bearings. These The Pump Handbook Series

19

FIGURE 3 B2

R2

D2 B1

R1

D1 h1

.14 .12

Cm2 .10 U2 .08

Ve U1

.06 .04 .02 10

15

20

25

30

Discharge Angle β2

.32 .30 .28 .26 .24 .22 .20 .18 .16 .14 .12 .10 .08 10

15

20

25

30

Inlet Angle β1

35

40

Incipient recirculation. Minimum flow is approximately 50% of incipient flow, while minimum intermittent flow is approximately 25% of incipient flow. See text under “Recirculation Calculations” for details

larger suction, thereby reducing the inlet losses. A couple of factors become entangled when this is done. A larger pump means operating back on the pump curve. Minimum flow must be considered. Is the curve stable? What about temperature rise? If there is already an NPSH problem, an extra few degrees of temperature rise will not help the situation. The thrust and eye diameter will increase, possibly causing damaging recirculation. When trying to solve an NPSH problem, don’t take the easiest way out. Look at other options that may solve a long-term problem and reduce operating costs.

RECIRCULATION Every pump has a point where recirculation begins. But if

20

this is the case, why don’t more pumps have problems? Recirculation is caused by oversized flow channels that allow liquid to turn around or reverse flow while pumping is going on (Figure 2 shows recirculation zones). This reversal causes a vortex that attaches itself to the pressure side of the vane. If there is enough energy available and the velocities are high enough, damage will occur. Suction recirculation is reduced by lowering the peripheral velocity, which in turn increases NPSH. To avoid this it is better to recognize the problem in the design stage and opt for a lower-speed pump, two smaller pumps, or an increase in NPSHA. Discharge recirculation is caused by flow reversal and high velocities producing damaging vortices on the pressure side of the The Pump Handbook Series

vane at the outlet (Figure 2). The solution to this problem lies in the impeller design. The problem is the result of a mismatched case and impeller, too little vane overlap in the impeller design, or trimming the impeller below the minimum diameter for which it was designed. Recirculation is one of the most difficult problems to understand and document. Many studies on the topic have been done over the years. Mr. Fraser’s paper (Ref. 1) is one of the most useful tools for determining where recirculation begins. In it he describes how to calculate the inception of recirculation based on specific design characteristics of the impeller and he includes charts that can be used with a minimum amount of information. An example of Fraser calculations, which show the requirements to calculate the inception of suction and discharge recirculation, is shown in Figure 3.

RECIRCULATION CALCULATIONS Figure 3 indicates the userdefined variables and charts required to make the Fraser calculations for minimum flow. Information to do the detailed calculations include: Q = capacity at the best efficiency point H = head at the best efficiency point NPSHR = net positive suction head required at the pump suction N = pump speed NS = pump specific speed NSS = suction specific speed Z = number of impeller vanes h1 = hub diameter (h1 = 0 for single suction pumps) D1 = impeller eye diameter D2 = impeller outside diameter B1 = impeller inlet width B2 = impeller outlet width R1 = impeller inlet radius R2 = impeller outlet radius F1 = impeller inlet area F2 = impeller outlet area β1 = impeller inlet angle β2 = impeller outlet angle

The above information is obtained from the pump manufacturer curves or impeller design files. The impeller design values are usually considered proprietary information. KVe and KCm2 can be determined from the charts in Figure 3. With all of the above information at hand, suction recirculation and the two modes of discharge recirculation can be determined. As previously mentioned, Fraser has some empirical charts at the end of his paper that can be used to estimate the minimum flow for recirculation. A few of the design factors of the impeller are still required. It is best to discuss recirculation with your pump manufacturer before purchasing a pump, in order to reduce the possibility of problems with your pump and system after installation and start-up.

pumps. Chemical Engineering (1984).

SUMMARY Minimum flow can be accurately determined if the elements described above are reviewed by the user and the manufacturer. Neither has all the information to determine a minimum flow that is economical, efficient, and insures a trouble-free pump life. It takes a coordinated effort by the user and the manufacturer to come up with an optimum system for pump selection, design, and installation.

REFERENCES 1.

F.H. Fraser. Recirculation in centrifugal pumps. Presented at the ASME Winter Annual Meeting (1981).

2.

A.R. Budris. Sorting out flow recirculation problems. Machine Design (1989).

3.

J.J. Paugh. Head-vs-capacity characteristics of centrifugal

The Pump Handbook Series

4.

I. Taylor. NPSH still pump application problem. The Oil and Gas Journal (1978).

5.

I.J. Karassik. Pump Handbook. McGraw-Hill (1986). ■

Terry Wold has been the engineering manager for Afton Pumps for the last four years. He has been involved in pump design for 25 years. Mr. Wold graduated from Lamar Tech in 1968 with a bachelor’s degree in mechanical engineering and is currently a registered engineer in the State of Texas. Thanks to P.J. Patel for his comments and assistance in preparing the graphics.

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PUMP AND SYSTEM TROUBLESHOOTING HANDBOOK

NPSH Required for Reciprocating PD Pump BY JIM MILLER Many times I am asked to predict the flow of a positive displacement pump operating under conditions with insufficient inlet pressure to ensure complete filling. In these circumstances the mixture may be considered two-phase, with either fluid vapor or some other entrained gas (air) present. Is there any literature dealing with the theory necessary to evaluate the filling capability?

Q:

Adequate suction pressure for a reciprocating positive displacement pump is an often misunderstood concept. Pump manufacturers are asked the minimum suction pressure that can be used for both continuous and intermittent service for a particular pump. The net positive suction head (NPSH) required may or may not be provided. NPSH required takes into account the fluid properties of water (usually), pump dynamics, and pump design. When analyzing a pump’s operation at very low suction pressure, several pump parameters must be studied.

A:

NPSH REQUIRED The subject of NPSH for such pumps is also frequently misunderstood. The definition in the Hydraulic Institute Standards of minimum NPSH required is established for a specific pump, pump speed, and discharge pressure by reducing the suction pressure until one of the following conditions occur: 1.

a 3% drop in capacity

2.

clearly audible cavitation noise

Several things are wrong with operating a pump at the NPSH required suction pressure. By definition, the pump is cavitating when condition 2 occurs. Loss of capacity according to condition 1 means that

22

the liquid chamber is only partially filling as the result of cavitation. NPSH required accounts for 1) Fluid properties a) vapor pressures

2)

∆volumeplunger = ∆gas volume + ∆liquid volume ∆gas volume = v1 - v2 p = v1 - __1 v1 p2 p1v1 _______ p1 + ∆p

b) gas saturation pressure

= v1 -

c) fluid compressibility

∆liquid volume = βυ0∆p

Mechanical design

∆volumeplunger =

a) chamber volume b) valve design

3)

p1v1 _______ + βυ0∆p p1 + ∆p

= v1 -

c) piping system

Solve for ∆p, the pressure

Operating conditions a) pump speed

(βυ0)∆p2 + (βυ0p1 + v1 ∆volumeplunger) ∆p + (-∆volumeplungerp1) = 0

b) discharge pressure c) fluid temperature The meaningfulness of the NPSH required will become questionable if any of these parameters change significantly for a given application.

PUMP LIQUID CHAMBER It is difficult to predict whether a pump will operate with some degree of vapor or gas breakout. Given a known volume and pressure of the gas or vapor escape, the solution is relatively easy to solve. The difficult issue is the amount of gas initially formed in the pump chamber in the first place. For the pump to continue to operate, the cylinder pressure must exceed the discharge manifold pressure at the end of the plunger’s discharge stroke. The following approach has been used to solve for pressure at the end of the discharge stroke: Given: ∆volumeplunger = plunger displacement The Pump Handbook Series

a = βυ0 b = βυ0p1 + v1 - ∆volumeplunger c = -∆volumeplungerp1 ∆p =

-b ± √(b2 - 4ac) ________________ 2a

For a typical 2-in. plunger by 6-in. stroke pump the following values have been calculated where: ∆volumeplunger = 18.85 in.3 β = 0.000003/psi for water liquid chamber volume = 100 in.3 υ0 = 100 (1 - gas fraction), (in.3) v1 = 100 x gas fraction (in.3) p1 = 30 psi As can be seen in Figure 1, this pump operating at 1,000 psi discharge pressure will become vapor locked when the gas volume is equal to 19% of the chamber volume. The graph suggests that a reciprocating pump will operate at very low volu-

CONCLUSIONS Minimum NPSH required suction pressure should not be used for a continuous-duty pump application. As general practice, suction pressure should be 10 psi above the minimum NPSH required for the pump. This additional pressure will significantly reduce the potential for pump cavitation problems such as damaged plungers, valves, and pump liquid ends. The minimum NPSH required could be used as a conservative suction pressure to assure that the pump will continue to operate in an intermittent service.

FIGURE 1 100000

Maximum Cylinder Pressure - psig

metric efficiencies, but in fact it will stop pumping when it falls to the 7585% volumetric efficiency range. This is probably because when the gas or vapor phase develops, it very quickly exceeds the 19% volume of the liquid chamber.

10000

1000

100

10 0 0.05 Liquid Chamber Gas Fraction

0.1

0.15

0.2

0.25

0.3

Maximum cylinder pressure versus liquid chamber gas fraction.

Jim Miller is president of White Rock Engineering in Dallas, TX. He has degrees in chemical engineering and business administration from the University of Texas at Austin.

The Pump Handbook Series

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PUMP AND SYSTEM TROUBLESHOOTING HANDBOOK

Pump Suction Conditions BY ROSS C. MACKAY f a wide receiver has the right a function of the system design on speed and good hands, all that’s the suction side of the pump. needed from the quarterback is Consequently, it is in the control of to throw the ball accurately, the system designer. and the team will probably gain To avoid cavitation, the NPSH good yardage, maybe even a available from the system must be touchdown. greater than the NPSH required by Believe it or not, much the the pump, and the biggest mistake same is true of a pump and its sucthat can be made by a system designtion conditions. If it has the right er is to succumb to the temptation to speed and is the right size, all provide only the minimum required that’s required from the quarterat the rated design point. This leaves back is to deliver the liquid at the no margin for error on the part of the right pressure and with an even designer, or the pump, or the system. laminar flow into the eye of the Giving in to this temptation has impeller. proved to be a costly mistake on If the quarterback’s pass is off many occasions. target, badly timed, or the ball’s In the simple system as shown turning end over end in the air, in Figure 1, the NPSH Available can the receiver may not be able to be calculated as follows: hang on to it, and there’s no gain on the play. When that happens, the quarterback FIGURE 1 knows he didn’t throw it properly and doesn’t blame Ha the receiver. Unfortunately, that’s where the comparison ends. The engineering Hvp ”quarterbacks” tend to Hs blame the pump even when its their delivery that’s bad! Hf Just as there are techniques a quarterback must learn in order to throw accurately, there are rules which ensure that a liquid arrives at the impeller eye with NPSHA = Ha + Hs - Hvp - Hf the pressure and flow characteristics needed for reliable operation. where RULE #1. Ha= the head on the surface of the PROVIDE SUFFICIENT NPSH liquid in the tank. In an open Without getting too complicatsystem like this, it will be ed on a subject about which comatmospheric pressure. plete books have been written, Hs= the vertical distance of the let’s just accept the premise that free surface of the liquid every impeller requires a miniabove the center line of the mum amount of pressure energy pump impeller. If the liquid is in the liquid being supplied in below the pump, this order to perform without cavitabecomes a negative value. tion difficulties. This pressure Hvp= the vapor pressure of the liqenergy is referred to as Net uid at the pumping temperaPositive Suction Head Required. ture, expressed in feet of The NPSH Available is suphead. plied from the system. It is solely

I

24

The Pump Handbook Series

Hf=

the friction losses in the suction piping. The NPSH Available may also be determined with this equation:

NPSHA= Ha + Hg + V2/2g - Hvp where Ha= atmospheric pressure in feet of head. Hg= the gauge pressure at the suction flange in feet of head. V2= The velocity head at the 2g point of measurement of Hg. (Gauge readings do not include velocity head.)

RULE #2. REDUCE THE FRICTION LOSSES When a pump is taking its suction from a tank, it should be located as close to the tank as possible in order to reduce the effect of friction losses on the NPSH Available. Yet the pump must be far enough away from the tank to ensure that correct piping practice can be followed. Pipe friction can usually be reduced by using a larger diameter line to limit the linear velocity to a level appropriate to the particular liquid being pumped. Many industries work with a maximum velocity of about 5ft./sec., but this is not always acceptable.

RULE #3. NO ELBOWS ON THE SUCTION FLANGE Much discussion has taken place over the acceptable configuration of an elbow on the suction flange of a pump. Let’s simplify it. There isn’t one! There is always an uneven flow in an elbow, and when one is installed on the suction of any pump, it introduces that uneven flow into the eye of the impeller. This can create turbulence and air

entrainment, which may result in impeller damage and vibration. When the elbow is installed in a horizontal plane on the inlet of a double suction pump, uneven flows are introduced into the opposing eyes of the impeller, upsetting the hydraulic balance of the rotating element. Under these conditions the overloaded bearing will fail prematurely and regularly if the pump is packed. If the pump is fitted with mechanical seals, the seal will usually fail instead of the bearing-but just as regularly and often more frequently. The only thing worse than one elbow on the suction of a pump is two elbows on the suction of a pump— particularly if they are positioned in planes at right angles to each other. This creates a spinning effect in the liquid which is carried into the impeller and causes turbulence, inefficiency and vibration. A well established and effective method of ensuring a laminar flow to the eye of the impeller is to provide the suction

FIGURE 2

of the pump with a straight run of pipe in a length equivalent to 5-10 times the diameter of that pipe. The smaller multiplier would be used on the larger pipe diameters and vice versa.

FIGURE 3 Air Pocket

RULE #4. STOP AIR OR VAPOR ENTERING THE SUCTION LINE Any high spot in the suction line It is worthwhile noting that can become filled with air or vapor these vortices are more difficult which, if transported into the to troubleshoot in a closed tank impeller, will create an effect simisimply because they can’t be lar to cavitation and with the same seen as easily. results. Services which are particuGreat care should be taken larly susceptible to this situation are in designing a sump to ensure those where the pumpage contains that any liquid emptying into it a significant amount of entrained does so in such a way that air air or vapor, as well as those operentrained in the inflow does not ating on a suction lift, where it can pass into the suction opening. also cause the pump to lose its prime. (Figure 3) A similar effect can be FIGURE 4 caused by a concentric reducer. The suction of a pump should be fitted with an eccentric reducer positioned with the flat side uppermost. (Figure 4). If a pump is taking its Any problem of this nature may suction from a sump require a change in the relative or tank, the formation positions of the inflow and outlet of vortices can draw if the sump is large enough, or air into the suction the use of baffles. (Figure 5) line. This can usually be prevented by proRULE #5. viding sufficient subCORRECT PIPING ALIGNMENT mergence of liquid Piping flanges must be accuover the suction openrately aligned before the bolts ing. A bell-mouth design Suction are tightened and all piping, on the opening will valves and associated fittings reduce the amount of should be independently supsubmergence required. ported, so as to place no strain This submergence is on the pump. Stress imposed on completely independent the pump casing by the piping of the NPSH required by reduces the probability of satisthe pump. factory performance. The Pump Handbook Series

25

ing strains are simply passed through to the pump.

FIGURE 5 Inflow

To Pump Suction

Inflow

To Pump Suction

Under certain conditions the pump manufacturer may identify some maximum levels of forces and moments which may be acceptable on the pump flanges. In high temperature applications, some piping misalignment is inevitable owing to thermal growth during the operating cycle. Under these conditions, thermal expansion joints are often introduced to avoid transmitting piping strains to the pump. However, if the end of the expansion joint closest to the pump is not anchored securely, the object of the exercise is defeated as the pip-

26

Baffle

RULE #6. WHEN RULES 1 TO 5 HAVE BEEN IGNORED, FOLLOW RULES 1 TO 5.

Piping design is one area where the basic principles involved are regularly ignored, resulting in hydraulic instabilities in the impeller which translate into additional shaft loading, higher vibration levels and premature failure of the seal or bearings. Because there are many other reasons why pumps could vibrate, and why seals and bearings fail, the trouble is rarely traced to incorrect piping. It has been argued that because many pumps are piped incorrectly and most of them are operating quite satisfactorily, piping procedure is not important. Unfortunately, satisfactory operation is a relative term, and what may be acceptable in one

The Pump Handbook Series

plant may be inappropriate in another. Even when ”satisfactory” pump operation is obtained, that doesn’t automatically make a questionable piping practice correct. It merely makes it lucky. The suction side of a pump is much more important than the piping on the discharge. If any mistakes are made on the discharge side, they can usually be compensated for by increasing the performance capability from the pump. Problems on the suction side, however, can be the source of ongoing and expensive difficulties which may never be traced back to that area. In other words, if your receivers aren’t performing well, is it their fault? Or does the quarterback need more training? ■

Ross C. Mackay is an independent consultant who specializes in advanced technology training for pump maintenance cost reduction. He also serves on the editorial advisory board for Pumps and Systems.

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PUMP AND SYSTEM TROUBLESHOOTING HANDBOOK

Where’s the Prime? BY ROBERT KREBS, CONTRIBUTING EDITOR professional engineer came to the Pumps and Systems booth at the Water Environment Federation Exposition in Anaheim last October. He asked to talk to someone regarding a pumping problem. I happened to be there and volunteered. Here is his story. Two pumps are installed, as shown in the Figure 1 sketch. The pumps are taking suction from a constant level water source. There is minimal velocity (much less than 1 ft/sec) in the stream flowing by the pump suction line. The foot valve in the suction line prevents flow from the pump to the source on pump shutdown and on initial priming when the suction line from the pumps to the source are manually filled with liquid. The pumps alternated in service and the ”off“ pump lost prime and would not start when activated. The engineer surmised that the ”off“ pump’s suction line was being partially dewatered by the ”on“ pump and would not prime on the next start. The check valves in the suction lines were added to the system. Periodically, the pumps still lost prime and would not function. The application sketch (Figure 1) was prepared and discussed with the engineer. The suction line size and the pump selection were satisfactory for the flow and pressure required. The system had been checked for suction side leaks. The shaft sealing system was so designed that air was not entering the pump casing through the stuffing box. In reviewing the installation as it was originally designed (before check valves were added in the suction pipes), I figured that the foot valve was probably slowly leaking and/or the reduction in suction pressure at the ”off“ pump was bringing air out of solution at the lowest pressure location—the ”off“ pump casing. The pressure at either pump suction was always below atmospheric pressure. During shutdown, air could form in the casing. On startup, there was not

A

sufficient liquid in the casing FIGURE 1 to prime the pump. The water being pumped is the effluent from a waste water To System Pump 1 treatment plant. The final processes in these systems saturate the water with air. Check Entrained air reacts to even Valves slight pressure reductions by 10 ft. coming out of solution and Pump 2 into a gaseous phase. Adding the suction side check valves to the system 5 ft. trapped liquid in the suction Foot Valve line to the pump and stabilized the ”off“ pump suction pressure, assuming the Chlorination Water Pump System check valves did not leak. The fact that the ”off“ pump would periodically still lose prime when started liquids with high levels of flashing hints that the foot valve is leaking gases or dissolved air. Pump designs sufficiently to reduce the liquid volfor suction lift applications should ume in the suction line enough to include stuffing box shaft sealing so prevent pump priming. that atmospheric air is not drawn Single suction lines for each into the pump. pump from an atmospheric pressure How can the problem in this source are recommended for all installation be rectified ... what is installations – flooded suction or sucthe most cost effective solution... tion lift. On applications from a process vessel, a single suction line FIGURE 2 for two pumps must be designed to assure uniform flow to either or both pumps. Solid bearing liquids require special consideration. Foot valves are not recommended as a conservative design practice Fill Port in an unattended application of a critical pumping task because interLow Liquid mittent or continuing leakage is a Level potential risk. Foot valves of all Priming designs have high friction loss which increases suction side losses. Chamber Any solids in the liquid increase the probability of leakage, and foot valves should not be used in solids bearing liquids. Suction lift system designs involving liquids with high vapor pressures or high levels of disFoot Valve solved air should be limited to short lifts depending on individual conditions. In another column we Centrifugal Pump Priming will discuss calculating net positive suction head available (NPSHA) on The Pump Handbook Series

27

3.

FIGURE 3

Control Valve

Pressurized Fluid

To Waste Venturi Ejector

Liquid Sensor

Vent Air Valve

Automatic Priming System

and how can the problem be avoided in a new installation? Retrofit methods and design options can include: 1. separate suction lines to each pump with foot valves 2. installation of a priming chamber with the existing piping

28

utilization of submersible pumps located in the liquid sump 4. installation of an automatic priming system 5. use of a self-priming end suction centrifugal pump A reasonable assumption is that the existing foot valve is leaking now or will leak in the future. Installing a priming chamber could be the most cost effective solution (Figure 2). The suction line is raised to the top of the priming chamber which is initially filled through the fill port. On shutdown, should leakage occur through the foot valve the priming chamber liquid level will drop no lower than the suction inlet. The priming chamber volume must be several times (3-5) that of the suction pipe volume. When the pump starts, the liquid level in the priming chamber drops, and the pressure in the priming chamber will fall below atmospheric pressure. As pumping continues, atmospheric pressure on the source surface will force liquid up the suction pipe and into the priming chamber assuring continued pumping. This would be a positive fix to the existing installation.

The Pump Handbook Series

For the engineer’s existing situation, the less costly solution would be separate suction lines with non-metallic ball check valves and seats as the foot valves. The pumps should again be checked out to be sure the shaft sealing method is not allowing air to enter the pump. If shaft packing is used, a water sealed lantern or seal ring may be necessary, and seal water would have to run continuously. If the foot valves leak, a simple vacuum priming system might be added. If pressurized water is available, an ejector priming system could be a minimal cost addition. Compressed air or steam can be substituted for water in the ejector priming system. One form of an ejector priming system is shown in Figure 3. Other forms of priming systems use vacuum pumps to reduce pressure and remove air in the pump and suction lines. If this application is new, the designer might consider submersible pumps, self-priming pumps, or end suction centrifugals with an automatic priming system. The pump choice would be set by the flow and pressure required. ■

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PUMP AND SYSTEM TROUBLESHOOTING HANDBOOK

Pump System Design - Part 1 BY ROBERT KREBS, CONTRIBUTING EDITOR plant engineer for a Southern cooperative utility sent us a letter recently expressing the concerns we hear so many times from individuals who have the occasional need to work with pumps and systems. Each time these people get involved with pumps, they must retain system design, pump types and sizes, friction loss calculation, and piping design. The five questions in his letter addressed these concerns and asked for sources to help the infrequent user. To address this need, we will begin, with this issue, a series of columns that will start with simple system designs and move toward the more complicated configurations and pump selection. Although some of this may be ”old hat“ to the more experienced reader, there may be others in your organization who could profit from this information, so pass it along to them. Let’s take a trip through a pump system. First, there has to be a need for a pump—a need to raise the energy level of a liquid to accomplish a purpose. A pump system takes liquid from a source and delivers it to a discharge point. It consists then of a source and destination point, valves, a pump or pumps with drivers and piping as shown in Figure 1. There are friction losses in the piping (suction and discharge), the pipe fittings and the valves which increase with flow rate. Tables of friction loss are available in hydraulic handbooks and in the engineering section of vendor catalogs. The pump requires a certain minimum pressure at its entrance or suction point. This minimum pressure increases with flow and pump speed. This is called the net positive suction head (pressure) required (NPSHR) above the vapor pressure of the liquid. Therefore, the system must provide at least that net positive suction head available (NPSHA) above the vapor pressure of the liquid at the design flow conditions.

A

All pump sys- FIGURE 1 tems can be categorized by liquid destination as either open-transfer or closed-circulating systems. Figure 1 is an open transfer system. The source and destination may be open to atmosphere or closed and subject to different pressures. Figure 2 is a closed-circulating system. The source and destination are the same. At Station A, work is performed on or by the liquid, and the liquid is returned to the source. The FIGURE 2 pump is supplying the energy to circulate the liquid through the system. An example would be a heat transfer system. Both open-transfer and closed-circulating systems may have multiple destinations as in Figures 3 and 4. The destinations examples in Figure 3 may all be at different pressures and flow rates, requiring pressure regulation and flow control at each destination. To provide for continuous pump operation a control valve bypass to the source may be necessary. Other options to achieve continuous operation are variable speed drive and differential system pressure operation. In the closed cir- FIGURE 3 culating system, Figure 4, the destinations may also be non-uniform as to flow and pressure and could require pumping to return to the source. Multiple pump stations may be included in open or closed pumping systems (Figures 5 and 6). Figure 5 is a multiple pump single The Pump Handbook Series

A

29

destination open system. Both pump stations (PS) 1 and 2 are designed to deliver liquid to destination A. In Figure 6, PS1 is providing flow to both A and B. PS2 could be designed to deliver to A and PS3 to deliver to B. It is obvious that the arrangement of pump systems is virtually unlimited. However, the analysis of almost all types can be simplified to some form of these six examples. The system is analyzed to determine the proper location and requirements for the pump. The system design will provide the pump supplier with the criteria needed for the application. The pump supplier will respond with an offering. Designing the system is illustrated by taking an open-transfer system and working through the example. The same approach would be used on a closed-circulating system. The designer gathers data on the liquid properties, source and destination location, pipe routing and lengths, the need for special regulatory considerations, and flow rates. A flow sketch, as in Figure 7, is made and the data for design entered. The designer then constructs a hydraulic flow picture of the system resistance called a system head curve. Figure 7 is an open-transfer system with the system resistance curve. The source pressure (Ps) and suction static head (hsu) are resisted by the frictional pipe resistance of the suction side (hfs). The pump adds sufficient energy to the liquid at the design flow rate to overcome the sum of the static head (h-hsu) and pressure differential (Pd-Ps) and the total pipe friction loss. The pump produces a differential pressure called the total head H1 at the design flow rate Q1. For this transfer system, the total head H1 can be expressed as

˙ FIGURE 4

FIGURE 5

2

A 1

FIGURE 6

2

H1=discharge side pressure – suction side pressure + friction losses

A

H1=[Pd+h] – [Ps+hsu] + [hfs+hfd] Note that the pump produces a differential pressure—discharge pressure at the pump discharge minus the pressure at the pump suction. The suction pressure must be calculated

30

1

3 B

The Pump Handbook Series

to determine that adequate NPSHA is available. In this example, suction pressure would be equal to Ps+hsu – hfs. The total head will vary with pipe size and with progressive fouling of the interior wall of the pipe in use. Since the pump can only operate at the intersection of its performance curve with the actual system head curve, determination of an accurate system resistance is key to a good design. Example 1 adds numbers to Figure 7. The solution will be discussed next month. Get out your handbook and give it a try. The solution will be based on the Darcy-Weisbach pipe friction formula and minor loss coefficients. You may use Chapter 3 of the Cameron Hydraulic Data book or the Engineering Data Book from the Hydraulic Institute (Table III B-4, lst Edition) that has almost the same values. Example 1 Substitute the information below into Figure 7. Select a pipe size(s) and construct a system head curve. Pipe schedule 40 steel (new). Liquid pumped water 70 F atmospheric pressure 14.7 psia. Design flow 200 gpm. Ps = 10 PSIG Pd = 30 PSIG Pipe lengths-100 ft suction-1000 ft discharge Fittings as shown-std radius els-swing check & ball valves h = 100, hsu 20, hst 80. ■

FIGURE 7

Pd hst

Ps

h

hfd hfs hsu

Smaller Pipe H1

System Head

Total Head

Larger Pipe hfs + hfd

H

hst + Pd - Ps Q

The Pump Handbook Series

Q1

31

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PUMP AND SYSTEM TROUBLESHOOTING HANDBOOK

Pump System Design - Part 2 BY ROBERT KREBS, CONTRIBUTING EDITOR

L

ast month we discussed pump TABLE 1. PIPE FRICTION LOSS FOR WATER (APPROXIMATE) system design. We also proSteel Pipe Schedule 40 Minor Loss Factor K for Fittings vided an example of a single Friction ft/100 ft/Velocity Hd (V2/2g) ft (Avg. for 3"-6") h(ft)=K x V2/2g pump, single destination Flow transfer system for you to solve for gpm 3 inch 4 inch 6 inch Description K pump performance requirements. 50 0.7/0.1 0.2/0.02 0.03/0.01 45 el 0.27 Table 1 gives friction loss data for 100 2.4/0.3 0.6/0.1 0.09/0.02 90 el 0.51 three pipe sizes and minor loss coeffi150 5.1/0.6 1.3/0.2 0.2/0.05 Ball Valve 0.05 cients for fittings in that pipe size range. The calculation, Figure 3, sum200 8.9/1.2 2.2/0.4 0.3/0.08 Check Valve 1.7 marizes the 4-inch pipe results with 250 14/1.8 3.5/0.6 0.4/0.1 TEE Through 0.34 the system head curve drawn for 3300 20/2.6 4.9/0.9 0.6/0.2 TEE Branch 1.02 and 4-inch pipe. 400 34/4.7 8.5/1.6 1.1/0.3 Entrance Loss 0.05 Reviewing the calculation, the Exit Loss 1.0 minor loss coefficients (K) were added together for the suction and FIGURE 1. SINGLE PUMP MULTI-DESTINATION SYSTEM discharge sides. The table was constructed to tabulate the total suction and discharge side friction losses as A C B a function of flow (gpm). Friction factor values per 100 ft and the velocity head values were then multiplied by the pipe lengths and total K values, respectively. The plotted system head curves for 3- and 46 2 10 ft inch pipe are the calculated total losses added to the static head (hsu). If this pump system were 1 5 3 7 designed for 200 gpm with 4-inch 4 pipe on suction and discharge, the PS 1 System Conditions total head would be 110.1 ft or, with Flows: Destinations A, B, C, D, Each 100 gpm water (stp) 3-inch pipe, 183.7 ft. Although Pressure: A, B, C, D, Each 40 psig D rarely used in water service, 3.5Source: Open 0 psig inch steel pipe is available. The sucPipes: 2,4,6 = 200 ft, 3,5 = 100 ft, 7 = 300-1,500 ft tion side losses should be carefully Suction = 20 ft examined. In this example, the static suction head (hsu) is 20 ft and the source pressure (PS) is 20 psig (46 ft). The pressure available FIGURE 2. MULTI-PUMP SINGLE DESTINATION SYSTEM at pump suction then is 66 ft, less the friction loss on the suction side or 55.8 ft for 3-inch pipe and 63.4 ft for 4-inch 10 ft pipe. The NPSHA should be adequate. 10 ft The optimum pipe size 2 PS 1 PS 2 will take into consideration the A installed cost of the pipe 1 3 (which increases with pipe System Conditions diameter) and pump power Flows: PS 1, PS 2, Each 200 gpm water (stp) (which increases with Pressure: A = 32 psig increased friction in smaller Sources: Open 0 psig diameter pipe). A reasonable Pipes: 1 = 500 ft, 2 = 400 ft, 3 = 100 ft plan to start with would limit Suction: PS 1 & 2 = 20 ft friction loss at design flow to 2-

32

The Pump Handbook Series

5 ft of friction loss per 100 ft of pipe. High pipe velocities are not recommended on the pump suction side. Velocities of 5 - 10 ft/sec are suggested. Discharge lines of longer lengths are normally in the same range. Realize that sudden interruption of flow can result in large pressure surges (water hammer) of about 50 psi per ft/sec of interrupted flow velocity on suction and discharge piping. In the example (Figure 3), 200 gpm equates to 8.7 ft/sec in 3-inch pipe and 5 ft/sec in 4-inch pipe. The respective friction losses are 9 and 2.3 ft/100 ft of pipe. 4-inch pipe should be used for this application. The pump design would call for 200 gpm at 110 ft with 4-inch pipe. We will go back to this example when we start pump selection. Let me have your questions and comments. Nobody is perfect—no one has all the answers. The more complicated systems require an initial approach analysis. Recall the analysis objective to establish the performance characteristics in selecting a pump (or pumps) to meet the pump system requirements because the pump can only operate at the intersection of the pump (or pumps) performance curve(s) and the system head curve. Figure 1 shows a pump station (PS) with a multi-destination system. The pump and destinations A, B, C and D are all at the same elevation. Destinations A, B, C and D would be equipped with flow and pressure control to deliver the liquid on demand. The maximum flow requirement would be with all four destinations open and demanding their maximum flow. This would set the capacity of pump 1. At flows less than maximum, a by-pass control valve in line 8 would open at some pressure above the set point and bypass flow to the source. Alternately, the pump could operate on an on-off cycle if the frequency of cycling was acceptable for the presumed electric motor driver. Or a gas-pressurized (hydro-pneumatic) tank could be sized to operate over a fixed pressure differential with intermittent pump operations. Still another option would be variable speed control for the dri-

ver. However, with large and frequent variations in flow, variable speed control may still require system storage or by-passing to avoid problems. How to decide on a pipe size? The easiest way for the experienced designer is to use one of the many expert pipe programs for personal computers. For the casual user (once or twice a year), this may not be the answer. Most computer programs are easy to understand when used frequently or after intensive training and use. Another way is to approximate a result by designing for the worst condition, that is, the highest

flow and pressure requirements. Intermediate and low flow conditions affect pump selection, but not system design in clear liquid systems. The worst condition scenario for the system (in Figure 1) would be with all four destinations at maximum flow and pressure. However, there could be pump application problems with this approach. Those will be covered later, when we discuss pump selection. Friction losses depend on flow and pipe size. In this example (Figure 1), the longest pipe string is for flow to destination C through pipes 1, 3, 5 and 7. With maximum flow to each

FIGURE 3. CALCULATIONS Minor Losses - Suction 90 el(0.51) Ball Valve(0.05) Entrance(0.5) Discharge 90 els([email protected]) Check Valve(1.7) Ball Valve(0.05) Exit(1.0) Suction Side 4 inch Discharge Side 4 inch 100 ft Total Total 1000 ft Total Total Total Flow Pipe K(1.06) Suct. Pipe K(3.77) Disch. Friction Hd 50 0.2 0.0 0.2 2 0.1 2.1 2.3 100 0.6 0.1 0.7 6 0.4 6.4 7.1 200 2.2 0.4 2.6 22 1.5 23.5 26.1 300 4.9 1.0 5.9 49 3.4 52.4 58.3 400 8.5 1.7 10.2 85 6.0 91.0 101.2

400

3 in Pipe 300

Head Ft

System Head

200

4 in Pipe 100

0

Static Head

100

The Pump Handbook Series

200 Flow gpm

300

400

33

destination, the friction loss for each pipe section can be manually calculated and the pipe size adjusted where necessary. The pressure required at each destination plus the pipe friction back to the pump (with all destinations receiving liquid at maximum flow) give pump discharge pressure for each destination. The highest of these pressure values would be the pump discharge pressure. Pump total head would be maximum discharge pressure plus friction loss to the maximum discharge destination less positive suction pressure. Figure 2 illustrates a multi-pump station, single destination, open transfer system. The general case would be where psi or pipe 2 or both could deliver flow up to some maximum value to location A. The complexity of the analysis of multi-pump station problems increases rapidly as the number of PS and destinations

34

increase. Again, the expert computer systems would be the solution choice. But, let’s examine a manual solution method. The objective is to determine the performance requirements for psi and PS 2. With the required pressure known at A along with the required flow from psi and PS 2, a pipe size and pressure drop for pipe segments 2 and 3 can be calculated. The discharge pressure of PS 2 and the pressure at the intersection of pipe segments 2 and 3 can then be calculated. This permits setting the pipe size and determining the friction loss in pipe 1 to determine psi discharge pressure. This problem has, again, conveniently assumed psi and PS 2, the pipe and location A at the same grade elevation. Normally there are elevation differences that must be factored into a solution.

The Pump Handbook Series

A system head curve should be made for all applications, where practical. For either of the two cases (Figures 1 and 2), preparing a manual system head curve is complicated. Using one of the expert software systems and varying the flow would provide a range of performance for each PS in a system. For single destinations and single PS systems, a spreadsheet template can be developed to produce system head curves and pump(s) performance curve(s). The system conditions shown for Figures 1 and 2 are to be used as two problem examples. Try working them out. We’ll give solutions next month. Use the friction loss values in the table. Let’s ignore the fittings. Getting the method right is more important. ■ Until next time, when we will start pump selection, which will further define the system.

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PUMP AND SYSTEM TROUBLESHOOTING HANDBOOK

Pump System Design - Part 3 BY ROBERT KREBS, CONTRIBUTING EDITOR segment 3 would have a lower friction loss. Since each PS has 200 gpm flow if only one PS is operating, the friction loss of pipe segment 3 would be equal to 200 gpm flow or one half the flow with both PS operating. Since pipe friction is proportional to the flow squared, the friction in pipe segment 3 with only one PS operating would be about one-fourth (actual is 2.2 ft from the chart) of that with PS1 and PS2 operating. This can become an important factor in pump selection. The pipe sizes are based on the same reasoning as the previous example. Figure 3 illustrates the solution from Example 1 in the March article. The system head curves for 3-inch and 4-inch pipe are shown. Assuming a design flow of 300 gpm, the system head curve (static head plus pipe friction) shows a total head

Anticipating a question that may arise—what about the pipe sizes selected? In choosing the pipe sizes, lower velocities are selected with reasonably uniform head losses for major segments. With instantaneous interruption of flow, surge pressures (water hammer) can occur under conditions present in this design. So the systems should be checked, at least by a rigid pipe analysis, for potentially dangerous surges. The solution for Figure 2 shows the head loss with both pump station 1 (PS1) and PS2 operating. PS1 would be rated at 200 gpm, plus the head loss in pipe segments 1 and 3, minus the suction side static head plus the suction line loss, for a total of 83.8 ft. Similarly, PS2 would be rated at 200 gpm at 81.6 ft. What are the conditions if only one pump station is operating? Line

FIGURE 1. SINGLE PUMP MULTI-DESTINATION SYSTEM A

C

B

2 3" – 4.8' 6 3" – 4.8'

10 ft.

3

1

7 3" – 7.2'

2' 2.

PS 1 400 gpm

– 4"

5 4

4. 9'

6" – 5.5' 6" – 0.2'

4" –

L

ast month we demonstrated more complex pump systems and the methods to arrive manually at pump performance requirements. Figures 1 and 2 (from March’s article; reproduced showing solutions) illustrate the head loss for the pipe size indicated at maximum flow conditions for each pipe segment. For instance, in Figure 1 the maximum total head loss the pump must produce would be the sum of the losses in pipe segments 1,3,5 and 7, which is 19.8 ft. To this amount add the 40 psig (92.4 ft) destination pressure for a total discharge head of 112.2 ft. The 10 ft static head minus the suction side head loss, 0.2 ft, is subtracted from the total discharge head. The pump would be rated at 400 gpm (100 gpm each for A,B,C,D) at 102.4 ft. This would be the maximum pump requirement. What is the performance requirement if only one destination required flow? The lowest head requirement would have destination A operational, with the balance of the 400 gpm flow recirculating to the source. Flow through pipe segments 1 and 2 at 100 gpm would have friction losses of 0.5 ft and 4.8 ft, respectively. When added to the static head at A (92.4 ft), the discharge head becomes 97.7 ft. Subtracting the net suction head (9.8 ft), the required pump total head is 87.9 ft versus 102.4 ft for all four destinations receiving flow. Each pump must be capable of producing 400 gpm at 102.4 ft and operating to 87.9 ft, where the flow will be greater depending on the design of the bypass system. To summarize, the maximum and minimum pump performance conditions have been determined. Since any other operational condition will fall between these values, a pump suitable for the two conditions should perform in this application. Exact site conditions, periods of minimum and maximum flow, variations in destination pressure and flow would also influence pump selection.

3" – 4.8' D

FIGURE 2. MULTI-PUMP SINGLE DESTINATION SYSTEM

10 ft. 10 ft.

PS 1 200 gpm

4" – 0.4' 1

4" – 0.4' The Pump Handbook Series

2 4" – 8.8' PS 2 - 200 gpm 4" – 11.0'

A 3 4" – 8.5'

35

of 138 ft. A pump must be selected with a performance curve that intersects the head and flow conditions. What type of pump? Since the liquid is water—the pressure flow conditions are not rigorous—we will assume the site conditions of the pump station permit a horizontal electric motor driven design to be used. Take my word for it for now—a single stage end suction centrifugal pump will do the job. If you are still interested, in the future I will present a generic method (using various manufacturers) on how to select a technically correct pump. Figure 4 is a manufacturer’s published performance curve for a pump that satisfies the requirement of 300 gpm and 138 ft. What can we learn from this curve? The pump is a 3x2x6.5 (that is, 3-inch suction, 2-inch discharge and 6.5 inches maximum diameter impeller). The varying diameters (from 4.9 inches to 6.5 inches) reflect machining or trimming the outer diameter of the impeller. The operating speed for this pump curve is 3500 rpm. The hydraulic efficiency of the pump is 72% to 87%, and the NPSH required varies with flow rate from 4 ft to 16 ft. The manufacturer has calculated the brake horsepower and plotted lines of constant horsepower coincident with standard motor sizes of 5 hp through 15 hp. The brake horsepower required at any point on the performance curve can be calculated by the formula:

BHP=QxH/3960xEff For our condition (300 gpm at 138 ft and 86% Eff from curve), this equals 12.2 BHP. What impeller diameter should we select? Figure 4 shows the performance of a 6.5-inch and 6.1-inch diameter impeller. The flow that will occur appears at the intersection of these curves and the system curve. The flows are about 315 gpm and 275 gpm for the 6.5-inch and 6.1-inch diameter impellers, respectively. As the piping system continues, deposits can reduce the pipe’s internal diameter and will degrade the smooth surface of the new pipe. This aging

36

effect increases fricFIGURE 3: GRAPH OF CALCULATIONS tion loss. Thus, more head is required to produce 400 the same flow. I would specify the 6.5-inch impeller. How accurate is the calculated 3 in pipe 300 head? If all factors are considered in Head the system head Ft System Head analysis, the calcuPerformance lated head is proba3x2x6.5 pump 200 bly within 10% and 6.5 inch impeller is most probably over the actual 4 in pipe head. What about 6.1 inch the source and destination conditions? 100 Static Head Figure 4 assumes New Static Head them to be constant. That is rarely the case. The cen0 trifugal pump deliv0 100 200 300 400 ers a differential pressure. As this pressure reflects changes at the source or destination, The operating head range will flow will change. The curve labeled vary with the number of pumps in new static head assumes a change in service at the pumping station. The source/destination pressure. Now if pump or pumps selected to operate the 6.5-inch diameter impeller is inmust be able to perform on the sysstalled, it will operate at approximatetem head curve within the operating ly 350 gpm under these new head range. The term “shut-off head” conditions. If the two static head defines the pressure at 0 flow. It is conditions represent the maximum important that the performance curve variation that can occur in an applirise uniformly to the shut-off head. ■ cation it is called the “operating head range.”

FIGURE 4: PUMP PERFORMANCE Pump 3x2x6.5 2000 10

180

15 h

p

hp

72

6.5" 7.5 hp Total Head 140 6.1" (ft) 5.6" 5 hp 100 4.9"

Speed 3500 rpm

82

86

86 82

87 72

60

4

100

16

12

8 200

300

Gallons per Minute

The Pump Handbook Series

NPSHR (ft) 400

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PUMP AND SYSTEM TROUBLESHOOTING HANDBOOK

Pumping System Piping BY ROBERT KREBS, CONTRIBUTING EDITOR

‘T

here is something wrong with the pump.” Sound familiar? From my 40 years of experience, I often find that the pump is blamed for a “system” problem. Yet many, perhaps most, system problems can be traced to suction side source or piping conditions. This month, I will describe some recommended source and suction side piping conditions, and analyze an installation. Listed below are some desirable source and suction piping conditions for shaft pump installation. 1. The source should provide sufficient pressure to provide the required NPSH plus a margin of 2 ft (8 ft for positive displacement pumps) at the pump suction.

FIGURE 1

2.

The source should be designed to avoid vortex formation and air or gas entrainment from turbulence.

3.

The suction piping should be straight, as short as possible, selfventing to the pump or the source, or to waste, with a low liquid velocity, preferably 5 ft/s.

4.

If each pump cannot have a separate suction pipe from the source, the header piping and the immediate pump suction pipe require special attention to ensure minimum turbulence of the liquid stream into the pump.

5.

Valves should be placed away from the pump suction so that flow to the pump is free from the turbulent effects they may cause.

6.

Reducing fittings should be of the eccentric type mounted with the straight side up. Steel and similarly designed reducing fittings should have a straight section before the pump suction.

20" Pipe

16" Pipe

Vertical shaft, submersible and other submerged pump designs will have some different requirements. The objective of these six items is to facilitate a design which avoids turbulence in the source and piping, and channel the liquid to the pump suction with as little change in direction or cross-sectional area as possible. For example, a consulting engineer sent me a sketch of an installation (Figure 1)

Vertical turbine pump installation The Pump Handbook Series

with these questions: Will it work? What do you think? The details of the installation show one of two water system high service vertical turbine pumps. The pumps are taking suction from a clear well of finished water with adequate suction pressure through the 16-inch suction line. Each pump is a 4-stage vertical turbine rated at 1800 gpm at 240 ft head at 1180 rpm. The pumps operate at constant speed. The pump suction bell is located at approximately the same elevation as the top of the 16-inch suction line. The velocity in the suction line is nearly 3 ft/s. In the early 1960s, I served on a Hydraulic Institute committee that formulated some recommendations for sump designs and piping. These have since been published in each edition of the Hydraulic Institute Standards. In analyzing this installation, it is helpful to review the current (14th) edition, pages 126-133. In analyzing any installation, there is a hesitancy to be critical of design. Making general statements about a single situation is not recommended. However, when experience shows that a design is suspect, as in this case, it merits further study. It is possible that this design would work fine, but I do not think so. Here is my analysis. With the system as shown, the 3 ft/s velocity will interfere with the liquid making the 90-degree turn into the suction in a uniform manner. The result is uneven filling or loading of the first stage impeller. The suction flow pattern will possibly result in vibration and premature wear of the impeller. To explain, consider the flow through a 90-degree elbow (Figure 2). The streamlines, which represent the liquid flow, are closer together at the larger outer radius and further apart at the inner shorter radius. In the open pump entrance, the same condition will occur. More flow will tend to go to the outer or right side of the pump suction, as shown in Figure 1.

37

FIGURE 2

FIGURE 3

Flow through a bend How should this design be modified? If possible, the pump suction should be raised by 2 bowl diameters to allow the flow to straighten. The suction pipe should be increased to 20 or 24 inches. I favor an approach velocity (line velocity in this case) of about 1 ft/s with this design. What can be done if the system is built and the problems mentioned have occurred? I would suggest parallel turning vanes, as sketched in Figure 3, to help turn the liquid and distribute uniform flow to the first stage impeller. ■

38

Turning vanes

The Pump Handbook Series

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PUMP AND SYSTEM TROUBLESHOOTING HANDBOOK

Have You Broken Any Shafts Lately? BY ROBERT KREBS, CONTRIBUTING EDITOR rom my experience, inertia or acceleration head is even less understood among pump users than another hydraulic transient—surge or water hammer. Inertia head should not be ignored because it is too frequently a cause of operating problems. To explain, let’s look at the example in Figure 1. The centrifugal pumps are driven by constant speed electric motors. The control scheme calls for a pump to operate and a motor to start. The motor and pump come up to full speed in a few seconds, and liquid flows through the discharge pipe. Right? Wrong? Maybe? The 10,000 ft discharge line is filled with liquid. This liquid mass must be accelerated from standstill to design velocity (3ft/s). By Newton’s law (F=mA), this requires an accelerating force ∆F = M (∆V/∆T) over a finite time period (∆T). So what happens? When the motor starts, pump discharge pressure increases to meet system resistance. From Figure 2 the accelerating force is the difference between the zero-flow head of the pump (Hso) and the static system head (Hs). This force is exerted as a pressure on the liquid in the discharge line. The line velocity will gradually increase to the design point. The dotted line in Figure 2 illustrates the immediate pressure response to the start-up signal. There may be a “pause time” at zero before the flow progresses to the design point—the intersection of the pump curve with the system head curve. This phenomenon can be viewed at any pump station where acceleration head is sufficient to cause a pause time. A pressure gauge on the pump discharge will, with pump start, indicate pressure in excess of the anticipated operating pressure, and after some time the pressure gauge reading will lower to an operating pressure.

F

You may be saying, “What’s the big deal?” The answer is pump problems. Centrifugal pumps are usually furnished with an enclosed impeller and a volute-style casing. When operated to the left of the best efficiency point capacity (lower flow), the liquid pressure in the casing varies around the casing periphery, and the resultant unbalanced radial force in the casing acts on the face of the impeller (between the shrouds) and deflects the shaft while increasing the radial bearing load. Since the force is always in the same direction, the shaft is experiencing a reversed bending stress with each rotation. If this stress exceeds the endurance limit of the shaft material, a premature fatigue failure of the shaft will result. Since the overstress condition only occurs on start-up, can it cause damage? Since stresses imposed in reversed bending are cumulative, it depends on the number of starts, the pause time, and the amount of stress imposed. It should be mentioned that loads imposed on the radial (normally inboard) bearing will reduce bearing life. Also, the deflected shaft will increase packing leakage or decrease mechanical seal life. Fatigue failures in pump shafts can also occur in rotary pumps

under some operating conditions, but that’s another story. The approximate pause time can be calculated from the formula: ∆T = (0.031) [LV / (Hso – Hs)] which is derived directly from Newton’s law. Figures 1 and 2 identify the variables. If a pump starts one time per hour, and the pause time is 10 seconds, and the pump rpm is 1800 (30 Hz), then each day the pump will experience 7200 revolutions of the shaft in the stressed mode. Most fatigue failures occur in 1 million cycles or less, or about 140 days of operation. Figure 3 is one possible plot of liquid line velocity from pump start to design velocity. This path function can only be estimated from point to point with the formula. Obviously, pump flow will increase to some value in excess of the manufacturer’s limiting minimum flow prior to reaching design velocity. Pump manufacturers place an operating range of minimum and maximum flow limit lines on the performance curves to ensure satisfactory operation. Operating at less than the minimum recommended flow can cause many problems.

FIGURE 1. PUMPING SYSTEM

147 Ft Hs

L = 10000 Ft V = 3 Ft/Sec

The Pump Handbook Series

39

Until next time. . .

FIGURE 2. PUMP & SYSTEM PERFORMANCE

Pause Time

Recommended Pump Operating Range

Hso 240

Sy

Hs 147

urv ste m C

Pu mp Pe rf. C

e

ve ur

Inertia head problems are most frequently found in systems with single-stage, end- and double-suction, overhung impeller centrifugal pump designs. Such designs often have a high static head and low capacity. To pass large solids, nonclog impellers are designed with a minimum of 3-inch spacing between the shrouds, providing a large area for the unbalanced pressure to generate a large deflecting force. For many years I have used a calculated 10 seconds as the threshold pause time to require some system design change. Again, the design service conditions must be considered. Other types of systems using diaphragm, blow case or pneumatic ejector pumps should always be checked for inertia head and sufficient air pressure accelerating force. How to design around the problem? For centrifugal pumps, a steep curve with a high zero flow head will maximize the available force for overcoming inertia. A solution is to bypass a portion of the discharge flow to the source until the remaining residual force can start liquid flowing, at which time the bypass closes. For clear liquid applications there are special valve designs. Variable speed drives with infrequent starts will also help. ■

Head Ft

Actual Pump performance From Start-Up

Flow GPM

FIGURE 3. SYSTEM INERTIA EFFECT

Pipe Design

3 Velocity Ft/Sec 2 1 0

40

The Pump Handbook Series

5

10

Time Sec

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PUMP AND SYSTEM TROUBLESHOOTING HANDBOOK

Pumps in Parallel - When One Is Not Enough BY ROBERT KREBS, CONTRIBUTING EDITOR

HEAD

1 or 2 is operated alone, the capacity each would produce is the capacity shown at the intersection of the pump curve with the system curve. With both pumps operating, the capacities are added at the same pressure. The resulting curve gives a new intersection point on the system curve for the total capacity. To find the flow contribution of each pump, trace back at the head pressure to the intersection with pump curves 1 and 2 at point B and determine each pump’s performance. Care should be taken that the smaller pump is not forced to operate at a flow less than manufacturer’s recommendations. If the pumps are positive displacement, their capacities are added together at a given pressure. The resulting intersection point with the system curve gives the total capacity.

In the case shown in Figures 2a and 3, the system head curve permits both pumps to contribute to the flow when both are operating. In some cases for centrifugal pumps, the smaller pump or pumps may be unable to move any liquid with a larger pump operating because the system resistance exceeds the pump shutoff head. Figure 3 illustrates this concept by changing pump 2 to pump 3 (dotted line curve). Pump 1 may be selected because it can provide the flow for a high percentage of requirements with a lower horsepower than pump 3. Applying a variable speed drive to pump 3 may accomplish the same thing. Figure 4 illustrates the importance of insuring that the pumps will operate individually in a system, as well as in parallel. The two pumps operating intersect the system curve

FIGURE 1

2

2 3

1

1

FIGURE A

FIGURE B

SYSTEM HEAD

A PUMP 1 OR 2

PUMP 1+2

PUMP DESIGN

PRESSURE

S

ometimes two pumps are required in the same service. This month we will look at system designs incorporating two pumps which allow flow contribution from one or both pumps. In designs such as these, the pumps are said to be operating in parallel. In a simple case, two identical pumps are selected to provide the flow (Figure 1 ). If they are centrifugal pumps, their performance on a head-capacity chart will appear as in Figure 2a. The system head curve locates the operating head and capacity for operating pump 1 or 2. When two pumps operate in parallel, their capacities at any head are additive, so the curve 1 + 2 shows the intersection of the system head curve with both pumps operational. Note that the intersection with two pumps operating is at a higher head than with only one pump operating. To find the flow being pumped by each pump with two operating, trace back at that head to the single pump curve and read the flow for each pump at Point A. Since, with two pumps operating in parallel, the system head rises with increased flow, each pump produces less flow than when it operates by itself. What if the pumps are positive displacement? Figure 2b illustrates such an example. FIGURE 2 The same rule applies— adding the flows of each pump at the same pressure provides the performance of the two pumps operating in parallel. Note the almost constant flow rate with pressure. The slippage or bypass flow increases with pressure and decreases with increased viscosity. Pumps of different size operating in parallel are also common. Figure 3 illustrates the system. If the pumps are centrifugal, the performance will be as shown in Figure 3. If pump

PUMP 1 OR 2 FLOW

The Pump Handbook Series

PUMP 1+2

FLOW 41

within their flow capability. However, the system resistance at the capacity limit of the identical pump cannot be reached with one pump. Manufacturers place both minimum and maximum capacity limit lines on centrifugal models. The maximum capacity limit or runout condition is usually 10-15% beyond Best Efficiency Point (BEP) capacity. Some pumps will pump more; however the NPSHR for centrifugal pumps increases as a power function (to the 1.5-2.0 degree) with flow and may exceed the NPSHA resulting in cavitation. Noise and vibration are other hazards of operating at a runout condition. Positive displacement pumps used in parallel should have similar maximum design pressures. Also, if internal and adjustable pressure relief valves are present, they should be set to correct pressures. In these examples, we have conveniently used identical suction and discharge piping for each pump from the source to their common discharge line. For instance, in Figure 1, if pipes 1 and 2 are identical centrifugal pumps, the pressure at the discharge of each pump would be the same. But what if pipe 1 is much smaller in diameter than pipe 2? The only point of common pressure for the two pumps would then be at the intersection of pipes 1 and 2 with the discharge line 3. What about the flow? The higher friction loss in pipe 1 would meet the pressures at the 1-3 intersection. The pump 1 discharge pressure would correspondingly increase, and the flow from pump 1 would decrease. What about the flow from pump 2 under this condition? Note that the system head curve is independent of the pipe or pump. Note also that pumps 1 and 2 operate independently. The system curve is the control. Pump 2 continues to provide flow at a rate limited by the system resistance. Some guidelines for pumps in parallel: • Use pumps in parallel as a backup in critical services.

42

FIGURE 3

HEAD

3

B

B

1+2 2

1

FLOW FIGURE 4

HEAD

MAX FLOW

1+2

1 OR 2

FLOW

•

Use two or more pumps in parallel to meet fluctuating flow demands or consider variable speed control.

•

Check how the pumps selected will operate in the system individually and/or in parallel if necessary to provide for adequate reserve.

•

With several pumps operating in parallel, even if the largest pump is out of service, the remaining units should be able to pump the maximum flow.

•

With identical pumps in parallel service, use elapsed hour meters on each pump, along with alter-

The Pump Handbook Series

nation, to assure that each pump is regularly exercised to balance the service life. It should be clear that the accurate determination or calculation of system resistance for the flow range is the important step. ■

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PUMP AND SYSTEM TROUBLESHOOTING HANDBOOK

Pumps in Series - For More Pressure BY ROBERT KREBS, CONTRIBUTING EDITOR

W

hen matching a pump to the system curve, one may find that the pump options available and suitable for the service cannot meet the pressure requirements at the design flow. The design system pressure consists of static or elevation head plus the friction head or losses and any velocity head changes at the design flow rate. If the application requires a positive displacement pump, the pump manufacturer’s literature should offer varying capacity pumps designed for a range of maximum pressures. However, meeting the design capacity may require two or more pumps in parallel to achieve the desired flow. (See ”Pumps in Parallel” in the August issue.) If the application calls for a centrifugal pump and a single stage pump will not produce the required pressure at the design flow, it is common to use a two-stage or multi-stage pump, i.e., two or more impellers and casings on a single drive shaft, each producing the same flow to raise the pressure a predetermined amount. For many services the search ends there. Centrifugal pump types that are manufactured in two- or multi-stage designs include end suction, double suction, vertical turbine and regenerative turbine pumps. Some examples of designs available only in single stage are non-clog designs for handling solids laden liquids, some slurry pump designs, and rubber lined pumps. For these pumps the pressure requirement may be met by employing two or more units in series. (Crude oil and petroleum products pipeline pumps are highly specialized and not included in this discussion.) The August article on parallel pumping described how the flow rates of two operating pumps are added at the same head or pressure to produce a composite performance curve. Conversely, when two (or more) centrifugal pumps are operated in series within the same pump sta-

FIGURE 1. TWO CENTRIFUGAL PUMPS IN SERIES Source (Constant Pressure)

Destination 2

1

Design Pressure

System Head

Head Pressure

Pump 1 + 2

Pump 1 or 2

Design Flow

Flow

FIGURE 2. TWO CENTRIFUGAL PUMP STATIONS IN SERIES Surge Tower

PS2 PS1 System Head PS2 PS2 Head Pressure

Design Pressure

Total Head PS1 = H1 Total Head PS1 = H1 + H2

H2 System Head PS1 PS1

Design Flow

H1 Flow

The Pump Handbook Series

43

tion, the heads or pressures produced by each pump are added at the same flow rate to produce the composite performance curve. The simplest configuration for centrifugal pumps operated in series incorporates two identical pumps connected so the discharge of the first enters the suction of the second, which then discharges to the system. Since the pumps are identical, each produces an equal pressure. Figure 1 illustrates the performance of two identical pumps with the correlating system head curves. Pump two should be designed for the higher suction and discharge pressures since the pressure at the stuffing box of this pump will be higher due to the higher suction pressure. The inter-connecting piping should be straight and no smaller in diameter than that of the pump suction. The series concept is complicated for pumps in separated pump stations (PS). Figure 2 illustrates a two PS in series operation. The suction pressure

44

at the PS2 location must be carefully controlled, and PSl must provide adequate pressure at design flow to deliver liquid to the PS2 suction. PS2 must then deliver the design flow to the destination. Ideally, PS2 is located so that the total head of PS1 and PS2 is the same. If PS2 is located some distance from PS1, a surge tower or standpipe at the second pump station suction might simplify the design (see Figure 2). Several years ago, I designed a water supply system that carried the raw water some 16 miles from the source to a treatment plant located at an elevation more than 500 ft. above the source. To accommodate the limited pipe design pressure, three booster pump stations were located so that their total heads were equal. Since there was no storage, water flowed directly from one pump station discharge pipe directly into the suction of the pumps in the next pump station. Calculations for surge pressure effects and starting considerations

The Pump Handbook Series

called for an array of control valves and sensing systems. If flow must move directly from PS1 discharge to PS2 suction, the total head of each PS must be accurately calculated at the design flow. PS1 should be designed to provide the pressure for liquid to reach PS2 with adequate suction pressure to meet the PS2 NPSHR at the design flow. PS2 should be designed for its system requirements. As Figure 2 illustrates, the problem becomes two separate applications. These separated pump stations connected in series may produce a maintenance prone system. To reduce the problem potential, the same impeller design should be employed for PS1 and PS2. Impeller diameters and rotating speeds may be different, but the curve shape will then have the same characteristics. If the application is suitable, the versatile vertical turbine pump design is the pump of choice for this type application.

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PUMP AND SYSTEM TROUBLESHOOTING HANDBOOK

Vertical Pumps with Integral Thrust Bearings BY TERRY M. WOLD

T

his article is an addendum to Herman Greutink’s articles “Vertical Turbine Pumps” and “Troubleshooting Vertical Turbine Pumps” (Pumps and Systems, February 1994). All suggestions in these articles on installation, troubleshooting, and maintenance are applicable to this discussion of vertical pumps with integral thrust bearings.

FIGURE 1

VERTICAL IN-LINE PUMPS WITH INTEGRAL THRUST BEARINGS Vertical in-line pumps with integral thrust bearings were originally developed to facilitate in-line pump maintenance. In-line pumps have an Achilles’ heel when it comes to maintenance. Since the driver bearings are also the pump bearings, alignment of the pump and motor shafts is critical to the life of the pump, especially the mechanical seal and the close-running clearances of throat bushing and wear rings. In-line pumps are usually of the volute design, and therefore a radial load is present. This results in a shaft deflection which is magnified by a combination of both final alignment and machining tolerances. The use of a close-coupled design could eliminate this difficulty, but then there are the complications from seal removal, motor availability, product temperature and limitation on sealing device types. Vertical in-line pumps with integral bearings are easier to maintain and thus have reduced maintenance costs. Another use for these pumps became apparent after initial users had installed and operated a few units. Installation savings were realized because these pumps used a relatively small foundation (that is, less space) materials, and time for construction. The vertical in-line pump with integral thrust bearings could do most of what the traditional horizontal pump could do and save money doing it. Vertical in-line pumps with integral thrust bearings have these advantages:

Cross section of in-line pump with integral thrust bearings 1.

2.

The pump has its own bearings for handling thrust loads. Therefore, high thrust vertical motors are not required, although a “P” base motor is recommended due to tighter face and runout tolerances, in addition to the more uniform rabbet fit and bolt hole patterns. A flexible disc type coupling can be used instead of a solid axialsplit spacer coupling, which can add to alignment problems if not correctly manufactured, handled, and installed by trained mechanics. The Pump Handbook Series

3.

Since the radial bearing of the pump is closer to the impeller, shaft deflection is minimized and mechanical seal life is extended.

4.

Foundation preparation and cost are greatly reduced.

5.

These pumps can be added to an existing facility without massive piping changes or relocation of other equipment.

Factors that need to be evaluated when considering the purchase of an in-line vertical pump with thrust bearings should include:

45

•

Bearing lubrication—oil mist or oil sump lubrication are recommended.

•

Limitation of seal types—check stuffing box space dimensions and distance to the nearest obstruction.

•

The rotating assembly should be removable without disturbing the pump case or driver. (This is an API 610 requirement.)

•

Check the length of the coupling spacer. You should be able to disassemble the pump without removing the driver.

•

Check to see how the pump can be handled during maintenance procedures. Consideration for lifting should be addressed.

the bearings will exceed these savings. •

•

VERTICAL TURBINE PUMPS WITH INTEGRAL THRUST BEARINGS Vertical turbine pumps with integral thrust bearings are becoming more accepted in the United States. This is because only a few motor manufacturers will build a high thrust motor capable of handling the loads imposed by a multistage vertical turbine pump. When evaluating the purchase of a vertical turbine pump, you should consider the following: •

46

For thrust balance considerations, the lower the thrust the longer the bearing life. Minimum bearing life should be 25,000 to 40,000 hours B-10, whether the bearings are in the driver or the pump. Some users have specified 100,000 hours B-10 bearing life. To achieve this, the pump must be thrust balanced or a larger or different type of thrust bearing must be used. Special thrust balancing is expensive. If the bearing size is increased, the balls may skid because of inadequate thrust load. If a different type of bearing is installed, the lubrication becomes complicated or the maximum allowable rpm is reduced. Although a higher pump efficiency can be attained without thrust balanced impellers, the cost of maintenance due to the higher thrust loads on

•

Pay close attention to the number of alignment fits incorporated in the pump design. The probability of misalignment increases with more fits. • Check how the adjustment for impeller clearance is made. Is there a one-piece shaft going through the pump bearings, which is then connected by a flexible coupling to the driver? Or is there a rigid adjustable coupling below the bearings and a jack shaft and coupling above the bearings to couple the driver? In general, the one-piece shaft is superior. The extra maintenance to service the seal FIGURE 2 when one shaft is used far outweighs the frequency of seal repair due to the extra coupling and shaft alignment required when using a two-piece design. Those attending the last API 610 8th Edition meeting agreed that the singlepiece shaft is the preferred design. This suggestion may become part of the 8th edition. Check how maintenance is performed on the bearing housing or mechanical seal. The bearing housing is located between the driver and the stuffing box. Therefore, removal of the mechanical seal is performed by removing the bearing assembly first. You should be able to remove the seal and the bearing assembly without disturbing the pump or motor.

•

Oil mist or oil sump are the preferred methods of lubrication. In most cases an ISO 32 or ISO 68 nondetergent turbine oil is sufficient.

•

The recommended thrust bearing arrangement is two 40° angular contact bearings arranged in a The Pump Handbook Series

back-to-back configuration with machined bronze cages. With the exception of a momentary upthrust at startup, it is preferred that a pump has downthrust when operating at design condition. Maximum oil temperature is dictated by the properties of the oil used. If the recommended ISO 32 or 68 oil is used, maximum oil temperature should be 150°F. This can be greater if a higher viscosity oil is applied, but two factors must be checked. First, the minimum viscosity at operating temperature should be 70 ssu. Second, if a higher viscosity oil is used, the

Motor Support Coupling Bearing Housing Stuffing Box Pump Head

Column

Bowl Impeller Barrel

Cross section of vertical turbine pump with integral thrust bearings

pumping characteristics of the oil circulation device will be drastically impaired. Arctic applications may require an oil heater or a lower viscosity oil, although a tropical environment may require a higher viscosity lubricant. •

Vibration should be measured on the bearing housing at the location of the bearings. It should not exceed 0.15 in/s between minimum flow and operating condition.

•

A conventional vertical turbine pump has a shorter overall height and, with a properly sized spacer coupling, the seal can be removed without disturbing the pump or driver. However, alignment becomes a more difficult and critical aspect.

CONCLUSION Vertical pumps with integral thrust bearings have an expanding role in industrial applications. This type of pump will not solve inherent

The Pump Handbook Series

hydraulic problems but will increase the mean time between failure in most applications, as long as the correct operating and maintenance procedures are followed. These pumps may have a higher initial cost, but repair savings may make them well worth the investment. ■ Terry M. Wold is Engineering Manager at Afton Pumps, Houston, TX.

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PUMP AND SYSTEM TROUBLESHOOTING HANDBOOK

Why doesn’t the pump pump? BY: J. ROBERT KREBS, P.E. everal years ago I was confronted by the following situation. The system, diagrammed in Figure 1, was a sewage pumping station with three constant speed centrifugal non-clog pumps of the same size. Pump 1 would not pump; pump 2 would pump now and then; and pump 3 worked fine. As shown in the figure, there were two inlets to the source wet well. Source A was a gravity sewer, and pipe source B was a force main destination from another constant speed pumping station which periodically delivered flow to the source wet well at a rate of 550 gpm. The first question we ex-plored was, why did pump 1 not pump? Was the impeller not turning or damaged? Was there no water in the casing? Was there a clog in the impeller, or in the suction/discharge pipes or valves? Were there other conditions we hadn’t considered that could be causing the trouble? A visit to the station provided these hints. Pump 1 would begin pumping, and then the flow would

S

FIGURE 1

gradually stop, as if there was no liquid in the casing. Pump 2, after starting, would react the same way when there was flow from the pipe B source. Finally, a crackling cavitationlike noise came from all three pumps. An inspection of the station and equipment ruled out clogging. The pumps were new, so impeller damage was unlikely. The pump shaft turned, and it was doubtful that the impeller was slipping on the shaft. Other possible culprits were grouped as follows: 1. inadequate NPSHA for pump NPSHR at the design flow 2.

air leakage to the casing

3.

operating total head much less than the design total head (i.e., the pump not operating on its curve)

Sewer (elev. 938) A Concrete Filet B

4. Wet Well

6" Suction

2 1

3

Pump Station

48

FIGURE 2 Cap with "I" hole tapped in top – standard pipe thread. Cover plates CL Elev. 950 8" cross 8' -0" wet well 8" D.I.P. Force main Float controls



,

 , 

entrained air coming out of solution and causing the pump to lose prime.

A second inspection demonstrated that the minimum wet well level would ensure adequate NPSHA with respect to NPSHR of the pumps at design flow. In addition, the pumps had double mechanical seals which were lubricated with seal water, and there were no apparent air leaks in the pump station. Finally, a check of total head versus design head showed

The Pump Handbook Series

Concrete filet

Elev. 935

Bottom Elev. 933

that the pump was operating at a reasonable point on the curve. Eliminating the first three possibilities left entrained air as the most suspect cause of the faulty pump operation. The discharge line from the upstream pumping station, pipe B, entered the station near the top of the wet well and was then directed down into the well through an 8-in. cross fitting in a closed pipe to prevent excessive surface aeration in the well. To avoid the 15 ft siphon effect, the designer provided a 1-in. opening in the cross cover plate as shown in Figure 2. However, no liquid came

from this opening when the upstream pump station operated. When there was no flow from the upstream pump station, the water level in pipe B was the same as the level in the wet well. And, since the well liquid surface was at atmospheric pressure, the liquid level in pipe B was also at atmospheric pressure because of the 1-in. opening at the top of the pipe. However, when the upstream pump station delivered its rated flow of 550 gpm into the station through pipe B, the pressure situation changed. Since the pressure was clearly atmospheric at the 1-in. opening, the pressure just inside the opening must have been less than atmospheric to induce the no flow condition. What was happening? To answer this question, we first analyzed the

forces acting on the liquid in pipe B. The 15 ft liquid column downward velocity is assisted by the gravity effect of the elevation difference but resisted by the pipe friction. These forces balance at about 3000 gpm, and at lower flow the downward gravity force dominates. Clearly, then, the pipe was not running full. And, because the down flowing liquid reduced the pressure in the vertical pipe B to below atmospheric, air entered the system through the 1-in. opening. The end result was that a huge volume of high velocity air entered pipe B. This air left the pipe in close juxtaposition to the pump 1 suction pipe so that when pump 1 was operating, substantial quantities of entrained air would replace the liquid in the impeller and casing, causing the pump to lose prime. Removing

The Pump Handbook Series

the cross cover plate could reduce the turbulence of the air, but not the quantity of air. The operating conditions of the upstream pump station were then inspected to confirm that the siphon effect would not seriously reduce the operating pressure of the pump in that station. Closing the 1-in. opening on the cross provided the best solution. That was done, and it worked. ■ Until next time . . .

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PUMP AND SYSTEM TROUBLESHOOTING HANDBOOK

Pumping Downhill BY J. ROBERT KREBS, P.E., CONTRIBUTING EDITOR

W

hy pump downhill? Doesn’t water run downhill naturally? While it’s true that in an open to atmospheric pressure system (closed conduit or open channel) the force of gravity causes the liquid to seek the system’s lowest unrestricted elevation above sea level, examples of downhill pumping systems are all around us. The pumping of effluent (treated waste water) into the ocean through pipes that are frequently miles long provides one good example of a downhill pumping application. Such ocean outfalls require pumping to overcome the pipe friction loss at the desired flow rate. Heating and cooling systems in commercial buildings are another example. These systems are generally of a closed-circulating type and the pumps are sized to compensate for friction loss. A more complex situation is presented by raw water and waste water pump stations. These systems often require pumping from the source to a higher elevation followed by pumping to a lower elevation destination. Many storm water systems, for example, pump over a high point, such as a river levee, then discharge to a lower level, thus creating an initial priming head and a lower operating head. Let’s look at some essentials of good system design for pumping downhill. Figure 1 illustrates a common arrangement, a raw water supply or waste water transfer system operating through an undulating terrain of peaks and valleys. The highest elevation in the system (point A) determines the static head which the pump must overcome at start-up. However, when the pipe is full, the static head will be a negative value based on the elevations noted in the figure. Theoretically, the siphon effect will supply flow, up to some value (Q), once the

50

FIGURE 1. DOWNHILL PUMPING SYSTEM INITIAL CONCEPT

Elevations ft. 120

A

100

B

90

C

95 80

85

D Head ft. +30

0 -10

System Resistance Static - Priming Operating Q

Flow -gpm.

FIGURE 2. DOWNHILL PUMPING SYSTEM FINAL DESIGN

A

VARVs C B

Surge Relief Valve

Head ft.

Design Condition

+30

0 -10 pipe is filled. In fact, since the discharge destination, point D, is at a lower elevation than the source for the pump, liquid could (depending on pump and control valve types) The Pump Handbook Series

Flow - gpm continue to flow through the system after the pump stops. If the pump discharge pipe employs a swing or ball check valve, liquid could flow through the stopped

pump into the pipe, a condition that is to be avoided. An actuator controlled plug or ball valve, substituted for the check valve, will act as an isolation valve and ensure that flow to the discharge pipe stops, as it should, with the pump. Clearly air must be evacuated from the pipe as it fills. Good design dictates the use of properly sized air release valves (ARVs) at the high points, A, B and C. The ARVs regulate the initial filling rate of the discharge pipe and any subsequent refilling of the pipe after draining. The flow rate can be controlled by throttling the pump discharge or reducing speed, if variable speed is available. The liquid velocity setpoint will vary with pipe diameter and the steepness of slope (i.e., sharp or gradual) approaching the high elevation points. The objective in regulating the liquid velocity is to permit the trapped air in the pipe to leave as the pipe fills. In addition, raw and waste water systems generally contain varying amounts of entrained air. This air may come out of solution and migrate to the high elevation points in the discharge pipe. The ARV selected for the application must also be capable of removing this air from the system.

With the ARV installed, the pipe filled and an open discharge at point D, the control valve on the pump discharge will close when the pump stops, prohibiting the water from entering the pipe. Under these conditions, at the discharge point D water will flow via gravity to a lower elevation, creating a lower pressure (or vacuum) in the pipe. Depending on the elevation differences, a vapor cavity could form in the pipe, developing a pressure differential that could possibly exceed the pipe’s threshold of collapse. If the ARVs selected are combination vacuum and air release valves (VARVs), these valves, at A, B, and C, will admit air from the atmosphere when the pump stops. This partially drains the liquid at point D. The VARVs must be sized to release air at the high points with subsequent pump start-up and to allow air into the pipe to relieve the vacuum formation. As an alternative, a remotely controlled valve at point D could be closed on pump shutdown. At this point in the design, the system should be inspected for surge pressure. The system designer should be aware that interrupting the flow in any system could cause the generation of dangerous pressure surges or water hammer. Provisions to attenuate these surge pressures for both normal and power failure shutdowns

The Pump Handbook Series

are essential. System analysis for water hammer is complicated and best handled by specialists. We are now ready to select a pump for our downhill pumping system. First, consider the operating conditions. Figure 2 illustrates a specific system design with system head curves for both startup/priming head and operating conditions. A conventional swing check valve is employed at the pump discharge. Surge relief may be provided by a controlled closing check valve and/or a separate surge relief valve. The points A, B and C are equipped with combination vacuum and air release valves (VARVs). A constant speed pump may be selected if it can meet the following criteria: • the ability to provide flow for the initial pipe filling at the stationary priming head •

the ability to maintain the design flow at the operating head.

If these criteria can’t be met, a variable speed pump should be selected. ■ Until next time . . .

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PUMP AND SYSTEM TROUBLESHOOTING HANDBOOK

Selecting a PumpThe Right Start Leads to the Right Finish BY J. ROBERT KREBS, P.E., CONTRIBUTING EDITOR

I

magine you have just received a request to select a pump for a special project (and perhaps you have). There are literally hundreds of companies offering one or more pump types to help the end-user solve that special application problem. And, the overwhelming number of choices and questions may make selecting a pump a dizzying proposition. Let’s try to sort a path through this maze. Since the pump business focuses on application engineering, it is imperative to begin with a thorough systems analysis. A faulty system design will certainly lead to improper pump selection– and poor operation. Locating the pump in the system should be an early consideration. In addition, it is clearly essential to determine how the pump will function in this system. Is the system functioning to transfer or circulate the fluid? Are the pump requirements continuous or intermittent? The system analysis should start with an evaluation of the process fluid. Does the solution contain solids? If so, what is the nature and concentration of the solids? What is the viscosity, specific gravity, vapor pressure and process temperature of the liquid? Next, the analysis should consider the process variables, such as: • net positive suction head (NPSH) requirements • •

•

52

the possibility of entrained air or gas source and destination pressure and temperature conditions and the anticipated degree of variability in these conditions

The pump site and the environmental conditions for the project are also important parameters in the system analysis. If the pump will be operated in a clean, well ventilated and protected area, the equipment demands will clearly be different from those for a site out of doors in variable weather conditions. Water booster and waste water pump stations, which may be located by necessity in flood plain areas, also impose unique site constraints. These stations often require special piping designs to avoid solids buildup, crystalline formations and/or to allow gas or vapor to escape. Additionally, determine whether the pump will be at a site that is heated or cooled. Safety and hazard considerations associated with the location should be included in the analysis. Finally, weight and space limitations as well as noise and vibration concerns must be taken into account. Although system control and pump driver decisions are generally made after the pump selection, the site and environmental conditions may also affect these decisions. If so, these system control provisions should be included in the system analysis. A completed design drawing will serve as an overall check of the system. One may think of the pump in terms of three separate design parameters: hydraulic, mechanical and materials of construction. The system design will impose requirements for each of these parameters. Now your job is to match the pump design to these requirements. This will assure a good working pump and system. ■ Until next time . . .

flow rate and whether the flow is continuous, variable or intermittent.

The Pump Handbook Series

System and Pump Requirements Environmental Considerations • location – temporary or permanent • hazards – electrical, mechanical or vapor • site – flooding or equipment protection Liquid Properties • chemical name • percentage of solids-their size range • viscosity • specific gravity • temperatures • vapor pressure • pH Hydraulic Design • capacity (min/max) • pressure-discharge and suction • static head-suction and discharge • NPSHA above vapor pressures • design (rated) capacity and differential pressure Mechanical Design • expected mean time between failures • L-10 bearing life hours • intermittent or continuous service • variable or constant speed • casing design pressure • hydrostatic test pressure Materials Design • consideration for abrasive or erosive wear • corrosion allowance • failure hazard concerns Drive and Control • electric motor or engine • air or hydraulic power • variable or constant speed • control of flow or pressure • valves, sensors and actuators

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PUMP AND SYSTEM TROUBLESHOOTING HANDBOOK

Selecting a PumpWhat Type Should It Be? BY J. ROBERT KREBS, P.E., CONTRIBUTING EDITOR FIGURE 1. PUMP TYPES FOR LOW VISCOSITY LIQUIDS 1,000 Centrifugal Commecial Sizes Axially Split Case

ve nt iti me s e Po lac p s Di

Head in Feet

T

he pumping system design is completed. Now the pump selection process begins. The principal pump operating criteria evolves from the system design itself. By examining these criteria, we can establish the hydraulic, mechanical and materials design requirements that the selected pump must provide. Once these requirements are fixed, it is possible to determine the type of pump that could be considered in the application. This is usually the time to decide whether to use a positive displacement or a kinetic (usually centrifugal) machine. For example, pumping a low flow rate of a highly viscous clear liquid at high pressures would indicate selection of a positive displacement pump. Conversely, higher flow rates of a low viscosity clear liquid at modest pressure would suggest a centrifugal impeller pump design. There are many other considerations – pulsed versus continuous flow, single or multi-stage centrifugals, slurry or solids bearing versus clear liquids, and as previously mentioned, viscosity – all of which will influence the type of pump selected. Complicated process system designs may require the pump to operate over variable or several specific flow and pressure conditions. These variations may definitely affect the type of pump. For instance, an injection pump supplying a blending agent at a constant flow rate to a pipe at a location of varying pressure would be an application for a positive displacement design. Commercial pumps are available for most low viscosity clear liquid applications. The figure from my pump training course is a head capacity chart illustrating the range of commercially available single stage centrifugal pumps of the end suction and axially split (double suction) designs. Note that the versatile vertical turbine design may be used in the same range of flow and heads

100

10 End Suction Vertical Turbine

1 10

100

1,000

10,000

Flow in GPM

in single or multi-stage design. Portions of the higher pressure area, including those marked for a positive displacement selection, may be suitable for special designs, such as high speed centrifugal and regenerative turbine pumps. Increasing viscosity of Newtonian type liquids requires special consideration. Non-Newtonian liquids are a special case not included in this discussion. Centrifugal pumps may be used for viscous liquids. I have used 5000 SSU as a maximum allowable viscosity for centrifugal pump application. Increased viscosity has a measurable effect on capacity-head performance and severely affects hydraulic efficiency. Positive displacement pumps are more tolerant of viscous liquid application and are normally used for this type of service. The hydraulic effiThe Pump Handbook Series

ciency of a positive displacement pump, depending on the type, is little affected by increasing viscosity at modest pressures. It becomes more apparent as the pump selection process progresses that the properties of the pumped liquid yield important selection criteria. ■ Until next time ...

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PUMP AND SYSTEM TROUBLESHOOTING HANDBOOK

Selecting a Pump — Suction Pressure By J. Robert Krebs, P.E., Contributing Editor ost pump problems occur on the suction side, ranging from difficulties in suction pumping to sizing and layout. Suction pressure can also be problematical—and that is the area we look at this month. There is no reason to have that enigmatic acronym NPSH (Net Positive Suction Head) a part of the suction pressure puzzle, when it is a valuable clue to it! There are two values of NPSH.

M

FIGURE 1. NET POSITIVE SUCTION HEAD AVAILABLE hp hvp

FLOODED SUCTION +h

CL

CL hf

SUCTION LIFT -h hp

There is no reason to have that enigmatic acronym NPSH a part of the suction pressure puzzle. NPSHA (A for Available) is the pressure you, as the designer, provide in your system. It must be more than the NPSHR (R for Required at the anticipated flow rate). The pump manufacturer provides the NPSHR on curves for the selected pump. The value of NPSHR for a given pump increases with flow rate in the usable centrifugal pump operating range. NPSHR decreases somewhat as the impeller eye diameter is increased (larger suction). NPSHR increases with operating speed (rpm). Figure 1 sketches a simple application with the formula for calculation of NPSHA. A centrifugal or rotary pump is applied here, since reciprocating positive displacement pumps require a somewhat different approach. Now that you have calculated

54

hvp = liquid vapor pressure (ft) hf = friction loss at flow rate

hvp NPSHA = hp - hvp ± h - hf (units ft)

your NPSHA, how much does the pump require? You could start leafing through manufacturers’ catalogs searching for a pump that meets your need—or you could try another approach. Many years ago it was discovered that the parameter suction specific speed was related to the impeller design NPSHR. Suction specific speed has the same variable arrangement as specific speed with NPSH replacing system head. So, how can suction specific speed be used to help you in designing your system? The Hydraulic Institute Standards (14th ed., Pg. 107) features a chart of NPSHA versus capacity for various operating speeds at a suction specific speed of 8500 (English units). The value of 8500 is considered to be an attainable design number for commercial sizes of radial and mixed flow impeller designs. I should note that many manufacturers produce higher suction specific speed design impellers, usually classified as “low NPSH impellers.” This discussion is limited to pumps available as a commercial product for general application. The HI chart noted above is for single suction impellers. If a double suction The Pump Handbook Series

pump is to be selected, the capacity may be doubled. For example, it a 2000 gpm single suction pump at 1800 rpm may require 20 ft NPSHR, the same 20 ft requirement in a double suction pump would permit a flow of 4000 gpm. I have used this chart for many years, both in my consulting practice and training courses. I have found that for a given pump selection it is a reliable guide to how much NPSHA the design will need — and I do not have to search through manufacturers’ curves. Consider an example with water at 2000 gpm, 150 ft total head and 24 ft NPSHA. From the selection chart in last month’s column, an end suction single stage, single or double suction, or vertical turbine single or multi-stage pump should be available. At a speed of 3600 rpm (synchronous speed), the HI chart shows a requirement of just over 50 ft for all single suction selections. A double suction pump, (using half the capacity 1000 gpm), would require over 30 ft. Reducing the speed to 1800 rpm would reduce the NPSHR to about 20 ft, an acceptable alternative. Just trying to make pump selection a little easier. ■ Until next time...

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PUMP AND SYSTEM TROUBLESHOOTING HANDBOOK

Troubleshooting Pump Performance Degradation BY LUIS F. RIZO

COST of head loss: (.68) (62.4 lb/ft 3) (1ft 3/7.48 gal) ($.02/lb) (60 gal/min) (1440 min/day) = $ 9,802.47/day where: .68 = specific gravity of the fluid 62.4 lb/ft3 = density of water The potential for nearly $10,000 in additional revenue per day justified our efforts to improve

FIGURE 1. TWO-STAGE TOP SUCTION TOP DISCHARGE PUMP

FIGURE 2. REPORTED FIELD PROBLEM 1000 OEM Curve Problem Report 950

900 TDH (ft.)

C

urrent market forces often dictate significant downsizing and cost reductions. With the resulting scarcity of resources for expansion and reliability upgrades, industrial processors continually seek means to maximize capacities with existing equipment. Under these conditions optimum performance of rotating equipment is a must! Deterioration in pump performance and the subsequent economic consequences of production loss can be addressed with little or no capital expenditures by applying knowledge of pump and system design principles. A recent experience with the performance deterioration of a 1 1/2 x 2 ESN 2 stage, top suction top discharge design Wilson Snyder pump (Figure 1) provides a good example of the value of this know-how. Operating at 115 gpm and 890 ft of differential head (TDH), the pump was not duplicating the OEM’s test curve performance (Figure 2). The pump was intended to operate at 115 gpm and to produce 925 ft of TDH. After a few simple calculations, the economic incentives to pursue a solution to this problem became clear. The lower head of the trouble condition was costing 60 gpm of flow. Since the product is sold at $0.02/lb, a calculation can be made based on the fluid properties:

850

800

750 700 0

25

50

75 100 125 Capacity (gpm)

pump performance using expense budget funds. Through a brainstorming session and careful analysis of the pump operating symptoms and system conditions, the items listed in Table 1 were identified as possible causes for the performance degradation. This list was used as a guide for the inspection and testing of the pump. When field testing a pump, consider the following key factors: • All gauges should be calibrated, even new gauges. New gauges can be as much as 25% off calibration.

The Pump Handbook Series

150

175

200

•

Gauge readings must be corrected to the centerline of the pump suction.

•

When using computer-collected flow readings, note the actual temperature and pressure of the product and correct the readings from standard temperature and pressure conditions to actual conditions. In hot bottoms services, the flow corrections are especially significant.

•

Measure the running speed of the pump for each set of readings. These must be corrected

55

TABLE 1 Likely Causes

Comments

1. Internal recirculation due to improper operations

The pump is operating and controlled at or near BEP. Internal recirculation is unlikely.

2. Ring clearance not to specifications

Possible cause. Field testing reveals that the head degrades with flow, and delivers OEM performance at shut-in conditions.

3. Obstruction in piping system

Poor piping design can cause severe performance problems. This pump has a significant reduction in the suction and discharge piping.

4. Damaged impeller

Not likely. The pump flow is off the OEM curve, but delivers test curve head at shut-in conditions.

5. Undersized throat or reduced interstage passage

Possible cause. If build up/foreign material is lodged between stages, significant performance reduction occurs.

6. System design created obstruction

The discharge piping was reduced from 4 to 1-1/2 in. at less than 5 pipe diameters from the discharge nozzle.

7. Use of non-OEM parts

Possible cause, but not likely. Records indicate the existing impeller is OEM.

8. Running speed does not match test speed

The pump is directly coupled and the running speed is close.

9. Errors in factors used to correct flows to STP

Possible cause. When adjusting flows to STP the operating temperature effect on the material density must be taken into account.

•

Verify the impeller diameter to correct for differences using the fan laws.

Figure 3 shows the results of the field test. These data revealed a significant reduction in the pump’s head as the performance approached 30 gpm. This drastic pressure drop with flow increase is generally indicative of an obstruction to the flow of liquid somewhere downstream from the impeller(s). An obstruction’s resis-

56

FIGURE 3. FIELD TEST VS. OEM CURVE OEM Curve Problem Report Field Test

1000 950 900 TDH (ft.)

to the OEM’s curve nominal speed as printed on the performance curve.Manufacturer’s curves are corrected to a nominal speed through the range of the performance curve.

tance increases as the square root of the flow rate. Consequently, at shut-off, when the flow rate is zero, the TDH matches that of the OEM test curve. Table 2 describes conditions identified as possible causes for the pump performance exhibited by the field test along with corresponding evidence obtained by mechanical inspection or testing. From these findings, the significant and rapid changes in piping diameters near the discharge nozzle appeared to be the major contributor to the sharp drop in pressure. A modification of the piping is the recommended solution. Good engineering flow measurement practice (Ref. 2) recommends that a straight run of 5 to 10 pipe diameters be designed at the discharge of the pump to allow for the process of recovery. Piping changes and restrictions interfere with the process of converting velocity head into pressure head, and this effect continues to occur at up to 5 to 10 piping diameters from the discharge nozzle. In this case a 2”x4” reducer, 4” check valve and 4” block valve were removed and then reinstalled downstream of a 10 pipe diameter straight piping run. After these modifications were made, the pump was re-tested.

850 800 750 700 0

25

50

The Pump Handbook Series

75

100 125 Capacity (gpm)

150

175

200

TABLE 2 Description of Condition

Discussion of Applicability

1. Undersized throat of the volute or reduced interstage passage

This condition was not present when the casing was inspected. No build-up or blockage was found.

2. Ring clearance not to specifications

The rings were inspected and the clearance measured/restored to OEM recommendations prior to re-installing.

3. Obstruction in piping system

No physical evidence of obstruction was found. However, as mentioned previously, the discharge piping was reduced from 4 to 1-1/2 in. at less than 5 pipe diameters from the discharge nozzle.

4. Use of non-OEM parts

Inspection revealed all serialized OEM parts.

5. Errors in factors used to correct flows to STP

All corrections made to the data reflect actual conditions.

The results, shown in Figure 4, demonstrate a significant improvement over the previous performance. Throughput was increased by 30%, and due to system characteristics the additional flow was directly translatable to additional product to sell. This performance improvement was attained using a very small amount of business funds relative to the cost of replacing the pump. However, every system is not the same, and other systems may require additional investments of time and resources to produce similar results. Foremost, you must be familiar with the system, its operating mode and how it is controlled. ■

REFERENCES

FIGURE 4. PERFORMANCE AFTER MODIFICATIONS

1.

Yididiah, S. Y. Centrifugal Pump Problems, Causes and Cures. Petroleum Publishing Co., 1980.

3.

Cheremisinoff, N. P. Fluid Flow, Pumps, Pipes and Channels. Ann Arbor Science, 1982

1000

950

900

Luis F. Rizo is Reliability Engineering Manager for G.E. Silicones in Waterford, NY. He is also a member of Pumps and Systems User Advisory Team.

TDH (ft.)

850

OEM Curve Problem Report Field Test TDH After Mods.

800

750

700 0

25

50

75

100 125 Capacity (gpm)

150

175

The Pump Handbook Series

200

57

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PUMP AND SYSTEM TROUBLESHOOTING HANDBOOK

Vibration Amplification AT THE JOB SITE, 8 A.M. Unit 4 is off the line for a regular maintenance check. Bill and Mike, the maintenance engineers, take me to the installation and then to the shop where a completely assembled unit is ready for mounting. We discuss the maintenance records—not one 600 hp vertical turbine condensate pump has operated for a full year without a failure. Resonance is our first suspicion. The structures used for pumps and their mountings have natural frequencies that cause vibration and noise amplification when excited by a resonant frequency. My experience consulting on a case at Pacific Power & Light in Glenrock, WY helps to illustrate the excessive maintenance problems created by selfinduced pump resonance.

NOT SO MUSICAL A tuning fork will always resonate at the same tone, and same overtones, when struck. This is true of all rigid structures. Although natural frequency resonance is employed in musical instruments and in our own vocal cords, it must be avoided in rotating machinery applications. This contrast in attitudes toward resonance frequencies is evident even in the language we use to describe them. Frequencies referred to as the major tone in music are called the first critical for pumps. Likewise, the first overtone in music is the second critical in a pump structure. Table 1 lists some sources of vibration energy in pumps. Residual unbalance and impeller vane passing (impeller vane pulse when passing a cutwater guide or diffuser) are the two primary causes of exciting forces in centrifugal pumps. Spring mounting of reciprocating equipment is often selected so that the first critical natural frequency is only 10% of the operating rpm. However exciting

58

TABLE 1 Sources of vibration energy originating the pump or driver 1. Residual unbalance 2. Vane pass 3. Oil whip 4. Misalignment resulting in radial/axial shock 5. Motor windings 6. Worn frictionless bearings forces in centrifugal pumps are relatively low, and the natural frequency may approach 75% of an exciting force. As a rule, avoid the first critical by +/– 25% and the nth critical by +/–[(25)1/n]%. The following expression can also be applied:

Fn (.75)1/n ≤Rn≥ Fn (1.25)1/n where: Rn = frequency range to avoid for the nth critical Fn = n critical frequency MEASURING NATURAL FREQUENCIES Back at Unit 4, our first step was to measure the natural frequencies of the pump installation. To facilitate the test, we installed the pump with the discharge head set on a wood blocking about two feet off the can. This set-up gave us access to the column with a rubber hammer. The lowest critical frequency we could determine was the second. Table 2 lists the critical frequencies as deter-

TABLE 2 Measured critical frequencies-Unit 4 Critical Frequency # 2 3 4 5 6 7 8

Vibration Frequency (cpm) 600 1,800 3,600 6,800 11,500 18,000 30,000

The Pump Handbook Series

mined by our method. A log-log plot of the points confirmed the accuracy of the measurements. The third critical, at 1800 cycles per minute (cpm), confirmed our suspicions. This was the interfering frequency which had caused the 12” column suspension failures. These failures had occurred despite the plant’s strict observation of alignment precautions and the conversion of the intermediated shafts to a single larger shaft. Resonant frequencies can be displaced by altering the stiffness of the suspension assembly. For Unit 4, I suggested a three-armed spider between the can and the flange at the bottom of the top column, an arrangement that would yield an approximate 23% increase. When I made my recommendation to the people at Pacific Power & Light, I received a small dose of good natured ”static” for proposing a solution that seemed so insultingly simple. Clearly, these physical principles are not new to most pump operators, but often their application isn’t as obvious.

CONTROL OF CRITICAL FREQUENCIES Control of static structural, resonant critical frequencies, such as those experienced with Unit 4, is a field responsibility. Because the manufacturer is not generally involved in decisions pertaining to mounting rigidity and alignment across couplings, the field designer must provide the final defense against unnecessary amplification of vibrations. On the other hand, control of dynamic (rotating) structural resonant critical frequencies is a manufacturing responsibility. Because three-phase motor drivers are available in limited choices of rpm, trial and error over time eliminate most rotating velocities that might cause trouble. However, variable

Ed’s Rules of Thumb frequency drivers are now becoming quite common. These drives can create unforeseen and excessive vibration levels at ”inbetween” frequencies. Control adjustments may permit the blocking of a troublesome frequency, but the required gap often produces undesirable effects on pump characteristics. The use of inexpensive variable speed drives can create costly problems in the field, so a thorough evaluation of potential vibration sources is recommended before installation. Low rpm, heavy overhung trash pump impellers and a limited number of impeller vanes, coupled with a single diffuser/cutwater can produce an expensive coincidence of excitation and impeller assembly/ cantilever critical frequency amplification. Impeller assembly critical frequencies can be controlled only at the pump design level. However, I’ve discovered that the mathematical computations and empirical field checks on critical frequencies are not sufficiently close in either foreign or domestic pumps. Whether the volute is water filled or dry, will make little difference in measurement of an impeller assembly critical. A vibration analyst can easily make the determination either before or after shipment.

BREAKING CADENCE The pumps on Unit 4 had undergone seven years of high maintenance due to self-induced resonance previous to our diagnosis. But, resonant frequencies are not only damaging for pumps. For more than 2000 years groups of marching men have broken step when crossing a bridge to avoid the damage possible if the cadence of their marching might happen to match the critical frequency of the bridge span.

ABOUT THE AUTHOR: Ken Hawkins is the owner of Vibration Control, an independent consulting firm in Overland Park, KS. He has worked with pumps since 1950. ■

Editor’s Note: From time to time, the Shoptech section of Pumps and Systems will feature Pump Rules of Thumb from Ed Nelson, a noted pump consultant and member of our Editorial Advisory Board. These Rules of Thumb are derived from a blend of engineering principles and experience. They are intended to provide a rapid assessment of conditions while troubleshooting a pumping system problem, and may be used as a supplement to general or detailed instructions furnished by the manufacturer. Each machine is different in design or construction but the fundamental principles apply to all of them.

FACTORS IN SUCTION PIPING LAYOUT Suction piping can be the cause of major pump damage—especially with double suction impellers. To reduce these problems with the suction piping apply the following practices:

RULES OF THUMB 1. Employ suction piping one or two pipe sizes larger than the pump nozzle. Suction lines should never be smaller than the pump suction nozzle. 2. To prevent cavitation in the pump, suction line velocities should not exceed 10 ft./sec. Consider 5-6 ft./sec. as a maximum for new systems. 3. Consider the pressure drop across permanent suction strainers. 4. Install valve stems and tee branches perpendicular to, not parallel to, the shaft. 5. Employ an absolute minimum of five pipe diameters of straight run before the suction flange. Seven pipe diameters is the preferred minimum.

HOT SERVICE PUMPS—WARMING STREAM FLOWS A single stage, double suction pump should be warmed up from “cold” to “hot stand-by” by passing hot liquid from discharge to suction for at least two hours. The warming stream flow rates shown below are a minimum and will require adjustment for a specific installation.

RULES OF THUMB Pump Discharge Nozzle Size 4”-6” 8”-10” 12”-14” 16”-18” 20”

Operating Temperature vs. Warming Stream Flow Rate (gpm) 200°F 450°F 700°F 4 5 6 7 8

5 6 8 10 11

6 7 10 13 14

About the Author: William E. (Ed) Nelson is a turbomachinery consultant based in Dickinson, TX. He is the author of more than 50 technical papers and a contributor to several handbooks on pump operation and maintenance. Previously he spent 36 years with Amoco Oil in various engineering, materials management and maintenance positions. Mr. Nelson is a registered professional engineer in Texas. The Pump Handbook Series

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PUMP AND SYSTEM TROUBLESHOOTING HANDBOOK

Selecting a Pump? Define Efficiency First! By J. Robert Krebs, P.E., Contributing Editor

A

60

FLOW OUT Q2 P2 Q1 = Q2 = Q

POWER IN (PI)

DRIVER

PUMP COUPLING

EFFICIENCY (ED)

EFFICIENCY (EP) POWER IN TO PUMP (PIP)

POWER OUT (PO) ED = PO/PI HP = 0.746KW OVERALL EFFICIENCY = HHP/PI OVERALL EFFICIENCY = ED X EP

FLOW IN Q1 P1

HHP = Q (P2 - P1) / 1714 Q (GPM) P (PSI) EP = HHP/PIP

Figure 1. Power calculations — pump and driver

100

CENTRIFUGAL AND AXIAL FLOW PUMPS Over 10,000 gpm

90 Efficiency, percent

ll users want the highest efficiency from their pumps. This is why having an accurate definition of efficiency prior to selection is so important. An operating pump transfers fluid adding energy (power) to the fluid as it passes through the pump. The power developed by the pump (HHP) compared to the power to the driver (PI) gives the overall efficiency of the pump and driver. Figure 1 illustrates a pump and driver coupled to a system. The driver takes in power. It can be an electric motor, a liquid or gaseous fueled engine, a hydraulic or air powered driver. The driver has an efficiency that varies with its power output. The power to the driver (PI) divided by the power from the driver (PO) is the driver efficiency (ED). The coupling between the pump and driver and any other additional mechanical components, such as a gear or belt drive, also have losses and absorb some of the driver output power before it reaches the pump. Gear and belt drives and some coupling designs have a measurable power requirement. The component manufacturer can provide an estimate of the loss. To simplify this discussion, we will ignore losses between the driver and the pump. The power to the pump provides the fluid flow and pressure. The energy terms of flow and pressure combine to equal the power out. [Power from the pump divided by driver power to the pump is the pump efficiency.] The calculations for efficiency are indicated in Figure 1. Pump efficiency stated by the manufacturer refers to water at the pumped liquid at STP (standard temperature and pressure) conditions. An increasing

10,000 gpm

80 500 gpm 200 gpm

70

1000 gpm 3000 gpm

100 gpm

60 50 40 500

1000

2000

3000 4000

SPECIFIC SPEED Ns =

RADIAL-VANE

10,000 15,000

RPM √GPM H3/4

FRANCIS-VANE

MIXED FLOW AXIAL FLOW

Figure 2. Pump efficiency versus specific speed and pump size The Pump Handbook Series

viscosity of pumped liquid reduces the efficiency of all pump types—centrifugal more so than rotary types. Centrifugal pumps are used for pumping viscous liquids to about 3000 SSU (Seconds Saybolt Universal)—about 600 cSt (centistokes) on a routine basis. Methods are available for approximating the degradation effect of viscosity on head, capacity, and efficiency (the Hydraulic Institute is a source for this information). Rotary pumps (gear, screw, circumferential piston, and similar designs) when pumping a viscous liquid, given proper suction and application conditions, also have reduced efficiency. Since a more viscous liquid reduces slip, the flow rate and head developed in rotary pumps may be minimally affected. Manufacturers should be consulted for specific applications. Figure 1 also suggests that the driver—and pump—efficiency must both be considered to determine overall efficiency. In my experience, too frequently the driver efficiency is overlooked.

The overall efficiency is the product of pump and driver efficiency. If, at a certain pumping condition the pump is 80% efficient and the driver is 70% efficient, the overall efficiency is 56%. With the common electric motor driver, the overall efficiency is called the “wire to water efficiency” (WWE). The WWE is defined as the hydraulic horsepower (HHP) from the pump divided by the measured electric power to the motor—kilowatts (KW) expressed as horsepower (HP). So what is a good pump efficiency? For centrifugal pumps, I have long used the chart in Figure 2 (originally published by Worthington), which determines efficiency from pump size (flow) and impeller shape as predicted by specific speed. This chart teaches that a single stage (or the first stage) centrifugal pump becomes more efficient as flow increases for radial vane design impellers. Now let’s take a look at the chart (remember—low viscosity liquids only, such as water). For a pump rating of 200 gpm at 100 ft total head,

The Pump Handbook Series

operating at 3500 rpm, the specific speed calculates to 1565 and predicted efficiency is 70%. [Note the impeller design is radial vane.] If the conditions were 2000 gpm at 100 ft total head and 1750 rpm, the specific speed is almost 2500, and the estimated efficiency would be 80% or more. The impeller design is in the Francis vane area of radial flow design. Rotary type pumps are quite efficient, 70–80%, when pumping low viscosity liquids. They adapt well to pumping highly viscous liquids. Manufacturers should be consulted for specific power requirements. Being able to estimate efficiency quickly is useful in preliminary system design. If the liquid properties and flow and pressure are known, the driver size and pump size can then be approximated along with building sizes and power system requirements. ■ Till next time...

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Selecting a Pump: Will it Operate Where You Want it to? By J. Robert Krebs, P.E., Contributing Editor Positive Suction Head). Curve A was constructed with maxiou need one or two pumps for Another limitation is available mum static head and a pipe resistance an application. The system suction head (NPSHA). If the head is coefficient equivalent to aged (14–17 analysis is complete and overstated, the pump will run at B, years) pipe. As the pipe ages in use, you’ve arrived at a flow rate rather than A in Figure 1. Is there resistance to flow normally increases. and pressure (head). Let’s assume enough NPSHA? What horsepower Curve B represents new pipe at the pumps are constant speed motor is required? Low to medium specific the minimum expected static head. driven centrifugals and you, the speed design centrifugal pumps have Curve C is for the same static head as designer, have used a conservative an increasing horsepower requireB, but with the pipe aged as in A. approach to your calculations. ment with flow rate. If the system Assuming that the system design The key point—you have decidresistance is lower than calculated, calculations for head are accurate, ed the Q (gpm) and H (head) knowwill the pump require a higher sucand the pump speed is as shown on ing that the pump only operates at tion pressure or a larger motor? the curve, the pump will operate the intersection of its performance These can be big problems with with new pipe at the curve B intercurve with the actual system resiseven small pumps. How can this be section corrected for the actual static tance curve. controlled? head. As the pipe ages, the intersecIt is the rare system designer To start, the system designer must tion point will be between the curves whose calculations are so precise be very sure that all pump and piping A and C. In other words—there are a that the operating pump actually perelevations are correct. Next, the number of system curves with proforms at the predicted conditions. designer in his calculation of system gressively higher heads and comConsider a single source/single resistance should use both new pipe mensurately lower flows. destination open transfer system to a and aged pipe resistance factors. The How do the pumps and systems pressurized vessel. Both the source aged pipe factor must reflect the anticinteract? From the performance and destination pressures can differ ipated actual life of the system. Pump curve, as flow increases, so do brake from design. The centrifugal pump selection and motor (driver) sizing horsepower and required NPSH (Net motor driver speed varies slightly with increasing horsepower delivered (centrifugal pump CUBIC METERS PER HOUR Model test curves are usually 300 400 500 6312-3D made for a constant speed). 1750 R.P.M. 200 The pipe friction losses 12 60 40200 vary with type, age and 30 Impeller No. 8 condition of the pipe. The 12″ 50 Y-4677 160 20 NPSH-R pumping of viscous liquids, Number of Vanes 4 11″ slurries, paper stock and 10 A 2 40 C similar combinations is not Max. Sphere 120 0 0 B considered in these com3″ 10″ 30 ments. Discharge Size Figure 1 illustrates a 6″ 80 9″ 60 H 20 performance curve from a Suction Size P 6″ or 8″ manufacturer’s catalog. 50 H L P 40 H 20 H Inlet Area 40 P Super-imposed on it are a 30 H 10 P 25 H P 28.27 sq. in. L schematic of an open transP fer system and calculated 0 system head curves A, B 0 400 800 1200 1600 2000 2400 and C. Curve A intersects L= Limit Line U.S. GALLONS PER MINUTE the maximum diameter impeller performance at the design flow rate and head. FIGURE 1

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FEET

METERS

METERS

NPSH-R

FEET

%

74

77 76 % %

79 78 % %

TOTAL HEAD

60 64 % 68 % 71 % 74 % 76 % 77 % % 78 % 79 %

Y

The Pump Handbook Series

must reflect these concerns consistant with the anticipated life of the system. There are many other operating range problems that cannot be discussed in just one column. To mention a few: 1. What do minimum/maximum

flow rate lines on a performance curve mean? 2. How does one determine NPSHR (with flow)? 3. When do motors overload? 4. How accurate is the manufacturer’s estimate of functions

The Pump Handbook Series

(NPSHR–efficiency)? Each of these questions deserves a lengthy answer. Maybe we can address them later. ■ Until next time...

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PUMP AND SYSTEM TROUBLESHOOTING HANDBOOK

Performance Curve vs. System Curve By Phil Mayleben losses are present in this system. Velocity head losses occur when the area in the flow loop changes suddenly. In this simple arrangement the system head curve is zero when there is no flow. Figure 3 shows a similar system with a net change in elevation. In this case the pump must provide an additional amount of energy equal to the change in elevation before flow can begin. This difference in elevation, expressed in feet, is identified as hs and is normally referred to as static pressure. The “normal” system head curve

pump’s purpose is to transfer mechanical energy from a motor into hydraulic energy within a fluid. This energy is then dissipated as the fluid moves through a system. The pump’s head capacity (H-Q) curve defines how much energy is available at a given flow rate and shaft speed. A sample head capacity curve at 1800 rpm, supplied by the vendor, is shown in Figure 1. For convenience, the energy level or head is expressed in units of feet. A simple system is shown in Figure 2. Only friction and velocity head

A

shown in Figure 1 represents the energy required to overcome static pressure, friction and velocity head losses present in a system. The system head curve establishes the pump’s operating condition. The point where the system head curve and the pump’s head capacity curve intersect defines the point where the pump will operate at a given speed. This point is identified as A1 in Figure 1. For a pump to operate in a given system, it must be capable of delivering a head that is greater than the system static pressure (hs). Note in

100 90

Normal System Head Curve Throttled System Head Curve

80 C1 70

B1

Pump H-Q @ 1,800 RPM

HEAD IN FEET

A1 B2

60 Pump H-Q @ 1,600 RPM

A2

50

Friction (h1 - hs)

40 B3

Pump Best Efficiency Point

A3

30

Pump H-Q @ 1,200 RPM

20

Static Pressure (hs)

10 0 0

100

200

300

400

500

600

700 800 900 FLOW IN USgpm

FIGURE 1. Sample head capacity curve

64

The Pump Handbook Series

1000

1100 1200

1300 1400

1500

Figure 1 that if the pump’s shaft speed is reduced from 1800 rpm to 1600 rpm, the head and flow capability decrease. If the speed of the pump is reduced much below 1200 rpm, it will no longer be able to pump in this system because its head is approaching 30′, which is equal to the static pressure (hs). The pump Best Efficiency Point (BEP) occurs when the ratio of hydraulic power to input power is at a maximum. The BEP for the pump in Figure 1 occurs at 1000 gpm and 70′ at 1800 rpm (point B1, and then drops to about 667 gpm and 31′ at 1200 rpm (point B3). Note that at 1800 rpm the operating point (A1) occurs at a flow that is greater than the pump’s BEP (B1)). As the speed is reduced to 1600 rpm, the BEP (B2) and operating point (A2) are almost identical. By the time the speed is reduced to 1200 rpm, the operating point (A3) is significantly to the left of the pump’s BEP (B3). The system head curve should be carefully calculated by the system designer. If it is in error, the pump may not operate at the intended duty point. When possible, pumps should be selected so that the BEP is close to the system head curve. In Figure 1 this occurs at about 1600 rpm. Operating a large, high power pump too far away from its BEP can cause pump damage, excessive wear and high vibration levels. The risk of damage is reduced as speed is decreased. The location of the pump’s BEP, and recommendations on how far away from BEP one can operate a pump safely, should be supplied by the pump vendor. If a pump operating point based on Q1, H1, N1 and HP1 is known, then a new operating point at a new speed N2 can be estimated. The following relations, known as the affinity laws, can be used to step known pump performance to a new speed. Flow: Q1/Q2 = N1/N2 Head: H1/H2 = (N1/N2)2 Power: HP1/HP2 = (N1/N2)3

VELOCITY HEAD LOSS FLOW

DISCHARGE VALVE SUCTION VALVE FIGURE 2. Simple pump system

hs

FLOW

DISCHARGE VALVE

SUCTION VALVE FIGURE 3. Pump system with elevation change

In these equations Q is the pump flow rate, H is the pump head, N is the shaft speed, and HP is the power required by the pump. Subscript 1 refers to known values, and subscript 2 applies to calculated values. The curves at 1600 and 1200 rpm in Figure 1 were calculated from the 1800 rpm curve by using these equations. It is also possible to calculate a system head curve similar to Figure 1 if a single flow-head point is known. At point A1 in Figure 1, the total system head (h1) is 67′. The static pressure (hs) is 30’. The frictional part of the system head curve (h1 - hs) is 37′. Because frictional and velocity head losses are proportional to the The Pump Handbook Series

square of the flow rate, an equation for calculating a new point (h2) on the system head curve can be written as: h2 = (q2/q1)2 x (h1 – hs) + hs Note that lower case letters are used for q and h in this equation so that the system head and flow values are not confused with pump parameters. The equation assumes that hs is constant and would not apply to a system with variable static head. For an example of variable static head, consider a system in which the pump is used to empty a tank. Because the elevation of the water feeding the pump is constantly dropping as the tank is drained, the whole system head curve increases

65

as the pumping proceeds. The rate of change in hs depends on the geometry of the tank. The most efficient way to decrease the flow delivered by a pump is to reduce the shaft speed. This is easy if some sort of variable speed control device is available. Unfortunately, many pump installations are driven directly by constant speed AC induction motors. In this situation the only way to adjust the capacity is to increase the system head against the pump artificially by throttling a valve on the discharge side of the pump. While this addresses the need to reduce flow, it also wastes energy. Referring again to Figure 1, it is possible to force the pump’s operating point from A1 to point C1 by partially closing a valve on the discharge side of the pump. The pump then produces a head of about 71.5′ at 910 gpm in front of the discharge valve, but the difference in head between points C1 and A2 is lost across the partially closed valve. Thus, only the energy available at point A2 reaches

66

the system. Although this flow control method works, it is comparable to turning your home heating system on full blast and then regulating the inside temperature by opening the windows! It should be reiterated that many large high-power pumps cannot be operated at low flows without sustaining damage. The following additional precautions should be observed when pump speed changes are anticipated in a system. First, be aware that high vibration levels could develop at certain speeds if you are unfortunate enough to encounter a natural frequency. Vibration problems due to a natural frequency are more common in vertical pump units. If the pump vendor is made aware of the need for variable speed pumping capability when the equipment is purchased, he can offer suggestions to avoid any possible natural frequency problems. If anticipated speed increase is necessary, one must determine that the pump is adequately designed to withstand the increased loads, and

The Pump Handbook Series

that sufficient horsepower is available from the pump driver. Notice in the above equations that horsepower increases with the third power of the speed ratio. An increase in speed also requires that an adequate suction pressure or NPSH (net positive suction head) margin is still available for the pump. Inadequate NPSH can result in significant loss of performance as well as noise, vibration and impeller damage. Usually, pump NPSH requirements can be expected to increase with the square of the speed in the same manner as head. Again, the pump vendor should be able to assist in these areas. ■

REFERENCES 1. Stephen Murphy, “Variable Speed Pumping,” Pumps and Systems Magazine, April 1993. 2. A. J. Stepanoff, Centrifugal and Axial Flow Pumps, 1957, John Wiley and Sons, Inc. Phil Mayleben is employed by ITT A-C Pumps (Cincinnati, OH).

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PUMP AND SYSTEM TROUBLESHOOTING HANDBOOK

Calculating Shaft Deflection By J. Robert Krebs, P.E., Contributing Editor he article on shaft stiffness factor calculations by Dan Besic in the February Pumps and Systems caught my attention. He is “right on” in considering all of the factors in the shaft stiffness formula when comparing the relative shaft stiffness of two or more similar pumps. It is understood that the deflection formula for a simple cantilever beam (shaft) is y=FL^3/3EI or since inertia I=πD^4/64 for a round shaft and the modulus for steel is E=28 to 30×10E6psi (I use the larger number), then y=FL^3/44.20D^4 (y in inches ×10E–3 or mils). This formula does not account for the overhung

T

shaft weight and the shaft length and diameter between the bearings, which also contributes to shaft deflection in overhung impeller design (end suction and double suction) pumps. The shaft stiffness factor is a convenient short-cut for comparing the relative stiffness of two pump shafts provided the materials of construction and the other design criteria, such as shaft diameters, are identical or essentially similar, as in ANSI pumps. Use of the shaft stiffness factor has become popular for comparing two ANSI pumps of the same size. Perhaps a review of the use of shaft

RPM

x

Stuff box face

P

R2

L2

L1 D1

D2

Y R1

P L1^2 Y= 4420

L&D P Y

[

L1

L2 +

D1^4

D2^4

]

L1 - x Yx =

Y

(approx.) L1

Inches Pounds Inches x 10E-3 or Mils.

FIGURE 1. Calculating shaft deflection The Pump Handbook Series

stiffness factor will be sparked by recent advent of several so-called super ANSI designs. For more than 40 years in the pump business, I have used a simple formula that includes both bearing spans and their shaft dimensions, along with the Hydraulic Institute recommended radial thrust factor to calculate shaft deflection. As chairman of the HI Technical Committee on radial thrust, I oversaw original publication of this work in the Hydraulic Institute Standards 12th Edition (1969). The latest HI Standards (1994) has expanded this work (pages 103-105) and is the source I now use for this discussion on shaft deflection. Today, most pump manufacturers use a modified finite element analysis (FEA) approach for the shaft deflection calculation. The formula I use (Figure 1) produces a slightly lower deflection than the FEA method. I calculated the maximum (zero flow) shaft deflection at the impeller centerline for two manufacturers’ 4×3×10 ANSI pumps at 3500 rpm using the formula in Figure 1. Pump manufacturer D-1 calculated to 2.6 mils and for manufacturer D-2 a value of 1.8 mils. Since the face of the stuffing box is about one-half the distance from the centerline of the impeller (maximum deflection) to the inboard bearing (zero deflection), the relative values are about 1.3 and 0.9 mils respectively at the face of the stuffing box. Manufacturer D-2 published data at the face of the stuffing box is 1.1 mils. For horizontal shaft pumps the weight of the impeller must be added to the calculated radial thrust load. In this example, the impeller weight of 13.6 lb increases deflection by almost 20%. Does giving effect to the shaft length and diameter between bearings make a substantial difference in shaft deflection? I estimated the max-

67

imum radial thrust for the 4×3×10 at 71 lb; adding this to the 13.6 lb impeller weight increases the stiffness factor deflection to 1 mil for the stainless steel pump in the February article or exactly one-half of the total deflection calculated from the formula in Figure 1. I would submit that using a deflection formula that includes shaft dimensions between the bearings is justified. What is not understood sometimes by the pump user is the importance of specifying maximum shaft deflection on a purchase request or request for quote (RFQ). The conventional volute casing

68

centrifugal pump has the highest radial thrust load; therefore, it has the largest shaft deflection at zero flow and practically zero radial thrust load at the best efficiency pump flow. Excessive shaft deflection, more than 2 mils at the face of the stuffing box (packing or seal gland), will shorten packing or seal life, subject the shaft to cumulative reverse bending (fatigue) stress, that could cause shaft failure and increase the bearing load with a commensurate reduction in bearing life while elevating inboard bearing operating temperature. So what’s the answer? Controlling shaft deflection and bearing

The Pump Handbook Series

loads to appropriate levels is, in my opinion, the most important factor in assuring a quality pump mechanical design. The RFQ to the supplier should ask for calculated shaft deflection and bearing life calculations for the pump proposed at the operating condition (or an operating head range). If you want to put it into more precise and detailed specifics, drop me a line. I will send the details so that you can properly specify shaft deflection and bearing life to assure a quality pump mechanical design. ■ Until next time ...

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PUMP AND SYSTEM TROUBLESHOOTING HANDBOOK

Pumps in Parallel with Variable Speed Drive By J. Robert Krebs, P.E., Contributing Editor

L

range to avoid problems. If operated below the minimum recommended flow, increased noise, vibration and recirculation, as well as increased radial thrust with accompanying increased bearing loads, will assure increased bearing operating temperatures and shorter bearing life. We could also mention increased shaft deflection with shorter seal life or increased packing leakage and shaft breakage. Figure 2 illustrates the maximum and minimum speed of one of two identical pumps and the system head curve against which they must operate at some speed between minimum (N1) and maximum (N4). The control system is programmed to start a second pump, also variable speed-driven.

There are two control-operating options, lead-lag or load-share. In lead-lag, the speed of the second or lag pump is speed-adjusted by the control system to handle only additional flow from the source. In Figure 2, if the lead pump is operating at speed N3 when the lag pump is activated, the latter will increase in speed to some speed, say N2, at which point it will start pumping the excess flow that the lead pump cannot handle. The curve A represents the performance of the lead pump at speed N3 plus the lag pump at speed N2. The horizontal line notes the flow from the lag pump, also noted as QA, less than the minimum flow line. This is one of the dangers of lead-lag operation. In this example, in the lead-lag

A B C Head Ft.

ast August this column discussed pumps operating in parallel. Jim Marean of Fluid Kinetics in the Buffalo area asked me to introduce variable speed into the discussion of parallel operation. I also received several calls requesting amplification of minimum and maximum flow restrictions for centrifugal pumps and how they affect the pump. These are inter-connected questions. This discussion will concern itself initially with two (or more) identical centrifugal pumps operating from the same source, with each motor driver equipped with a variable speed frequency drive control. Figure 1 illustrates two identical pumps, each operating at full speed. Note that when the two pumps operate in parallel, each one produces less flow, but together the two produce more flow than one pump operating by itself. Figure 1 shows the effect of a high friction loss system “A” on the total flow of two identical pumps in parallel. Note that as the design system friction loss decreases (system B & C), the contribution to total flow from adding a second pump increases. What is illustrated in Figure 1 is applicable to all centrifugal pumps. When special impeller designs, such as non-clog and solids handling, are considered, operation at reduced flow must be carefully analyzed. The manufacturer may recommend a minimum flow in these designs as a percentage of capacity at Best Efficiency or incorporate a minimum flow line on the performance curve. Minimum and maximum flow lines, when noted on performance curves by the manufacturer, restrict the pump flow to that operating

Pump 1 + 2

Pump 1

Contribution of Second Pump to Total Flow

Flow Q FIGURE 1. Two pumps in parallel The Pump Handbook Series

69

Min. Flow Line

Head Ft.

System Resistance

Max. Flow Line A N4 B N3 N2 QA

N1

QB

A = Lead-Lag B = Load-Share

Flow Q FIGURE 2. Variable speed pumps in parallel—lead-lag and load-share operation

mode, the lag pump must have a minimum speed in excess of N2 to assure minimum flow in excess of the minimum flow line. Of course, the same restriction applies to the lead pump. That is, the control system must recognize the minimum variable speed setting for each pump to assure operation in excess of the minimum flow value at each speed. Some lead-lag systems operate with the lead pump attaining maximum speed before a lag pump is started. In those cases, most often the lead pump continues to operate at maximum speed while the lag pump varies in speed to pump the excess flow. The same comments apply –

70

the lag pump speed must assure flow in excess of the minimum flow line at all speeds. If the pump performance curve being considered does not show a minimum flow line, question the manufacturer. For non-clog designs, both radial and mixed-flow impeller, there is a minimum flow. Load-share is my preferred method of operation. Referring again to Figure 2, at some selected speed (N2), the lag pump joins the lead pump in operation. Both pumps will operate at the same speed to maintain the source signal control flow. The curve B in Figure 2 represents the performance of both pumps at speed N2. The flow from each pump is QB – outside the

The Pump Handbook Series

minimum flow line. If the source signal calls for more flow, both pumps respond with increased speed to supply the flow. If less flow is required, both pumps slow down. At some lower speed, when both pumps are approaching the minimum flow, one pump will shut off and the remaining unit will adjust its speed to the flow signal. At station design, pump size selection and load-share may be used to operate the station pumps at their optimum efficiency, minimizing power costs. Where considerable variation in flow occurs at a pumping station, more than one size pump may be installed. It is important that the parallel operation of unequally sized pumps, whether variable or constant speed, be checked out to assure their proper operation in parallel. What about maximum flow lines on performance curves? Flow substantially beyond the Best Efficiency Point capacity, commonly called a run out condition, introduces increased NPSHR, may require additional horsepower and frequently produces substantial noise and vibration. Centrifugal pumps should not be operated at severely restricted or run out conditions. Depending on the design, specific speed and type of pump, an operating range of 25-110% of BEP capacity is a rule of thumb for smaller, say through 8″ discharge size, radial flow impeller design pumps. As pump size increases, the allowable operating flow range from BEP becomes more restricted. For mixed-flow design impellers, the operating flow range may be from 45105% of BEP capacity for 14″-18″ and larger discharge sizes. ■ Until next time ...

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PUMP AND SYSTEM TROUBLESHOOTING HANDBOOK

Sizing Pumps for Complex Systems: Part I By J. Robert Krebs, P.E., Contributing Editor dence in using an expert program. In Figure 1 the problem is sizing pumps 1 and 2. The variables are flow rates (Q1 and Q2), the discharge pressures (P1 and P2), the destination pressure (P3), and the liquid elevation differences (and pressure for a closed system). Z1, Z2 and Z3 elevations are from a datum reference elevation (normally the shaft centerline of one of the pumps). Pressures P1 and P2 are also a function of the proper pipe size and length for pipes 1, 2 and 3, and their corresponding friction losses (F1, F2, F3). The suction side friction losses are assumed negligible in this analysis. Clearly, source pressures, if above atmospheric, must be included. There are three possible operating conditions—(A) both pumps operating or off, (B) #1 on and #2 off, and (C) #1 off and #2 on. Calculating the pump discharge pressures, P1 and P2, at given flow rates, Q1 and Q2, is the first step in sizing a pump. Variations in pressure and or flow rate provide the pump(s) operating head range, important criteria in the final selection. Using the information shown in Figure 1, we will solve for the pressures P2 and P1. For example, the pressure at P2 will be the algebraic sum of destination pressure (P3), friction losses (F3 and F2) and the differential elevation (Z3–Z2), all expressed in the same units (ft of liquid flowing or psi). This can be

or the pump system designer, the rare case is an open transfer, single-source single-destination system. If only real life were always that simple! In practice, a multi-source single-destination system such as the one shown in Figure 1 is common. The system is descriptive of a blending or injection system, where source 2 is adding a predetermined amount of material to the source 1 flow. Some examples would be pH balancing, polymer feed addition or the introduction of blending agents. Pumps in these services can be centrifugal or positive displacement or a combination, as determined by flow and pressure, and obviously, either source flow can vary. For the experienced system designer, the problem in Figure 1 is familiar, but I will use it in explaining a manual approach that can be applied to solving difficult problems, such as those that follow in later issues of this column. The manual solution is a desired method in this instance. Expert pipe network programs fast become a more expedient approach as the system complexity increases, but I have always liked the cozy feeling of knowing how to solve the problem manually. Computer programs are only as good as the data supplied and the ability of the user to analyze the results. Perhaps learning how to approach the solution of a complex system problem will increase the less experienced designer’s confi-

F

Pump 2 Q2

Z2 P1

Z1

P2

2 1

3 P3

Pump 1 Q1

Z3

Pipe 1-300 ft Q1 = 200 GPM Pipe 2-100 ft Q2 = 50 GPM Pipe 3-300 ft P3 = 30 PSI (69.3ft) Liquid water at 60°F

FIGURE 1. Multi-source single destination system

expressed as follows: P2=P3+F3+F2+(Z3–Z2) P1=P3+F3+F1+(Z3–Z1) F3 calculated at Q1+Q2 flow For the three conditions A (both #1 and #2 pumps operating) use the expressions above B (#1 pump on, #2 off) P2=P3+F3+(Z3–Z2) P1=P3+F3+F1+(Z3–Z1) F3 calculated at Q1 flow C (#1 pump off, #2 on) P2=P3+F3+F2+(Z3–Z2) P1=P3+F3+(Z3–Z1) F3 calculated at Q2 flow Calculating the pressures for the three conditions gives the operating pressure ranges for pumps 1 and 2. Normally, the pipe lengths and flow rates (or flow ranges) are fixed by the system design. The engineer can adjust pump design pressures by changing pipe size. This can simplify pump selection by reducing the pump(s) operating head range for the various operating conditions. In Figure 1 we have listed the pipe lengths and Q1 and Q2 set flows. Consider Z1, Z2 and Z3 all zero. (There is just enough pressure to get liquid to pumps 1 and 2 at their set flow rates.) Why not determine the optimum Schedule 40 pipe size to minimize the variation in pump pressure for the three conditions of operation? If you have an expert program, you can verify your manual calculations. Everybody seems to have access to the Cameron Hydraulic Data book, so we can use its friction loss numbers for new pipe. I will discuss the results briefly in next month’s column, when we will also look at another, more complex problem. ■ Until next time…

The Pump Handbook Series

71

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PUMP AND SYSTEM TROUBLESHOOTING HANDBOOK

Sizing Pumps for Complex Systems: Part II By J. Robert Krebs, P.E., Contributing Editor ontinuing the multi-source single-destination discussion begun last month, we will study the problem of four pump stations pumping to a single destination system as shown in Figure 1. Going from a two source to a four source single-destination system greatly complicates the determination of individual pump discharge pressures when one or more pump stations (PS) are not operating. The system illustrated is a common problem in wastewater pumping. For example, perhaps stations 2, 3 and 4 were added to the original PS1 with discharge line pipes 1, 3, 5 and 7. The addition of each PS affects the operation of the PS already on line and may require equipment or piping changes with each additional PS. The method used in last month’s column will again be employed to calculate the pump discharge pres-

C

An important consideration— the friction loss in any line is calculated as the sum of the flows in that line. For instance, F7 would be calculated as the sum of flows Q1+Q2+Q3+Q4 if all stations were operating, and for the same condition F3 would be calculated as the sum of flows Q1 and Q2. As an example, consider pump 2 out of service and pumps 1, 3 and 4 operating. The friction losses F3, F5 and F7 would be calculated without the flow Q2, and the pressure at P2 would be the pressure at the intersection of pipes 2 and 3 (P5+F7+F5+F3+(Z5-Z2) with Z2 corrected to pump 2 centerline. It is obvious that to balance the flow and size pipes for this example is much more complicated than for the two source system. I use a simple expert pipe network program for this type of problem. When I entered into the computer the flow values, pipe sizes and lengths

sures manually for the four source single destination system. Recall that in the two-source, single-destination system, there were two basic equations to calculate the discharge pressure of the two pump sources. With a four source singledestination system, there are four equations, one for each pump discharge pressure. EQN1 P1 = P5+F7+F5+F3 +F1+(Z5-Z1) EQN2 P2 = P5+F7+F5+F3 +F2+(Z5-Z2)* EQN3 P3 = P5+F7+F5+F4 +(Z5-Z3)* EQN4 P4 = P5+F7+F6 +(Z5-Z4)* * The pump discharge pressure is to be corrected to pump centerline.

P2

P4 (2)

PUMP 2

Z2 P1

Z1

2

Z4

6

PUMP 4

(4)

1

3

(2)

(2.5)

(6)

5

7

(6)

P5 Z5

PUMP 1

Z3

4 P3 (3)

PUMP 3

FIGURE 1. Multi-source single destination system

72

The Pump Handbook Series

Pipes 2 & 4 = 100 ft Pipes 1, 3 & 6 = 200 ft Pipe 5 = 400 ft Pipe 7 = 500 ft Liquid Water 60°F (2.5) = Pipe Size Inches

Q1 = 50 GPM Q2 = 100 GPM Q3 = 200 GPM Q4 = 300 GPM P5 = 69.3 ft

160 ft. 120 ft.

P2 Pump 2 Q2

2 - 100 ft - 1.5 IN

Head

Z2

PUMP CURVES

PUMP 1

80 ft. PUMP 2

P1 Z1

1 - 300 ft - 3.5 IN

Pump 1 Q1

3-300 ft - 4.0 IN

40 ft. P3

Z3

Pipe 1-300 ft Q1 = 200 GPM Pipe 2-100 ft Q2 = 50 GPM Pipe 3-300 ft P3 = 30 PSI (69.3ft) Liquid water at 60°F

FIGURE 1. Multi-source single destination system

shown in this example, I saw that the results using four pump curves I selected closely matched my manual calculations using new pipe friction factors. Try it! If you have any questions, let me hear from you. The last two columns have focused on open transfer systems. Next month we will try a more complex closed system. Incidentally, I would like to thank the many readers who called and wrote for more details on specifying shaft deflection and bearing life after reading the May article. ■ Until next time ...

JULY ISSUE—PART I SOLUTION Using the two equations for manual solution of the pump discharge pressure requires pipe size selection for the three pipes. To calculate the friction losses, Figure 1 gives the pipe sizes I chose. With pump flow set at the given value, if pipes 1 and 2 are sized so that their friction losses are approximately equal, the effect on total head with either pump out of service will be minimal. With the pump curves used (see Figure 1B), the effect on head of either pump working versus both working was 2-4%. Capacity varied 50-60 gpm and 230-237 gpm for

The Pump Handbook Series

0 ft. 0

100

200 300 Flow Rate GPM

400

FIGURE 1B. Pump curves

pumps 2 and 1 respectively. In the actual system design, centrifugal pump performance curves would be used. Pump curves with similar characteristics should be chosen. For example, if a pump is selected that has a drop in total head of 50% from shutoff (zero flow) to best efficiency flow, then other pumps should be similarly selected. Also, it is better to select the larger flow pumps before smaller flow units. The pump curves I chose show the effect of one and two pump operation.

73

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PUMP AND SYSTEM TROUBLESHOOTING HANDBOOK

Sizing Pumps for Complex Systems: Part III By J. Robert Krebs, P.E., Contributing Editor ontinuing the discussion on complex systems in this column, we will examine closed systems. Recall there are two basic pumping system types – open or transfer systems and closed or circulating systems. Either type may have single or multiple sources and single or multiple liquid destinations. The least complicated closed system would be a single source pumping to a single destination and returning all of the pumpage to the source. An example might be pumping liquid from a controlled temperature source to a destination where some of the liquid thermal energy is absorbed or dissipated and the liquid is then returned to the source. The liquids used in such systems are often specially formulated heat transfer liquids. One well known application is chilled water systems in commercial buildings and similar equipment used in industrial settings. Adding two destinations will make the closed circulating system just described more complex. Figure 1 illustrates such a single source multidestination closed or circulating pump system. We will consider the destinations A, B and C identical. They could be batch process or continuous flow vessels, each with a flow control valve (FCV) to regulate the flow rate to the process energy requirement. A relief valve (RV) bypasses excess flow at a set pressure. I realize other accessories are needed to complete the system. However, the system in Figure 1 will serve our purpose for this discussion. The question is how to size a pump (or pumps) to provide a correct supply of liquid. The designer would start with the maximum energy requirement (BTU/hr), the thermal capacity of the liquid, the allowable temperature gradient to develop flow rate. Next is the question of how to calculate the pressure required for the pump(s). The heat transfer liquid manufacturer would provide the liquid physical properties and pipe pressure loss charts to facilitate the total head calculation. To simplify the sys-

C

74

tem calculations, if you wish to try a sample problem, assume the liquid is water. Remember that in these system designs our objective is to establish a method, if possible, to calculate the total head requirement manually. As the systems become more complex, network software is the only way to play “what if” and establish the boundary conditions quickly. Referring to the figure, assume that only destination C is receiving liquid. The pressure gauge P1 would then read the sum of the pressure drops from pipe friction through the system at the flow rate of the pump, which will include bypass flow pressure from the relief valve (RV) or P1=F11+F10+F9+F8+F7 +F6+F3+F2+F1 for only destination C operable. No elevation head differentials are considered. The same formula arrangement would apply to any single destination operation by incorporating the appropriate pipes. For a manual calculation for two, three (or more) destinations that are identical, I have found that adding the boundary friction losses and taking the average of the destination losses closely approximates the results from an expert network software program. If all three identical destinations

were operating, the pump gauge pressure would read the sum of the boundary friction losses plus the average losses across the three destinations plus any pressure drop across an operating relief valve. P1=F11+F10+F9+F8+F7 +[F6+F5+F4]/3+F3 +F2+F1 for three identical destinations operating. If the destinations are substantially different in friction loss, using the maximum loss unit would produce the greatest head requirement. I hear a chorus of voices rising, saying that this is an excellent application for a variable speed pump, and it is. With constant speed operation, the RV would be bypassing substantial flows with only one destination active or with one or more destinations operating at less than maximum flow. The pump is operating at a constant flow and pressure or a fixed horsepower. A variable speed pumping system, with the motor speed control sensing the total FCV requirement, would permit the pump to operate over a speed range based on demand, and it would decrease the pump power required. ■ Until next time ...

9

10

8

7

4 RV

6

5

A

B

C

P1 FCV 11 1

2

Q

FIGURE 1. Single source multi-destination closed system The Pump Handbook Series

3

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PUMP AND SYSTEM TROUBLESHOOTING HANDBOOK

Pumping with Air By J. Robert Krebs, P.E., Contributing Editor ir is used as the motive force in air motor driven pumps, some forms of diaphragm pumps, pneumatic ejectors (which are enjoying an increased popularity for laboratory, blending, drainage and sampling systems) and in air lift pumps. The special air lift pump we will discuss in this month’s column has been around a long time. Some of the earliest designs were for pumping water from wells. Figure 1 shows the general design. A vertical pipe extends into a liquid reservoir. The pipe is fitted near

A

the bottom with a diffuser mechanism that will let a compressed gas (air) into the liquid column. The gas bubbles will rise through the static liquid column in the pipe, displacing liquid. The resulting gas-liquid mixture has a specific gravity (specific weight) less than that of the liquid. The liquid-gas mixture rises in the pipe in proportion to the submergence of the pipe in the reservoir. When the reservoir submergence head exceeds the static lift and pipe friction losses, the liquid is discharged from the system and

AIR AIR UNDER PRESSURE DISCHARGED LIQUID

STATIC LIFT

SUBMERGENCE

AIR DIFFUSER

FIGURE 1. Air lift pump The Pump Handbook Series

the spent gas is wasted. The energy level of the discharged liquid is the potential energy of its elevation above some destination to which the flow may be directed by gravity. In the most common application of lifting water, the gas would be air—supplied at an appropriate pressure and quantity to pump the desired volume of liquid at the desired rate. Looking at the air lift pump design, it is clear that there is the advantage of no moving parts. Also, the air is compressed at a location far removed from the pumping action, where equipment can be protected as required. Liquids containing solids, slurries and abrasive solids such as sand or mine tailings move freely through the pump without problems. The simple design can be made in almost any shop. Laboratory and process applications are an interesting area where the pump might be used. Common applications in smaller waste treatment plants include pumping return and waste activated sludge. In digesters the mixing of biological anaerobic sludge can be accomplished with a special air lift pump, where the gas used is methane under pressure. The methane is produced by the process. Sequencing batch reactors (SBR) and other treatment processes also have applications. Some additional applications include drainage of wetlands and the obvious transfer of water to a discharge point. Note that the applications mentioned all involve low heads and relatively high volumes. In matching the product to the application, one must consider product and system constraints. In my opinion, the best application area is for relatively high flows and low heads, where the efficiency and con-

75

trol is best. Also, the solubility of the gas in the liquid must be considered. This could be a plus or minus in various lab and process applications. Working with static lifts of 4-6 feet with 65-80 percent submergence will give a pump design of reasonable efficiency. Pipe friction losses that negate the submergence effect should be minimized. Something should be said of the disadvantages of the air lift principle. Flow control can be difficult. The pump percent submergence design and head, and the air pressure and diffuser design, will determine how well the flow can be controlled, though varying the flow accurately will still be a problem. The gas bubble size is dependent on the dispersing method. The most common design is a series of holes around the periphery of the pipe with a collar covering the holes and the air pipe connected to the collar as

76

shown in Figure 1. Bubble size is important because the bubble rise rate increases with bubble size, as does the drag force on bubbles rising through the liquid. As the liquid-gas column accelerates to some steady state condition, the relative motion between the bubbles and the rising water column will decrease remarkably. The bubbles’ internal pressure will decrease as the liquid-gas mixture rises. This will increase the bubble diameter and the mixture velocity. This natural action will help explain the problem of flow control. This is a brief review of a useful product. The literature sources I checked presented minimal information on pipe sizing, flow rates and pressure. It seems that the equipment manufacturers that make pumps of this type for their own use are the custodians of the design information. From the limited data found, I

The Pump Handbook Series

was able to formulate an application, sizing and design approach. Before I publish this approach, I would like to confirm the theory against practice. I would ask readers of this column to pass along their design criteria. If you do not want your name mentioned when I acknowledge your contribution, please say so. In return, I will attempt to consolidate the data to provide a general design form for air lift pumps. ■ Until next time... To send or fax information regarding the design criteria; Pumps and Systems Magazine Attn: Bob Krebs 123 N. College Ave., Suite 260, Ft. Collins, CO 80524 Fax (907) 221-2019

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PUMP AND SYSTEM TROUBLESHOOTING HANDBOOK

Troubleshooting Centrifugal Pumps ffective troubleshooting should be an integral part of any plant’s equipment reliability and maintenance program. Before you examine the system, however, take time to review maintenance records. The performance history of your pumps will point you in the right direction. Also, pump operators can provide clues to what is happening. Troubleshooting should focus on these areas of investigation:

E

1. THE FOUNDATION Poor foundations, grouting and baseplate design often cause problems.

2. DRIVER Vibrations of the driver (motor, steam turbine, gearing) can be transmitted to other components.

3. MECHANICAL POWER TRANSMISSION

pling hubs, can also be transmitted to other parts of the pump and/or system. Users should note any incorrect positioning of the driver and pump such that the distance between shaft ends (DBSE) exceeds the axial flexing limits of the coupling.

4. THE DRIVEN PUMP Pump design has a major influence on the hydraulic interaction between the rotor and the casing and consequently the problems encountered. Misconceptions about pump thermal-growth can create problems.

5. SUCTION PIPING AND VALVES Improper design and layout of suction piping and valves can create flow disturbances such as cavitation, intake vortexing or suction recirculation.

6. DISCHARGE PIPING AND VALVES

Excitations from the coupling area, especially because of driver misalignment or eccentrically bored cou-

Unfavorable dynamic behavior of piping resulting from loads traceable to dynamic, static or thermal causes (including resonance excita-

tion) can create trouble.

7. INSTRUMENTATION FOR CONTROL OF PUMP FLOW Pressure pulsations can result from control system-pump interaction during start-ups, periods of low flow and valve changes.

8. MAINTAINING ALIGNMENT Once the alignment is established, dowels into the baseplate must hold the pump in alignment. Troubleshooting centrifugal pumps begins with observing operating conditions at the site. While a myriad of problems can exist with any pumping system, here are some of the more typical scenarios, possible causes and corrective actions. ■ Ed Nelson has more than 40 years of experience with industrial pumping systems as a former end-user at a major petrochemical company, and today as a turbomachinery consultant. He is a member of the Pumps and Systems Editorial Advisory Board.

Cavitating-type Problems A cavitating sound is heard in a pump that does not normally cavitate, and it is not clear whether it is pumping into the system.

POSSIBLE CAUSES

CORRECTIVE ACTION

The suction piping layout is poor. There may be too many ells in too many planes or not enough straight run before the suction flange of the pump.

Redesign piping layout, using fewer ells and laterals for tees, and have five or more diameters of straight pipe before suction flange.

Suction piping configuration causes fluid to rotate adversely when approaching impeller.

Install enough straight run of suction piping, or install vanes in piping to break up prerotation.

Flow rate is high enough above design that NPSHr has increased above NPSHa.

Reduce flow rate to the level for which the pump was designed.

Pump is operating at a low-flow, producing suction recirculation in the impeller eye, resulting in a cavitation-like sound

Install bypass piping back tosuction vessel to increase flow through pump. Note: Bypass flow may have to be as high as 50% of design flow.

The suction screen is clogged.

If screen is present, remove and clean it.

Piping gaskets with undersized IDs have been installed—a common problem in installations of small pumps.

Install properly sized gaskets.

Pipelines are constricted because of buildup of corrosion materials.

Replace deteriorated pipe.

The Pump Handbook Series

77

Capacity-type Problems Pump does not have enough capacity. No significant noise is coming from it.

POSSIBLE CAUSES

CORRECTIVE ACTION

Pump suction is below atmospheric pressure, causing air leaks into system.

Eliminate air leaks with appropriate actions.

Wear ring clearances are excessive (closed impeller design).

Overhaul pump. Replace wear rings if clearance is about twice design value for energy and performance reasons.

Impeller-to-case or head clearances are excessive (open impeller design).

Reposition impeller for correct clearance.

The discharge block valve is partially closed.

Open valve completely.

Any of the following conditions may have increased friction in the piping to the discharge vessel: 1. Gate has fallen off the discharge valve stem. 2. Check valve spring is broken. 3. Check valve flapper pin is worn, and the flapper will not swing open. 4. There is collapse of lined pipe. 5. The control valve stroke is improperly set, resulting in too much pressure drop.

Do the following: 1. Repair or replace gate valve. 2. Repair valve by replacing spring. 3. Overhaul check valve. Restore proper clearance to pin and flapper bore. 4. Replace damaged pipe. 5. Adjust control valve stroke as needed.

Suction and/or discharge vessel levels are not correct.

Calibrate level controllers as needed.

Motor is running backward or impeller of double suction pump is mounted backward. Discharge pressure developed in both cases is about one-half design value.

Check for proper rotation and mounting of impeller. Reverse motor leads if necessary.

Entrained gas from the process is lowering NPSH available.

Reduce entrained gas in liquid by process changes as needed.

Mechanical seal in suction system under vacuum is leaking air into system, causing pump curve to drop.

Change percentage balance of seal faces or increase spring tension.

There is polymer or scale buildup in discharge nozzle areas.

Shut down pump and remove scale or deposits.

Vortex formed at high flow rates or low liquid level. Does the vessel have a vortex breaker? Does incoming flow cause surface to swirl or agitate?

Reduce flow to design rates. Raise liquid level in suction vessel. Install vortex breaker in suction vessel.

A variable speed motor is operating too slowly.

Adjust motor speed as needed.

Bypassing is occurring between volute channels in a double volute pump casing due to a casting defect or extreme erosion.

Overhaul pump. Repair damaged areas.

Axial positions of impeller(s) are not centered with diffuser vanes. Offset of several impellers will cause vibration and lower head output.

Overhaul pump; reposition individual impellers as needed. Reposition rotor by changing thrust collar locator spacer.

When the suction system is under-vacuumed, the spare pump has difficulty getting into system.

Install a positive-pressure steam (from running pump) to fill the suction line from the block valve through the check valve.

Some pump designs incorporate an internal bypass orifice port to change head-flow curve. However, high liquid velocities often erode the orifice, causing the pump to go farther out on the pump curve. The system head curve increase corrects the flow back up the curve.

Overhaul pump. Restore orifice to correct size.

The volute and cutwater area of casing is severely eroded.

Overhaul pump. Replace casing or repair by welding. Stress- relieve after welding as needed.

A replacement impeller does not have a correct casting pattern, so NPSH required is different.

Overhaul pump. Replace impeller with correct pattern.

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The Pump Handbook Series

Motor Overload Problems POSSIBLE CAUSES

CORRECTIVE ACTION

Pump is circulating excessive liquid through a breakdown

Overhaul pump, replacing parts back to suction needed. bushing or a diffuser gasket area.

There is polymer buildup between wear surfaces (rings or vanes).

Remove buildup to restore clearances.

There is excessive wear ring (closed impeller) or cover-case clearance (open impeller).

Replace wear rings or adjust axial clearance of open impeller. In severe situations cover or case must be replaced.

Open impeller has slight rub on casing. This usually occurs in operations from 250–400°F due to piping strain and differential growth in the pump.

Increase clearance of impeller to casing.

The minimum flow loop has been inadvertently left open Close minimum flow loop or control valve bypass valve. at normal rates, or bypass around control valve is open. Discharge piping is leaking beneath liquid level in sump-type design.

Inspect piping for leakage. Replace as needed.

One phase has low amperage due to electrical switch gear problems.

Check out switch gear and repair as necessary.

Specific gravity is higher than design specification.

Change process to adjust specific gravity to design value, or throttle pump to reduce horsepower requirements. (Note: This will not correct the problem with some vertical turbine pumps which have a flat horsepower required curve.)

Pump motor not sized for end-of-curve operation.

Replace motor with a larger size, or reduce flow rate.

A replacement impeller was not trimmed to the correct diameter.

Remove impeller from pump. Turn to correct diameter.

The Pump Handbook Series

79

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PUMP AND SYSTEM TROUBLESHOOTING HANDBOOK

Water Hammer – Containing the Surge By J. Robert Krebs, P.E., Contributing Editor formulas and the surge signals are reproduced in an accompanying presentation this month. It should again be stressed that the rigid pipe theory approach and the 12 surge signals provide a quick check to judge the need for a more complete elastic pipe theory analysis. They are not a substitute for the more detailed approach. Figures 1 and 2 illustrate gravity and pumped flow systems respectively with the piping design the same in each.

n the November1995 and January 1996 issues this column identified the cause and effect of water hammer in piping systems. A gravity flow and a pumped system were used to illustrate conditions conducive to the development of potentially dangerous pipe surge pressures and to show how the maximum pressures and cycle time can be calculated. Twelve surge signals were also provided to serve as a guide in determining the need for a closer look at surge pressure potential. Both the

I

Q=4000gpm, V=6.4fps, VC=4000fps, T=6.7sec H=795 ft + 100 ft = 895 ft (387psi) HS=100 ft

Reservoir On-off valve

2.5 MI (13333 ft) 16" ID pipe 150psi Design Plus 100psi Surge Allow

SRV

Reservoir

Figure 1. Gravity flow system – surge relief

100 ft

SRV

Destination

CCV

Head FT

Source

CCV

Pipe Same as Figure 1 T = 6.7sec H = 795 ft + 200 ft = 995 ft (431psi)

Max RPM

200 100

RED RPM

0 0

Q GPM

Figure 2. Pumped flow system – surge relief

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The Pump Handbook Series

4000

The values of the cycle time period, maximum surge pressure and pipe design criteria are listed in each figure. One should first examine the surge signals and decide which, if any, apply to the problem. Then select the method of surge attenuation and the size of the device(s), if needed, or proceed to a complete elastic pipe analysis. For Figure 1, items 2, 3, 4 and 6 of the surge signals are applicable. Power outages (item 2) are the overriding rationale for most surge protection devices used. Normally, emergency power systems cannot respond quickly enough to avoid activation of a pressure surge. Valve close (and opening) time (item 3) can be controlled in the system design. This can reduce, even eliminate, the surge pressure in normal cycle operation. If pneumatic or hydraulic valve actuators are used, a pressured fluid supply (enough for several valve operations) can reduce or eliminate the danger from power outages. Pipe and accessory equipment design pressures (item 4) with surge allowance must be higher than the calculated total pressure including surge. Note that if repeated surges are possible, the repeated cumulative stresses could produce a fatigue failure. The endurance limit of materials, including plastics, must be considered. The use of stronger piping and equipment can be cost effective compared to surge suppression equipment. Items 4 and 6 can be considered together. Calculated results for maximum pressure and cycle time period are shown. The calculated maximum pressure exceeds the pipe design pressure with surge allowance. Increasing the valve closing time is an option. If the on-off valve has a linear closing characteristic and could be closed uniformly, in say 15 sec, to reflect a 75% drop in flow (or, say, 2 ft/sec followed by a 5 sec interval to close), the maximum surge

pressure would be significantly reduced to about 150 psi. For the Figure 1 gravity flow system, regulating the valve closure time could satisfy all concerns except power outage. A surge relief valve is the obvious answer for that concern and would most likely be less expensive than changing the pipe design specifications. The surge relief valve (SRV) pressure setting would allow the SRV to open as necessary on power outages. Two manufacturers recommend either a 6- or 8-in SRV for this application. The Figure 2 pumped system would be compared with the 12 surge signals. Items 2, 3, 4 and 6 are again not in agreement with the design. A power outage (item 2) would interrupt flow abruptly as the pump or pumps stopped requiring some form of system surge pressure protection. A controlled closing check valve (CCV in Figure 2) would permit an extended closing time (item 3) as in Figure 1, when a control valve was substituted for the onoff valve. As the valve closes, the additional pressure drop across the valve reduces the flow rate. At some preset time or valve position, the valve closes and the pump stops. A similar effect can be accomplished by using a variable speed driver and reducing speed to reduce head and flow prior to shutoff of this pump. Both methods are shown on the pump and system curve in Figure 2. The power outage concern can be covered by an SRV as shown in the Figure 1 solution. A surge relief valve would be sized at 10 in by one manufacturer. Higher design strength (item 4) pipe and fittings could be considered. However, another consideration (item 6) of pipe velocity and the accompanying friction loss should be considered. If 18" ID pipe were considered, the velocity at 4000 gpm comes down to 5.04 ft/sec, reducing the friction loss to 56 ft and the maximum surge pressure to 626 ft or a total pressure of 682 ft (295 psi). The lower friction loss will also be reflected in smaller driver 232 BHP vs. 190 BHP and a substantial power savings of $10.44 per million gallons

pumped (at $0.08 per KwH). To summarize the Figure 2 solution, I would consider using the 18" pipe with the 150 design and 100 psi surge allowance. While a control valve or controlled closing check valve and constant speed drives are possibly a less expensive solution, I would prefer using a swing check valve (perhaps cushioned) and a variable speed driver (VSD). The control valve method of reducing flow imposes higher pressures on the pump, as is obvious from the pump system curve. Conversely the flow rate and pressure are reduced as speed is reduced with the VSD. There is no single protective device to solve all water hammer problems; rather, there are many devices that can absorb or eliminate dangerous surge pressures. To name some that are regularly used – in line check valves for low head systems, bladder and hydro pneumatic tanks (Helmholtz resonators) and surge towers. There are few designs that are as straightforward – simple might be more correct – as these two examples. These articles have been intended to provide an introductory explanation of possible solutions to water hammer problems in a system. Finally, system operating conditions are all important. Initial filling of the piping system must be considered with the operation of vent and air release valves. Oversized surge relief valves can promote dangerous pressure surges, strengthening the case for multiple-sized valves with separate pressure settings in larger systems. ■ Until next time ...

Maximum Surge Pressure – Cycle Time Calculations H = Vc x V/g + Hs Vc = Sonic velocity ft/sec V = Steady state liquid velocity, at flow interruption g = Gravitational constant ft/sec2 Hs = Static or operating head ft The Pump Handbook Series

H = Maximum pressure in ft of liquid flowing T = 2L/Vc

L = Pipe length ft T = Cycle time in sec

Twelve Surge Signals 1. Pipe or fittings broken or cracked (existing system) 2. Power outage – emergency power normally not a solution 3. Valve closes in less than cycle time period 4. Maximum calculated surge pressure plus static or operating head approaches the pipe design including surge allowance 5. Large elevation changes (34 ft or one bar) over short pipe lengths 6. Higher pipe line velocities, say, above 5 ft/sec 7. System renovation with greater flow through an existing system 8. A low head system that permits continued flow through a pump after shutdown 9. Dead ends in piped systems that can promote higher than calculated pressure surges through resonant conditions 10. An undulating pipe line topography with lower head encouraging vapor cavity formation 11. Long suction lines for pumped systems 12. Improperly sized vent and vacuum valves

Pumping With Air Many readers responded with their information on air lift pumps. To the ten of you who wrote and the many of you who called my office with their helpful contributions a sincere thank you! From my position, putting it all together is not as straight forward as I had suspected. As soon as it is completed, each person who provided information will have the results and a column will be written. Thanks again!!

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PUMP AND SYSTEM TROUBLESHOOTING HANDBOOK

Reducing Noise In API Process Pumps Tests and observations point to geometry change as solution. By Leon K. Stanmore

n offshore drilling operations, API process pumps sometimes have to be mounted on structural steel platforms. Because the pump assembly is not attached to a concrete foundation and the baseplate is not filled with grout for rigidity and vibration damping, the movement and noise generated by a centrifugal pump are of special concern. Steel structures supporting rotating equipment not only transmit but amplify even the smallest vibrations or noise. In one such application it was determined that the maximum vibration had to be kept below 0.10 ips and that no discernible hydraulic noise would be permitted. Two 8x10x19 horizontal API 610 process pumps were brought in from the field and set up on a test floor to investigate procedures for minimizing random noise. The units were identical top suction and top discharge API 610 centrifugal pumps with overhung single entry impellers. The pumps are normally installed on an offshore platform. Both have ample NPSHR. The pump supporting base is made of welded structural steel, and there is no grout to provide damping or rigidity. End suction, top discharge pumps have been used successfully in this situation since 1956. However, top suction designs were added later, and few pumps with the bell-mouth design were built and tested. The pumps in question, tested at duty conditions of Q=2150 gpm, H=314 ft and N=1770 rpm, generated intermittent bursts of broadband noise. Reviews of test logs did not indicate any unusual pump behavior.

I

Figure 1.

“X”

“C”

Figure 2.

PROBLEM DISCUSSION When pressure fluctuations are produced by liquid motion, the sources are fluid dynamic in character. Fluid dynamic sources include

82

“Y”

Figure 1. Before geometry change Figure 2. After geometry change The Pump Handbook Series

Secondary flow patterns that can produce pressure fluctuations in centrifugal pumps include stall, recirculation (secondary flow), circulation, leakage, cavitation and wake (vortices) at inlet guide vane (splitter).

PROCEDURES

cient to accommodate the uneven flow. The area marked "C" in Figure 1 is a pocket in the case casting where the velocity is very low as the flow changes its direction by 90º. This area was considered a possible trap for entrained air. The observed noise could be best described as random "shearing" action of water. The pump was tested from closed valve to open valve flow. See curve 4590 for flow, head efficiency and NPSHR. Careful observations were made as to the noise intensity. With the flow reduced to 1,150 gpm, the noise was slightly less but still not acceptable.

B=100 (R3-R2) =12% R2 hence, this was considered acceptable for a volute type pump. The impeller eye was reduced from a diameter of 9 1/2" to 8 5/8", thus raising the velocity of fluid moving through the eye from 10.76 fps to 13.37 fps. Subsequent tests indicated that intermittent noise was reduced by 50%. Both pumps were the customer's property and a little late on delivery, so we did not have the time to install windows in the suction bellmouth for visual observation and study. Again, we examined the splitter in the section bell-mouth and determined that its design and its position relative to the impeller vanes at inlet were not optimized. The distance between the splitter edge (Figure 1, "Y") and the impeller eye was considered insuffi-

RESULTS OF GEOMETRY CHANGES It was reasoned that the splitter required a geometry change from Figure 1 to Figure 2. All edges were therefore rounded, and the pump was retested. The intermittent hydraulic noise disappeared. The NPSHR went up 1.6 ft with the 8 5/8" eye ring. Flow head characteristics were unaffected, the H-Q curve was stable, and there was no change in pump efficiency. Vibration in the vertical plane, as measured at the radial pump bearing, went up from 0.04 to 0.07 ips at the speed of rotation, 1770 rpm.

RECOMMENDATIONS AND OBSERVATIONS Based on the observations and NPSH in Feet 40 NPSH Req'd

20 0

EFF

100 90 80 70

400 TDH VS CAP

350

60

300

50

250

40

% Efficiency

200 400 BHP @ SP GB = 822

200

0

500

1000

1500

2000

2500

3000

0 3500

Horsepower

One of the pumps was opened and examined for any possible restrictions that could trap air. The pump case and impeller were examined for dimensional compliance, casting quality, machining, balancing and cavitation or mechanical damage. During the investigation it was discovered that the suction flow splitter had square edges instead of radii facing and trailing the flow. This would create a disturbed flow when in contact with the square edge splitter. Furthermore, the square edge of the 6-3/4"splitter (edge X in Figure 1) would create turbulence as well as a change of momentum. If we consider the splitter as a stationary blade, then a careful examination of the splitter in the suction bell-mouth led us to believe that the splitter design and its position relative to the impeller inlet were not optimized. It was felt that the boundary layer on the splitter, when reaching the square edge of the

splitter "separation" points, caused the wake, together with the freestream, to extend into the impeller eye. Greater distance between the splitter's vertical edge and the impeller inlet was considered important. The impeller outlet angle was β=24º, and the francis vane inlet angles were shroud 17º, hub 30º. There were 7 vanes, and the gap between the impeller's outside diameter and the case cut water was expressed as:

Total Head in Feet

turbulence, flow separation, cavitation, water hammer, flashing and impeller interaction with the case at the water or inlet guide vane (splitter). The pressure and flow pulsation may be periodic or broadband in frequency and excite either the piping or the pump into mechanical vibration. Mechanical vibrations can then radiate acoustic noise into the environment. Pump pulsating sources are classified as: • discrete frequency components produced by the pump impeller • broadband turbulent energy resulting from areas of high velocities • impact noise consisting of intermittent bursts of broadband noise caused by cavitation, flashing, water shearing action, water hammer and pockets of entrapped air at the suction inlet near the impeller • flow-induced pulsation caused by periodic vortex formation when the pump fluid passes an obstruction.

Gallons Per Minute Figure 3. Performance curve of 8x10x19, 1770 rpm API process pump The Pump Handbook Series

83

data collected in this application, it is recommended in general that any changes in the suction inlet area where a splitter is located should be evaluated first by checking the velocity flow, cascade and geometry. This should be followed by comprehensive tests at all flows in the range of application. It is essential that uniform velocities are present from the suction flange through the passages where the flow will accelerate into the impeller eye. It appears that other pump sizes with the same splitter configuration worked fine, but in our case the results were unacceptable. From this, it is clear that it is necessary to verify any geometry changes by full speed hydraulic tests at all flow conditions. If the pump had been fitted with Plexiglas windows in the area of the suction splitter and the impeller vanes at inlet, it would have been possible to observe the separation of flow at the splitter, as well as the wake extending into the impeller eye. One also might have seen entrained air in the low velocity area, "C" in Figure 1. Generally, though, a full range of flow testing can provide the same results if it is conducted thoroughly. ■ Leon K. Stanmore is a centrifugal pump consultant with more than 40 years of experience in research, design, development and analysis. He is past chairman of the Reciprocating Pump Section of the Hydraulic Institute and a task force member of API Standard 685 on Sealless Centrifugal Pumps.

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The Pump Handbook Series

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PUMP AND SYSTEM TROUBLESHOOTING HANDBOOK

Pump and Turbine Rerate Problems Tests and observations point to geometry change as solution. By Leon K. Stanmore he boiler feedwater system at the Novacor Chemicals worldscale ethylene plant in Sarnia, Ontario included horizontal multistage barrel case pumps and single stage impulse steam turbine drivers with mechanical governors. As part of a major effort to eliminate a production bottleneck, the pumps and drivers were rerated to provide more flow. The entire pump inner case assemblies (including rotor and inner casing) were replaced with higher capacity elements. New governor valves and nozzle rings were installed on the turbines to produce more power. And an electronic governor that adjusted the pump speed to maintain constant header pressure was added to the control scheme. The new equipment was installed in September 1994. Nothing went well. The pump suffered high blade pass frequency vibrations. The solution required unique hydraulic modifications and inboard bearing bracket stiffening to eliminate resonance. The turbines tripped mysteriously and broke governor valve stems. A valve resonance problem was solved by installing a valve with a different geometry. Teething problems with the electronic governor control system also had to be resolved. These problems became very high profile when the newly rerated boiler feedwater pumps tripped and caused a total plant shutdown resulting in a 14-hour flare. Not only did lost production reduce revenues, but the Ministry of the Environment initiated a review of our entire flaring history. Although it took a year of difficult effort, the pumps and turbines are now operating satisfactorily. Following is a chronology of these problem events and the resulting solutions.

T

BACKGROUND The boiler feedwater pumps and turbines had operated reliably for the past 18 years. Overall vibration levels were around 0.2 in/sec. (All vibration readings have units of inches per second, zero to peak, calculated and were

taken with magnetic base accelerometers and a portable dual channel analyzer using a Hanning window). We normally operate two steam turbine driven pumps (designated A and B). An electric driven spare pump remains on hot standby but we avoid operating it because it results in high demand charges from the power company. If any pump is unavailable the Operations group gets nervous because the entire plant could be shut down if a second pump trips for any reason. This means a significant loss of production. Furthermore, the long restart process generates a major flaring incident with its environmental impact. Thus, if one of these pumps, turbines or auxiliaries needs maintenance, it is considered an emergency requiring a 24-hour priority repair crew. A debottlenecking project was implemented to coincide with a planned major maintenance outage. The project modifications necessitated an increase in boiler feedwater flow. The obvious solution was to add a fourth identical pump rated at 1375 U.S.gpm and 2000 psig at 4310 rpm. Barrel pumps are expensive, however, and need costly foundations, high pressure piping modifications, steam piping modifications and a large installation space. To reduce costs, the actual flowrates were carefully examined. This current project required an additional 300 U.S.gpm. Any foreseeable debottlenecking project would require only an additional 750 U.S. gpm (including a contingency of 200 U.S. gpm). A fourth pump would have provided too much flow. Thus, we investigated the possibility of rerating the existing pumps to deliver 1750 U.S. gpm. Fortunately, the pump manufacturer had a standard inner bundle that could provide the flow at 4390 rpm and would be interchangeable with the existing barrel. The existing seals and bearing could even be reused. The steam turbine manufacturer also advised that the existing turbine could be rerated simply and inexpensively. The Pump Handbook Series

In all, the modifications were only 3/4 the capital cost of new equipment. Once installation costs were considered, the rerate option was the obvious choice.

MODIFICATIONS PERFORMED The rerate resulted in the following equipment modifications: Pumps. The pumps are 6x8x10 horizontal 9 stage API 610 barrel style pumps with opposed impellers and an internal crossover for thrust balance (Figure 1). The complete inner assembly consisting of the rotating element and the inner volute case were replaced. The outer barrel, bearings and seals remained the same. These are the highest capacity internals that would fit into the barrel. Turbines. The turbines were API 611 single stage impulse type design that used 500 lb inlet and 50 lb exhaust steam (Figure 2). The power rating was increased from 2100 to 2800 hp. This required a new nozzle ring and governor valve. In addition, an electronic governor conversion, new-technology dry gas steam seals installation and a general overhaul were performed. These changes approached the maximum power rating of the design. Controls. The old control scheme consisted of a hydraulic mechanical governor that maintained a constant speed of 4310 rpm and individual boiler flow control valves. For energy optimization reasons, a new electronic governor control scheme was installed that allowed the flow to vary by regulating the turbine speed from 3700 to 4610 rpm while the header pressure remained constant. Coupling. A new lightweight flexible disc coupling was installed with a higher power rating. Minimum Flow Valves. New higher capacity and updated modulating valves were installed. Although we had difficulties with all five items, this article discusses only the major problems with the turbines and pumps in a chronological order. For simplicity, the turbine issues are grouped separately from the pump

85

Figure 1. Pump cross section

problems, but you will need to try to imagine everything happening simultaneously. Also keep in mind the difficulty in troubleshooting and performing accurate analysis in an operating plant environment. Compared to a test stand, we didn't have enough instrumentation, we couldn't control all the variables, and we had to avoid disrupting plant operations.

EARLY OCTOBER: TURBINE STARTUP Both turbines had been sent to the manufacturers' service center to ensure that the work was done properly. They were installed and started up without any problems. Vibration levels were acceptable. But then, two unexplained trips on the B turbine occurred. No cause was ever found for these incidents. Vibration checks continued to show normal readings.

EARLY NOVEMBER: B TURBINE PROBLEM 1 We then experienced four governor valve failures over a weekend. On Friday, the governor valve stem broke into three pieces. The guide bushing was found inside the turbine casing and reused with a new stem from our warehouse. On Saturday the guide

bushing worked loose and machined itself through the seat. The valve stem was worn and the governor valve pin had sheared. A new valve seat and bushing were installed, but since the spare stem had already been used, the existing stem had to be refurbished. The valve was installed on the stem with a pin of unknown origin. On Sunday the valve pin sheared, and the bushing was out of place. During this failure there was an upset with the electronic governor of the other pump resulting in a loss of 1500 lb steam header pressure and a total plant shutdown. On Tuesday the valve stem was so badly worn it was grabbing the bushing and limiting valve travel. By this time, all new OEM parts had arrived on site, and a new seat, bushing, stem, valve and pin were installed. A mechanic noticed that the old bushing was magnetic and the new one was not, suggesting a material error. The turbine was placed back into service and restarted without incident at normal bearing vibration levels.

LATE NOVEMBER: B TURBINE SOLUTION 1 The manufacturer was surprised at first because governor stem failures

Figure 2. Turbine cross section

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The Pump Handbook Series

were extremely rare for this model turbine. But they immediately came up with a solution. The original design had a nitrided 416SS bushing and a nitrided governor valve stem. In some rare cases, the differential expansion between the steel cage and the bushing resulted in a loss press fit. Their solution was a nitrided 304SS bushing that had expansion rates higher than steel and was self-locking. This material became the OEM standard in early 1982. Since the new parts conformed to the new standard, the manufacturer advised that this would solve the problem. The only nagging doubt was why the original bushing had lasted for 18 years. The overhaul records confirmed that the bushings were checked and found to be in excellent condition. The explanation offered was that the rerating resulted in increased velocity across the valve. This resulted in higher vibration that loosened the bushing.

LATE DECEMBER: B TURBINE PROBLEM 2 Just before Christmas an operator passing by heard an unusual squeal emanating from the turbine. By the time we arrived with our vibration instruments, everything was quiet and vibrations were normal. We decided to do some testing. The B pump was put on manual and the speed was varied through the operating range. The A pump was left on automatic and its speed adjusted to provide the required flowrates. A dual channel analyzer was programmed for automatic data capture every 100 rpm. A resonance was discovered around 4000 rpm and reproduced the squeal. The vibration energy fluctuated but was extremely high, and occurred with a specific spike around 162,000 cpm (Figure 3). It was measurable all over the turbine but was highest around the governor valve area (in the 3 in/sec and 10g range). When the turbine was operating at resonance speed, one could see vibrations causing the trip lever to walk off the knife edges. We concluded that this resonance was the root cause of the mysterious trips and the broken governor valve stems. The knife edges were not in particularly good condition. They were replaced and the turbine placed on manual at 4200rpm, where vibration levels were normal and more analysis began. Similar vibration resonance patterns were noticed on the A pump but with much smaller vibration amplitudes.

MID JANUARY: B TURBINE SOLUTION 2

LATE JANUARY: B TURBINE PROBLEM 3 We felt that the valve position theory was sound, and since the old 3030º valve was still in our warehouse, we decided to reinstall it and reevaluate. We were extremely disappointed with the results. Although the vibration levels were marginally lower, the resonance persisted at the same frequency and now peaked at 4100 rpm.

EARLY MARCH: B TURBINE SOLUTION 3 We performed another test, mea-

0.8 3600

In/s pc

Our analysis pointed to problems with the nozzle ring/governor valve combination. The manufacturers' rerate calculations consist of a simple table showing maximum power, plus a margin, versus flow-passing capabilities of various nozzle ring and valve/seat combinations. Our horsepower requirements nudged us into the largest valve/nozzle ring combination available for this turbine. However, further investigations revealed that this table was based on standard mechanical governor data with a maximum valve travel of 0.6". The new electronic governor used a control valve actuator with a permissible travel of 1". We had awarded the governor conversion to the turbine manufacturer specifically to avoid this sort of confusion, but it happened anyway. The new two-stage governor valve had an aggressive 40º front taper and a 45º rear angle. It could actually produce 3052 hp at a stroke of 0.43". The manufacturer had a valve with a gentle 20-30º valve angle that could provide the power with 0.675" stroke. This valve had proved to be "more stable" in other applications, but delivery was six weeks. Interestingly, our original 30-30º valve could meet the power requirements with a 0.625" stroke. The theory was that a different valve would operate at a different position, one that would change the velocity distribution and the exciting force and would thus move the resonance point away from our operating range. Unfortunately, the manufacturer could not calculate these resonance relationships, so we could only postulate and use the trial-and-error method.

0.4

4000

4400

0.0 1.5K

150K

rp

m

300K cpm

Figure 3. Waterfall spectrum of OB turbine bearing before modifications (vertical direction)

suring inlet and outlet pressures and temperatures, steam chest pressures, valve stroke, vibration levels and speed in an attempt to correlate stroke to critical pressures. This was unsuccessful. It appeared that we had to put the turbine on manual away from the resonance and forgo the energy saving benefits and the major justification for the electronic governor conversion. But having come this far, we decided to try one more time and install the more stable 20-30º valve, which had finally arrived on site. Note that the 2030º valve is identical to the 40-45º valve, except for the machining of the valve angle. The 30-30º valve is a different design and casting (Photo 1). Although our enthusiasm was dampened, we installed the 20-30º valve and made a complete test throughout the speed range. We were pleasantly surprised to find that all traces of the problem disappeared. Vibration levels were 0.1 in/sec with no signs of resonance. We were able to turn over the unit to Operations without restrictions.

LATE JUNE: A TURBINE PROBLEMS As you will see in Part 2 of this article, we were also having problems with the pumps. Internal leakage had caused the speed to increase steadily over the months until the equipment was operating at 4500 rpm. The pump problems also produced higher vibration levels, and these suddenly began to increase exponentially. In an attempt to minimize damage until the repair parts arrived, we asked Operations to reduce the discharge header pressure setpoint to the absolute minimum. This was successful, and overall vibration levels did decrease. Shortly thereafter, the B pump was repaired and the operating speeds The Pump Handbook Series

of both pumps dropped dramatically to 4005 rpm. This just happened to coincide with the turbine governor valve resonance speed that we had discovered earlier and repaired on the B turbine. We were planning to replace this valve on the A turbine with the new 20-30º design during the pump outage. Overall vibration levels near the steam chest had been increasing but had doubled in a week to the 20 g range. We suspect that over the months, the bushing had worn and was now rapidly deteriorating. In an attempt to move the turbine out of the resonance range and limp along until the pump parts arrived, we asked Operations to raise the header pressure to increase the speed out of the resonance range. Turbine vibration levels did diminish, but several operators expressed bewilderment about what we were doing.

LATE JULY: A TURBINE STARTUP A new governor valve, stem, pin, seat and bushing were installed on the A Turbine. It was now identical to the repaired B unit. The turbines started smoothly with overall vibration levels below 0.1 in/sec., and there were no signs of resonance.

PART 2 You may recall that Part 1 dealt mainly with turbine issues as part of the Novacor Chemicals boiler feedwater rerate project. For simplicity, Part 2 discusses primarily pump issues. In reality, however, the challenge was considerably more complicated because pump and turbine problems were interrelated.

EARLY OCTOBER: PUMP STARTUP As soon as the pumps were started, we knew we had problems.

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0.3

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Figure 4. Typical frequency spectrum of B pump before modifications 1X

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Figure 5. Typical frequency spectrum of A pump before modifications

On minimum flow bypass, overall vibration levels exceeded 0.8 in/sec in the vertical direction on the inboard bearing housing of both pumps. And, even after more normal flows were established, overall vibration readings were 0.4 to 0.6 in/sec. When the situation was analyzed, it became clear that all the energy was coming from peaks that were at 5x and 7x running speed (Figure 4). Note that the new pump internals had a seven-vane impeller on stages 5 and 6 and five-vane impellers on the other stages. The B pump had higher vibration levels while the A pump also exhibited lower energy peaks at 3x, 4x and 6x running speed, indicating some kind of looseness (Figure 5). There was little vibration on the outboard bearing. This is perhaps explained by the fact that the throttle bushing stabilized that end of the pump. We approached the manufacturer for assistance. Its representatives'

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first reaction was that this was not a problem and we should not be concerned. They provided some literature advising that operation with high frequency blade pass vibrations did not pose a problem for long term reliability. Our position was that a new pump should meet the specified API 610 vibration levels for acceptance.

We thought that the evidence clearly showed a blade pass problem. But the pump manufacturer suspected a resonance problem in the bearing support. Apparently, problems of this nature had been reported previously. They agreed that vibrations were caused by vane pass pressure pulsations, but the real problem was that these pulsations were being amplified by a bearing housing/bracket resonance. Their solution was to increase the natural frequency of the inboard bearing housing. We agreed to do some testing to validate this theory. Unfortunately, we did not have a calibrated hammer, so the results are not highly accurate. They did indicate the presence of a resonance, however. The B pump was shut down, and a vibration probe was placed on the inboard bearing housing in the vertical plane. A baseline signature was recorded. Then the housing was struck with a rubber mallet numerous times. The frequency analysis showed a definite spike occurring at 25,200 cpm (Figure 6). The process was repeated with the vibration probe in the horizontal direction. A spike was also noticed at 23,175 cpm but at half the amplitude. This correlated with the field vibration measurements that always showed

0.4

In/s pc

In/s pc

LATE JANUARY: PUMP INBOARD BEARING RESONANCE TESTING

VERTICAL

7X

LATE DECEMBER: WORSENING PUMP PROBLEMS Vibration data was continuously supplied to the manufacturer for review and comment. Overall vibration levels steadily increased while the amplitude of the 5x and 7x peaks appeared to change over time. We reported operating concerns to their engineering group all along but felt it was important to formalize our position. We therefore filed a warranty claim. This elevated the problem to their senior management, and we began receiving more active attention. The Pump Handbook Series

0.2

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A – BASELINE SPECTRUM

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5X

0.6

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B – NUMEROUS IMPACTS WITH RUBBER MALLET – 32 AVGS

Figure 6. Results of resonance testing of inboard pump bearing housing

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4100

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A – NO WEIGHTS

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B – 100 LB WEIGHT ADDED Figure 7. Waterfall spectrum showing effects of adding weights to pump inboard bearing housing

higher readings in the vertical direction. For our operating speed range of 3700-4500 rpm, the 5x range is 18,500 to 22,500 cpm, and the 7x range is 25,900 to 31,500 cpm. This could explain the strange variance of the 5x and 7x spikes in our field testing. Weights were added to the bearing housing in an attempt to alter its natural frequency. A long threaded rod was installed in the bearing housing lifting eyebolt tap, and first 65 lbs and then 100 lbs of lead weights were bolted in place. The pumps were run through the operating speed range and vibration levels were compared to the baseline. We discovered that 100 lbs significantly changed the overall vibration levels, and the magnitude of the 5x and 7x frequency spikes as the speed varied (Figure 7).

EARLY MARCH: PUMP REPAIR RECOMMENDATIONS The manufacturer reps felt that only bearing bracket modifications

were needed to reduce vibrations to acceptable levels. Their calculations showed that if the mass of the bearing housing were increased by 400 lbs, it would move the natural frequency completely outside the operating range. But we felt this was an excessive amount of mass to add physically. Therefore, it was decided to increase the stiffness of the bearing bracket. Their calculations showed that a 150% increase would also solve the problem. Their proposal was to disassemble the pump, remove the inboard end cover and cut off the welded bearing housing support. They would supply a new support that was 250% stiffer than the old one, weld it onto the old end cover, stress relieve and reinstall. The total turnaround time to disassemble, send the parts to their repair facility and reassemble was 2-3 weeks. If these bearing bracket stiffening modifications did not work, they would then pursue hydraulic The Pump Handbook Series

modifications. We had a problem with the proposal. First, an outage of this length was unacceptable to our Operations Department. Second, plant operations had already been severely disrupted by all the trips, shutdowns and tests that were performed. The entire pump would have to be disassembled to gain access to the inboard end cover. The costs associated with pump disassembly are very high, and we did not have the luxury of a lengthy test program. We needed a better solution. Resonance test results indicated that bearing housing modifications were necessary. But we felt that we should also address the hydraulic problem. The first issue was impeller stack-up. Multistage pumps with identical trailing edge vane positions are known to amplify blade pass vibration energy. Although no records were kept, we were assured that the impeller keys are carefully positioned to avoid vane line-up, and that this item is always checked as part of their normal quality inspection process. Another problem area was the impeller blade tip-to-volute tongue gap. API 610 recognizes that high

volute tongue

stage impellers a

b a = 1/2 in. b = 1/4 in.

Figure 8a. Impeller/volute trim detail (stage impellers)

volute tongue

1st stage impellers

Figure 8b. Impeller/volute trim detail (1st stage impeller)

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LATE MAY: PUMP OPERATING PROBLEMS We had planned to do the A

90

pump first because it showed signs of mechanical looseness. However, we noticed a deterioration in pump performance. The pumps were slowly speeding up. At first we thought the process designers had made a mistake, and we required more flow than planned. However, in December we began noticing a strange but moderate cavitation noise in the balance line near the first elbow. Over months of negotiations and testing, this noise became louder. We started to trend vibration levels on the balance line near the throttle bushing and noted a steady increase. By the end of May, the vibration energy was 10 g. In early June it was 20 g. When all the parts finally arrived in the first week of June, it was more than 30 g. By then, we had decided to repair the B pump first and expected to see some internal damage. We were not disappointed.

EARLY JUNE: B PUMP REPAIR Severe cavitation damage was discovered when the pump's outboard end cover was removed. The throttle bushing and sleeve are overlaid with welded stellite material. The sleeve on the shaft had contacted the stationary bushing. They had welded themselves together and spun in the support. This opened a leak path, and

the interstage pressure of 1000 psi had cavitated the 3/4" housing almost to failure (Photo 1). During startup we experiences some system upsets that might have caused excessive rotor movement and contact. Another possible explanation was improper installation. We discovered that the bearing housings had not been realigned after the new internals were fitted. To reduce costs, the manufacturer's service representatives were not employed during the original installation. But they were on site for all rebuilds. The end cover was shipped to the manufacturer's repair center to have the washed out holder cut off and a new ring welded on. A new throttle bushing was installed, and the repaired end cover was back on site within 3 days. The new suction end cover with the heavier bearing supports was installed (Photo 2), and the bearing housings were carefully aligned. The new pump bundle was fitted without incident.

MID JUNE: B PUMP STARTUP We had found mechanical damage, corrected the impeller stack-up, made hydraulic modifications and stiffened the support. The pumps had to work because there was nothing left to be done. But quite honestly, we were nervous because of the history

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In/s pc

energy pumps require special provisions to reduce vane passing frequency vibrations, and the standard specifies that the gap be at least 6%. Our as-built gap of 6.8% was marginally acceptable. We asked the manufacturer to develop a repair recommendation,and they increased this gap very creatively. Instead of trimming the impellers and underfiling to restore the head, they trimmed the impellers at an angle. The angle cuts also were in the opposite direction of the volute trims. The theory was to create a "scissoring" action that would not only increase the effective gap but change the energy distribution over the volute tongue to lessen the magnitude of the pulsations. The eye sides of the impeller tips were cut back about 1/4" and the volute tongues 1/2" in the opposite direction. The first stage was a double suction, and application of that theory resulted in "V" cuts (Figures 8a and 8b). We decided to pursue hydraulic modifications and bearing bracket stiffening simultaneously. To minimize repair time, we purchased a new end cover. Unfortunately, it was an expensive forging that had a fourmonth delivery cycle. The pump manufacturer agreed to pay for the bearing bracket modifications. We agreed to pay for a new end cover and the hydraulic modifications. All modification work would be done at the pump manufacturer's local service center. We had purchased a complete inner bundle as a spare to upgrade the electric pump at a future date. This unused bundle was the first to be modified. When it was opened, we discovered that the impossible had occurred. The vanes of the two 7-vane impellers were exactly aligned. Four of the 5-vane impellers were exactly aligned. The three other 5-vane impellers also were aligned in a different plane. This situation required that all impellers be removed and their keyways welded up and recut. Also, two impellers seized on the shaft during disassembly, and it had to be undercut and chrome plated.

0.2 7X 5X

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25K

3700 35K cpm

Figure 9. Waterfall spectrum of B pump inboard bearing after modifications The Pump Handbook Series

rp

m

of events. The B pump was run up to minimum governor of 3700 rpm with the electric pump still on and the A pump on automatic. The B pump was manually run up in 100 rpm increments ,and vibration spectrums were taken. Then the electric pump was shut down and the speed lowered in 100 rpm increments, and vibration readings were taken. Success was nearly complete. Overall vibration was down to acceptable levels. However, spectrum analysis showed that the vibration energy was still coming from 5x and 7x peaks, albeit at much lower levels. As speeds were lowered and the pump was pushed back on the curve, the 5x spike dropped off, but the 7x peak increased to 0.4 in/sec at 3900 rpm, indicating that there was still some sort of resonance, Figure 9. Also, the repaired B pump now operated at 85 rpm faster than the A pump, indicating that the underfiling did not restore the head expected. Results were short of perfect, but the unit was turned over to Operations without restrictions. Shortly after, the pump was shut down for a minor turbine repair, and the bearing housing resonance test was repeated. The previous vertical resonance had all but disappeared.

MID JULY: A PUMP REPAIRS The bundle removed from the B pump was sent to the service center for inspection and hydraulic modifications before installation into the A pump. With all the cavitation to the end cover, we thought there would be major internal pump damage as well. Surprisingly, there was little rotor degradation except for the seized throttle bushing. There was slight wear on the interstage bushing. Both were replaced. The other wear rings were in excellent condition. The impeller wear rings were specified with API 610 clearances of 0.018 in., but the pump manufacturer insisted that the throttle and interstage bushing clearances remain at their standard of 0.012 in. Again, we found an impeller stacking problem: four 5-vane impellers lined up. Two impellers also seized on the shaft during disassembly. A temperature probe, accidentally installed in the unused vibration probe location, cut a groove into the burnished surface.

All these difficulties were corrected for, and a heavier bearing bracket was installed on the old end cover. The pump was then shut down and disassembled. We searched for some looseness that could explain this pump's abnormal vibration signature. We were almost relieved to find that a loose journal bearing fit in the top inboard housing had caused severe damage to the babbited surface. The measured clearance was more than 0.004in. At speeds exceeding 4000rpm, the manufacturer advised that their normal practice is to shim this clearance to 0.0005 in. Either the fit had worn or the shims were accidentally left out during reassembly.

LATE JULY: A PUMP STARTUP The rest of the pump rebuild and turbine governor valve, seat and bushing installation went without incident. The pump bearing housings were carefully realigned. Our crew was very efficient by this time. Startup was enlivened by a plugged oil passage in the pump inboard housing, a siutation that wiped the journal bearing. We were pleased to report satisfactory results. But again the startup tests showed signs of a resonance at 3900 rpm with 5x and 7x spikes approaching 0.3 in/sec. And at speeds below 4000 rpm, small but noticeable 6x and 8x peaks started to appear. At higher speeds everything smoothed out, and overall levels were below 0.2 in/sec. The pump manufacturer suspects some sort of acoustical resonance is taking place. Again, everything was not perfect, but the unit was turned over to Operations without restrictions. The bundle that was removed from the A pump was sent to the manufacturer's service center for inspection and hydraulic modifications before being returned as a warehouse spare. Again, there were impeller vane line-up problems. There was also some erosion wear on one side of the throttle and interstage bushings. Obviously, we had similar problems to the cavitated pump and were just lucky that there was no physical contact. Both items were replaced. The Pump Handbook Series

CONCLUSIONS This rerate was presumed to be a straightforward replacement of proven parts with complete assurances from original equipment manufacturers. What went wrong and what can be done to ensure it does not happen again? Unfortunately, there is no simple answer to this question other than the fact that rerates of complex turbomachinery can be difficult, particularly when resonances are involved. These can be extremely difficult to predict, isolate and correct. To solve these problems, extensive effort and cooperation between our technical staff and the manufacturers was necessary. Also, we needed a great deal of patience and understanding from our Operations and Maintenance departments. As a result of our efforts, we did manage to solve some complex technical problems together, and the pumps and turbines are now operating satisfactorily. Following are some specific recommendations resulting from this project: • This experience reinforces the need for shop testing wherever possible. It can eliminate many field problems and allow more accurate analysis and easier modifications. We actually tried to procure a test barrel for the pumps, but it proved to be prohibitively expensive. We should have insisted that the turbine be sent to the nearest OEM facility with a test loop and paid the premium. • We will change our multistage pump specifications to require an inspection hold point for impeller stack-up verification. We had hired a third party inspector and used the OEM's superb quality plan and still missed this important item. • For all critical equipment, we will now specify an engineering audit of manufacturer's rerate calculations and selections. We used to do this only for our large compressors. • For major rerate work, use of the OEM service representative is recommended as inexpensive insurance.

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• We need a more open relationship with our OEM suppliers. This is especially true for rerate work , in which reliability is more important than capital costs. We thought that we had unique problems but discovered afterwards that other users had faced similar issues. With a more open dialogue before the order was finalized, we could have minimized the adverse impact.

ACKNOWLEDGMENTS I would like to recognize and thank all those who participated in this project. Special thanks go to Jim Gardiner, a rotating equipment engineer who has since left Novacor and is now working for North Atlantic Refining Co., and to Mike Dufresne and Fred Robinett of Sulzer Bingham and to Rob Kirkpatrick from Elliott Turbomachinery Canada. ■ Editor's Note: This article is reproduced with permission of the Turbomachinery Laboratory, from Proceedings of the 13th International Pump Users Symposium, Turbomachinery Laboratory, Texas A&M University, College Station, TX, pp.89-95, Copyright 1996. John P. Henderson is an associate engineer with Novacor Chemicals' ethylene plant in Sarnia, Ontario. He is a rotating equipment specialist responsible for technical support, troubleshooting, reliability improvements, vibration analysis and new equipment review and specification. He received a B.S. degree in mechanical engineering from Carleton University in Ottawa and is a Registered Professional Engineer in Alberta.

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The Pump Handbook Series

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PUMP AND SYSTEM TROUBLESHOOTING HANDBOOK

Looking for Hidden Dangers By Gary Glidden

a major repair expense and extended down time. In the utility industry, down time is money. In the petrochemical industry a failed pump may result in far more than simply spilling some water on the floor. Fire, explosion and death could occur. Figure 1 shows an inspection report on a large boiler feed pump. Under the results, item #1 is the suction impeller. If this impeller had come apart, it would have propelled metal fragments through the pump. Because the pump runs at 5600 rpm, the damage and mess would be considerable. The cracks on this particular pump were visible to the naked eye and easy to spot. Item #5 is a different story, however. Without the wet particle inspection and the eddycurrent test we would never know these cracks existed. I'm not going into what causes cracking. This subject would be worth an article itself. What I want to stress is the importance of having a good inspection system in place.

There is an old saying..."Pay me now, or pay me later." This fits the need to inspect to a tee. ■ Gary Glidden is crew leader of the pump shop at Houston Lighting and Power and a member of the Pumps and Systems User Advisory Team.

PHOTO COURTESY OF HOUSTON LIGHTING AND POWER

Throughout our lives we face hidden dangers. When we were young, it was the monster under our bed. In our late teens it was a patrol car in the rear view mirror. Today, we work on pumps with more hidden dangers than a mine field. One of the biggest is cracking in various pump components: shafts, impellers and couplings on the rotating end; cases, nozzles and fittings on stationary parts. When do you look? How do you look? And who should look for these potentially dangerous situations. Maintenance staffs are expected to do more with less in this era of "right sizing." Every step of our process must be proved cost effective and essential for operation. Some maintenance steps are being dropped entirely, and others are being cut back. So, with fewer financial resources available and a shorter time period to work on equipment, where to we concentrate? Some problems, such as worn or broken parts, bearings and seals – are obvious, of course. But other difficulties are harder to detect. It is impossible to see a difference of .002" at a bearing fit but it could determine whether the pump keeps running or not. We cannot tell by visual inspection a difference of .005" or .006" over specs in an impeller wear ring clearance, but it could affect pump performance. So why do we visually inspect pump parts? It is easy to say everything looks okay, change the bearings and seals and put the pump back together. When you do this, what risks are you taking? If it is a small ANSI pump, there isn't much of a risk. Damage can result, but the cost would be small. On the other hand, if we are dealing with a large boiler feed water pump, damage from a hidden crack could result in

Example of a boiler feed pump's impeller damage that started as a minute crack. The Pump Handbook Series

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Figure 1. Sample inspection report

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The Pump Handbook Series

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PUMP AND SYSTEM TROUBLESHOOTING HANDBOOK

The Correction and Prevention of Low NPSH Conditions Understanding the source of the problem will identify the way out. By Ross C. Mackay

Introduction Patient: “Doctor, it hurts when I do that.” Doctor: “Well, don’t do that.”

T

NPSH Required by Pump

To Cavitation

Figure 1. NPSH balance diagram

problems, it is important to recognize that the physical symptoms that identify cavitation difficulties also reveal other conditions which are totally unrelated. But these will be discussed in a future article. To stop or prevent cavitation, we have two options: 1. decrease the NPSH required by the pump 2. increase the NPSH available from the system

The Pump The NPSH required by the pump is a function of the hydraulic design of the eye area of the impeller. The level of NPSH required can be reduced by increasing the eye area of that impeller. However, this can render the impeller more susceptible to suction recirculation, with almost identical symptoms as cavitation. Consequently, I would recommend that the pump manufacturer be conThe Pump Handbook Series

(COURTESY OF PATTERSON PUMP CO.)

his old vaudeville line can still make a point with regard to low Net Positive Suction Head (NPSH) conditions in a pumping system. The very existence of such a condition is forgivable only in the few situations where it truly cannot be avoided at the system design stage. Low NPSH situations are occasionally a necessary evil in process conditions, or in a batch operation in which the function of the pump is to empty a tank. In the latter situation, the pump is bound to cavitate towards the end of the batch, where the NPSH steadily decreases below the value required by the impeller for satisfactory operation. In considering the variety of ways out of a low NPSH, or cavitating condition, we should first understand the source of the problem – but without getting into an exhaustive treatise on the subject. Pumps cavitate when the NPSH required by the impeller is greater than the NPSH being made available from the system in which it operates. Under these conditions the liquid vaporizes in the eye of the impeller, and the bubbles created then collapse in a series of implosions on the vanes. In trouble-shooting cavitation

NPSH Available from System

sulted in any change of this nature. Some manufacturers will sometimes have an alternative impeller, with a lower NPSH requirement, available for that particular pump. Another possibility that has been used with varying success is the suction inducer. It consists of a small axial flow screw arrangement in the eye of the impeller to give the liquid a pre-rotation. This raises the pressure just enough to prevent vaporization from taking place in the eye of the impeller. The inducer is not always available as very few pump manufacturers offer this option. The only other choice in reduc-

Photo 1. Double suction impeller pump

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Ha

H vp

DRIVER

Hs Hf

DISCHARGE HEAD ASSEMBLY

SUCTION

PACKING BOX

DISCHARGE

BARREL

BOWL ASSEMBLY

OIL FIELD PRESSURIZATION PUMP HIGH PRESSURE MULTIPLE PUMPS CAN BE USED IN SERIES BARREL OR CAN PUMP

Figure 2. Vertical canned pump

ing the NPSH required is a complete change of pump style or size. Here are some options: • A lower speed pump requires less NPSH but will need a larger impeller to handle the same pumping conditions. • As a result of the two eye areas, a double suction impeller design (Photo 1)needs only twothirds of the NPSH required by a similarly sized single suction design. • A vertical canned pump (Figure 2) can have additional static suction head built into the column length to allow for the needed NPSH. • A number of lower flow pumps operating in parallel will also need less NPSH. Another solution involves the use of a booster pump upstream of the main pump. The booster unit must operate over the same capacity range, but it can develop a lower head. In this arrangement it needs to

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Figure 3. NPSH diagram

develop only the amount of head that is needed for the NPSH required by the main pump. It can, therefore, be a low head and/or low speed pump, both of which need less NPSH. The reader will notice that of the seven options affecting the pump side of the equation, one option can create a similar adverse condition, one option is not widely available, and the other five options require new pumps. Perhaps we should consider the system side of the equation.

The System The NPSH Available from a system consists of the following four factors only. Consequently, at least one of these must be changed to increase the NPSH Available (Figure 3). NPSHA = Hs + Ha - Hvp - Hf the static head over the impeller centerline (Hs) the pressure on the surface of the liquid (Ha) the vapor pressure of the liquid (Hvp) the friction losses in the suction line (Hf) It is apparent that the effective cures are those that increase the first two factors in the equation or decrease the last two.

Static Head To increase the positive static head is a simple (?) matter of lowering the pump or raising the suction tank, or raising the level of the liquid inside the suction tank. While the physical movement of the tank or The Pump Handbook Series

pump would often be an expensive proposition, the raising of the tank levels may be relatively cheap and simple. However, lowering the pump can often be more economical when taken together with any other changes that are being effected if more than one problem is being corrected. For example, if the suction piping arrangement is being changed to stop the creation of turbulence in the inlet, the pump could be moved to a lower floor of the building. This would straighten the inlet piping, eliminate the turbulent flow and increase the static head.

Surface Pressure Surface pressure is a little tricky to change if the suction source is the Atlantic Ocean or some other body of water that resists being controlled by mere mortals. It might be possible, however, to enclose a man-made tank and pressurize it, or even introduce a nitrogen blanket. Both of these possible solutions are subject to the dictates of the operating system. For example, increasing the pressure inside a deaerator would defeat the whole function of that vessel and thus must be judged impractical. But since this pressure is one of only four factors in the NPSH formula, it is worthy of some consideration in certain installations.

Vapor Pressure The only way to reduce the vapor pressure of a liquid is to reduce its temperature. In many instances this is operationally unacceptable and can be ignored. Also, the extent of the temperature change needed to provide an appreciable difference in NPSH usually renders this method inappropriate.

Friction Losses Owing to the fact that pump inlet piping is notoriously bad in the vast majority of industries throughout the world, this is the area where significant improvements can often be realized. However, I must caution you against the tendency to shorten the length of suction piping simply to reduce friction losses. While this will be effective, it could deny the liquid the opportunity of a smooth flow

path to the impeller eye. This, in turn, could cause turbulence and result in air entrainment difficulties that create the same symptoms as cavitation. To avoid this, the pump should be provided with a straight run of suction line in a length equivalent to 5 -10 times the diameter of the pipe. The smaller multiplier should be used on the larger pipe diameters and vice versa. The most effective way of reducing the friction losses on the suction side is to increase the size of the line. This can make a dramatic difference in the losses, to the extent that more than a 50% reduction in friction can be realized by replacing a 12 inch line by a 14 inch line. Exchanging a 6 inch line for an 8 inch line can reduce the friction losses by as much as 75%. It must be acknowledged, however, that changing the pipe size also changes the size of all valves and fittings. Reductions in friction losses can be achieved even with the same line size by incorporating long sweep elbows, changing the valve type and reducing their number. One final item that shows up with surprising regularity is the suction strainer. It is widely used in the

commissioning stages of a new plant. Unfortunately, there are many times when it is a forgotten piece of equipment, and blockage in the strainer basket gradually increases and raises the friction loss to an unacceptable level. If a strainer is required in a system on a continuing basis, it should be located downstream of the pump, and the pump should be one that can handle the solid sizes expected. Obviously there are limitations to this concept, but it can be used more frequently than current practice would have us believe.

Most Effective The most effective cure for cavitation is the one which is both economically and practically viable. It is important to consider all possibilities and not dismiss any of them out of hand just because they are going to cost a “lot of money.” If the cost of repair is, say, $25,000, and the present cost of nursing the pump through its cavitation problems is $5,000 annually, it follows that it would take five years to retrieve the cost of the repair. However, if you are planning to run that pump for an additional 10 years after that date, the savings will amount to $50,000

The Pump Handbook Series

by the end of the 15-year period. By the same token, if the cost of the repair is only $2,000, the decision becomes much simpler as the payback is achieved well within the first year. From my experience in plants all over the United States and in many other countries around the world, it has become evident that many pumps are cavitating for the want of one or two feet of NPSH. As a result of this, I have found that the relatively simple change of raising the level of the liquid at the suction source to be an extremely effective and economical cure. Again, it is a little difficult if the suction source happens to be the Atlantic Ocean. The best cure of all is, of course, prevention. If the patient doesn’t ‘do that’ in the first place, it will never hurt!■ Ross Mackay is a consultant in pump reliability. He has more than 35 years of experience in the international pump community and conducts the Mackay Pump School throughout North America.

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? Torsional Vibration Linked to Water Pumping System Failure PUMP AND SYSTEM TROUBLESHOOTING HANDBOOK

By Troy Feese

municipality purchased several units to pump water on an emergency basis for fire fighting. These units are located around the downtown area and can provide water through portable pipes and hoses in case a major catastrophe cuts off the primary water supply to fire hydrants. Each unit consists of a V-12 diesel engine driving a five stage vertical centrifugal pump through a right angle gearbox with a 1:1 ratio. The 1600 hp engine has an operating speed range of 1200 to 1800 rpm. The engine flywheel connects to the gearbox through a drive shaft that has universal joints on each end. This article describes how failures of the water pumping system were linked to torsional vibration. It also shows how a solution was developed using computer analyses. Field tests were performed to verify that the solution was adequate to prevent additional failures.

A

The Problem The system experienced failures of the input gear and cooling fan. Several of the bolts that held the gear to the input shaft broke after only 13 hours of operation. It was speculated that this failure was due to excessive torsional vibration or improper fit that may have caused the transmitted torque to be unevenly distributed among the bolts. The cooling fan on top of the gearbox also experience several failures. The first time the fan blades broke, it was thought to be related to possible problems with the material or manufacturing process. The fan is constructed by pressing the

98

general shape out of the sheet metal, and then the blades are twisted to the proper pitch angle. If a small crack formed at the base of the fan blades, a high stress riser would be created. However, this type of fan has been used successfully at other installations. These fan blade failures could have been caused by high torsional vibration. Although it was possible to use additional bolts to hold the gear more securely to the input shaft, there was con- Figure 1. Interference diagram cern that if torsional vibration levels were too severe, then indicate torsional critical speeds. The some other portion of the system two circled points in Figure 1 show would fail resulting in more damage. where the 4.5x and 3.5x engine harTherefore, a detailed torsional analymonics intersect the fifth torsional sis was performed. The wet pump mode - at approximately 1400 and impeller inertial (25% greater than 1800 rpm respectively. The fifth tordry) was used to account for the water. sional mode was of concern since the The dynamic torque produced by the shape showed twisting in the input engine was calculated from the cylingear shaft and oscillation at the coolder pressures and inertia forces acting ing fan. Also, the engine damper is not on each of the six throws. Appropriate as effective for this mode. Steady-state stress concentration factors were forced calculations were performed to applied to the gear and pump shaft predict the level of torsional vibration sections. The damping in the system versus engine speed. The alternating was evaluated to include the engine shear shaft were above the endurance viscous damper properties as well as limit of the shaft material. The force the effects of the bearings, packing analysis also showed that the cooling and pumped fluid. fan would experience high torsional The torsional analysis indicated oscillation. that the engine produces significant dynamic torque at 2.5x, 3.5x, 4.5x and Torsional Analysis Yields 6x running speed. The interference Solution diagram in Figure 1 show the calculatBased on the torsional analysis ed torsional natural frequencies and results, the system needed to be modharmonic speed lines associated with ified to reduce the excessive torsional the engine. The intersection points The Pump Handbook Series

vibration and stress levels. A parametric study was performed. This showed that using a “torsionally soft” coupling between the engine and gearbox would help isolate the engine harmonics from the rest of the system. A coupling with a rubber element in shear was selected that would have a low torsional stiffness (less than 300,000 in-lb/rad) and provide additional damping to the system. Since this coupling bolts directly to the engine flywheel and supports the drive shaft, the system could be easily retrofitted in the field. The analysis indicated that with the rubber coupling the torsional vibration and stress levels in the gear and pump would be significantly reduced compared to the original system. Field measurements of the system were performed before and after the rubber coupling was installed to verify that the torsional vibration would be reduced enough to prevent additional failures. Strain gage measurements could not be made on the input gear shaft due to limited accessibility inside the gearbox. However, the torsional analysis showed a correlation between the predicted dynamic torque in the drive shaft and the alternating shear stress in the input gear shaft. Therefore, strain gages were attached to a uniform section of the drive shaft, and the signal was transmitted using all FM telemetry system. The torsional oscillation of the cooling fan were measured using an HBM torsiograph. The vibration data were gathered over a two minute period as the engine speed was increased from 1200 to 1800 rpm. During the tests the pump was operated with recirculated water. Speed rasters of the dynamic torque measured in the drive shaft for both configurations are shown in Figures 2a and 2b. The most significant engine harmonics from the speed rasters and the overall dynamic torque

Figure 2a. Original system – dynamic torque in drive shaft

Figure 3a. Original system – dynamic torque in drive shaft

Figure 2b. Modified system with rubber coupling – dynamic torque in drive shaft

Figure 3b. Modified system with rubber coupling – dynamic torque in drive shaft

(all harmonics combined) are plotted versus engine speed in Figures 3a and 3b. For the original system, the overall dynamic torque in the drive shaft was as much as 56% of the transmitted torque. However, with the rubber coupling installed the overall dynamic torque was reduced to 16% of the transmitted torque. The measured torque data compared favorably with the predicted results from the computer analyses. For example, in Figure 2a the 4.5x engine harmonic passed through a resonance near 1400 rpm, and the amplitude of the 3.5x harmonic increased as the engine speed approached 1800 rpm. The stress levels in the input gear shaft would be reduced by approximately the same amount as the dynamic torque in the

drive shaft. The torsional oscillation at the cooling fan was also reduced. Based on the test results, the rubber coupling was recommended as a permanent solution. No failures have been reported since the coupling was installed.■

The Pump Handbook Series

Troy Feese is a Project Engineer at Engineering Dynamics Incorporated in San Antonio, Texas. He has six years of experience performing torsional and lateral critical speed analyses and rotor stability analyses and in evaluating structural vibration problems using finite element methods. Mr. Feese is a member of ASME, and the Vibration Institute, and he is a registered Professional Engineer in Texas.

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?

PUMP AND SYSTEM TROUBLESHOOTING HANDBOOK

Modeling Pump Intake Noise Huge circulating water pumps are “scaled-down” to solve a mystery. By J. P. Messina, Consultant could remain an unsolved curiosity. One day, however, a wo 60-inch condenser circuSurface lake storm created large surlating water pumps, designed vortex face vortices in the forebay. for 126,000 gpm and 34 ft of Severe vibrations of the Pump head and driven by 1250 hps suction pumps and motors resulted. bell 195 rpm motors, were installed in a Operators would have shut Screen nuclear generating station (Figure 1). down the pumps if someone Initially the performance of these had not started the traveling pumps was not acceptable even screens, whereupon the rumthough rated flow was being delivSurface blings and vibrations quieted ered. During normal operating condivortex only to resume after the tions the pumps ran smooth, but screen cleaning was stopped. they occasionally produced rumbling High vibrations persisted durnoises that were irregular and puzing storm conditions and the Figure 2. Forebay surface vortices zling. A typical 15-minute recording problem had to be resolved. revealed 18 distinct rumbles from turer was confident that the noise one pump, 9 from the other. The was not mechanical since it was Type of Pumps Used noises were faint at first, grew to a irregular and low in frequency and Pumps for this service are usualpeak, then gradually disappeared. not related to pump speed. The forely vertical, wet-pit, mixed-flow difThese sounds occurred during single bay water level was never below fuser types – a relatively inexpensive and multiple pump operation. design minimum of 12 ft, 6 in above design that requires minimum space. It was initially believed that the impeller eye, an elevation more However, such pumps are not very these occasional rumblings did not than adequate to prevent cavitation. accessible for maintenance. The signal a serious internal problem and If necessary, the pumps would have pumps in this case were a vertical, been able to pull a suction lift. Small dry-pit design employing a surface vortices occasionally volute casing rather than a appeared in the forebay around the diffuser and were more dividing wall and stop log supports at accessible for mainteMotor the entrance to the pump suction nance. chambers (Figure 2). However, noise Maxium Both pump types lake level occurred even when these vortices employ a suction bell that were not present. Surface vortices, if uniformly directs and strong, can funnel air and cause the accelerates the water into water to prerotate as it enters the Stop-log the impeller eye. For optiPump slot pump impeller. This produces noise mum performance it is and less than optimum pump perforScreen crucial that the suction mance. bells for these pumps be The only probable causes adequately submerged and remaining to be considered by the correctly located in a proppump manufacturer were air accuerly designed suction pit. mulating under the pump floor Pump suction bell around the outside of the suction Initial Diagnosis bell, air in the pump, or subsurface Figure 1. Two 60″ vertical volute pumps and The pump manufac-

A Noisy Problem

T

condenser circulating water intake

100

The Pump Handbook Series

Flow Meter

Stilling Tank

Model Forebay

Suction Bell

Pump Window

Figure 3. 1/10 scale model

vortices. A vent was installed to release air by making a hole in the pump floor and adding a vertical pipe open at the top above high water level in the forebay. However, no air was found. The pumps did employ balancing rings on top of the impeller – a low pressure area that could conceivably collect air. This area was drilled into and vented, but this did not take care of the problem.

Action Taken

10 Underwater Vortexing rate, no./min.

The idea of modeling pump intakes is very controversial. While it is important to maintain geometric similarity between model and prototype, it is equally important to attain dynamic similarity. For a more in-depth explanation of modeling laws, see the Pump Handbook, section 10.2, 2nd edition, published by McGraw-Hill. Since the pump intake is a combination of open and closed conduits, the model flow and velocity should satisfy both the Reynolds and Froude number laws. Unfortunately, at any set of test conditions both laws cannot be satisfied at any one time. If the model were tested to have the same Reynolds number as the prototype requiring 10 fps under the pump floor, very turbulent flow would result in the forebay, preventing the formation of surface vortices that solely or partly could contribute to a noise problem. If the intake were solely an open pit, Froude similarity would be the way to go. A more conservative test favored by some investigators for an open pit would be to test at the same velocity as the prototype – or 1260 gpm model flow. It was therefore decided to test the model for a flow range of 398 gpm Froude flow to 1260 gpm equal prototype velocity flow, not to a 12,600 gpm Reynolds flow.

Froude velocity

9

Prototype velocity

8 7 6 5

Average of test points

4 3 2 1 0

400

1200 2000 Flow rate, gpm

2800

Figure 5. The number of vortices forming per minute nearly equals the number of rumbles in model pump’s test velocity range.

700 gpm, which is 1.7 times Froude velocity and .55 times prototype velocity, vortices appeared under the pump floor and around the suction bell. They started horizontally from either side wall and were drawn up into the pump suction bell. They looked like flashes of lightening, appearing one or two at a time (Figure 4). Figure 5 shows the test results and plots the number of vortices appearing per minute vs. flow in gpm. It was interesting to note that between Froude and prototype velocities the vortex rate averaged about 1 or 2 a minute, or 15 to 30 in 15 minutes. The two prototype pumps were recorded producing 9 and 18 rumbles in a similar 15 minute period. Was this just a coincidence? The pump manufacturer did not think so.

The last possible cause of the noise to be considered was underwater vortices, but it could not be determined if these indeed were occurring since the area in question was hidden under the pump floor. Because of the pumps’ size, large and costly baffling would be required to correct Model Observations a subsurface vortices problem, and at best it would probably be a hit and In the flow range tested no surmiss procedure. Furthermore, there face vortices appeared. At a flow of was no proof that underwater vortices existed – or, if they did, that they were producing the noises. Those involved decided to build Motor a relatively inexpensive 1/10 scale Maxium model of one pump chamber (Figure lake level 3). Only the suction bell of the pump was modeled, and circulation of water was provided by an external Axis of underwater vortex Stop-log Pump slot pump and stilling tank. The model Horizontal Suction bell was made of wood with Plexiglas Screen baffle extension windows and measured 7 ft long, 4 ft high and 2 ft wide. While the forebay, the suction bell and its chamber, and water depth dimensions would Pump suction bell be 1/10 actual size, what should the model flow and velocity be? That Figure 4. Formation of underwater Figure 6. Formation of underwater vortex as was a critical question. vortex and underwater baffling suggested from model observations The Pump Handbook Series

101

Figures 4 and 6 show how the underwater vortices formed and suggest where baffles might be added and/or alterations made to the suction pipe. Sixteen different modifications were tried. Each was evaluated based on the number of underwater vortices formed per minute over an extended range of flows from 300 to 2150 gpm. The simplest and most effective modification was moving the back wall right up against the suction bell. This produced no underwater vortices. If this modification were to be made in the field, however, it was feared that the suction bell would no longer serve its original purpose, which was to direct flow to the impeller eye evenly. The suction bell would then act more like a short radius elbow, hydraulically unbalancing the flow, and this could result in rough running and less than optimum performance. An equally effective test modification would have required extending the suction bell to .4 bell diameters from the floor (presently 1 diameter), moving the side and back walls closer to the bell, i.e., 1.8 diameters width (presently 2.4 diameters) and 1 diameter from the bell centerline to the back wall (presently 1.5 diameters). These modifications would be more in keeping with the recommendations of the Hydraulic

1.5D D

D

1.8D 2.4D

D .4D

Figure 7. Installed vs. Hydraulic Institute suction pit dimensions

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Institute Standards (Figure 7). Making these changes in the existing suction chamber walls in the field would not have been practical.

Recommended Modifications

Surface vortex Pump suction bell

Baffling added to stop surface vortexing

Screen

One modification tested which showed no underwater vortex forSurface vortex mation and still retained a good flow distribution to the pump impeller was eventually recommend- Figure 8. Location of surface vortices and surface ed to the customer. This baffling required extending the appeared. It can be concluded that suction bell to half a diameter above the underwater vortices caused the the floor and placing a horizontal noises, and this was due to the steel plate at bell bottom elevation in improper spacing of the suction bell the rectangular area between the bell relative to the floor and walls. It was and back wall (Figure 4). The steel also concluded that noise can be bell extensions and plates were made modeled in the range of prototype in sections to facilitate installation in and Froude velocities.■ the field. The horizontal plate would be supported on columns. J. P. Messina is a pump and A further recommendation was hydraulics consultant in Springfield, NJ the addition of a steel, anti-surface and co-author of the Pump Handbook vortexing, horizontal surface plate published by McGraw-Hill. He has preand vertical skimming wall to be sented numerous papers on various placed between the stop log dividing pump topics as well as taught courses walls and opposite walls. This would on centrifugal pump theory, construcbe attached to a gang of several stop tion and operation. He can be reached logs which, when lowered in place, at (973)379-5483. would float on the surface as depicted in Figure 8. It was theorized that during a storm the three sets of screens could become overloaded and discharge unequally if left to collect screenings. This could create an unsymmetrical flow pattern into the pumps. The flow would then approach either of the stop log dividing walls at an angle and cause surface vortices. Also, these surface vortices could feed into an underwater vortex, increasing its strength and adding air to the pump. Because the model only contained half of the forebay and screens, the unsymmetrical flow due to an unevenly loaded screen could not be observed causing surface vortices.

Results The recommendations were followed, and the noises completely dis-

The Pump Handbook Series

?

PUMP AND SYSTEM TROUBLESHOOTING HANDBOOK

Motor Bearing Failures in Cooling Tower Water Pump Vibration analysis unlocks solution. By Cliff Hammock, Technical Services Unlimited n May 13, 1997 our company was called upon by Colonial Pipeline to perform a complete vibration analysis on the circulating water pump for its #2 cooling tower. The unit had repeatedly experienced motor bearing failures caused by excessive vibration. Detailed vibration examination of the vertical pump, including spectral and phase analysis along with impact and coastdown testing, revealed that the motor vibration was caused by resonance. The motor speed of 720 rpm was very close to a structural natural frequency, which was amplifying the vibration to very high levels. The vibration was reduced from 1.4 inches/seconds (ips) to .06 ips by providing additional stiffness to the top of the motor. This shifted the natural frequency to 827 CPM, which is above the motor’s running speed, and thus solved the resonance problem.

O

Definitions ips = inches per second, is a unit of measure of velocity. The velocity is the rate of change of the displacement of a moving part. CPM = cycles per minute, a measure of frequency. Hertz (Hz) which is cycles per second (cps), is another commonly used measure of frequency.

BEFORE STIFFNESS

Amplitude (in/sec) Phase (degrees) AFTER STIFFNESS

Amplitude (in/sec) Phase (degrees)

Vertical Opp Drive End

Horizontal Opp Drive End

Vertical Drive End

Horizontal Drive End

.42 105

1.47 251

.23 102

.62 255

Vertical Opp Drive End

Horizontal Opp Drive End

Vertical Drive End

Horizontal Drive End

.19 23

.06 88

.09 26

.03 81

Table 1. Vibration amplitude and phase data – before and after modification

When we arrived at the site, we ran the newly rebuilt motor uncoupled to evaluate the severity of the vibration problem and to try to determine the source of the vibration. As shown in Table 1, the asfound vibration was indeed excessive and highly directional. The Alarm 1 level for a vertical pump of this size (12-20 ft) is .6 ips, and the Alarm 2 level is .9 ips. The level of vibration at the top of the motor in the horizontal direction was at 1.47 ips, far exceeding the Alarm 2 level. For purpose of agreed-upon orientation on a vertically mounted motor, the direction in line with the discharge piping is considered vertical, and the direction in line with the motor junction box is considered horizontal. The vertical vibration readings were much lower in amplitude than the horizontal, as shown in Table 1. If this was only an unbalance condition or some other mechanical defect, the vibration The Pump Handbook Series

would be similar in both directions. Having an amplitude ratio from one direction to another of greater than 3:1 is often an indication of a resonance problem. The phase data collected on the motor also indicated a resonance problem. Table 1 shows that the motor was in phase from top to bottom in both directions, but the phase from vertical to horizontal was out of phase by about 150 degrees. The phase data were taken from two vibration transducers mounted in perpendicular planes (90 degrees apart). If this were purely a mechanical defect, the most likely scenario would have the phase from one direction to another approaching 90 degrees. A resonance problem allows the phase to shift. After collecting the as-found data, we shut the pump motor off and monitored the overall vibration as the speed decreased. The amplitude versus speed plot is shown in Figure 1. Note that the vibration had

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motor running speed. An impact test is per15 formed – while the equipment is not running – by striking the 1 machine with a soft object such as a rubber hammer and mon0.5 itoring the natural vibration that occurs. The natural frequen0 cies of vibration will 300 360 420 480 540 600 660 690 705 720 Speed (RPM) show up as peaks in Overall Vibration the spectrum. This may not be the most Figure 1. Amplitude vs. speed taken during technically correct coastdown method of determinMED - Colonial Pipeline-Clg Twr #2 ing natural frequenC11-P002 -V01 Motor Outboard Bearing 0.07 cies, but it is quick Analyze Spectrum 13-MAY-97 09:30 and has yielded very 0.06 PK = .0974 LOAD = 100.0 good information on RPM= 720. RPS= 12.00 many occasions. We 0.05 did impact tests in 0.04 both vertical and horizontal directions, and 0.03 the resulting spectra are shown in Figures 2 0.02 and 3. It is evident that the natural fre0.01 quencies in the verti0 cal directions are Freq: 690.00 0 1000 2000 3000 4000 5000 6000 Ordr: .958 slightly higher than Frequency in CPM Spec: .04160 those in the horizontal Figure 2. Impact test on motor outboard bearing in direction because of the vertical direction added stiffness in the MED - Colonial Pipeline-Clg Twr #2 discharge piping. Also C11-P002 -H01 Motor Outboard Bearing 0.05 note that the coupling Analyze Spectrum 13-MAY-97 09:32 access area is in the PK = .0745 LOAD = 100.0 horizontal direction, 0.04 RPM= 720. RPS= 12.00 which yields less stiffness. In both direc0.03 tions the motor speed is very close to the natural frequency, thus 0.02 amplifying the vibration. 0.01 The natural frequency of a piece of 0 equipment depends Freq: 677.85 0 1000 2000 3000 4000 5000 6000 on the mass (weight) Ordr: .941 Frequency in CPM Spec: .04115 and the stiffness of the Figure 3. Impact test on motor outboard bearing in system. The amount the horizontal direction by which resonance amplifies vibration an absolute drop of over 1 ips over a depends on the damping of the 23 rpm range of speed. This was equipment. The formula for the natanother indication that resonance ural frequency is: was present. After the shutdown, impact testing was performed to Fn = √(k/m), determine if a structural natural frequency was occurring at or near the where k = stiffness and m = mass. Overall Vibration vs Speed

104

702.67

As you can see, the natural frequency can be increased by increasing the stiffness or decreasing the mass, or reduced by increasing the mass (weight) or decreasing the stiffness. In this case it would be difficult to add mass to the structure or decrease the stiffness, but additional stiffness could be added to the top of the motor. The motor will always have some inherent vibration due to any unbalance associated with the rotor. In the situation with the cooling tower water pump, that unbalance, no matter how small, was being amplified by the fact that the motor was in a resonant condition. If the motor speed could have changed, there would no longer have been a forcing frequency to amplify the natural frequency. However, for this application changing the speed was highly impractical. In an attempt to move the natural frequency above the motor speed, we decided to add stiffness to the top of the motor. This is preferable to operating machinery below a structural natural frequency, so that each

677.85

PK Velocity in In/Sec.

PK Velocity in In/Sec.

Representation of Actual Data

The Pump Handbook Series

Photo 1. The jacking screw arrangement was designed so that the stiffness could easily be controlled and optimized in both vertical and horizontal directions.

Photo 2. Close-up of jacking screw – a relatively simple and low cost solution.

MED - Colonial Pipeline-Clg Twr #2 C11-P002 -H01 Motor Outboard Bearing

0.06

Analyze Spectrum 14-MAY-97 15.19

PK Velocity in In/Sec.

PK = .0795 LOAD = 100.0 RPM= 720. RPS= 12.00

827.14

0.05

0.04

0.03

0.02

0.01

0 0

1000

2000 3000 4000 Frequency in CPM

5000

6000

Freq: 840.00 Ordr: 1.167 Spec: .04350

sive. While this particular solution may not be the answer for all vertical pump problems, it does show that, with effective use of technology, answers can be had even for long term problems.■ Cliff Hammock is President of Technical Services Unlimited, Inc., in Vidalia, Georgia, a company specializing in maintenance engineering consulting services, including vibration analysis, laser shaft alignment and balancing. He is certified by the Vibration Institute as a Vibration Specialist II and is Chairman of the Georgia Chapter of the Vibration Institute.

Figure 4. Impact test results after adding stiffness in the horizontal direction MED - Colonial Pipeline-Clg Twr #2 C11-P002 -H01 Motor Outboard Bearing Max Amp 1.22

713.08

PK Velocity in In/Sec.

Plot Scale 1.4

11-JUL-97 09:25

0

0

600

1200 1800 2400 Frequency in CPM

3000

13-MAY-97 08:48 Freq: 713.08 Ordr: 1.000 Spec: .05516

3600

Figure 5. Vibration before and after stiffening

time the motor is started, the motor speed will not pass the natural frequency range, thus amplifying vibration during each startup and shutdown. The solution was a jacking screw arrangement designed so that the stiffness could easily be controlled and optimized in both vertical and horizontal directions (Photos 1 and 2). The results of impact testing following the stiffening is shown in Figure 4. The natural frequency was shifted in both the vertical and horizontal directions. The horizontal direction natural frequency was shifted from 677 CPM to 827 CPM, which is greater than 10% above the motor speed. The vertical direction was shifted from 702 CPM to 765 CPM.

The before and after stiffening vibration levels are shown in Figure 5. There was a dramatic decrease in vibration in the horizontal direction, from 1.4 ips to .06 ips. The vibration phase data also significantly changed. The phase from top to bottom in still in phase, but the phase difference from vertical to horizontal is now 65 degrees. This is much closer to the expected 90 degree phase shift than the as-found 150 degree phase shift. Because resonance is common on vertically mounted machinery, I have seen similar situations with other vertical pumps. Although this was a significant problem that resulted in numerous costly failed motor bearings in the past few years, the fix was relatively simple and inexpenThe Pump Handbook Series

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PUMP AND SYSTEM TROUBLESHOOTING HANDBOOK

Cavitation in a Nutshell Here’s a simplified approach to determining whether you have a cavitation problem and what you can do about it. By Jeff Hawks, Buckeye Pumps, Inc.

his article takes a different approach to the subject of pump cavitation. Rather than focusing on the technical theory of fluid in motion, it offers a hands-on explanation of ways for users to determine whether or not they have a cavitation problem and how to find a practical solution. Simple formulas and definitions are found at the end of the article to aid further system testing.

T

What Is Cavitation? Cavitation is the formation of partial vacuums in a flowing liquid as a result of the separation of its parts. There are two types. Suction side cavitation is by far the most common form (probably 90% of all known events). Discharge cavitation is significantly less common. Suction side cavitation is a restriction on the suction side of the pump system which does not allow sufficient fluid to enter the pump and be discharged. The pump reacts to pressure on the discharge side and produces a higher flow of liquid than can be drawn in on the suction side. Suction side restrictions or atmospheric pressure decreases the flow to the pump, particularly in suction lift applications. The pump produces a higher flow of liquid than can be supplied to it, due to suction side restrictions. Discharge side cavitation is a restriction on the discharge side of the pump system which constricts the fluid flow out of the pump. Since liquid can’t escape, it is recirculated in the pump casing, damaging the outer edge of the impeller and the casing, or the casing ring if it is present. A common cause of suction side

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cavitation relates to the vapor pressure of the liquid. Liquids boil at specific temperature and pressure points. For instance, we know that water will boil at 212ºF at sea level. Carbon tetrachloride boils at 170ºF. Benzene boils at 176º F. Dowtherm will not boil or vaporize until it reaches 494.3ºF. Ethylbromide boils at 101ºF, etc. When liquids turn to a gas (boil), they will cause cavitation in a pump. Solvent transfers from outside tanks often can become a problem in the summer when ambient temperatures heat the liquids to their critical vapor pressures. The liquid doesn’t have to boil to be a problem. If it gets close to boiling and the supply to the pump is restricted, cavitation can be, and often is, the result.

How to Recognize Cavitation Cavitation is relatively easy to recognize. In its mildest form, it produces a sharp pinging noise that has often been likened to the sound of corn kernels or gravel going through the pump. If you suspect cavitation in your pump system but are not sure because you don’t hear that noise, put the blade end of a screwdriver on the pump casing and the handle end up to your ear. This will enhance your ability to hear any such noise inside the pump. Another sign of cavitation is that the discharge pressure gauge on the pump system will fluctuate wildly over a 5-to-10 psi range at a high rate of speed indicating uneven discharge flow. One must be careful to put a new gauge on the system and check the gauge tap opening to ensure the gauge is operating correctly. A properly operating system will give a

The Pump Handbook Series

steady pressure gauge reading with little or no variation during pump operation. Cavitation has numerous undesirable side effects. Because the pump is not operating in its proper hydraulic balance, it is subject to internal stresses that cause shaft deflection and premature bearing and seal wear. These are two other symptoms of cavitation. If bearings and seals constantly are being replaced in a particular pumping system, severe misalignment or cavitation is the probable cause.

What Causes Cavitation? The five most common reasons for cavitation are: 1. The pump was oversized by the specifying engineer or pump salesperson. Oversizing the pump occurs because the specifying person does not conduct a detailed system analysis to determine the proper head pressure and flows required for the application. Even when calculations are performed, there may be a tendency to “fudge” the numbers to be “safe.” In actuality, when the pump is first started up, the discharge pipes are new; therefore, the losses in the system are less than originally calculated. The resultant pump oversizing is the most common cause of cavitation. 2. The second most common reason for cavitation is a change in the system demands. This can be illustrated by a spray system where a given number of nozzles is used and the back pressure against the pump to force water through the nozzles at the desired flow rate would be 100 lbs. At 100 lbs., this theoretical pump may discharge 100 gallons per

minute on the performance curve. As the nozzles wear out, the openings through which the water passes are eroded. More water is allowed to flow through the nozzles, thus lowering the head pressure against the pump. The pump attempts to pump more and more liquid, but supply can’t keep up with demand. Now the pumping system produces only 50 lbs. of pressure at the discharge of the pump, and the flow through the pump, depending on the characteristic shape of the centrifugal performance curve, may be 300 to 500 gallons per minute. The pump is no longer operating in its best efficiency range due to a change in the system performance requirements, which may very well appear one day seemingly “out of the blue.” A common complaint is the pump was working fine yesterday, running well for years, and suddenly it begins to cavitate! 3. The third most common cause of suction side cavitation is on a suction lift or a pump whose suction side supply comes from a pit below the centerline of the pump. In this situation, debris within the pump can block the suction and restrict the appropriate amount of fluid it needs to operate at peak efficiency. Also, leaks can develop in the suction line, and air is thereby introduced into the pump. 4. As stated earlier, temperature combined with marginal suction supply can cause cavitation. Changes in the process or unusual swings in atmospheric conditions are the most commonly observed reasons. 5. Lastly, as discharge lines in the system corrode or plug, pump discharge output is restricted, and discharge cavitation can occur. Check valves not operating properly on either the pump discharge or suction side also contribute to a state of cavitation.

How to Verify Cavitation Beyond the obvious and characteristic noise described earlier and the erratic discharge pressure gauge, an inspection of the impeller in a centrifugal pump also will reveal the

effects of cavitation. It is important to note that under proper operating circumstances impellers simply do not wear out. If it appears as if “iron worms” have eaten through the center of an impeller, there is suction side cavitation. If you notice damage around the outer diameter of the pump impeller, and in the casing, this is probably evidence of discharge cavitation. Cavitation is not unique to centrifugal pumps. Cavitation is the formation of partial vacuums – or bubbles – in a flowing liquid as a result of the separation of its parts. When these partial vacuums collapse, they pit or damage parts of whatever they contact, particularly the metal surfaces or the elastomeric surfaces of a pump. In other words, cavitation affects every type of pump — centrifugal, progressing cavity, gear pumps, sliding vane pumps, airoperated diaphragm pumps — or any other device that applies energy to fluid. The laws of physics apply to all pumps and to all systems. To prove cavitation, install a combination gauge (one that reads in vacuum and psi) on the suction side of the pump and a discharge gauge on the discharge side of the pump and take the readings. The discharge pressure, plus suction pressure or vacuum, will be the operating pressure at which the pump is performing. To avoid doing extensive calculations, assume that 1 inch of mercury on the vacuum gauge equals 1.33″ of head, and remember that 1 psi equals 2.31 feet of head. (Centrifugal pump curves measure discharge in feet of head, NOT in psi.) To illustrate, if there is a suction lift condition and a vacuum pressure reading of 5 inches of mercury on the suction combination gauge, convert that to 5 feet of head. If the discharge pressure gauge reads 100 lbs, multiply that by 2.31 and see that the discharge pressure in feet of head is 231 feet. Add the 5 feet of suction head to 231 feet, and you will find that the pump is operating at 236 feet of head. If there is a positive head condition on the suction gauge and it reads plus 10 psi, multiply 10 x 2.31. The result is 23.10 feet. Deduct that The Pump Handbook Series

23.10 feet of head from the discharge pressure of 231 feet of head, and you will determine that the net pump operating point is 207.90 feet of head. Refer to the rotating shaft speed of the pump to find the pump’s operating performance curve. Then determine from these readings where the pump is operating on its performance curve. Make sure that the pump performance curve matches the motor speed. Motors can be switched from one rpm to another. To refine these measurements further, take amp readings on the motor inlet leads and convert them to brake horsepower. This will enable you to pinpoint the horsepower performance on the pump curve, which will be a double-check on the pressure readings taken earlier.

How to Temporarily Correct a Cavitation Problem This is fairly easy for suction side cavitation. If there is a valve on the pump’s discharge side (and there should be), close the valve slowly until the cavitation noise disappears. Conversely, if opening the valve to full open makes the noise disappear, this probably is discharge cavitation. Other restrictions downstream may cause the problem to continue even with the valve open. One may think closing the valve will restrict the flow of liquid to the system. In reality, the system is filled with fluid separated into gas bubbles, which may restrict the full flow of liquid which you think you are getting. By returning the pump to its correct operating condition, you produce a steady stream of gas-free fluid and this will render the most efficient flow of material from the pump that can be expected under its current operating conditions. In the case of discharge cavitation, it may be necessary to recirculate some of the liquid from the discharge side of the pump back to the supply of liquid. Do not recirculate liquid directly to the pump’s suction side as this will not alleviate the problem. With discharge cavitation, it is necessary to bypass some of the fluid out of the discharge line so that

107

the pump operates as if it is producing more flow than it really is. Fluid flow will continue, but now the pump will stop self-destructing. While the above temporary fixes will work for centrifugal pumps, they will not work as well for positive displacement pumps. Any temporary corrective action for positive displacement pump cavitation should only be done by an experienced pump technician or field engineer. A Word of Warning: Never restrict the flow on the discharge side of a positive displacement pump because it can cause personal injury or damage the pumping system.

How to Correct Cavitation So It Does Not Reoccur The only way cavitation can be eliminated is to analyze your system precisely and determine head pressure and flow requirements. This evaluation produces a system head curve that can be used to determine the correct size and type of pump to

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do the job. Many times merely trimming the impeller or changing the speed of the pump corrects the problem. Perhaps 50% of the time it is necessary to replace the pump with one more properly suited for the existing system. The remaining cavitation problems can be corrected by altering piping and/or supply elevation, or by temperature regulation. System changes can be made, including cleaning out the pipes, removing obstructions, or replacing worn components. These measures will solve the cavitation problem with little or no expense. One should also consider overall system performance and try to enhance the system when correcting the cause of the cavitation. The key is to reduce the overall cost of operation by reducing maintenance costs and improving efficiency through proper equipment sizing. Identifying a cavitation problem, understanding the physical causes and knowing how to deal with the problem can help minimize costly downtime and optimize pump performance.

The Pump Handbook Series

Helpful Formulas and Definitions Ambient Temperature The normal temperature at any given location at any given time. Atmospheric Pressure is 14.7 psi or 33.9 feet of water under standard conditions at sea level. Implosion The collapse or inward bursting of a bubble. The Net Positive Suction Head (NPSH) The total suction head in feet of liquid (absolute at the pump centerline or impeller eye) less the absolute vapor pressure (in feet) of the liquid being pumped. Vapor Pressure The pressure at which liquid will begin to vaporize. This pressure is relative to the temperature of the liquid.■ Acknowledgments for contributions to this article: Steve Cooper, Dick Bonesteel and Ed Plummer, Buckeye Pumps Inc.; Andy Fraher, Flygt Pumps; David Doty, Moyno Industrial Products; and Dana Maselli, IngersollDresser Pumps.

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PUMP AND SYSTEM TROUBLESHOOTING HANDBOOK

Hydraulic Pressure Reactions in Pump Piping Systems By Jack Claxton, Patterson Pump Company

are must be taken in the design of a typical municipal pumping station or similar applications to ensure that the piping system does not impose excessive loads on the pumps. Excessive loads create the potential for increased maintenance expense during the life of a pump. They also can increase downtime and reduce pump life expectancy. Since the purpose of a pump station is to house the pump and associated equipment, which, all totaled, may represent a very significant investment, conventional wisdom dictates providing an environment as close to ideal as possible for the pump. This will help achieve the best possible return on the investment in the total structure. A station designer following generally accepted good practice would not plan to load a pump with forces approaching those associated with, and best restrained, by pipe anchors. Yet improper attention to the piping design in the vicinity of the pump can cause pump loads approaching this magnitude and in effect use the pump as a pipe anchor – a situation that is far from ideal. This article is intended to alert readers to an often overlooked aspect of piping design that can create loads of pipe-anchor magnitude on pumps. That consideration is the hydraulic pressure reaction force of a piping system.

C

Stretching and Pulling Simply stated, all piping will stretch some due to pressure in the pipe. The amount of stretch becomes significant in a long or axially flexible pipe. When such a pipe is connected

to a pump, and the other end of the pipe is restrained or resists this elongation in some way, the hydraulic pressure reaction force on the pump may be excessive. The forces produced approach those intended for pipe anchors and can be strong enough to move the pump out of alignment, overcome internal running clearances, and excessively load the pump casing, base and anchor bolts. A shear load and moment will also be transmitted to the pump foundation and the station floor. As can be seen, failure to consider this effect can create havoc and result in an unsatisfactory installation – one that is both detrimental to the life of the pump and unsatisfactory to the owner. This is one reason why the recommendation is provided in the current Hydraulic Institute Standards ANSI/HI 1.1-1.5-1994 for centrifugal pumps, page 119, as follows: “Suction and discharge piping must be anchored, supported, and restrained near the pump to avoid application of forces and moments to the pump.” A similar statement is found in ANSI/Hydraulic Institute Standards ANSI/HI 2.1-2.5-1994 for vertical pumps, page 65.

Complications from Expansion Joints and Flexible Pipe Couplings An extreme case of pump loading due to a hydraulic pressure reaction occurs when an expansion joint or flexible pipe coupling is used between a pump and an anchor or some other restraint with no tie rods to restrain the resulting force. In this case, the force on the pump will be The Pump Handbook Series

equal to the pressure times the projected area of the maximum inside diameter. As the standards state, “This force may be larger than can be safely absorbed by the pump or its support system.” In most cases the force will be larger than can be safely absorbed by the pump because the pump typically lacks the mass and strength of a pipe anchor that is designed to restrain such forces. Consider the fact that a pump must to a certain degree be “open” or “hollow” within due to the hydraulic passageways through which water must pass in order for the pump to pump! Another factor that can make the problem even worse is when the discharge pipe has been made larger than the pump discharge by using an increaser (not a bad practice in itself), so that the pipe coupling or expansion joint is on a larger pipe. This creates a larger force on the pump. By not keeping such loads off the pump, not only is a relatively valuable machine being heavily loaded unnecessarily, natural reactions to those loads are produced in the pump foundation and pump station floor. Again, reference is made to ANSI/Hydraulic Institute Standards, which recommend that a pipe anchor be installed between an axially flexible pipe coupling or expansion joint and the pump to absorb the axial force. In cases where proper anchoring cannot be provided, the use of adequate tie rods to protect the pump and expansion joint or pipe coupling is acceptable (as is also stated by the ANSI/Hydraulic Institute Standards).

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Tie Rod Problems This brings us to a second case of excessive loading due to the hydraulic pressure reaction – loads caused by poorly designed tie rods. In determining the adequacy of tie rods, axial deflection as well as stress must be evaluated. A common mistake is to size the tie rods based on allowable stress without considering deflection. High strength steel is often used for the tie rods, and a correspondingly high value of allowable stress is used to determine the size and number of rods. If this is the case, a relatively high value of deflection will occur because the modulus of elasticity that affects deflection remains essentially the same as that of carbon steel. Rods so designed will be relatively flexible axially compared to the forces imposed upon them. While such designs may be acceptable in many instances, when they are used near a pump that must remain aligned with its driver within thousandths of an inch, the result is an unacceptably high reaction force on the pump. As a short cut to analyzing the forces and deflections involved in such an application, a specifying engineer might decide to use tie rods based on a standard that does not consider axial deflection and its effect on pumps. Many design firms use internal company standards that may be inadequate in this way. Often tie rod designs encountered in municipal or similar pumping applications comply with or are comparable to American Water Works Association (AWWA) standards. These standards (AWWA M11, Third Edition, for example) are suitable for many applications in which hydraulic thrust forces must be restrained. But in applications near pumps, the 40,000 psi allowable stress design criteria upon which they are based allows unsuitable deflection values near a pump without a pipe anchor or other means to restrain the pump side of the piping. In addition to the high allowable stresses, the length of the tie rods

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that affects axial deflection is not specified. Tie rods used near pumps are best kept as short as possible. Furthermore, axial deflection is a function of the cross sectional area of the tie rod or pipe. The cross sectional area of the recommended tie rod design for a given pressure and pipe size is found (in AWWA MII) to be approximately one-fourth to onethird of the corresponding recommended pipe. This inherently introduces an axial flexibility at the tie rod location, as compared to solid pipe. The point is not whether or not to use such standards in the design of piping systems, but to verify that the piping system will not place excessive loads on the pump in any given design.

A Case in Point To illustrate the problem, one pump installation involved a 24” sleeve type pipe coupling with four 1-1/8” diameter tie rods 46-1/2” long. The pipe was connected to a manifold that restrained the axial deflection of the 24” pipe at that end. The pipe pressure of 246 psi produced a thrust force of approximately 55 tons. The calculated axial deflection across the rods was only 0.043” but actually measured 0.72” using dial indicators. This was most likely due to the bending of the lugs that held the rods. The adjacent double suction pump, which fortunately had not yet been doweled to its base, moved out of alignment with its motor, resulting in overheated inboard pump and motor bearings. I say “fortunately” here because had this load been restrained by the doweled pump and the pump had somehow got through the start-up period without the problem being manifested, the tremendous forces imposed upon it most likely would have had an adverse effect at some time.

Alternative Solutions One approximation to avoid excessive tie rod flexibility is to design rods to have the same axial

The Pump Handbook Series

rigidity of that of the piping. The total tie rod area that provides equivalent rigidity is AR=Ap (Ep/ER), where AR = total tie rod cross sectional area, in2, Ap = cross sectional area of the pipe metal, in2, Ep = the modulus of elasticity of the pipe material, and ER = the modulus of elasticity of the tie rod material. If tie rod material and pipe rigidity are the same, it is a simple matter to determine the approximate length of pipe to the nearest pipe anchor for a given value of allowable deflection. In the case of relatively rigid pipe, it may not be practical to obtain equivalence. In this instance, using as many rods as practical is advised, keeping them as short as possible to limit their effect. Another approach is to limit the allowable value of the tie rod deflection to a very small value – say 0.005”. This may be difficult, but 0.005” is comparable to allowable horizontal misalignment values of a pump with its driver. Once the tie rod design has been established, the entire run of pipe to the nearest anchor can then be analyzed to verify that the deflection is not excessive. (Short runs of piping are best.) Using pump and baseplate rigidity information provided by the pump manufacturer, calculations can then be made to ensure that the manufacturer’s maximum recommended nozzle loads are not exceeded. Other aspects of piping design that can adversely affect a pump and that are beyond the scope of this article will need to be considered also. These include piping misalignment, thermal expansion and contraction, and the weight of the piping and its contents.■ Jack Claxton is Vice President of Engineering for Patterson Pump Company in Toccoa, Georgia. He has 23 years of experience in hydraulic and mechanical design, as well as application and field troubleshooting for vertical and centrifugal pumps. He is a graduate of Georgia Tech and is active in the Hydraulic Institute.

Pipe Anchor

ARRANGEMENT 1: NOT RECOMMENDED An expansion joint or flexible pipe coupling between the pump and the nearest anchor allows a force equal to the pressure in the pipe times the area corresponding to the maximum inside diameter to be put on the pump. This force is transmitted from the nozzle to the casing, pump tie-down bolts, base, anchor bolts, pup foundation, and pump station floor. This arrangement is not recommended by Hydraulic Institute and pump authorities, because it is impractical to design pumps to withstand this force and the pump will essentially be used as a pipe anchor. This arrangement can produce reaction forces of such magnitude to cause catastrophic failure or reduced pump life, and is therefore not recommended.

Expansion Joint or Flexible Pipe Coupling (no tie rods) FLOW

Pipe Anchor

ARRANGEMENT 2: Better than Arrangement 1 but probably not as good as Arrangement 3 because of the likely use of undersized tie rods. Tie rods are frequently designed for an adequate safety factor considering stress and using alloy steel. The alloy steel allows higher design stresses, but gives no additional resistance to deflection compared to carbon steel because the modulus of elasticity of alloy steel is the same as that of carbon steel. This sets up a potentially excessive axial deflection situation due to the axial flexibility introduced into the piping arrangement by the tie rods. For a tie rod arrangement to be equivalent to a rigid pipe arrangement (Arrangement 3) in terms of axial rigidity, the following relationship must be met: AR = AP EP ER AR = total cross sectional area of the tie rods (sq. in.) AP = total cross sectional area of the pipe material (sq. in.) ER = modulus of elasticity of the tie rod material (psi) EP = modulus of elasticity of the pipe material (psi)

Expansion Joint or Flexible Pipe Coupling with inadequate tie rods

FLOW

( (

Pipe Anchor No expansion joint or flexible pipe coupling between pump & pipe anchor.

ARRANGEMENT 3: Better than Arrangement 1 and probably better than Arrangement 2 depending on the tie rod design but there still may be excessive axial flexibility in the pipe that will load the pump due to the long length of pipe between the pump and the nearest anchor. When the pipe is pressurized, it will stretch and load the pump.

FLOW

If an expansion joint or a flexible pipe coupling is desired, place it on this side of the pipe anchor

Pipe Anchor ARRANGEMENT 4: RECOMMENDED Recommended as per ANSI/H.I. 1.1-1.5-1994 (centrifugal pumps) or ANSI/H.I. 2.1-2.51994 (vertical pumps). “Suction and discharge piping must be anchored, supported and restrained near the pump to avoid application of forces and moments to the pump...” Also, “It is recommended that a pipe anchor be installed between an expansion joint and the pump to absorb the axial force.” In addition, keep “L” as small as possible.

FLOW

L Pipe Anchor

Expansion Joint or Flexible Pipe Coupling with adequate tie rods

ARRANGEMENT 5: ALTERNATIVE DESIGN Per ANSI/H.I. 1.1-1.5-1994 and ANSI/H.I. 2.1-2.5-1994 “When proper anchoring cannot be provided, adequate tie rods must be provided and properly adjusted to protect the pump and the expansion joint. Limit tie rod axial deflection to 0.005” for best results, and verify that the total axial deflection is not excessive.

FLOW

NOTES:

CLARIFICATION - PIPE SUPPORT Shown above are typical pipe supports which support the weight of the pipe and its contents but do not restrain the pipe along its axis. Supports of this type therefore are not to be considered as anchors as recommended by the Hydraulic Institute Standards.

1. For the sake of illustration, a horizontal double suction pump is depicted. The same principles will apply to other pump types. 2. For the sake of simplicity, only the discharge side is shown, although the same principles apply to the suction side. There will tend to be more problems on the discharge side in axially flexible piping arrangements due to the higher pressures encountered. 3. This drawing discusses piping arrangement recommendations to avoid problems caused by axial flexibility in the piping when subjected to pressure. Effects of misaligned or offset piping and thermal expansion are not discussed, which can also create pipe strain and thereby cause the pump to be subjected to a reaction force.

Figure 1. Piping arrangement recommendations. The Pump Handbook Series

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PUMP AND SYSTEM TROUBLESHOOTING HANDBOOK

Developing Meaningful Pump Failure Data Improving performance by categorizing failure data: one refinery’s success story. By Oleh Berezowskyj, Clark Refining and Marketing he purposes of collecting failure data are to find problem pumps that need to be addressed, to trend improvements and to compare results against those in other facilities. The body of work order data collected over several years, even in a small facility, will be large and varied. At the same time, comparison with other facilities is difficult because there are so many different ways of reporting failure information. The approach we tried was to combine similar failures together in categories. This enabled us to analyze a large number of failures at one time, it simplified some of the calculations required to determine Mean Time Between Failures, and it forced us to define failures in detail, which should facilitate better comparisons to other facilities. The information collected is summarized in Tables 1 and 2. Our facility is a small refinery with a capacity of about 57,000 bpd. We have the standard process units: crude/vacuum, FCC, alkylation, reformer, hydrotreater, coker and isomerization. And the refinery has the usual supporting units: boilerhouse, waste water treatment and tank farm. The most common type of pump in our facility is the horizontal single stage overhung impeller pump. It accounts for approximately 60% of the pump population. We do not have an accurate count of metering pumps. Therefore, they are not included in the pump count. The average driver-rated horsepower is 72.

T

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Pump Failure Categories Seal Failures Ball/Roller Bearing Failures Case Gasket Leak Overhaul of Packed Pump Material in Pump Corrosion/Erosion Internal Rubbing Sleeve Bearing Failure Infrequent Vertical Pump Failures Geared High Speed Pumps Positive Displacement Pumps High Vibration Total Non-Pump & Non-Failure Categories Coupling Failures Minor Repairs Packing Adjustment or Re-Packing Metering (Controlled Volume) Pumps

1994

1995

1996

1997

141 30 8 13 12 5 6 0 0 18 12 15 4 264

151 33 6 4 6 4 4 0 9 22 5 6 26 276

102 37 11 5 15 10 9 2 3 25 6 3 13 241

116 22 6 2 6 4 1 4 3 24 7 5 10 210

1994 27 19 68 65

1995 24 70 54 77

1996 24 80 42 59

1997 36 30 39 68

Table 1. Pump failure categories

Categories

1994 1995 1996 1997 No. of pumps

Seal Failures 2.3 3.4 3.0 2.5 Ball/Roller Bearing Failures 8.8 7.8 13.2 9.7 Overhaul of Packed Pump 5.2 16.8 13.4 33.5 Vertical Pump Failures 2.1 1.8 1.9 2.6 Geared High Speed Pumps 5.2 4.3 3.7 2.2 Positive Displacement Pumps 5.7 11.3 6.8 2.3 All Failures 1.6 1.9 2.1 1.7 Coupling Failures 12.7 14.3 14.3 9.5 Packing Adjustment or Re-Packing 1.0 1.2 1.6 1.7 Notes: (1) Vertical In-line pumps, geared pumps were not included. (2) Metering pumps, submersible pumps were not included.

346 290(1) 67 46 26 34 449(2) 342 67

Table 2. Mean time between failures

Data Collection The failure data presented here were gathered from the Work Order Module in our Computer Maintenance Management System (CMMS). Since our CMMS was installed in The Pump Handbook Series

October of 1993, the information covers the years 1994 through 1997. All pump repairs are handled by persons in the machinist craft. Operators generate the work orders for the pumps in their unit and tag the work

orders for the machinists. Each pump work order was reviewed by analyzing the repair or discussing the work done with the machinists. A work order was not counted as a failure until it was listed as completed in the Work Order Module. The completed date on the work order was then used as the failure date. This way the failure was analyzed and the corresponding manhours and material costs were reviewed before the work order was categorized. Furthermore, the previous month’s data did not have to be updated if the repair extended into the next month. We do not have standing work orders for minor repairs. A separate work order is written for most jobs in a unit, including adjusting packing and setting bearing housing oilers. Also, operators will generate work orders to investigate problems. Both of these factors substantially increase the number of work orders in our CMMS. The information in Table 1 was limited to items under the control of the machinists craft. This was done for two reasons. First, tracking repairs by another craft adds substantially more time and effort to failure analysis. Second, some limit should be set for the types of repairs that are counted as part of a pump. If one includes motors, should the data be limited to bearing failures? Or should problems with windings, start/stop switches, switch gear and wiring be included? Should piping, control valves and suction vessel repairs also be counted as pump failures? In our facility, electricians are responsible for all the motor work including replacing bearings. Consequently, motor repairs and motor bearing failures (including motors for vertical in-line pumps) were not added to Table 1. Submersible pumps were not included in the failure data either because the electrical craft is responsible for repairing these pumps. Similarly, any work on the suction and discharge piping connected to a pump, including cleaning suction screens and repairing check valves, is handled by another craft

and was not included in this data. About 20% of the work orders were not considered repairs or failures and were not included in Table 1. They were: • Situations that were investigated and no problems were found. • Work orders involving a problem unrelated to the pump. An example of this type of work order is a performance problem caused by operational difficulties or a bad check valve. Even if the pump was disassembled, the repair was not counted if there was no problem with the pump. • Duplicate work orders. • Work orders generated to buy parts. • Work orders generated for inhouse repairs of pumps or parts carried in warehouse stock. The original work order to repair the pump was counted. • Upgrades such as installation of larger diameter impellers and mechanical seal conversions. • Preventative Maintenance (PM) work orders such as changing oil and greasing. PM and inspection work orders scheduled during turnarounds were not included either. Reworks were not counted as separate failures. The pump had to run for at least a few days before we would count the second repair as a failure. For example, if a seal leaked on start-up (the most common situation), it was not counted as another seal failure but was considered part of the original work order.

Categories The classification of failures in Table 1 is by no means 100 percent accurate. First of all, the evaluation of a failure in many cases is subjective. For example a damaged bearing is found in a pump that is being repaired because of a leaking seal. Did the bad bearing cause the seal to leak? Or was the bearing damage coincidental? If the damage was severe enough, the repair was counted as a bearing failure. Secondly, a failure can belong to The Pump Handbook Series

more than one category. A number of our positive displacement pumps have internal product lubricated antifriction bearings. Should a bearing failure in this type of pump be counted as a regular bearing failure or as a failure of the pump itself? Since the bearing is internal to the pump and is in a different environment than a bearing in a centrifugal pump, we decided to count it as a pump failure. Finally, if the definition of a category is changed, then past data should be reclassified. In our case the only information we have regarding a repair is the work order description, which is filled in by the operator, and time and materials charged to the work order. Since memories fade quickly, it becomes difficult to reclassify work orders once they are more than a few weeks old. The data for minor repairs shown in Table 1 illustrate the problem. They show a significant drop in work orders in 1997. Initially any work order that the machinists investigated and did not find a problem was counted as a minor repair. Later we stopped classifying these work orders as repairs or failures and did not add them anymore to the minor repairs category. However, we did not go back and reclassify older work orders. On many of them it was impossible to determine from the description where the work order belonged. Seal Failures: Repair done to replace a leaking seal. If failed bearings or heavy rubbing was found, then the repair was not counted as a seal failure. This category has the most amount of causes. We have found at least 10 separate problems. Ball/Roller Bearing Failures: Repair because of vibration, noise, hot bearings or a locked rotor where damaged bearings were found. A bearing that was replaced because vibration analysis indicated a problem was counted as a failure even if the damage found on the bearing was minor. It was assumed the damage was the initial stage of a failure. Most bearing failures could be divided into two types: lubrication problems and surface damage. Failures from improper lubrication were the most dramatic. The retainers were cracked, smeared or missing

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Photo 1. Smeared bronze retainer resulting from lack of lubrication

Photos 4 and 5. Heavy erosion of case and stuffing box from catalyst carryover

Photo 2. Extrusion of inner race on 5307 thrust bearing caused by overheating

Photo 6. Corrosion of impeller in sour water service with heaviest attack at the grain boundaries. Photo 3. Spalling of outer race on 5309 bearing from thrust overload

(Photo 1). Sometimes the rolling elements and bearing races were deformed and discolored from excessive heat (Photo 2). This mode accounted for about three quarters of all the failures. Most of the time we suspected that low oil level was the cause, but this was confirmed in only a few cases. In the second type of failure the bearing had some surface damage which caused noise and vibration. The work orders for this failure were generated by the operators and by our vibration analysis program. The

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damage included spalls and dents on raceways and marks on balls and retainers. The source of the damage was not determined in many of these failures. The causes we were sure of are dirt in the oil, corrosion of the races and impacts from poor mounting practices. One other type of failure we have found is heavy spalling on bearings under excessive loads (Photo 3). This problem has occurred on only four pumps, all manufactured in the 1940s. Case Gasket Leak: Repair that required replacement of the case gasket or clean-up of the case gasket seal The Pump Handbook Series

area. About a third of all the failures have occurred on pumps in hydrofluoric acid service. The cause is possibly corrosion under the monel overlay in the casing. Other possible causes are improper installation, damage to the gasket surface and corrosion of the gasket surface. Overhaul of Packed Pump: Work order for packing sleeve replacement or stuffing box repair to correct excessive packing leakage. More than 60% of the work orders were generated for four pumps in boiler feed water circulating service, which we recently converted to mechanical seals. Material in Pump: Work order generated because of poor performance, locked rotor or high vibrations where material (metal parts, coke, etc.) was found in the impeller or suction screen or where the impeller was plugged with deposits from the product (lime, salts, etc.) Corrosion/Erosion: Work order generated because of poor performance or leakage through the pump casing where heavy corrosion or erosion of the casing, impeller or wear rings was found (Photos 4, 5 and 6 ). Three occurrences of impeller damage due to cavitation were included in this category. Internal Rubbing: Work order generated because of high vibrations, seal leakage, locked rotor or poor performance where heavy rubbing of wear rings and bushings was found upon disassembly. Most of these failures have occurred on impeller between bearing type pumps, single and multistage. We suspect loss of liquid in the pump is the most common cause of these failures. A few of the failures include pumps in our alkylation unit. These pumps have an unusual internal corrosion problem that causes the case wear ring to grow into the impeller wear ring and rub. Sleeve Bearing Failure: Repair where a sleeve bearing was replaced due to babbitt wiping. We have only seven pumps with sleeve bearings. All but one of the failures have occurred on two pumps in the same service. We are still investigating these failures. Infrequent: A category encom-

passing odd failures such as: • water jacket, water cooler or pump casing fractured from freezing • cast iron pump casing fractured from mishandling • failure of impeller bolting on a overhung pump caused by chemical attack Vertical Pump Failures: Work order generated because of vibration or poor performance on a vertical pump. A vertical pump included in this category is any with a long shaft and bushings. This includes sump pumps and vertical turbine pumps (both deep well and can type), but not vertical in-line pumps. Seal failures and corrosion/erosion of the pump were counted in their own categories. However, failures caused by restricted suction or material in the impeller were still counted in this category. These failures were categorized this way because damage to the bushings and shafts caused by inadequate flow appeared to be common in this type of pump. Geared High Speed Pumps: Failure of the gearbox bearings, gears or oil seal. Product seal leakage, corrosion/erosion, material in the pump and case gasket leaks were included in their respective categories. If a failed gearbox was found on a work order written to repair a leaking product seal, it was assumed the leak was caused by the gearbox vibration. The repair was counted as a gearbox failure. Positive Displacement Pumps: Any repairs. This category combined rotary and reciprocating pumps because we have only four reciprocating pumps. Repairs of leaking seals were counted in the seal failures category, but failures of internal product lubricated roller bearings were kept in this category. Most of the work orders for rotary pumps were for performance problems. All the work orders for the reciprocating pumps were generated for two pumps that are no longer in service. The problems listed on these work orders included performance, noise, packing leakage and oiler malfunction. High Vibration: Work order for high vibration generated by our vibration analyst (and sometimes by

operations personnel). The purpose of this category was to count predictive maintenance type repairs in which work was done (alignment, balancing) before a failure occurred. Most of the repairs (80%) required correcting misalignment and looseness (usually loose bolting) and did not require disassembly of the pump. Those that required disassembly — mostly unbalance and vane pass vibration — were kept in this category if no obvious problems (internal rubbing, corrosion/erosion, material in pump) were found. Work orders generated due to vibration analysis indicating a bearing defect were almost always counted in the ball/roller bearing failure category. Whenever a bearing was cut apart and analyzed, some kind of imperfection was found. The imperfection could be considered the initial stage of a failure. To be conservative, these work orders were added to the ball/roller bearing failure category. Poor performance for centrifugal pumps is not listed as a category because in all cases where a work order was generated because of poor performance one of the previous failure modes listed in the above categories was found to be causing the problem. Replacement of wear rings is not listed as a category either because all the replacements of wear rings to date were due to rubbing or corrosion/erosion, and the failures were counted in their respective categories.

Non-Pump and Non-Failure Categories Coupling Failures: Most of the couplings in our facility are urethane donut shaped elastomers. The most common failure mode for these couplings is cracking because of material degradation (“aging”). The rest of the couplings are gear, grid, metal disk and miscellaneous types of elastomers. The most common failure mode for gear and grid couplings is lack of lubrication (Photo 5). A few failures have been due to misalignment. Since the predominant failure modes are a function of the coupling condition, we feel coupling failures are independent of the driven and driving equipment. Therefore, they The Pump Handbook Series

were not counted as pump failures. They can be added back to the failure categories if deemed necessary. Minor Repairs: Any work that did not require substantial disassembly of the pump and where the pump was operable was counted as a minor repair. Minor repairs included adjusting or replacing oilers, tightening bolting, checking oil rings, repairing flush or cooling water piping, changing oil filters, replacing oil seals and adjusting or replacing belts. Packing Adjustment or Re-Packing: This type of repair, considered normal wear, was not counted as a failure. Metering (Controlled Volume) Pumps: Our metering pumps do not have equipment numbers, and we do not maintain the same type of records on them as we do on other pumps. Therefore, the work order information is not that accurate and was not included with the other pump failure data. Table 1 lists the number of work orders generated to do some kind of work on metering pumps. Also, in our facility the pipe-fitters craft can install metering pumps. In cases where a work order for a metering pump was assigned to the pipefitters, it would not be listed in the table. Portable air operated diaphragm pumps, which are handled the same as metering pumps in our facility, were included in this category.

Mean Time Between Failure (MTBF) Calculating traditional MTBF for an individual piece of equipment is relatively easy. From equipment history, take the first failure date and subtract it from the last failure date. This is the total time span for all the recorded failures. Count the number of failures and subtract 1. This is the number of operating periods of the equipment in the time span. Divide the time span by the number operating periods and you have MTBF. Note that MTBF cannot be calculated for equipment with zero or one failure. MTBF =

Time Span

(Last Failure Date - First Failure Date)

Number of Operating Periods (Number of Failures - 1)

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A variation of the MTBF calculation ignores the true time span. Instead, the time span of the equipment history database is used. The database time span is divided by the number of failures, not operating periods, to produce an approximate MTBF. This eliminates the need to find the first and last failure dates. With this method MTBF cannot be calculated for equipment that has never failed. MTBF = Equipment History Database Time Span Number of Failures

Number of Failures MTBF calculation for a group of pumps is somewhat more complicated. A group can be all the pumps in a unit or facility, a particular type of pump (centrifugal single stage overhung) or a particular configuration of a pump (pumps with seals). I have used two methods of calculating MTBF for a group. The first method, which is the one presented in the tables, is not a true MTBF. It is a number that represents an average time between failures of a component or pump type for an average piece of equipment. This type of calculation is easy and allows evaluation of failures the first year information is collected. Table 2 shows the MTBF for all the pumps in our facility and the MTBF for the most significant failure categories. MTBF was not calculated for categories with a small number of failures and large population such as case gasket leaks and internal rubbing. Using the seal failure category for 1997 as an example: MTBF = 346 pumps with seals/116 seal failures per year = 3 years. The second method calculates the MTBF for each pump and then determines the average of all the individual MTBFs. Since pumps with zero failures do not have a MTBF, a dummy failure is added to those pumps to permit calculation of the average. For this type of calculation the time span of the equipment history needs to be as great as possible in order to reduce the number of added failures. Calculations for only one

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year will not be accurate because a large number of dummy failures will have to be added to the database. This hinders calculating MTBF in the first year data are collected. Moreover, trending MTBF requires using failures from the beginning of the equipment history. This muddies the effect of recent improvements. Also, the second method is insensitive to problem pumps. As the number of failures for an individual pump increases, its MTBF decreases approaching zero. In a large population of pumps, a few pumps with very low MTBF (high failure rates) will not affect the overall average. Significantly different MTBF can be produced depending upon which method of calculation is chosen. For our facility, using data from the last four years, the overall pump MTBF was calculated individually and averaged. The result was 2.82 (15% additional failures were added to eliminate zero failures). On the other hand, the overall pump MTBF calculated using the total number of pumps divided by total number of failures divided by the time span equaled 1.82. An accurate count of pumps is necessary to calculate MTBF correctly and to compare accurately against other facilities. We inventoried the pumps in the whole refinery before evaluating the MTBF. In two other facilities we found that almost half of the pumps on the master pump list were removed or abandoned in place. Total number of pumps installed, not the number of pumps in operation, was used to calculate MTBF. This way we did not have to account for spare pumps operating in parallel with main pumps and for intermittent services such as pumps in the tank farm. Refineries and most other facilities are structured the same way — spare pumps for most services, single pumps for intermittent services. Thus, comparisons between facilities should still be accurate.

Summary The amount of repair data generated for pumps can be overwhelming. Trying to allocate one’s time and The Pump Handbook Series

the company’s resources effectively requires accurate information. Categorizing failures can reveal enlightening patterns and highlight problem areas. From Tables 1 and 2 we can conclude that: • Seal failures account for the majority of the failures and deserve the attention that they receive. • Improving lubrication will have a significant impact on pump failures. • Vertical pumps are a problem area that needs to be addressed. Comparisons from year to year in a particular category can readily show where improvement efforts, both equipment specific and plantwide, are making a difference. Table 1 shows a 30% reduction in seal failures from 1995 to 1996. Over the past few years we have been troubleshooting problem seals and improving materials, designs and flushes for individual services. Also, we have involved our machinists in correcting seal problems. Half of the reduction is due to seal modifications, and we believe the other half is due to higher quality repairs. There are many theories about failure reporting, and they all provide useful information. However, standardization — or at least an understanding of how the failure reporting was done — is necessary so facilities can compare against each other. The use of agreed-upon categories promotes development of the kind of detail needed to make meaningful comparisons. Moreover, with categories one can account for differences in failure data. For example, metering pumps are not included in the MTBF for all failures because our count of these pumps is not accurate. With categories we can still compare ourselves to other facilities because we can account for this difference in the data.■ Oleh Berezowskyj has held various positions as Rotating Equipment Engineer and Supervisor in corporate engineering and in refinery maintenance during the past 23 years. He is currently the Rotating Equipment Supervisor at the Clark Refining & Marketing refinery in Hartford, Illinois and is a Registered Professional Engineer in that state.

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PUMP AND SYSTEM TROUBLESHOOTING HANDBOOK

The Mystery of Cooling Tower Pump Noise by Steve Schmitz, Bell & Gossett

he problem of ”cavitating condenser water pumps” with adequate NPSH available is not uncommon. We have discussed this phenomenon with cooling tower manufacturers and other centrifugal pump designer/manufacturer members of the Hydraulic Institute. All agree that this condition occurs predominantly in cooling tower applications. We see on average one or two such cases per year. At present, several theories have been offered to explain the cavitation-like noise. None has been validated. We do know, however, the following:

T

1. The noise is very similar, if not identical, to classical cavitation (resembling the sound of marbles being pumped). 2. The phenomenon can occur with either a forced draft or induced draft cooling tower. 3. The noise tends to be more prevalent on negative suction pressure systems, but it will occur on positive suction pressure systems as well. 4. The introduction of small amounts of air to the pump suction often quiets the noise. This entrained air has little effect on a pump’s life expectancy. 5. Such small amounts of entrained air have little deleterious effect on other system components. However, each system must be analyzed for possible harmful effects. 6. Unlike classic cavitation, throttling the pump discharge to a lower capacity usually has little impact on the noise level. In an effort to determine the probable causes, we visited a site experiencing such complaints. We

conducted a detailed inspection and made an audio recording of the noise spectra for laboratory study. The analysis revealed the following: 1. There were no distinct frequencies. 2. The predominate noise measured was broadband, occurring above 300 Hz. Pump noise can have both liquid and mechanical causes. Both sources produce acoustic pressure fluctuations that can be transmitted as audible noise. For centrifugal pumps, mechanical noise is generally the result of component imbalance (impeller and/or coupler), coupler misalignment, components rubbing against each other, or improper installation of the base plate and/or motor. These problems generate distinct frequencies equal to rotational speed and/or its multiples (1,2,3). Because the noise spectra did not reveal distinct frequencies, we determined that this noise was not

mechanically generated. Liquid noise is produced directly by water movement and is fluid dynamic in nature. Turbulence, flow separation (vortex), cavitation, water hammer, flashing and impeller interaction with the volute cutwater are all examples of fluid dynamic noise sources. According to the Pump Handbook, 2nd Edition, by Igor J. Karassik, there are generally four types of pulsation sources in pumps that are the result of liquid noise: 1. discrete frequency components generated by the pump impeller 2. broadband turbulent energy resulting from high flow velocities 3. impact noise consisting of intermittent bursts of broadband noise caused by cavitation, flashing and water hammer 4. flow-induced pulsations caused by periodic vortex formation when fluid moves past obstructions and side branches in the piping system

Figure 1. Incorrect suction piping and reducers installed upside down can make it possible for air to collect, creating an obstruction. The Pump Handbook Series

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Eccentric Reducer

Figure 2. Cooling tower vortexing is the most common source of air in the pump

Discrete frequency, item 1, can be ruled out in this situation. As previously mentioned, we did not find distinct frequencies such as the vane passage frequency and/or its multiplies. The frequencies would have been present if an interaction had occurred between the impeller and volute cutwater. Items 2, 3 and 4 are generally identified as broadband noise and would occur in the 300 Hz and above frequency range as identified on our noise spectra. Therefore, we believe the noise was being generated by one or more of these liquid sources. The pump noise we heard was like that of cavitation. Pump cavitation results from the formation of vapor bubbles when the localized static pressure is lower than the vapor pressure of the liquid being pumped. To evaluate a pump for classic cavitation (NPSHR greater than NPSHA), close the discharge valve, thus pushing the pump back on its curve toward shutoff. The noise should diminish significantly if it is originating from classic cavitation because lower pump flows require reduced NPSHR. If the noise continues, the cause is likely to be entrained air. We determined that classic cavitation was not occurring, as the operating suction pressure measured 30 feet above vapor pressure. Thus, the NPSHA was approximately twice that required by the pump. For that reason, we knew the pump was not cavitating because of insuf-

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5 to 10 pipe diameters Straight Pipe

Figure 3. Proper suction piping

ficient NPSHA. If not classic cavitation, then, what was causing the noise? It is a well-documented fact that highly aerated cooling tower water can contain as much as 4-6% excess air. This excess air increases the potential for a noisy pumping installation. The excess air absorbed in the cooling tower comes out of solution as it flows through the piping and becomes entrained air. Suction velocities are often high enough to pull the air through. However, air sometimes collects in an area of the suction piping or the impeller eye itself, creating an obstruction (Figure 1). As the liquid passes through this restricted area, its velocity increases, creating an area of reduced pressure. At this point of reduced localized pressure, water vaporizes, with the resulting bubbles passing into the pump impeller where, as the pressure increases, they collapse and produce ”cavitation.” Several noise control techniques have been successfully employed in the past to mitigate excessive noise. They include: 1. Increasing or decreasing the pump speed to avoid resonances in the mechanical or liquid systems and/or reduce the pump NPSHR. 2. Increasing liquid pressures (NPSHA, etc.) to avoid cavitation or flashing; decrease suction lift. This could include raising the tower, lowering the pump or straightening the suction piping The Pump Handbook Series

(see “Other Contributing Factors”) to reduce friction losses. 3. Modifying the pump so that the clearance between the impeller blade tips and casing cutwater (tongue) or diffuser vanes is increased. 4. Injecting a small quantity of air into the suction of a centrifugal pump to reduce cavitation noises by providing a shock absorbing cushion that minimizes the impact of recondensation of water vapor within the pump’s impeller. Injection of small amounts of air can usually be accomplished quickly and easily in the field with minimal expense. Small amounts of entrained air usually cause no problem in the cooling tower/condenser circuit. Bell & Gossett therefore considers this alternative as desirable and recommends its application as a solution to many field problems or, at a minimum, as an analytical tool.

Other Contributing Factors In addition to the techniques outlined above to reduce or eliminate noise, attention must also be given to two other factors that can exacerbate the situation: vortexing of the liquid in the tower pan, which is the most common source of air within a pump (Figure 2), and the suction piping arrangement (Figure 3). Vortexing of Liquid in the Tower Pan The amount of entrained air caused by vortexing depends on sev-

Recommended Installation

Not Recommended

Long Radius Elbow

Inlet Parallel to Pump Shaft

Figure 4. Correct and incorrect elbow installation configurations

eral variables, but particularly the vortex size and the submergence level of the pump suction pipe below the water level of the pan. The most common method of eliminating vortexing in the tower pan is by including baffle assemblies that prevent vortexes from forming. Raising the fluid level in the pan to a sufficient depth can also solve this problem. Suction Piping Coupled with the vortexing phenomenon, or by itself, improper layout of the pump suction piping can be a significant contributor to the generation of pump noise. Friction losses caused by undersized suction piping can increase the fluid’s velocity into the pump. As recommended by the Hydraulic Institute, Standard ANSI/HI 1.1-1.51994, suction pipe velocity should not exceed the velocity in the pump

suction nozzle. In some situations pipe velocity may need to be further reduced to satisfy pump NPSH requirements and to control suction line losses. Pipe friction can be reduced by using pipes that are one to two sizes larger than the pump suction nozzle in order to maintain pipe velocities in the 5 to 10 ft/s range. Eccentric reducers used to step down to the pump flange from the larger suction piping can also be a culprit if they are used improperly. At the problem facility discussed earlier, the reducer was installed upside down, with the flat side on the bottom (Figure 1). If the liquid contains air (or vapor), as it did in this case, the air can become trapped in the sloped area of the reducer now located on ”top.” At a minimum this will obstruct the flow passage, causing higher velocities and thus localized vaporization. If transported into the impeller, the trapped air can create a momentary choking that could even result in shaft breakage. Elbows used on the pump suction flange, while convenient, can cause an uneven flow of liquid into the impeller if the elbow bend is along the axis of the pump shaft (Figure 4). If the elbow is a short radius design, its use may unintentionally create turbulence that produces entrainment that can, and does, worsen noise problems. The addition of a second elbow only increases the problem, especially if the elbow has been placed at a right angle to the existing elbow. Numerous technical publications, as well as the Hydraulic Institute itself, state that systems

should have a minimum of five pipe diameters of straight run of pipe before the pump suction flange to allow for a smooth unimpeded flow to the impeller (Figure 3). System strainers need to be located on the discharge side of the tower pumps and not on the suction side (Figure 5). On another project, the location of basket strainers directly in front of the suction flange on a large HSC pump resulted in an unexpectedly high pressure drop. This contributed to poor pump performance in that installation, as well as higher pump noise levels.

Conclusion It must be understood that each job site has its particular set of operational requirements and, therefore, there is no single solution to the mystery of cooling tower noise. The Hydraulic Institute has published a standard that provides recommendations for an ample margin of safety between NPSHA and the pump manufacturer’s published NPSHR. The safety margin would be a minimum of 1.7 times NPSHR or NPSHR plus 5 feet, whichever is higher. ■ Steve Schmitz is Senior Product Line Manager, Centrifugal Pumps & Engineered Products, at Bell & Gossett. He has been with the company for 14 years and has written many articles on pump application and operation. He is the concept originator and was a member of the development team for the successful ESP software program. He is also a member of the Hydraulic Institute.

Service Valve Triple Duty Valve

Strainer

Concentric Reducer

Long Radius Elbow

Figure 5. Strainers need to be located on the discharge side of the tower pumps and not on the suction side. The Pump Handbook Series

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PUMP AND SYSTEM TROUBLESHOOTING HANDBOOK

Controlling Surge and Pulsation Problems The plant may be imaginary, but the problems are real. Take the tour and learn the most common causes of problems like water hammer, and how you can prevent them. by Gary Cornell, Blacoh Fluid Control, Inc.

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here was no mistaking the sound—that rushing metallic thud. As the valve slammed shut, a pressure wave traveling at 4000 feet per second and more than five times normal system pressure instantly crashed into the one-way check valve protecting the main supply pump from back flow. Once again, the integrity of the system held, but everyone knew it was just a matter of time. This article addresses the all too common scenario of working near pumping systems that are on the edge of catastrophe. The causes, results and solutions to the problem of surge or “water hammer” will be discussed, as will the different but related problem of pump-induced pulsation. To look at the practical side of surge and pulsation, we will take a trip through an imaginary, but realistically depicted, modern manufacturing plant. To set the stage for our tour, we need to describe hydraulic conditions that set up the potential for these problems to occur. Both surge and pulsation in a liquid handling system are the result of uncontrolled pressure waves caused by an abrupt change in flow, either directional or volumetric. Liquids contained in an enclosed system (piping) have a physical volume; therefore, a mass can be measured and/or calculated. We can then determine the acceleration forces needed to move that given mass. Once in motion, the mass will stay in motion as long as enough force is

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applied to overcome friction loss plus any gravitational component. It can be said that hydraulic equilibrium is reached when the fluid is flowing in a laminar state. Since for all practical purposes liquids are not compressible, force or energy is not absorbed into the fluid but rather is transferred through it. The kinetic energy of the moving fluid will exert all the force it has acquired to resist any condition that tries to cause a change in its velocity. Depending upon the mass and velocity of the liquid and the rate of change applied to it, very destructive forces can be generated, leading to catastrophic component or system failure. The critical consideration for this discussion becomes the rate of change in energy for any given mass. For example, consider the analogy of a battleship moving at full speed. If the engines are stopped, the ship will travel five miles in bleeding off the kinetic energy through water friction before coming to a stop. No damage would be incurred because the massive amount of energy involved is allowed to change form slowly. However, if that same ship at full speed were to run squarely into an aircraft carrier, all the kinetic energy would be concentrated and would change form very rapidly—in a matter of seconds. Since the greater mass of the carrier could not absorb the kinetic energy as fast as it was being delivered, the energy would become concentrated at the point of The Pump Handbook Series

impact. In liquid transfer systems, kinetic energy is generally observed as pressure. Therefore when fluid velocity is changed, the result is an increase or decrease in pressure. Commonly, changes in velocity occur when pumps are started or stopped (either intentionally or due to failure), fluid flow direction changes abruptly, pipe diameters change abruptly or a quick-closing valve shuts.

Surge At least in potential, a quick-closing valve (generally one that closes in less than 1.5 seconds) represents the most dangerous condition. Since liquid efficiently transfers energy rather than absorbs it, a sonic or pressure wave is created. The intensity of the pressure shock wave is directly proportional to the speed of flow before the change in velocity and to the speed of propagation of the sonic wave created. Unrestricted, the pressure spike in the liquid will rapidly reach the speed of sound moving through liquid (approximately 4700 feet per second). This is more than four times the speed of sound through air. It is mathematically possible to calculate this pressure increase. For design purposes, the increase in pressure can be determined by the rule-of-thumb formula: P = 60xVS Where P = The increase in pressure over the steady-state system flowing pressure V = Flow in feet per second of

the fluid before valve closure S = Specific gravity For example, consider a 2” pipe carrying 60 gallons per minute, with a system pressure of 100 psi and a specific gravity of 1.2, and a valve closing in one second. The 2” pipe and gpm equates to a flow rate of 6 feet per second. In this scenario, the pressure increase above normal flowing pressure would be 432 psi, creating a total spike pressure of 532 psi. If the valve were shut in half a second, the peak pressure would double to 1064 psi. Conversely, if the valve were to close in 2 seconds, the peak total pressure would be half of 532 psi, or 266 psi. Note that the pressure increase is independent of, but cumulative to, the normal system pressure. For example, a system pressure of 50 psi flowing at 10 feet per second would have the same pressure rise for a given rate of change in velocity as a system pressure of 300 psi flowing at 10 feet per second. The proportional pressure change, however, would be a much higher percentage. Therefore, regardless of whether it is low or high pressure, the system may not be able to handle the pressure wave. Clearly, when a mass such as liquid in a pipe is in motion and its velocity is changed, there is the potential for a catastrophe.

Pulsation Unlike surge, pulsation is the rapid uncontrolled acceleration and deceleration of units of energy. In the context of this article, these units of uncontrolled energy are actually slugs of liquid moving through a pipe. The degree of pulsation in them is usually designated by frequency in Hertz and a pressure amplitude (See Figure 1). Outwardly, pulsation is usually observed as component vibration or rapid gauge fluctuation. On an oscilloscope it appears as a sinusoidal curve (waves of peaks and valleys). Pulsation can occur and/or be influenced by the specific harmonics of various components in a liquid transfer system. Piping, valves and mechanical movement—and system design itself, to a certain degree—

combine to influence measurable can be many multiples of that genvibration. erated by the pump alone. This is However, the system component more likely to occur at higher frethat instigates the pulse generation quencies such as those generated by is the pump—specifically, a reciproa centrifugal or multipiston pump. cating, positive displacement pump. The Tour This type of pump creates its motive force by repeatedly capturNow that we have explained fluid ing and expelling a predetermined movement and discussed the undeslug or volume of liquid. It does this sirable effects that changing dynamby using inlet and outlet valves, ics can produce, it is time to visit which account for rapid acceleraour imaginary plant (Figure 2). For tion and deceleration of fluid. our tour we’ve selected the Quality Pulsation, then, is a rapidly repeatPaint Company. QPC is one of the ing change in energy form. largest paint manufacturers in the Depending upon its frequency and United States, with eight plants from amplitude, the potential for catacoast to coast. They specialize in strophic system component failure water soluble paints using state of is very real. A simplistic comparison the art manufacturing equipment. would be to the weakening effect of Stop 1: Tank Farm Transfer the human arterial system caused by Pumps a constant elevation in heart rate and blood pressure. As we pass the guard shack, the The major pulse in a pumping first thing we notice is a tank farm. system will be at the frequency of QPC purchases titanium dioxide the plunger or piston speed times (TiO2) solution in bulk and stores it any multiplicity factor. By way of in eight 10,000-gallon tanks. The reference, a reciprocating pump’s TiO2 is transferred into the facility pulsation is generally described as as demand requires. For several reahigh amplitude, but low frequency, sons, including the abrasive nature as compared to the high frequency of the material and the sheer sensibut low amplitude frequency of a tivity and operational characteristics centrifugal pump’s impeller vane. required, the transfer pumps speciSeldom are liquid handling systems fied are air operated double designed with an analysis of the diaphragm units (AODD). Each tank total system harmonics that will occur. Whenever a reciprocating pump is called for, at the very least consideration must be given to the potential effect of the pulsation generated. In most cases, minimizing pump-generated pulsations will provide sufficient system protection. However, if the need is to eliminate total system vibration, rather than pump-generated pulsations, then both system and pump harmonics must be considered. In such cases, if there happens Pulsations of ± 30 psi have been reduced to ± 3 psi with to be a harmonic a SENTRY III Dampener match of frequencies, Figure 1. Undampened vs. dampened pressure pulsapressure amplification tions in a metering pump

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has its own pump supplying product to a common header pipe going overhead into the plant. Even though AODD pumps are the best choice for this application, they have a rather poor tolerance for high inlet pressure. When the inlet pressure approaches 12 to 16 psi, pump diaphragm life starts to deteriorate rapidly. Since QPC’s TiO2 tanks are 30 feet tall and their particular mixture has a specific gravity of 1.3, the static inlet pressure is 16.8 psi. (feet in height x .432 x 1.3 s.g.) AODD pumps are positive displacement, so they have inlet valve balls that act exactly like quick-closing valves. From our previous discussion we know that quick-closing valves create instantaneous pressure spikes several times higher than flowing pressure. Since 16.8 psi is above the pump manufacturer’s maximum allowable inlet pressure, we can predict that there will be problems. What happens in the pump is pretty straightforward. The spike created when the inlet valve balls close on the liquid chamber seeks to move to a lower pressure area. Since the inlet valve on the pump’s other chamber is simultaneously opening and the motive diaphragm is creating a vacuum, this becomes the low pressure area. The pressure spike rushes in and slams against the diaphragm. This distorts and stresses the diaphragm, leading to premature failure. Positioning We see that the pump is located at the bottom of and just two feet or

so away from the tank. Because it is this close to the tank, acceleration head is not a significant factor. One method of reducing high inlet pressure is to reduce the height of the tank, but this is not practical here. The AODD pump has several features that make it the best choice for this application. So what can what can be done? A practical and economical solution is to install an inlet stabilizer as close to the pump’s inlet as possible, but within 10 pipe diameters. Inlet Factors An inlet stabilizer is a hydropneumatic device consisting of a pressure vessel with an elastomeric bladder or diaphragm inside it that separates a compressed gas charge from the liquid being pumped. The inlet stabilizer acts literally like a shock absorber and receives the pressure spike created when the pump’s inlet valve closes. An inlet stabilizer is typically precharged to 50% of the static inlet pressure. When properly charged and sized, it will minimize the stress on the pump’s diaphragms by temporarily accumulating liquid and absorbing the pressure spike. This minimizes stress on the pump’s diaphragms. Discharge Factors Now that the pump has been protected, the discharge needs to be examined. The AODD pump produces a pulsating flow. As previously mentioned, pulsation is the rapid acceleration and deceleration of liquid caused by reciprocating action

Figure 2. The imaginary Quality Paint Company plant

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The Pump Handbook Series

Photo 1. Transfer pumps fitted with pulsation dampeners to alleviate shaking of overhead piping.

of the positive displacement pump in conjunction with the quick opening and closing of the pump’s discharge valves. This pulsing flow will be observed as pipe vibration as energy rapidly changes form. Vibration Concerns—Mechanical and Liquid There is also mechanical vibration caused by the shifting of pump components. And remember, if both mechanical and hydraulic pulses coincide, vibration will be increased by a multiple factor. Due to the overhead piping configuration, vibration is a serious problem. (Photo 1) There are several ways to minimize the hazards caused by this vibration. Larger diameter pipe can be used to reduce back pressure. Thick-wall pipe can be installed to delay the eventual pipe fatigue. And pipe braces (not hangers), used for support, can absorb some of the vibration. In effect, the system can be overbuilt to absorb and disperse the vibration caused by pulsation. Back pressure valves and orifices can provide some dampening effect but usually at the expense of efficiency. All of these solutions, however, simply address only the symptoms of pump vibration and pulsation. They will be costly and only delay inevitable component and/or system failure. The most economical and proper way to minimize a pump’s mechanical vibration is to use some type of flex coupling between the pump and the discharge pipe. This isolates the system and prevents component damage. Several companies supply this product. Sometimes, however, all that is required is a length of

hose reinforced to withstand system pressure. As a rule of thumb, the hose should be at least 15 pipe diameters long. The most economical and efficient way to dampen the liquid’s hydraulic pulses is probably a pulsation dampener. This device is similar in construction to the inlet stabilizer, but it must be pressurized to 80% of the liquid flowing pressure. The pulsation dampener will absorb the spike created by the rapid acceleration of the liquid during the discharge stroke of the pump. At the same time it will accumulate a small amount of the liquid. When the pump shifts, pressure is momentarily reduced as discharge pressure is lost and the dampener releases accumulated fluid back into the pipeline, filling the void created during the pump shift. This will minimize vibration. Also, liquid flow downstream of the dampener will now be in a continuous steady state rather than a pulsating start/stop mode.

Stop 2: From Mixing Tanks to Blenders As we continue our tour and enter the plant, we see large mixing tanks where the paint is blended. Since the main ingredient in water based paint is water, we see an 8” water line running overhead and the length of the plant. Coming off this main header are six 4” branch lines, one to each blending tank. Following the main line back, we observe that the water flow is produced from a holding tank by an 8” end suction centrifugal pump producing 1500 gpm at 180 feet of head. A one-way stop check valve prevents system back flow, protecting the centrifugal pump and creating a pressure tight seal. The centrifugal pump starts on demand. The demand is based upon pressure changes in the system when valves at the blending tanks open. These valves must be the quick-closing type because when a predetermined weight of water is let into the blending tanks, flow must immediately stop. From our earlier discussion on surge or ”water hammer,” we know that as a mass changes velocity,

there is the potential for rapid energy transformation. In this water feed system, we are faced with the following potential problems: 1.Rapid start-up of the centrifugal pump against a pipeline full of static fluid. 2.Rapid shutdown of the centrifugal pump by operator or by motor failure. 3.Rapid closing of the valves at the blending tanks. Although there can be other design factors of the system that either minimize or exacerbate such hydraulic problems, the items above represent the greatest potential for disaster. Rapid Pump Start QPC’s centrifugal pump will start automatically when the system pressure drops due to the opening of a blending tank’s valve. When this occurs, the pump will throw water into the 8” pipeline that is filled with a stationary water column. As the rapidly moving water collides with the stationary column, a pressure (energy) spike will occur. This is similar to an automobile running into a block wall. Great stress will be put on the system. To prevent this from happening, several options can be considered: Slow or soft start pump motors will introduce water into the system slowly and minimize the pressure spike. This is a costly but effective solution. Slow opening isolation valves will be closed when the pump starts up. They will open slowly, allowing gradual pressure and flow increases to equalize force in the stationary column of water. This solution can be complicated and expensive from a control point of view, but it will minimize the pressure spike. Surge tanks are closed vessels with air trapped in them. When the pump is started, initial water flow will enter the surge tank until system pressure is equalized. The surge tank’s major drawback is that the air trapped in it will be absorbed, creating aerated fluid that is undesirable in paint products. In addition, once a waterlogged condition exists, all cushioning effect will be lost. Surge suppressors are similar in The Pump Handbook Series

Photo 2. Metering pump installed with a pulsation dampener to inject chemicals into a process line.

design to surge tanks, but there is an elastomeric bladder inside separating the fluid from a compressed gas charge. By capturing the gas charge, the proper pressure required for the specific application can be maintained. The suppressor should be charged to approximately 85% of normal operating pressure. A surge suppressor is an effective and cost-efficient way to minimize the pressure spike that occurs during pump start-up. Rapid Pump Shutdown When the pump is turned off, the flow at the pump’s discharge stops quickly. This creates a low pressure area. Fluid column separation can occur as momentum temporarily continues to move the mass of water down the pipeline. As soon as friction acts to slow the forward motion of the liquid column, it will reverse and travel back toward the low pressure area at the pump discharge. The reversal will travel at the same speed as its initial forward motion. Depending upon the initial flow rate, the pipe gradient and the fluid mass, the pump casing will be stressed, with a potential for failure. Also, pump seal integrity can be lost, and the impeller can become warped. Additionally, systems with

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Photo 3. Pulsation dampened air operated diaphragm process pumps

Photo 4. Paint filling machine dispensing into five-gallon buckets. Accuracy and reduced splashing are achieved by the smooth flow provided by pulsation dampeners.

a check valve can create a water hammer effect with catastrophic transient pressure spikes when the reverse flow hits the one-way check valve. There are several solutions to this problem: Controlled-close check valves can be used to time the closing period. A dashpot device or motor control on the valve can be used to accomplish this. If we control the rate of change in the velocity of flow, we can minimize the amplitude of the pressure spike. This type of control valve works well in conjunction with check valves, but they must be maintained for reliability, and they can be an expensive option. Surge suppressors can be installed just downstream of the check valve or, if no check valve is used, at the

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pump discharge. Properly sized and precharged to 50% of operating pressure, a surge suppressor will accomplish two things. First, being 50% precharged, the vessel will accumulate fluid during pump operation, and this accumulation will be released at shutdown to prevent column separation. Second, the suppressor will absorb the pressure spike generated as the water column reverses against the check valve or pump. The suppressor is a straightforward and economical approach. Pump motor failure or loss of power is another contingency that must be considered. While modern electric pump motors are extremely reliable, failures do occur. Power outages and surges must also be taken into account. When the motor stops, the same condition exists as at pump shutdown. That is to say, column separation can occur and flow will reverse against the check valve. The worst possible situation is where power is only momentarily interrupted and then, just as the flow column is returning to the pump, the pump restarts. This would be similar to a head-on collision of two automobiles traveling at high speed. Options to control these situations are few. Certainly an interrupt or time delay switch could be installed at the motor to prevent an immediate automatic restart. Controlled close and open check valves will not be an option because they probably would not be operable or would not react fast enough. The best choice here would be a surge suppressor. The compressibility of the gas in the suppressor will react instantly to the transient pressure spike. Quick-Closing Valves Moving down the water feed line, we now need to determine what will occur when the valves at the blending tanks are closed quickly. For purposes of this discussion, we will define a quick-closing valve as one that shuts in 1.5 seconds or less. It is important to remember, however, that as valves get larger, they can still cause problems even if they close more slowly. The Pump Handbook Series

But what actually occurs to create a potentially catastrophic spike? When one of the valves closes quickly, flow there is instantly stopped, but the column of water behind it will still be moving. Think of a speeding train with the engine abruptly hitting a stationary blockage. The engine stops, but the cars continue moving forward. The only difference between a train and water flow is that the train is not contained, and the cars derail. If the pressure spike does not rupture the pipe or some other component, the compression wave created will reverse and travel back down the pipe toward the pump at the speed of sound. When the wave hits the check valve or pump, it will again reverse and continue to reverberate until something finally breaks or the energy dissipates due to friction loss. Even if nothing fails initially, the system is under repeated stress, a situation that will ultimately lead to fatigue failure. What can be done to prevent this? The pressure spike can be either released or absorbed. The following options are among the most accepted solutions: 1. Rupture discs 2. Pressure relief valves 3. Slow closing, timed valves 4. Surge suppressors Rupture discs are simply plugs that will break at a lower pressure than any other component in the system. They are usually a ”last resort” protection and are not normally used to control pressure spikes because the liquid released into the environment is often hazardous, costly, or just dangerous in being released at such a high pressure. Pressure relief valves are valves that will open at a predetermined pressure. As the pressure spike builds, a relief valve will open, and fluid will be released. Either a holding tank or a return pipe system must be installed to exercise this option. Valve sizing and relief pressure settings are critical to minimize a spike as quickly as possible. This approach can be costly, and relief valve reliability must be tested regularly. Slow closing valves will solve the

problem if there truly is no need to close the valve quickly. In QPC’s application, however, it is necessary to stop flow instantly, so this solution is not an option. Surge suppressors provide a solution that will absorb the pressure spike by momentarily accumulating the flow of liquid as the valve closes. Because of the speed of propagation of the transient wave created, the surge suppressor must be installed directly upstream from the quick-closing valve and in no situation further away than 10 pipe diameters. Since the full capacity of the suppressor must be available to accept the accumulation of liquid when the valve closes, the suppressor must be precharged to 95% to 98% of system pressure. It must also be properly sized so it can provide the proper compressed gas cushion and momentarily accumulate a predetermined amount of liquid. A surge suppressor installed at each quick-closing valve for this application will provide an economical and reliable protection. In addition to the water supply system, we observe a bank of metering pumps that are used to inject precise amounts of fungicide into the paint blenders (Photo 2). This is accomplished with small single-diaphragm metering pumps dispensing the chemical through a precise mass flow meter. Because of the reciprocating nature of the metering pump, we know the discharge flow will pulsate. In many applications this is not a problem, but because the flow meter cannot accurately measure a pulsating flow, a pulsation dampener must be installed at the pump’s discharge. If properly sized and charged, the dampener maintains the mean system pressure within 1%, which is sufficient for the meter to function properly.

Stop 3: Packaging the Finished Product From the blending tanks, AODD pumps transfer the paint to holding tanks before it goes to filling equipment. The AODD pumps are all fitted with discharge pulsation dampeners to prevent rubber hoses,

which connect the pumps to the tanks, from ”jumping” around the plant floor and endangering employees. Hose jumping results from pulsations created by the start and stop action of the pump. In addition, the jumping will wear holes in the reinforced hose wall as it rubs on the plant floor. Since it installed dampeners QPC has been able to reduce hose replacement greatly. As we pass the filling equipment for buckets and cans of paint, we observe several more pulsation dampeners, again installed in conjunction with AODD pumps (Photo 3 and 4). The filling machines cannot be accurate with pulsating flows. The manufacturer of the filling equipment installs dampeners on its equipment, so that whether buckets and cans are filled by flow meters or by bulk weight measure, the flow is laminar and measurable by the instrumentation.

The Future The last stop on our plant tour is the research and development area. QPC is committed to ongoing product development, and today technicians are testing a new proprietary paint for spray applications. They are using a piston type spray pump, which produces a reciprocating flow. Their goal is to produce a paint that will disperse well through spray nozzles for coating purposes. They will need to use a pulsation dampener with this pump for two reasons. First, system vibration at the pump’s high cycle rate can lead to system fatigue and eventual failure. Second, without the dampener, the spray pattern produced will be wavy and inconsistent as the spray rises and falls in synchronization with the pressure pulse from the pump. The inconsistent pattern will require over-spray to coat the test board. This results in paint waste and a non-uniform coverage. Properly sized and charged, a pulsation dampener can provide the required level of dampening to allow for a smooth and continuous flow from the nozzles. Although our tour is finished, there are many other pump systems The Pump Handbook Series

in a typical plant that would benefit from pulsation and surge control. Each should be analyzed with regard to potential hazards that may result from a change in velocity of the flowing fluid or from the pulsations created by a reciprocating pump. Examples include filter press systems, where the pulsating flow from an AODD pump can damage filter media or ”cake,” and in-line mixers, which benefit from a steady stream of injected fluid as opposed to slugs of liquids. Our tour was not designed to be a detailed analysis of every system, but rather to highlight the kinds of damage that can occur in any system that uses pumps and valves. No manufacturing plant need subject itself to the potential hazards of pulsation and surge. I have made no attempt to explain sizing, materials or other parameters involved in using the preventive devices reviewed. In some cases, complicated mathematical formulas are required to make such determinations. There are several good manufacturers of the various devices discussed. You should consult with these manufacturers or their distributors to determine the proper product(s) for your specific application. ■ Gary Cornell, president of Blacoh Fluid Control (Riverside, CA), has worked in the reciprocating pump industry for more than 20 years. He has a B.S. degree from California Polytechnic University, and has had several articles published on the subject of pulsation and surge control.

References Karassik, Igor, William Krutzsch and Warren Fraser. Pump Handbook. McGraw Hill (1976) Wachel, J.C. and S.M. Price. Understanding How Pulsation Dampeners Work. The American Society of Mechanical Engineers Pipeline Engineering Symposium (1988) Young, Winston. The Young Engineering Technical Manual. Young Engineering, Monrovia, CA

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Avoiding Gremlins and Alligators at Pump Start-up Is a checklist all you need? Or should you buy alligator repellent, too? by Bob Matthews, GCI-Texas any pumping industry start-up stories are tales of horror. Here are just a few of the most common problems and some solutions that can overcome them.

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Oil Leaks Oil leaks are a typical start-up nuisance that simply should not happen. Most oil leaks occur at the gaskets, lip seals, pipe plugs, bearing isolators, oil bowls, breathers, oil lines and O-rings. Most of these leaks can be avoided by using quality parts and superior installation techniques. The invisible oil leak is a tough problem to detect and solve. One reason this type of leak occurs is that an oil bowl or sight glass has been installed on the wrong side of the bearing housing. The rotation of the bearings and/or the oil slinger causes the oil to be higher on one side than the other. This gives a false reading, and the lubricator

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Effected Area

Cast Iron Housing

adds more oil than necessary. This excess oil in the bearing housing has two effects. The bearings will run hotter than normal, and there is the possibility of an external leak. Oil leaks caused by vacuum in the coupling guard can be a real bother and a headache to troubleshoot. These leaks occur when the coupling guard is mounted so close to the bearing housing that the centrifugal force of the coupling pulls air like a house fan. Oil leaks in totally enclosed bearing housings usually result from an air pressure imbalance between the oil reservoir and the housing. This normally occurs with the use of contact seals, such as magnetic or the Inpro VBX positive vapor seals. Magnetic bearing seals are obviously not vents for the housing, but neither are Inpro VBX bearing isolators. If an oil reservoir is used, it should always be a closed system like the ones available from either Trico or Oil Rite. An expansion chamber or a directional vent valve must be used on top of the pump to prevent moisture ingestion from

breathing. This arrangement will ensure that as the air inside the pump expands and contracts, the air pressure between the oil reservoir and the bearing housing stays in balance, thereby eliminating the leak. It is common to have oil leaks in horizontal pumps that, for space considerations, are stored in a vertical position. Contact seals prevent this type of leak. Always be sure to use good installation practices with any oil seal, taking care to avoid the sharp burrs that can make any good seal look bad. Start-up leaks on ANSI and other pumps occur at the gasket between the adapter frame and the bearing housing. These leaks can be from broken or folded gaskets or from trash holding the two faces open. However, this kind of leaking often results from metal crowning at the threaded area of the sealing surface. Crowning happens when the threaded bolt holes do not have enough of a bevel cut at the face, and the torque applied by the bolts forms a crown of metal around the

These raised areas can be indicated and removed with a mill file

Shaft

Figure 1a. Over-torque or tapped holes without chamfer will pull a crown out of metal on machined surfaces. These raised areas will cause many problems.

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Figure 1b. The raised areas can be easily identified and removed with a mill file. The Pump Handbook Series

hole (Figure 1a). This problem is easily resolved by adding a note to the rebuild sheet to check for this. Filing with a large flat mill bastard file or a carbon stone on the tapped sealing face will reveal the crown instantly, and it can then be removed (Figure 1b). RTV silicone can be applied when raised areas exist, but this often puts the two mating faces out of square, initiating a chain reaction of alignment problems. I’ve mentioned only a few problems here. Be sure to use your imagination and good troubleshooting techniques.

Leaks and Gaskets I have seen start-up leaks on split case pumps stopped by the use of thicker gaskets at the split lines. However, this causes the stuffing box to become elliptical and different problems to occur (Figure 2). Technicians should begin by checking the manufacturer’s maintenance manual for a specification on proper gasket thickness. You can also determine the proper thickness in the field by assembling the two housing halves using only half of the nuts. Once this has been done, you can mic the stuffing box at and perpendicular to the split line. Subtract the height from the width to determine the exact gasket thickness required.

Product Leaks Product leaks are a giant headache in centrifugal pumps. Many such leaks occur at the gaskets, flush lines, pipe plugs, O-rings, packing or mechanical seals. The majority can be resolved by using good quality parts, superior installation skills and technical adjustments during start-up.

Leaks Caused by Cracks Product leaks in cast iron pump cases and oil leaks in cast iron housings result when cracks develop due to over-torquing to compensate for gasket leaks. The most overlooked and easily detectable culprit is the impact wrench being set on the highest torque setting with antiseize on the threads. To top it all off, the torque wrench is applied to the bolts after the impact wrench. Amazingly enough, the torque wrench clicks! Yes, the wrench may click to indicate it applied the

wrench setting, but more torque actually exists because the impact has surely applied too much torque. Important tip: If the torque wrench does not move the fastener before it clicks, the torque is set too high. Even if over-torqued bolts don’t cause cracks that leak product at start-up, vibration can finish the work started by the impact wrench, resulting in leaks. There is a great difference between the torque wrench settings for dry threads and those painted with anti-seize products. Be sure to choose the proper torque for your conditions. Failure to do so will likely result in your being bitten by any or all of these start-up and extended-run bugs. If you are not sure, you can identify the torque applied by the impact wrench by loosening the bolt with a needle type torque wrench. Next, make a visual inspection of the bolt heads. To troubleshoot a historic problem of this type, look for the mushroom edge marks on one side of the flats made by the impact wrench and for depressions in the face of the casting under the bolt heads.

Lubricant Contaminated lubricant can also cause pumps to fail at start-up. Contamination can come in the form of water, steam, product, sweat, dust or dirt. Totally enclosing your bearing housings is a good first step to prevention. For example, in the chemical and refining process industry, most pumps are spared. One pump runs the majority of the time, and another runs hardly at all. The spare pump with a long idle time can have start-up problems due to oxidation in the bearings. This is quite common in humid areas like south Texas, where I live. As the pump breathes, condensation forms on the walls of the bearing housing. The result is rust. After totally enclosing the housing, the next step should be a switch to a synthetic oil. Finally, an even better idea would be to install an oil mist lubrication system for the bearings. Volumes of documentation show that the cost saving benefits of this idea far outweigh oil mist’s cost of installation. Some reasons are: •Less oil is used. •Air flows across the bearings, which helps to cool them. The Pump Handbook Series

Figure 2. The diameters of A and B must be equal. Gaskets which are too thick can cause problems. If running packing, the result is a leak. The quick indicator of this situation is that the gland will lock. If running mechanical seals, the bolt circle will be out of line and bind the seal’s internal parts.

•A small but constant supply of clean oil is applied to the bearings. •A slight positive pressure is held within the housing to stop the destructive breathing process. Oil misting will all but eliminate the start-up bearing failures described here. And if you have totally enclosed the housings, there won’t be any housekeeping problems due to stray mist.

The Break-in Period Lack of technical maintenance during the break-in period accounts for most of packed equipment’s start-up problems. At start-up and for several hours afterward, packed stuffing boxes must have the attention of qualified personnel. If packing is not adjusted carefully, burnout can occur, causing production delays. Proper selection and installation techniques for packing are very important. Naturally, if you

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use packing material that does not suit the application or if you install it incorrectly, start-up problems are inevitable. It sounds simple, but millions of dollars are wasted each year because packing is not selected and adjusted properly. I have found that the Anti-Keystone design packing is more consistent and easier to control than others during start-ups, resulting in longer run times.

Mechanical Seals Mechanical seals require start-up attention but often go unchecked. Inspection at start-up is imperative to ensuring the necessary flush rates. If you use a barrier fluid tank, someone must check for proper operation and effectiveness as the pump starts to run. Whether a seal flush or a barrier tank is used, qualified personnel must check the piping against the specified requirements to see that it is correct. Do not let anyone tell you it is all right to put barrier fluid in a double or tandem seal other than from bottom to top. Clear-housed seal chambers show that the trapped air does not leave the area during rotation but is pulled to the rotor, thereby starving the seal faces of lubrication and cooling. If there is a lack of available qualified maintenance personnel, control the start-up electronically.

Dry Running Once every so often you may hear of a pump started with no medium, or the suction is dry. A single seal will burn up, and this is a guaranteed product leak. Sometimes the barrier fluid is left out of the tank and the outboard seal of a double seal will fry. This turns the double seal into a single seal and damages it beyond repair. A packed pump, which depends on product for cooling and lubrication, burns up without liquid in the pump at start-up. This glazes the inner packing surface, and a leak will occur as operating pressure is reached. Level controls and other signal devices will prevent many of these problems.

Dry Products Dry products in pumps, products with solids, and products that solidify when left to dry out in the volute can create several start-up prob-

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lems: • The motor will pull high amps. • The shaft can break or bend. • Product will dry on the impeller, causing imbalance and vibration. • Seal faces become glued together, destroying them when horsepower is applied. Pumps that operate in such conditions require the development of a flush procedure to eliminate these failures. To say the least, it is a good idea to lock out the equipment and check for free rotation before start-up.

Vibration Here are just a few of the reasons why vibration will kick your butt at start-up. Balance specifications recommended by some manufacturers and some engineers are not tight enough. The better you balance a piece of equipment, the better it runs. The ultimate compliment a technician can get is for an operator or supervisor to say that the equipment is not running when in fact it is! Again, I cannot adequately express how important it is to have a check-off sheet for repairs. Do everything possible to make rotating assemblies run smoothly. Put that coupling hub on the shaft when balancing and take your tolerances past “just good enough” on the balance machine. Also, don’t overlook what I might term “flow imbalances.” When the mass of a rotor is larger on one side than on the other, even if it is within dynamic balance specifications, the push creates a medium force imbalance. Another example is where the impeller bore is off center and the vanes on one side are longer than on the other. This imbalance transmits right to the bearings as a radial load and shortens pump life.

Lockup A pump can lock up at start-up for many reasons. Lockups happen when bearings fail due to lack of lubrication. Leaks from neglected and loose oil drain plugs can cause this problem. The preservative lubricant used for pump storage must be compatible with the oil used during the pumping operation. Have your oil supplier’s engineer check the preservative to be sure it The Pump Handbook Series

will blend with the oil in the bearing housing. Thread compound or Teflon tape will seal threaded plugs and pipe fittings from most leaks. When using these thread sealers, do not use too much because excesses can get into ports under bearings and plug the housing’s oil returns. Too small clearances in case wear rings can also cause lockups. This happens when the rings are set wrong, when temperature changes cause expansion, when shaft runout is excessive or when materials are changed, such as going from carbon to steel to stainless. Understanding the process in which the pump is running is important. Technicians who lack this knowledge put themselves at a great disadvantage. Good technical information, skills and process knowledge are the keys to resolving these problems.

Unexplained Lockups Some bearing lockups at start-up are seemingly inexplicable. My coworker, Mike, told me of a bearing locking up with adequate oil in the housing. He immediately applied troubleshooting skills and found no cause. Mike changed from the first oil used to Royal Purple Synfilm. This worked, but who knows why? Sometimes you must go with your gut feeling. I greatly prefer to use engineered technology and experience to solve problems. When I asked for permission from Royal Purple to use their name in this article, they gave me an explanation for the phenomenon just described. “Synfilm” is the brand name used for their high film strength synthetic lubricant, which is derived from their proprietary “Synerlic” additive system. The bottom line is that Synerlic forms an ionic (electrostatic) bond with the metal surfaces, leaving a long lasting film on the bearing that is very good at startups. Reduction in the coefficient of friction when Mike added the Synfilm to the bearing housing enabled the machine to run.

Sleeve Bearings Sleeve bearings in vertical mediumlubricated pumps are a never-ending source of problems. Poor shaft support, dry starts, high friction and rapid

wear in abrasive services are some of the more common difficulties encountered. One solution is the addition of an enclosed tube around the shaft with either a fresh water flush or oil/grease lubrication. This textbook fix has a high initial cost and a continuing expense in the supply of fresh water or oil/grease. It can also contaminate the pumped medium. The solution I am familiar with is the installation of Thordon bearings. This company produces a range of medium-lubricated bearings offering low friction, dry start capabilities, good shaft support and excellent resistance to abrasion.

Running Backwards and Flowback Running backwards at start-up will spin some screw-on style impellers loose and jam them into the volute, locking up the pump. Some impellers can be installed backwards and still fit into the pump case and thus go unnoticed. I know this can happen because I was part of this mistake once. What a learning experience! Pumps started in reverse can have a variety of effects on a system, most of which are very costly and dangerous. When install-ing a pump, always verify that everything will rotate in the correct direction. Fluid back-flowing through a pump at start-up will often break the pump shaft, and it can cause

any number of other problems. Backflows through leaking check valves or bypasses with a large enough volume of medium will spin an im-peller backwards. Visual inspection of the pump before startup is necessary to insure that this problem does not occur.

Alligators These creatures have also been a problem at start-up. I know you must think I’m crazy, but this is a true story. A young farmer went to start his irrigation pump one day and noticed that a foundation support had been moved away from the driver engine. Unknown to him, a large alligator had moved into the neighborhood to feed on area rodents and had knocked out the support block. After repair and realignment, the engine was ready to start. The next morning as my farmer friend began to start his pump, he slipped on the morning dew, slid down to the suction pond, and came face to face with Mr. Gator. His escape was successful, and he will always remember that particular start-up. And I’ll bet you anything that this pump now has a start-up checklist!

Final Tips To summarize, I will suggest that you can improve start-ups in these ways: • Always have a hands-on mainte-

The Pump Handbook Series

nance crew present to tune the equipment until the operation is running smoothly. • Choose the best parts and lubricants to maintain long life when repairing pumps. • Go that “extra step” with tolerances. I don’t have engineering data to prove that tighter specs are better, but seeing longer life is proof enough for me. • Use good pump check sheets. This is necessary, not optional. • Make the rounds. Look the pump over before start-up. Confidence is important, and these prechecks save money. (Is the flush water on? Is the shaft going to rotate in the right direction? Is the oil level all right? Is the mist turned on? Is the pump case dry? Answer the applicable questions for your equipment.) • Find out if application changes have been made and if so what effect they might have. • Look out for that alligator! ■ Bob Matthews is a Rotating Equipment Consultant for GCI-Texas in Pasadena. He has more than 30 years of experience in mechanical applications, 10 of which were handson pump repairs. He is a graduate of Lamar University. Bob is a frequent contributor to Pumps and Systems. If you have questions about the article, call him at (713) 473-7802 or send a fax to (713) 473-4068.

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Experience with Replacement of Boiler Feed Pumps for Reliability Enhancement By Merwin W. Jones, Potomac Electric Power Company tility boiler feed pumps are among the most critical components in a power plant. They are used to provide feedwater to the boiler where it is heated to produce steam. The reliability of these large pumps affects the reliability of the entire plant because output must be reduced or stopped when the pumps are not able to supply water to the boiler. The following case analyzes how Potomac Electric Power Company (PEPCO) increased its efficiency and reliability through close monitoring of its boiler feed pump operations. Boiler feed pumps used in large power generating stations must be designed to operate reliably in the severe operating conditions of high heat, pressure and flow. A large station of 600 megawatts (MW) would typically have two parallel pumps, each rated at about 15,000 hp, pumping feed water at about 500°F. The discharge pressure of such pumps in a supercritical cycle is around 5,000 psi, with a flow of about 5,500 gpm. Interestingly enough, some utilities of comparable operating capacity are not equipped with two parallel pumps; rather they utilize a larger single pump to provide their feedwater. Most of these pumps are variable speed, and they are typically driven by mechanical drive steam turbines or electric motors through hydraulic couplings. Recent technology has introduced variable-speed electric motors, which have become an alternative for driving these pumps. PEPCO’s two Chalk Point Station

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Figure 1. Typical HDB pump

units, which were completed in 1964 and 1965, had access to limited supercritical boiler feed pump experience, as did its fellow utilities. Chalk Point’s primary pumps on Units 1 and 2 were an early generation supercritical design. These units, like most early models of any product, had design limitations that directly affected their in-service performance. Each unit generates 350 megawatts and is equipped with two parallel boiler feed pumps. The pump operating conditions are about 2,800 gpm with a discharge pressure of 4,800 psi. Operating temperature is about 360°F. These pumps, although efficient, experienced complications early, with three of the four pumps failing within one month of initial start-up. Over the years these pumps experienced low MTBF and high MTBR. Simultaneously, other utilities were experiencing a similar trend with these early model pumps. In 1979 a major boiler modification project was proposed for this station. The project would introduce additional pressure drop in the boilThe Pump Handbook Series

er, thereby requiring additional head to be produced by the boiler feed pumps. A study was done to reexamine the capabilities of the existing boiler feed pumps. The research indicated that performance of the existing pump was marginal in providing the required flow at a higher discharge pressure. Faced with these findings and forecasting future demand for generation, PEPCO took an uncommon initiative in the utility industry and proposed the replacement of existing pumps with new pumps, primarily to improve reliability. Before purchasing the new pumps, a task force was convened to determine the root causes of the pump failures and the corresponding requirements or system changes needed to assure optimal efficiency. The managers at Chalk Point used a research project prepared by the Electric Power Research Institute (EPRI) for this purpose. Based on a survey of a number of utilities, EPRI issued a report that correlated design features with reliability problems. The institute identified a number of potential problems, and these were compared to the Chalk Point pumps. Specific problem areas identified by EPRI that were common to the Chalk Point pumps included the following. • Flat rise to shutoff of the pump head/capacity curve • High specific and suction specific speed • Marginal NPSH available/NPSH required • Operation off Best Efficiency

Point (BEP) • Inadequate minimum flow The utilities surveyed for the study, including Chalk Point, also experienced difficulties with shaft seals, balancing mechanisms and interstage takeoff seals. The pumps were also suspected of having a problem with inlet eye flow recirculation due to modified impellers installed several years earlier. Most of these failures were related to the manufacturer’s design, particularly, close clearances used to enhance the pump’s performance. At the conclusion of the investigation, the task force recommended replacement of the old boiler feed pumps. PEPCO prepared a specification for new pumps that addressed all the issues raised during the task force investigation. Although plant design created limits such as maximum shutoff pressure or suction pressure available to the boiler feed pumps, each deficiency in the situation was examined carefully and was specified as accurately as possible. For example, the feedwater system was given a complete test at minimum through maximum flow to characterize the range of operating conditions. Small adjustments were made in the test results to account for proposed boiler modifications. Specific issues addressed in the specification and a discussion of the findings and approach used to eliminate them as a source of unreliability are given below. Constantly Rising Head Capacity Curve from Runout Condition to Shutoff Condition This is a standard requirement in pump specifications. However, the shape of the curve is also important. The existing pumps had a sharp rise from runout flow to about 75% flow, and a flatter rise at lower flows. As a result, it was found that a change in flow from about onehalf flow to shutoff resulted in an increase in head of only about 25 psi. This indicated potential instability at low loads. Fortunately, there was a slight margin in the pipe’s allowable pressure rating—a circumstance that enabled the pump

designer to increase the maximum allowable shutoff head slightly. This was not considered a major risk, because we were still within the range of stresses permitted by the Power Piping Code, and because these pumps are turbine driven and reduce head at lower flows by reducing the turbine speed. Specific Speed and Suction Specific Speed The existing pumps were found to have a high specific speed and a very high suction specific speed. In the EPRI surveys, pumps with high specific speeds were found to have a much greater failure rate. Although limits were not placed on these parameters in the specification, they were considered in the evaluation of new pumps. Suction specific speed was reduced from about 12,000 for the old pumps to 8,700 for the new pumps, largely attributable to a double suction first stage impeller. Specific speed was also reduced due to one additional stage in the new pumps. NPSH Available versus NPSH Required NPSHA was marginal for the application, at least at the highest flows, so a great deal of attention was focused on this area. The system test was used to establish an accurate NPSHA curve. Our investigation of new pumps showed that the criteria for establishing NPSH requirements varied from vendor to vendor. NPSH is often based on the amount of cavitation to reduce the discharge head by 1% or 3%. For this application the pump vendor was required to provide NPSH requirements based on a 0% head loss, and to perform tests at four different flows to confirm the NPSH curve proposed. To establish a 0% head loss NPSH level, the pumps were tested at the vendor’s test facility, and suction pressure was reduced until it affected the total head produced by the pump. Data points were established to show the NPSH at about 1% and 3% head loss and then extrapolated to 0% loss. To allow for system transients and upsets, the required NPSH was specified so as not to exceed 2/3 of The Pump Handbook Series

the available NPSH. Operation Off Best Efficiency Point Operation at less than the design Best Efficiency Point (BEP) is a common cause of premature failures of large feed pumps. The usual explanation for this condition is that during the design process margins are added for flow, piping pressure losses and other uncertainties in the application. The combined effect of these design margins results in a pump that is sometimes substantially oversized and therefore always operates below its design point. This concern was eliminated by using the results of the feedwater system test. The test established the exact system head resistance curve, so the replacement pumps could be designed to operate near the BEP. Inadequate Minimum Flow Although the existing recirculation system had adequate minimum flow to prevent flashing of the feedwater, the criteria in recent years had been changed to provide for a flow at which the first stage impeller operates in a stable flow regime. In the case of the selected pumps, the minimum flow increased from 550 gpm to 1100 gpm. This necessitated changes in the recirculation valves and piping. Balancing Mechanism The thrust loads on a multi-stage pump with a discharge pressure of about 4800 psi must be balanced by some force within the pump. Although small pumps generally use a thrust bearing to absorb these forces, a device of this kind is considered impractical for large boiler feed pumps. Most vendors have developed hydraulic mechanisms to balance the thrust. The existing pumps at Chalk Point were equipped with a balancing disk. This design relies on a small clearance to throttle the discharge flow into a balance chamber. Pressure changes open or close to maintain the location of the shaft with in a few thousandths of an inch. This mechanism was an important issue at Chalk Point. It was our opinion that the change in operation of this unit from full load

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during the day to about one-third load at night increased the rate of failure of this component. PEPCO has several pumps at other plants which utilize opposed impellers to balance the thrust. By aligning half of the impellers in one direction and the other impellers in the opposite direction, only minimal hydraulic balancing is required. PEPCO’s experience with these pumps in similar cycling duties was considered very good. The company also conducted many interviews with other utilities using similar size pumps. It was found that units which cycled in load had higher rates of failure with disk or balancing drum type thrust balancing than those in similar service with opposed impeller designs. Our evaluation considered this factor for new pumps.

at the vendor’s shop. The tests verified performance in terms of flow, pressure, NPSH and efficiency. The vendor was required to subject the replacement pumps to additional tests as well. These were proposed as a result of recurring failures to the old pumps, as well as in response to claims made by the pump vendor. For example, an interesting test was done to demonstrate the ability of the pumps’ shaft seals to withstand loss of injection water momentarily. Loss of injection water on a pump of this size usually results in severe damage in only a few seconds. During the test, the pump was set to operate with minimum NPSH, and the injection water was shut off for one minute. We required that the pump show no increase in vibration during this period, and the shaft seals were to be disassembled after the test and examined for any signs of rubbing

Testing The pumps were each given a full flow and pressure performance test

or wear. These pumps passed this test. A particularly notable event during the NPSH tests occurred at about 80% flow. The pump suction was throttled to reduce the pressure to 5.3 psig. In this condition the pump was passing 2,125 gpm at a discharge pressure of 4,200 psi. The pump remained quiet with no outward signs of cavitation.

NPSH “Meter” During the system tests it was found that hotter than normal water was sometimes entering the pumps. The operators generally understood that they should close one steam line to the feedwater heater ahead of the pump suction to control the water temperature, but the reliability of this approach was still a concern. To resolve this difficulty and to ensure that the pumps never experienced a flashing condition, an

BYRON JACKSON

90 80

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DATA BY MA DRAWN BY MA

DATE 21 MAY 81 VOL. HYDR. LAYOUT

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Figure 2. Results of the NPSH test performed on Chalk Point’s boiler feed pumps.

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MAX DIA.

280 TEST DIA.

PEAK EFF. PEAK EFF.

320 UNDER FILED COVER 642082 642082 642082

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360 STGS. 1ST 3 3 T- 39005-3

“NPSH meter” was installed on these units. The NPSH “meter” was basically a computer algorithm, installed in a plant’s computer, that monitored the suction conditions, that is pressure, temperature and flow. Based on temperature, the computer calculated the vapor pressure of the water. The net NPSH available was calculated by subtracting the vapor pressure from the suction pressure. The required NPSH was loaded into the computer from the pump manufacturer’s curve (Figure 2). When the unit was in service, the computer continuously monitored the flow to determine the required NPSH. Thus, the NPSHA was continuously calculated based on temperature and pressure. The computer subtracted the required NPSH from the available, and if there was insufficient margin, an alarm signaled the operator. The operator could then take whatever action was necessary to reduce the temperature or increase the suction pressure to solve the problem. This sys-

tem was tested after the pumps were installed, and it performed perfectly.

Long Term Results The four new pumps were installed in 1981 and 1982. Since that time they have operated an average of more than 16 years. During this period they have had routine overhauls at intervals ranging from two to four years. Two spare volutes and rotating elements are stocked at the plant. These are generally rotated into the pump to be overhauled, and the old volute assembly is rebuilt. The total forced outage time during those 16 years with four pumps has been 796 hours. The annual average forced outage rate is only 12.5 hours per pump per year. The historical average for the old pumps was 389 hours per pump per year. That is a reduction of an impressive 97% in the failure rate. This has enabled the units to generate an additional four million megawatthours of electricity during this 16-

The Pump Handbook Series

year period. These pumps have operated well since they were installed and continue to perform well.

Summary As the utility industry moves toward deregulation, the importance of having plants on-line to meet electric demand will become critical. Reliable equipment is essential to this goal. This project demonstrates that with proper attention to details and to forecasting equipment difficulties, a company can justifiably make a capital investment that will optimize future generation capacity. The role of the specifying engineer should be to provide the pump designer with the best information possible on operating conditions and allow the designer to meet them. This case is a prime example of that philosophy. ■ Merwin W. Jones, PE, is the Lead Engineer for the Mechanical and Civil Section of the Potomac Electric Power Company.

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Troubleshooting ANSI Pumps Using Limited Information by David Weidner, General Electric Silicones his article explains how to calculate where a pump is operating on its curve when you have little or no information, and how you can use that limited information to troubleshoot hydraulic problems with ANSI dimensioned centrifugal pumps. It also includes a checklist that you might find helpful when troubleshooting installations.

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The Typical Breakdown ANSI pumps are probably the most prevalent pumps in the chemical industry, and they certainly account for their share of headaches. How do these problems usually show themselves? Typically, maintenance personnel in the area get a call and a work order stating that a pump isn’t running correctly. So they disassemble the unit, check everything over, then re-install it. Occasionally, something is actually wrong: impellers are corroded or grossly misadjusted (especially if setting clearances off the casing), debris is clogging up part of a closed impeller or there is blockage in the suction piping. Sometimes nothing is found to be mechanically wrong, but the pump is still not operating where it should be (according to operations people). Now you must become a pump sleuth. If the pump is installed in typical fashion, you are lucky to have a discharge gauge that is readable to help you. Believe it or not, you can use this one simple piece of equipment as the first indicator in troubleshooting the problem and analyzing overall pump performance. Step 1: Get the Right Information Hopefully, when the pump was apart you were able to measure the impeller, or at the very least, you

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have a good record of the impeller size and a curve handy for the pump in question. If you don’t have either one of these, obtain the pump curve for your installation. The original vendor should be able to provide it, especially if you have a serial number. It is important to get the original curve for your particular pump because curves change over time. Lacking an impeller size, you can take an educated guess by deadheading the pump. This, of course, is not a practice you want to employ often or for a long period of time, since bad things can start to happen in the volute and seal chamber box areas.

age or, worse, do you bodily harm. At this point, knowing the liquid’s specific gravity (SG) and any suction conditions measurable will help greatly. Actually such information is necessary. Once you have the discharge pressure, convert that to a liquid head using the following formula (in feet from pressure):

Step 2: Visit the Site Once you have the pump curve and know the impeller size, take a field trip to the installation with area operations and maintenance personnel. You may want to also use this valuable time to query the folks at the site as to what has changed in the installation. You might ask, for example, why the pump was running fine for 10 years, but now it won’t pump? You can gather useful information in these walk-throughs that may otherwise not come up in a more formal setting. Someone might say, “Oh yeah, we used to recycle at 15 gpm, but now we don’t at all,” or, “Well, we couldn’t find that size impeller so we put a different one in to try to help.”

Step 4: Determine Impeller Size Once you have calculated the head, find it on the Y axis at the zero point of the X axis (indicating zero flow). You should be able to determine the impeller size based on this. Remember, impellers are typically trimmed in 1⁄8” increments, so there is a little margin for error. Since the discharge is blocked in, there aren’t any friction or velocity losses. The head should be pretty accurate.

Step 3: Figure Deadheaded Pressure and Compute Head When you arrive at the pump, get the discharge gauge changed, if you can, and record the pressure when the pump is deadheaded (discharge valve fully closed). Again, don’t dally while doing this as the procedure could result in equipment damThe Pump Handbook Series

Head=(Pressure x 2.31)/SG Subtract any head in a flooded suction condition and add any for suction lift. (This gives you the total differential head because the pump is blocked in.)

Step 5: Figure Where Your Pump Is on the Curve You now have the correct curve, but you need to find where the pump is operating along that curve. With only a discharge gauge this is a little difficult: friction head losses, velocity head and the like make it hard for one gauge to be used for figuring out where the pump is operating. What else can we use? How about amperage draw? For a given flow and head there is a definite horsepower requirement (hp=(flow x head x SG)/(3960 x efficiency of the pump). Therefore, using amp draw, we can work back-

wards to get approximations. It’s time to get the area electrician to help you. With the pump running, use a clamp-on amp meter and record readings on all three legs (we’re assuming three phases) feeding the pump motor. Obtain a voltage reading by using a multimeter. Average all three readings and use the formula: hp = (1.732 x volts x amps x motor efficiency x motor power factor)/746. The motor efficiency and power factor can be found on the motor nameplate. Now you have a horsepower figure that you can use. Enter the curve and follow it until it intersects with the calculated horsepower. (Curves aren’t standard and will look a little unusual, but they will provide the same information.) This procedure will give you a good idea of where the pump is running (or not). For example, let’s look at the pump curve shown in Figure 1 and use the following information:

• • • •

Suction vessel at 5 psig Liquid SG = .9 Liquid height above impeller 6 feet Discharge pressure at 23 psig (blocked in) • Amp draw (3 phase motor)= 4,5.5,6 at flow conditions • Motor voltage is 460v • .85 efficiency and .91 power factor (from motor nameplate). What is impeller size and what flow and head are we running at based on power input? With the discharge pressure at 23 psig equaling approximately 59 feet (after subtracting suction head of 17.55 feet), our formula looks like this: (5 x 2.31+6’) = 41.45’ Finding this result on the Y axis at zero flow indicates an impeller size of 63⁄4” (shown as #1 on the curve). The amp draw averages 5.2; therefore, hp = (1.732 x 460 x 5.2 x .85 x .91)/746 = 4.27 To correct for SG (pump curves

are based on SG of 1), we use the following formula: 4.27/.9 = 4.74 hp Plotting the horsepower shows a flow and head of something around 330 gpm and 34 feet (shown as #2 on the curve), which would equal around 20 psig on the discharge pressure gauge if the piping is sized correctly. The flow and head are slightly to the right of BEP, which isn’t great, but it is close enough as to not cause any problems. This information is very useful in troubleshooting the installation if there is a problem. So, with no flow instrument and only a discharge pressure gauge, we can determine where on the curve the pump is running. Hopefully, it is not way to the right of the Best Efficiency Point (BEP) and not below minimum flow for that pump design. Both possibilities will cause problems from a maintenance standpoint, and should be corrected as

Figure 1. Point 1 shows the impeller size, point 2 the calculated capacity and head. The Pump Handbook Series

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TROUBLESHOOTING CHECKLIST Rule Number 1: Do the easy stuff first. • What has changed from a maintenance or operational standpoint? • When was the pump pulled last and what was done? • Is operations using a different fluid or trying to get a different flow rate? • What happens if you raise the suction pressure? • Is the pump rotating in the right direction as viewed frim driver? • Is the pump running on its curve? • Check amp draw. High amps indicate high flow and head; low amps are a sign of a suction/ impeller problem. • Has it been properly primed? (Self primers need liquid to work, too.)

Suction Conditions • Is the suction plugged? (There can be foot valve problems, build-up around inlet and build-up in suction piping.) • Is the eye or any part of the impeller plugged? • Are there air leaks on the inlet (problem for suction lift applications)? • Is the number of impeller vanes correct for the application? • Is there an inlet vortex? • Is there enough straight run of pipe (5 to 10 pipe diameters) into the pump? • Is the suction pipe sized correctly? • Are any eccentric reducers installed flat side up so as to not trap any gas pockets and cause problems? Pump Problems • Is the impeller adjusted correctly? • Are abrasion or erosion causing loss of efficiency? • Are the cutwaters open and clean? • Are the cutwaters broken? • Is there enough straight pipe (5 to 10 pipe diameters) off discharge flange? Note: This is not an exhaustive list, but when faced with a problem it’s a good place to start. soon as possible to prevent hydraulic instability. So now what? What good is all this stuff? Let me explain.

A Real Life Example #1 Operations people in one plant I know of couldn’t keep up with required flow, and running two pumps in parallel didn’t help. All we had to go on were readings from a discharge gauge and the knowledge that the pump had just been taken out of service, cleaned and reinstalled with correct impeller clearances. The flows we were trying get should have been easy for one pump to achieve. We had verified impeller dimensions (both diameter and the number of vanes) when the pump was out for maintenance asked what had changed (nothing had), and looked at the suction conditions (all were good). We blocked in the pump

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and found that it developed the correct deadhead pressure for the installation (according to the pump curve). Operationally, it looked as if it was running with an impeller that was a few sizes to small. Using the amperage/horsepower calculations, we were able to determine that the pump wasn’t running where it should have been on the curve. A flow meter indicated low flow, and the discharge pressure gauge was reading lower than expected. The hp calculations, however, showed a pump running at a much higher head than what was showing on the pressure gauge. (The flow meter indicated that it could only be head that was raising the hp input to the pump.) This told us that there was a huge pressure drop somewhere between the pump and the pressure gauge. Operators didn’t

The Pump Handbook Series

want to shut down, but once we showed them the calculations and evidence, they agreed. This problem pump had a double cutwater to help balance radial forces, and one of them had been cleaned, but the other was about 75% blocked (as seen through a boroscope). This blockage was removed, and the pump came up right to the curve after reinstallation.

Real Life Example #2 Another call I got was for a pump that kept losing prime. It just wouldn’t pump. The walk to the installation, and the conversation that took place with plant operations people along the way, revealed that the impeller had been changed to the largest size because the old one was corroded and operations didn’t want to wait for maintenance to trim and balance the impeller. Thinking the larger size would give better results anyway, they installed it. Amp readings indicated a high flow rate—discharge pressure was essentially the same, with most of the flow recycled back to the suction vessel. After running for a while, the pump would lose all capacity for pumping. When the level in the vessel was raised as high as possible, the problem would go away. The conclusion: The pump flow was way too high (based on hp draw), and the vessel outlet to the pump was developing a vortex that was drawing gas into the pump eye, thus eliminating prime. We knew there was no vortex breaker in the suction vessel, and operations could not accept the high liquid level (The higher level prevented the vortex from forming.) When the impeller was removed and trimmed to the appropriate size, the pump ran great, and the operations crew was off the hot seat.

Conclusion It is helpful to use amps in determining how a pump is doing when you are ”flying blind” in a troubleshooting situation. This method helps indicate whether it is actually on the curve and where it is, and if it might be hydraulically unstable, causing flow and maintenance problems. "

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PUMP AND SYSTEM TROUBLESHOOTING HANDBOOK

Why Doesn’t Your SelfPriming Pump Work? Installation practices of these workhorses differ from standard centrifugal pumps. By Steven J. Hrivnak, Eastman Chemical Company

elf-priming pumps are used all over the world in many applications. They are used to pump fluids out of sumps, tankers and buried tanks. Yet, with their widespread use, engineers continue to install them as if they were standard pumps, failing to take into account the manufacturer’s installation requirements. If a self priming pump is installed like a standard pump, it will not “self-prime.” Then operations people will call with such complaints as: “We must always pressurize the supply tank to get the pump to prime.”; “We must vent the pump each time.”; “We must add liquid to the priming chamber every time we start the pump.” Every time that I have been asked to troubleshoot, the designers forgot to read the manufacturer’s installation instructions. Sometimes, only simple changes were needed to make the pump self prime. To understand why self-priming pumps must be installed differently from standard centrifugal pumps, it helps to first understand how they work. A pump is said to be primed when its casing and all suction piping are full of liquid. If the suction source is above the pump, this can be accomplished typically by opening the suction and discharge valves. But if the pump is above the liquid source, then the suction line and the

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casing must be filled with liquid. This is done typically by using a vacuum source, such as a vacuum pump or an ejector. A foot valve and an elevated source of priming liquid also work well. However, valves can leak over time, causing the pump to lose prime. The beauty of self-priming pumps is that they do not need either an external vacuum source or a foot valve and filling source. Also, no people are required to operate these auxiliary systems. A self-priming pump will re-prime if it becomes air bound. Built into the casing of self-priming pumps is a chamber that remains full of liquid after the pump shuts down. (Note: This chamber must be filled the first time the pump is installed and every time it is removed for maintenance.) This liquid is used to prime the pump. When the pump is started, the agitation action of the spinning impeller re-entrains the air in the suction line and moves it to a discharge air separation chamber. The air and liquid separate, and the air must vent out the discharge line. The liquid flows to the bottom of the chamber, where a small slot allows it to flow back to the eye of the impeller. This movement of air causes a small vacuum in the suction line, allowing liquid to climb the suction pipe into the pump with the help of atmospheric The Pump Handbook Series

pressure on the suction source. Full flow and pressure is not attained until all air is evacuated from the suction line. After the pump stops, liquid in the discharge piping flows back into the chamber, where it is stored for use the next time the pump needs to be primed. If these pumps are installed incorrectly, they will not self-prime. Several rules must be followed in order to have a trouble free installation. Take a look at recent photographs of a self-priming pump I was

Photo 1. Rail car unloading pump suction pipe (side view).

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asked to examine (this month’s cover photo and Photo 1 in this article). Can you see why it will not prime? 1. Remember that maximum liquid lift is about 25 feet. This is based on atmospheric pressure, which varies with elevation and the pump’s efficiency. If the liquid you are pumping has a specific gravity greater than 1, then divide 25 feet by the specific gravity to determine the effective lift you can expect. If the liquid has a high vapor pressure, vaporization can occur in the suction line at the point of highest vacuum. The lift for these liquids will therefore be lower. Make sure to check for NPSH available to determine your maximum lift. This pump is right at 25 feet. But NPSHA indicates that it should be all right. 2. Minimize suction elbows. Suction elbows should be limited to 1 or 2. More elbows cause more friction loss, which may reduce NPSHA and cause vaporization in the suction line. In our case we have 8. That is a strike against us. 3. Locate your pump near its supply source and try to limit your suction

piping lengths to under 25 feet. The longer the pipe is, the longer it will take to prime the pump. The actual length of the suction piping is 50 feet. Strike two! 4. Be sure the pipe size is the same size as the pump suction pipe. If you have a 3x2-13 pump, then your suction pipe size should be 3”. The larger the diameter of the pipe, the more air that must be evacuated and the longer it takes to prime the pump. The pump is a 3x2-13; therefore, its suction pipe should be 3”. But it is 4”. Strike three! I guess that means we are out! 5. Avoid suction pipe foot valves on self-priming pumps. Before the invention of self-priming pumps, a foot valve held the fluid in the pump in an effort to avoid priming each time the pump was shut down. On a self-priming pump, of course, a foot valve is not needed and only adds friction loss and reduces NPSHA. The example installation does not have one. 6. Provide a small diameter (about 1/4”) air bleed line. This should be from the discharge pipe located between the pump’s discharge flange and

the discharge check valve or isolation valve. Run this bleed line back to the suction source. Air at 77°F has a density of 0.074 lb/ft3. This is about 840 times less than water density. Let’s say that with water a pump can generate 100 psig. When it pumps air, it will generate approximately 100 psig/840 = 0.12 psig. Your pump cannot generate enough pressure to overcome the check valve and vent the air. So always install a vent line - preferably one that remains open all the time. Yes, you will lose a little capacity through that line (in my example about 8 gpm), but the alternatives are to have an employee open the line every time you start the pump, or to install an automatic valve. In our case there is no bleed line. Even if everything else is correct, this pump will never prime. Strike four! Another strike against us! Other things to look for: • On some self-priming pumps, the casing and impeller are a matched set. The 1/8” clearance between the impeller and the cutwater must be maintained in order for the pump to self-prime. If the impeller needs to be trimmed, you should consult your manufacturer for guidance. • Follow minimum suction pipe submergence guidelines. If the suction pipe submergence is too small, air will enter the pipe, and the pump could lose flow while it re-primes. This can become a vicious cycle. (See Figure 1.) • When a self-priming pump shuts down, the discharge pipe must be designed such that enough liquid flows back into the pump to fill the self-priming chamber. Therefore, do not put a 90° elbow on the pump discharge. Always allow for plenty of pipe before the discharge check valve and isolation valve. (The volume of pipe should be equal to the volume of the casing.)

Figure 1. Minimium suction pipe submergence guideline.

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• Time to prime a pump is a function of pump size, impeller size and suction pipe length. Please refer to the manufacturer for this information. If the liquid in the priming chamber is hot, it can vaporize if priming takes a long time. This can make the pump lose prime.

Photo 2. Close up picture of new discharge air vent line.

• If the pump still won’t prime, check the suction pipe for leaks. Even a small leak can prevent priming.

So what should we change on our problem pump? The first thing I would do is add an air bleed. Other things are still wrong, but this is the least costly thing we can try, and the other installation errors may be part of the problem; no bleed is a definite problem. Without a bleed, the pump will never re-prime. Since the NPSHA calculation says our suction lift should be limited to 27 ft, we are all right. When we added the bleed line, the pump did prime. However,

The Pump Handbook Series

the user installed a valve in the bleed line. Since I have taken the picture (Photo 2), I asked them to lock it open. The rest of the installation is not up to par, but the pump is now working. Although there is some forgiveness with some of the rules stated above, following them will ensure a trouble free installation that always self-primes. " Steven J. Hrivnak is a Principal Mechanical Engineer in the Chemicals Technology - Acid Development Department of Eastman Chemical Company. He is a member of Pumps and Systems’ User Advisory Team and a frequent contributor to the magazine.

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PUMP AND SYSTEM TROUBLESHOOTING HANDBOOK

In-Plant Perspective: Eastman Chemical Company Go inside for the scoop on how the reliability team at this plant increased MTBR from 6 to 57 months. By Steven J. Hrivnak, Eastman Chemical Company

n 1998, Eastman’s Kingsport, Tennessee facility had more than 9,000 work requests on centrifugal pumps; an average repair cost about $1,700. In fact, pumps have the highest maintenance cost of all rotating equipment at the chemical facility. The Kingsport site has more than 30,000 pumps; 8,000 of those are ANSI end suction pumps. So why do some of these pumps have an average 6-month life? Installation and operation! Over the years we have learned that proper installation and operation reduce work requests. Even though the pump population has increased, the work requests are down 6% overall since 1996. Eastman’s best practice for increasing mean time between repairs (MTBR) has been paying attention to installation and operation details. A pump is part of a system. That system contains a foundation, a baseplate, grout, piping, a pump and motor, alignment and controls. Ignoring any of the details of these items usually causes short pump life, which manifests itself in short seal life. Properly installed and operated ANSI pumps at Eastman’s facility now boast 57 months MTBR. This is how it was done.

I

Standards The first step in improving pump reliability is creating good standards.

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We developed standards for baseplate design, grouting, piping and equipment alignment. The main areas of the standards are covered below. Baseplates Like a house that you want to build to last a lifetime, you must provide a support system for your pumps. Fabricate or purchase a welldesigned pump baseplate. Good designs will install with ease, and a machined baseplate with alignment jack screws will make alignments easy. The following are the characteristics of a good pump baseplate: • The baseplate shall be fitted with one 4” grout fill hole uniformly distributed for every 10 square feet of baseplate surface and/or per subdivided section or raised welded cavity. • Vent holes 1⁄2” in diameter shall be provided for each bulkhead compartment at all corners, high points and perimeter edges of the bulkhead. Perimeter vent holes in the baseplate shall be on 18” centers maximum spacing. Any angle or “C” channel added for stiffness will require vent holes on both sides. • Vertical jacking screws 1⁄2” diameter minimum shall be provided around the baseplate perimeter 3” from each anchor bolt location to facilitate alignment of the baseplate in the vertical direction. The Pump Handbook Series

• Machined mounting surfaces for the equipment and driver shall have horizontal jack screw positioning bolts 3⁄8” diameter minimum. • All welding on the baseplate shall be completed prior to machining the equipment and driver mounting surfaces. • Machined mounting surface shall be coplanar to .002”. • Machined mounting surfaces shall extend 2” beyond pump and driver feet on all sides with a 125 micro inch Ra finish. When machining a bent plate as a mounting surface, maximum material removal is not to exceed 1⁄16”. • Provide 1⁄8” minimum shim adjustment under driver feet for alignment. • If a spacer coupling is to be used to couple the equipment to its driver, add an additional 1⁄4” axial clearance to the spacer length in the baseplate design. • Anchor bolt holes shall be 1⁄4” larger on the diameter than the anchor bolts. • Tapped equipment bolt holes in the baseplate should be of sufficient depth to allow for plenty of thread engagement. On 1⁄2” thick or smaller baseplates, weld 1⁄2” or 3⁄4” thick square plates under or on top of the baseplate to increase the thickness before drilling, tapping and machining. Tapped

Photo 1. Re-installed 450-hp diesel-driven fire water pump

•

•

•

•

depth shall be 11⁄2 times the fastener diameter. All bulkhead cross-bracing on the underside of the baseplate shall have a 2” x 6” opening to allow for grout flow from bulkhead to bulkhead. All corners of baseplate flanges shall have a 1” minimum radius. All surfaces that will be in contact with the grout shall be rounded to eliminate stress risers. Angle, channel or stud anchors shall be welded to the bottom surface to act as a shear key in the grout. Gussets shall be welded inside the driver superstructure to increase stiffness and vibration damping.

Most of the above baseplate design points can be found in Process Industry Practice (PIP) RESP002.

Grout Grout your baseplates. Baseplates need to be grouted to minimize the vibration that decreases seal life. Certain fundamental rules are being overlooked when it comes to foundations of freestanding pumps. The foundation mass ratio—the weight of the foundation in comparison to the weight of the machine, driver, base-

plate and the liquid or material in the machine—must be of a certain value to effectively prevent vibration. For rotary equipment such as pumps, this ratio should be 3 to 1. For reciprocating equipment, it should be 5 to 1. A freestanding pump cannot meet this requirement. You need to start grouting your problem pumps. Note that if you grout a pump to a slab, the slab mass acts as part of the 3 to 1 ratio. Usually, 6” of grout attached to the slab is adequate. Slab or Foundation Preparation and Protection • After concrete has been properly cured and obtains its design strength (typically 28 days) scarify the existing slab or foundation surface that will come in contact with the grout using a small chipping hammer to expose aggregate (1” minimum depth). High-pressure air clean the chipped area to remove all dust and grit. • Slab or foundation protection — The cleaned, scarified surface of the foundation shall be protected from water, oil, dust and contamination. The preferred method is to place the grout directly against the clean foundation. (For epoxy grouts, if the time interval between The Pump Handbook Series

cleaning and grout placement is sufficient for the contamination of the prepared area by water and oil, the scarified area can be coated with the grout manufacturer’s recommended primer approximately 3 to 5 mils thick. Under no circumstances should generic epoxy paint be used.) If the prepared area becomes contaminated, clean before grout placement. • Keep the prepared area for epoxy or corrosion-resistant grouts dry. Epoxy grout will not bond properly to damp or uncured concrete. • For cementitious grouts, the prepared area should be wetted prior to grout placement per the manufacturer’s recommendation. • The prepared area shall be kept within the manufacturer’s temperature limits. The area is to be shaded from direct summer sunlight 24 hours before pour and 48 hours after. Anchor Bolt Sleeves • Clean out any foreign material in the anchor bolt sleeve, if used; fill it with a non-bonding urethane foam.

Grout Thickness and Flow Lengths • In the absence of information from the grout manufacturer, minimum thickness of grout shall be 1” with a maximum grout flow length of 1 foot. For each additional foot of flow length, the foundation to baseplate clearance should be increased by 1⁄2”. The maximum allowable grout flow length shall be 5 feet. • The maximum grout thickness shall be per the manufacturer’s written instructions and recommendations. Reinforcing Steel • In cementitious and polymer grout pours 4” and deeper, concrete reinforcing steel shall be provided in accordance with ACI 318-89 Section 7.12 requirements for temperature and shrinkage.

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torque may break them. If grout adheres to the leveling jack screws, they cannot be removed after the grout cures. Grout is designed to support equipment weight and absorb vibration; leveling screws are not.)

JACKSCREW

ANCHOR BOLT

MOUNTING PLATE GROUT 2” ~ ROUND x 3⁄8” THICK SANDBLAST EDGES AND BOTTOM OF MOUNTING PLATE TO OBTAIN A 3 MIL. PROFILE

(REBAR NOT SHOWN FOR CLARITY)

WRAP AS MUCH EXPOSED THREAD AS POSSIBLE WITH DUXSEAL TO PREVENT BONDING OF THE GROUT CONCRETE BASE

Wood Forms • Coat the inside of the wooden form’s vertical face with one coat of form oil for cementitious grouts or three coats of paste wax for epoxy grout (to prevent sticking). Forms should be waxed before installation to prevent accidental application of wax to surfaces where the grout is to bond. • Seal all form joints with silicone rubber sealant caulk to prevent grout leakage.

ENSURE ANCHOR BOLT SLEEVE, WHEN SPECIFIED, HAS BEEN FILLED WITH NON-BONDING URETHANE FOAM BEFORE POURING GROUT

FORM

SEAL JOINTS WITH RTV

⁄8” ~ BLEED HOLE AT 8” O/C

3

1” CHAMFER STRIP FORM SEAL JOINTS WITH RTV

TYPICAL GROUTING FORM

Figure 1. Typical grout detail

Reinforcing steel shall conform to ASTM A615, Grade 60 and shall be fabricated and placed in accordance with ACI 301 and Specification 03001, unless otherwise specified on the construction drawings. Consult the manufacturer for reinforcing epoxy grouts.

Hold the baseplate in position against the jack screws with the anchor bolts. • Anchor bolt and jack screw preparation — Prior to pouring the grout, wrap the baseplate leveling jack screws and anchor bolts with putty to prevent the grout from adhering. Coat the equipment fastener and coupling guard bolts with anti-seize and install them all the way into the baseplate to keep grout from clogging the threaded bolt holes and adhering to the bolts. (If grout adheres to the anchor bolts, applying proper

Installation of Baseplate (Figure 1) • Place 2” diameter by 3⁄8” thick stainless steel pieces with sharp corners removed on the scarified concrete under the jack screw locations. Level the baseplate using the jack screws against the 3⁄8” thick pieces. The preferred method is to level the baseplate without the TYPE OF EQUIPMENT LEVELING equipment installed. If the ANSI pumps 3600 rpm and less, positive equipment must be mounted displacement and reciprocating pumps with while leveling, pre-grout the horsepowers below 100. 3⁄8” thick pieces to stop moveAPI pumps ment under the increased load. Level the machined area Turbomachinery, and reciprocating of the baseplate to within manequipment above 100 horsepower. ufacturer’s specifications. If specs aren’t available, use the parameters shown in Table 1. Table 1. Baseplate leveling recommendations

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The Pump Handbook Series

Grout Placement • Place and mix the grout per manufacturer’s instructions. An 18” to 24” long by 4” head pipe and funnel is recommended to facilitate grout placement. Put a flange on one end and physically hold the flange down over the grout hole, or build a wood bracket to hold it in place. Pour the grout through the funnel into the head pipe. (The head pressure will help the grout flow and push the air out from under the baseplate.) Before removing the head pipe, plug all grout vent holes with rubber or wood stoppers. After you remove the head pipe, plug the fill holes with plywood. This holds the grout in place and prevents air pockets beneath the plate.

REQUIREMENTS OF FOOT PADS Use a standard bubble level. A maximum elevation variation across the length of the baseplate of 0.010” is expected. .002” per foot of length of the baseplate or less. .0005” per foot of length. Use machinery analysis in plant engineering to assist leveling the baseplate. The mounting surfaces should be coplanar within 0.002”.

Bolt Diameter, in.

Bolt Area, in.

Torque, ft-lbs

0.625 0.750 0.875 1.000 1.125 1.250 1.375 1.500

0.226 0.334 0.462 0.606 0.763 0.969 1.160 1.410

60 100 180 270 390 550 720 950

tion line shall be constructed to provide a spool section for strainer removal. • Discharge valves shall be located directly after the required straight run. Suction valves shall be located directly ahead of the straight run. The suction valve port area shall be equal to or greater than the suction line area. Suction piping shall be sized for 5 ft/s (bulk velocity) or less. Hot condensate service and service for other liquids within 10°F of the boiling point at suction pressure shall be sized for 2 ft/s or less. • Reducers in the horizontal suction lines shall be eccentric and located directly after the suction valve. If the liquid is within 20°F of boiling, place the eccentric reducer at the suction flange of the pump.

Table 2. Recommended anchor bolt torques

• Vibrators — Do not use a vibrator on polymer concrete or epoxy grouts; it will separate the aggregate from the resin. When using cementitious grout, use a high frequency vibrator to remove entrapped air bubbles. Curing • Let the grout cure in accordance with the manufacturer’s specifications. Typically, cementitious grouts are moist cured and epoxy or corrosion resistant grouts are air cured. Leveling Jack Screw Removal • Remove jack screws after the grout cures. Degrease the jack screw holes and, in non-corrosive areas, fill them with silicone caulk. In corrosive areas, fill the holes with a mix of silica powder and the resin used in the grout. If jack screws cannot be fully removed, retract them 1⁄2” and saw them off flush with baseplate. (This makes the baseplate rest on the grout, not the screws.) Anchor Bolt Tightening • After the grout has cured, tighten the anchor bolts to the applicable torque values (Table 2). For published grouting guidelines, see the PIP Specification REIE686, chapter 5.

pull piping in for final bolting. The following are some typical rules. • Discharge and suction lines shall be straight in the runs adjacent to the pump flanges for a minimum of 5 pipe diameters or 18”, whichever is greater. This minimizes the possibility of cavitation and smoothes out the flow entering the impeller. The pump suc-

PIPE HANGER OR SPRING SUPPORT, DIRECTLY ABOVE THE PUMP VENT VALVE, IF REQUIRED

SPOOL PIECE FIELDFIT, 5 PIPE DIA. STRAIGHT RUN OR 18”, WHICHEVER IS GREATER

SPOOL PIECE, FIELD - FIT, 5 PIPE DIA. STRAIGHT RUN OR 18”, WHICHEVER IS GREATER DISCHARGE PRESSURE GAUGE, 3 PIPE DIA. FROM DISCHARGE VALVE

SUCTION VALVE, 1 PIPE SIZE LARGER THAN THE SUCTION FLANGE OR FULL PORT

SUCTION PRESSURE GAUGE, 3 PIPE DIA. FROM SUCTION VALVE

ECCENTRIC REDUCER

Piping (Figure 2) Pipe Strain Pumps are not designed to be pipe hangers; nor are they designed to

PIPE SUPPORT, 11⁄2 T0 3 FEET FROM PUMP NOZZLE

Figure 2. Typical pump piping detail

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143

Figure 3. Pump distress at various flows off the BEP

• Lateral suction piping shall be supported and guided no closer than 1.5 feet to a maximum of 3 feet from flanges. Vertical discharge piping shall be supported from above using spring hangers. Hangers shall be sized so minimal vertical loads are placed on the pump nozzle when it is cold. Piping shall be guided to help prevent transfer of piping moments to pump nozzles. All pump loads shall be below those determined by the pump manufacturer and verified by calculations for both hot and cold service. Exceptions: Support design may differ from the above description if acceptable loading is verified by calculations and approved by the appropriate engineer. Special care should be exercised for lines with fast-acting and shock-inducing check valves.

discharge stub ends and flanges with gaskets to the pump; use four bolts. The spool pieces connected to the pump shall be fieldfabricated and field-fit. Tack weld the piping to the stub ends and match-mark the flanges. Finally, remove the spool piece and final weld the piping to the stub ends.

A

C

G FEET 700

144

E

D

B GPM 10,000

CS 10,000

CQ 1.00

CHP 1.60 1.50

4000

5000

400 2000

0

1.40

1000 1.30

2000

200

0.95

F 1000

0.90

100

400 200 100

0.05

1.20

500 40

1.15

20 200

0.80

Installation • Install piping to within 5 pipe diameters of the pump. Grout the pump baseplate and align the motor using the reverse indicator method (as a minimum) to within .002” TIR. Bolt the suction and

• Rework the spool piece if any of the following conditions exist: 1. If the flanges cannot be aligned by hand (not hand tools) for the insertion of the bolts 2. If a set of flange faces are not parallel to each other to within 1⁄32” across the length of the faces

0.10

1.10

6

40 100

0.70

10

1.05

0.15

20

0.60

0.18

15

Figure 4. Viscosity correction factor The Pump Handbook Series

50

30

1.00

400 350

Head in Feet

300 HQ — Two Pumps in Series — Add Heads Vertically

250 200 150

HQ — Single Pump

100

HQ— Two Pumps in Parallel Operation — Add Capacities Horizontally

50 0 0

200

400

600

800

1000

1200

1400

1600

1800

Flow in GPM

Figure 5. Parallel pump operation

3. If the spool piece, which should be 5 pipe diameters or greater in length, is more than +/- 1⁄8” of true length • The pump shaft deflection shall be checked during flange bolt tightening. Coupling alignment must remain within .002” TIR. (See also PIP REIE686, chapter 6 for typical piping practices.)

Alignment Do quality motor alignments. Alignments reduce failures! Maintaining proper alignment is impossible without correct pipe installation and a solid foundation. Proper coupling alignment for pumps requires a 0.003” offset with a 0.0005” angular tolerance at 1,750 rpm and a 0.0015” offset with a 0.0003” angular tolerance at 3,500 rpm. As a minimum, use reverse dial indicator; preferably use a laser.

Operation Musts Keep flow above minimums and under maximums. There are two types of minimum flows for pumps: mechanical and thermal. Let’s discuss mechanical minimum stable flow first. Stop pump dead-heading by using a minimum flow bypass or a power monitor. Vibration and shaft movement is lower at the pump’s best efficiency point (BEP). See Figure 3 for a list of problems that

occur when operating off the BEP. Excluding the new low flow models, standard ANSI pumps have the following typical minimum flows: at 1,750 rpm, minimum flow is about 20% BEP flow; at 3,500 rpm, minimum flow is about 30 to 50% BEP flow. If production needs less flow than that, you must install a re-circulation line to re-circulate enough flow back to a well pipe in the supply tank. This will keep the pump above its minimum flow point and can be accomplished using an orifice plate, a control valve or a re-circulation valve. Maximum flows are 110 to 115% BEP. Size your pumps to operate within these limitations to reduce vibration and shaft movement. Thermal minimum flows for pumps are something you may need to consider. In all centrifugal pumps some of the input energy is turned into heat due to fluid friction. The temperature increase is very dependent on the flow rate of the liquid through the pump. At BEP flows, this temperature rise is insignificant. But at shut off (no flow), the temperature rise is so substantial that damage will occur to the pump as well as the surrounding area. Because of this, a thermal minimum flow must be determined. On most pumping applications, though, the thermal minimum flow is well below the The Pump Handbook Series

mechanical minimum stable continuous flow. Thermal minimum flow becomes more prevalent when: 1. The temperature of the liquid being pumped is close to its boiling point 2. The temperature of the liquid being pumped is close to the temperature where the liquid’s properties may change 3. When a pump is re-circulating its flow back to the inlet of the pump and heat is allowed to build up In most cases, a 10°F temperature rise is acceptable unless NPSH is critical; in these cases the temperature rise should be kept to less than 5°F. The following equation is used to calculate the minimum thermal flow rate: Q=

5 x BHP x CHP T x CP

Where: Q = minimum thermal flow rate, in gpm BHP = non-viscous performance curve horsepower at shutoff Cp = specific heat of liquid ∆T = maximum allowable temperature rise, degrees F CHP = viscous horsepower correction factor from Figure 4. Pick a ∆T and calculate the minimum thermal flow rate in gpm.

Parallel and Series Pump Operation (Figure 5) In general, pumps operated in parallel give increased gpm capacity, while pumps operating in series give increased head. When operating pumps in parallel, it is important to keep a few rules in mind: 1. Both pumps should be identical in manufacturer, model, size and impeller diameter to avoid mismatch problems. 2. Each pump should be equipped with discharge check valves to keep the stopped pump from spinning backward in the event of a pump shutdown, which typically causes damage.

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3. Suction and discharge piping should be arranged as symmetrically as practical so that both pumps have the same Net Positive Suction Head available (NPSHa) and the same suction and discharge piping friction losses. 4. Flow orifices or meters should be provided to verify that each pump is providing flow above its minimum continuous stable flow. When operating pumps in series, be sure the pumps are matched in flow; the heads will be additive.

How to Start and Stop Pumps Some pump problems occur because of the way we start and stop them. If a pump is to supply liquid to an already pressurized piping system, the suction and discharge valves are open, and there is no check valve in the discharge line, the pump will start spinning backward. If the pump is started while it is spinning backward, the coupling and shaft can be damaged. If the impeller is screwed onto the pump shaft, spinning backward can unscrew the impeller and cause impeller, casing and shaft damage. The following are general rules, but please refer to your equipment manuals for specific instructions. Starting a Centrifugal Pump 1. If the pump is not full of liquid, open the suction and discharge valves to let the pump fill with liquid. Close the discharge valve and re-open it to about 1⁄4 open. Skip to Step 3. If the pump is full, skip to Step 2. 2. Open the suction valve all the way and open the discharge valve to 1⁄4 open. 3. Start the pump motor. 4. Open the discharge valve the rest of the way. Do this within one minute of starting the motor. Stopping a Centrifugal Pump 1. Close the discharge valve.

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2. Turn the motor off as soon as possible (within one minute). 3. Close the suction valve. Starting a Canned Motor or Mag Drive Pump After maintenance, air must be purged from the pump and bearings. To purge the pump and bearings of air, do the following: 1. Start with the suction and discharge valves closed. Open the suction valve all the way. Open the discharge valve 1⁄4 of the way to bleed air from the pump. 2. Start the pump and run it for 15 to 20 seconds. Repeat this jog starting three times. This enables the air to escape from the motor, bearings and pump. 3. Start the pump. Open the discharge valve all the way (again, within one minute of starting the pump). Starting a Full Canned Motor or Mag Drive Pump 1. Starting with the suction and discharge valves closed, open the suction valve all the way. Open the discharge valve 1⁄4 of the way. 2. Start the pump motor. 3. Open the discharge valve as soon as possible (within one minute). Stopping a Canned Motor or Mag Drive Pump 1. Close the discharge valve to 1/4 open. 2. Turn the motor off as soon as possible (within one minute). 3. Close the discharge and suction valves.

Other Considerations Pump Rebuilds Rebuild your pumps back to factory tolerances. Ask your pump manufacturers to supply you with the tolerance measurements you need to take to ensure your shaft is not bent, your bearing fits are correct, shaft end play is not a problem, etc. Many The Pump Handbook Series

pump problems are hidden in worn parts. Self-Priming Pumps Self-priming pumps are used to pump out of sumps, tankers, buried tanks, etc. Built into the casing of these pumps is a chamber that remains full of liquid after the pump shuts down. This liquid is used to prime the pump. When the pump is started, the agitation action from the spinning impeller re-entrains the air in the suction line and moves it to a discharge air separation chamber. The air and liquid separate and the air vents out the discharge line. The liquid flows to the bottom of the chamber, where a small slot allows it to flow back to the eye of the impeller. This movement of air causes a small vacuum in the suction line that enables the liquid to climb the suction pipe into the pump. Full flow and pressure is not attained until all air is removed from the suction line. After the pump stops, liquid in the discharge piping flows back into the chamber. It is stored for use the next time the pump needs to be primed. If these pumps are installed incorrectly, they will not self-prime. Several rules must be followed to ensure a trouble free installation: 1. Maximum liquid lift is about 25 feet. This is based on atmospheric pressure, which varies with elevation and the pump’s efficiency. If the liquid you are pumping has a specific gravity greater than 1, divide 25 feet by the specific gravity to determine the effective lift you can expect. If the liquid has a high vapor pressure, vaporization can occur in the suction line at the point of highest vacuum. The lift for these liquids will therefore be lower. Make sure you check for NPSH available to determine your maximum lift. 2. Minimize suction elbows. Don’t use more than one or two suction elbows. More elbows cause more friction loss, which can reduce NPSHa and cause vaporization in the suction line.

3. Choose the pump location carefully. Locate your pump near its supply source and try to limit your suction piping lengths to shorter than 25 feet. The longer the pipe is, the longer it will take to prime the pump. 4. Your suction pipe size should be the same size as your pump suction pipe. The larger the diameter of the pipe, the more air must be evacuated, and the longer it takes to prime the pump. 5. Avoid suction pipe foot valves on self-priming pumps. Before the invention of self-priming pumps, a foot valve held the fluid in the pump in an effort to avoid priming each time the pump was shut down. On a self-priming pump, these are not needed and only increase friction loss. 6. Provide a small diameter air bleed line. This should be about 1⁄4” from the discharge pipe and located between the pump’s discharge flange and the discharge check valve or isolation valve. Run this bleed line back to the suction source. Air at 77°F has a 3 density of 0.074 lb/ft . This is about 840 times less dense than water. Let’s say that with water a pump can generate 100 psig. When it pumps air, it will generate approximately 100 psig/840 = 0.12 psig. Unless your discharge piping is empty and your system does not contain a check valve, your pump cannot generate enough pressure to overcome the check valve and vent the air. Always install a vent line that remains open all the time. Yes, you will lose a little capacity through that line, but the alternatives are to have an employee open the line every time you start the pump or to install an automatic valve. Other Things to Look For • On some self-priming pumps, the casing and the impeller are a matched set. The clearance

•

•

•

•

between the impeller and the cutwater is 1⁄8”, and it must be maintained in order to self-prime. If the impeller needs to be trimmed, you must adjust the casing too. Follow minimum suction pipe submergence guidelines. If the suction pipe submergence is too small, air will enter the pipe and the pump could lose flow while it re-primes. This can become a vicious cycle (Figure 6). The discharge pipe must be designed so that enough liquid flows back into the pump to fill up the self-priming chamber when the pump shuts down. Therefore, do not put a 90° elbow on the pump discharge. Always allow for plenty of pipe (volume of pipe to be equal to the volume of the casing) before the discharge check valve and isolation valve. If the pump still won’t prime, check the suction pipe for leaks. Even a small leak can stop the pump from priming. The time needed to prime a pump is a function of pump size, impeller size and suction pipe length. Please refer to the manufacturer for this information. If the liquid in the priming chamber is hot, long priming times can cause the liquid in the priming

chamber to vaporize and make the pump lose prime.

Summary Paying attention to installation details or standards and verifying that the pump is not being operated in a way that it was not designed for will reduce work requests significantly. You, too, can have four yearsplus pump and seal life! ■

References 1. Duriron Pump Company, Pump Engineering Manual, 5th Edition. 1980. Steven J. Hrivnak is Principal Mechanical Engineer at Eastman Chemical Company’s Tennessee facility. He has a B.S. degree in Mechanical Engineering from Virginia Polytechnic Institute and State University. His responsibilities include process improvement of the coal gasification plant, working with a wide variety of rotating equipment, and focusing on reliability. He also leads the pump diagnostic team. Mr. Hrivnak has developed and taught on-site pump reliability courses and was instrumental in developing current pump, piping, baseplate and grout standards. He has written and published several articles on pump reliability for Pumps and Systems.

Figure 6. Minimum suction pipe submergence The Pump Handbook Series

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PUMP AND SYSTEM TROUBLESHOOTING HANDBOOK

Root Cause Failure Analysis Do you know why your problem solving is ineffective? This new way of thinking can drastically improve your maintenance program. By Dean Gano, President, Apollo Associated Services

he term Root Cause Failure Analysis (RCFA) has been associated with equipment reliability for the past ten or fifteen years and is generally understood to mean find and fix the causes of equipment failure. Many companies have no structured failure analysis. Instead, they operate on the broke-fix system: If something breaks, they fix it, but they don’t ask why it failed. When the problem happens a few more times, they may look for a root cause and fix it differently. These organizations often have measuring systems that record how long it took to repair a failure. Maintenance personnel get very good at repairing the equipment, but the failures keep happening. Organizations that are striving to learn and improve often have some formal root cause analysis program, but their effectiveness is often lacking, and repeat events are not uncommon. Unfortunately, in most businesses today, the majority of failures will be repeated and they will be repeated simply because the employees have never learned effective problem-solving skills. The purpose of this article is to examine the causes of ineffective problem solving and to present a problemsolving methodology that anyone can use on any problem and find effective solutions every time.

T

Conventional Wisdom First, let’s examine what conventional wisdom is telling us. Conven-

148

tional wisdom would have us find the root cause and remove it, but after 13 years of studying problem solving I have discovered the fallacy of most root cause thinking. It is linear and categorical and thus fails to disclose all the causes needed to find effective solutions. There are many other reasons for our ineffective problemsolving skills, but let’s examine the two basic schemes that have been developed to find root causes.

Figure 1. The Ishikawa Fishbone Diagram

Categorization Methods Over the past 25 years this has been the most popular method, but it is no longer used in many industries because of its history of ineffectiveness. Categorization methodologies provide a predefined hierarchical outline, checklist, or “Cause Tree” from which a root cause is chosen. Through group discussion and consensus, a solution is attached to the defined root cause and implemented. Categorization is a likely choice because it is the “Operating System” of the human mind. We compartmentalize and categorize everything, including ideas. It is only natural, then, The Pump Handbook Series

that we would try to use this process for problem solving. What we fail to understand is that categories are not causes. That is, categorization places causes in predefined boxes, such as People, Procedures and Hardware. Unlike cause and effect relationships, there is no necessary causal relationship between the defined categories. Perhaps one of the most publicized categorical methods is the Ishikawa Fishbone Diagram. This method asks us to guess at possible causes of the event, usually in a brainstorming session with knowledgeable people, place the results of our guessing within predefined categories, like a fish skeleton, and then vote on which causes we think are the root causes. The fishbone diagram is shown in Figure 1. As you can see, causes are listed categorically instead of in time-based causal relationships of “A” caused by “B” and “C”. In addition to this weakness, Ishikawa taught that each cause must be supported by evidence, but like other methods, he did not provide any way to graphically represent evidence, leaving the user to input biases and unsubstantiated wishes. Using the Ishikawa Fishbone Diagram is as random as spear fishing in an icecovered pond. Make several holes in the ice and poke the spear up and down until you hit a fish. If you get nothing from one hole, go poke in another hole (category). Repeat the process until you have several fish and then vote on which one is the best looking.

If we focus on categories instead of cause and effect relationships, we fail to identify the specific causal relationships necessary to understand what happened. By not understanding the causal relationships, solutions are often ineffective and the problem is repeated. A classic example of this can be found in the category of human error. Categorical thinking may set this up as a cause and often dictates punishment as the solution. In the 13 years I have been studying problem solving, I find that placing blame and punishment is used about 20% of the time, and is effective in less than 1%. To stop at a categorical cause such as human error is to ignore the infinite set of specific causes behind the failure. Another fundamental problem with categorization schemes is that the hierarchy of causes is subject to the biases or “realities” of whoever creates the categorization scheme. The notion that any individual or group could possibly create a cause tree that would include all the possible causes of every event-based problem is so arrogant as to be laughable. This will become more obvious when we discuss the cause and effect principle later in this article.

Causal Relationship Methods These methods are based on the understanding that everything that happens has a cause, and by knowing these causes we can control them and hence find effective solutions. While a sound basis, this method is only as effective as the tools provided. All but one of these methodologies fails to understand the four elements of the cause and effect principle. Instead, most of these methods encourage the problem solver to tell a linear story and fill out some forms or charts to document the probable root cause. Without a structured representation of the causal relationships and the evidence to support the causes, individual biases, political power and ignorance prevent effective solutions most of the time. In order to better understand the cause and effect principle, let’s examine it more closely.

The Cause and Effect Principle For at least 5,000 years, mankind has used the notion of causation to express happenings. Unfortunately, we have failed to differentiate the immense power of the Cause and Effect Principle from the simple notion of causation. Causation tells us that everything that happens has a cause; the Cause and Effect Principle provides four basic characteristics that enable us to develop effective problem-solving tools. The four characteristics are as follows: 1. Causes and effects are the same thing. 2. Causes and effects are part of an infinite continuum of causes. 3. Each effect has at least two causes in the form of actions and conditions. 4. An effect exists only if its causes exist at the same point in time and space. Knowing that cause and effect are the same thing only viewed from a different perspective in time helps us understand why people can look at the same situation and see different causal relationships. They are actually perceiving different time segments of the same event. If we treat each perspective as a different piece of a jigsaw puzzle, we can stop the usual arguing and work on putting the different pieces together. Knowing that causes and effects are part of an infinite continuum of causes helps us understand that no matter where we start our investigation, we are always in the middle of a chain of causes. This helps us understand that there is no right place to start. Again, just like the jigsaw puzzle, we can start the problemsolving process anywhere and still end up with a complete picture. This avoids the usual arguments over who is right. Probably the most profound characteristic of the Cause and Effect Principle is that each effect has at least two causes in the form of actions and conditions. This teaches us that every time we ask “Why?” we should find at least two The Pump Handbook Series

causes and for each of these causes we should find at least four more causes, and from each of these four causes we should find at least eight, and on to sixteen, thirty-two, and so on (Figure 2).

Figure 2. The infinite set

With this understanding, we see that there is an infinite set of causes for each effect, limited only by our lack of knowledge. It is this element of the Cause and Effect Principle that is responsible for our pursuit of simpler strategies. We never knew how to deal with an infinite set of causes before now. This is what the Apollo problemsolving method is all about. Cause and effect relationships exist with or without the human mind, but we perceive them relative to time and space. From observation we see that an effect exists only if its causes exist at the same point in time and space. In Figure 3, an open fire exists because conditional causes came together with an action cause at a particular point in time and space. As we can see, three conditional causes: oxygen, oily rags, a match, and one action cause, a match strike, occurred at the same point. If these four causes did not exist at the same time and space, the fire would not exist. For example, if the oily rags where stored in a closed can, or if the match was struck at a different time, a fire could not exist. Understanding this characteristic helps us determine the validity of causal relationship. By understanding these four relationships we can devise some simple tools that will enable us to tap the awesome power of the Cause and Effect Principle.

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Figure 3. Example of cause and effect relationships

Effective Problem-Solving Tools For the past 13 years, I have been working with many different industries to develop a simple set of tools that everyone can use on any event-based problem. This methodology, which is as much a philosophy as it is a set of simple tools, does not require forms or checklists. When fully implemented, it empowers everyone in the organization to be an effective problem solver. It truly changes the way people think about the world around them and improves their ability to effectively communicate what they know. People begin to realize that things do not just happen, and that everything is caused to happen. As a result, a proactive attitude begins to develop, and conditional causes that set people up to fail are removed before they can cause a loss. This methodology, called Apollo Root Cause Analysis (ARCA), provides a four step process based on the cause and effect principle. Step 1. Define the problem by writing the What: Primary Effect (noun verb) When: Relative time of the Primary Effect Where: Location in system, facili-

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ty or component Significance: Why you are working on this problem?

Step 2. Create an Apollo Cause & Effect Chart (Figure 4) For each Primary Effect ask why. Look for causes in Actions and Conditions. Connect causes with “Caused By.” Support causes with evidence or use a “?”

of this later. An added feature unique to this method is the inclusion of evidenced-based causes. When you can’t find evidence, or can’t find the next cause, use a question mark. Evidence should be data we know from using our senses, not supposition. It supports the existence of the various causes. Without sensed evidence, we are subject to our own prejudices and preconceived ideas. It helps to be humble when pursuing causes and evidence. When you don’t know, admit it and use a question mark to signify this on the chart. If there is value in pursuing the unknown, then do so. If no value is perceived, you can make a conscious decision to stop, like we did after “Car Existed,” or leave the uncertainty on the chart, like we did after “Narrow Road.” If evidence is uncertain, use a question mark to express your doubts. Look closely at Figure 5 to see how the question marks have been used. To better understand the four-step process, let’s take a closer look at each step.

Problem Definition Let’s start by defining what a problem is. A problem is the gap between actual and desired. If the goal is to produce a product safely and efficiently and someone is injured in the process, or the cost of production exceeds the sales price, there is a gap

Step 3: Identify effective solutions Challenge the causes and offer solutions. Identify the best solutions—they must: • Prevent recurrence • Be within your control • Meet your goals and objectives

Step 4: Implement the best solutions When completed, an Apollo Cause and Effect Chart may look something like Figure 5. Please note the branched causes connected by the words “Caused By.” Using “Caused By” prevents storytelling by forcing our mind to go from present to past. More on the importance The Pump Handbook Series

Figure 4. Basic chart elements

between actual and desired. Every problem can be defined in terms of a gap. With this fundamental understanding of a problem, every problem should be defined within the context of a goal. Therefore, to define a problem, we must first know what the

goals are. Likewise, we must know what our present state is. Sometimes the goals are not clear. Have you witnessed groups within an organization, such as Maintenance and Operations, which appear to have conflicting goals? Do they end up seeing different problems? Does it lead to miscommunication and conflict? Do different groups within an organization really have different or conflicting goals? It may seem like they do, but if the organization is to be effective, these goals must be aligned. This perception of conflicting goals is the result of not understanding the goals. As an example, consider a football team. The goal of a football team is to win the game. At the highest level everyone’s goal—players, coaches and managers— is to win the game and everyone’s actions are focused on that goal. But let’s look at a different level. What is the goal of the

offense? What is the goal of the defense? At this next level, the offense’s goal is to score points. The defense’s goal is to prevent their opponent from scoring points. The offense and the defense, at this next level, have completely different goals: score points and prevent points from being scored. Different goals at this level, same goal at the higher level. Is there a conflict? No. At the highest level they have the same goal: win the game. Work groups and individuals should always consider their individual, group or team goals within the context of the overall goals. If individual goals are not aligned to the overall goals, sub-optimization can occur at the expense of the organization’s goals. How do we overcome this apparent conflict in goals? By defining the organization’s overall goals within the significance section of the problem definition. Every problem, incident,

opportunity or project should be defined within the context of the organization’s overall goals. A clear problem definition helps to get everyone “on the same page” when an incident occurs. It improves the understanding of what is really important to the organization—a crucial step toward improving the organization’s overall effectiveness and ultimate success in preventing problems from occurring. The four elements of problem definition are listed below. What Asking, “What is the problem?” can generate everything from a common response to a variety of answers from a group. By asking the question, we are beginning to clarify the gap between desired and the actual we are trying to prevent. People will respond from their individual points of view, which may

Figure 5. Example of Apollo Cause and Effect Chart The Pump Handbook Series

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or may not reflect how the overall organization sees the problem. If there are two or more responses, simply separate them with a comma. The “Significance” section will align the different perspectives in an organization by providing the primary effect(s) for the cause and effect chart. There may be more than one problem, and you are going to find it out here. When There are two components to this section. The first part of the “When” is the date and time. Sometimes it might be important to capture an incident in terms of which second the incident occurred in a process step or it may be sufficient to simply state “in the morning”—it depends on the nature of the problem. The second part of this section is the relative timing. It may help to think of this in terms of the contextual factors associated with this incident. You might ask, “Were we using this equipment for the first time? Is this the first time this person has performed this task or used this application?” This may provide important causal reference points once you’re into the analysis. At this point in the problem definition, we don’t know whether they are important or not, so we need to write them down. Where This section should define the physical or process location of the incident or problem. A consistent approach of starting with the higher levels of a system and stepping to the lower levels works well for developing a clear structure that captures the setting. The systems approach helps develop physical locations that accurately reflect the actual layout of a work group or facility. There is also a relative aspect of the “Where” that may help to capture the subtle, but potentially important causes, such as: “In the corner where the light doesn’t shine.” This prompts us to ask during the analysis phase: “Why no light?” Significance The “Significance” section asks the question at the beginning of the analy-

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sis process: “Why are we spending time and resources on this issue?” To take full advantage of the significance, it should reflect the overall goals of the organization. The incident perspective should not be from any one group or individual, like the maintenance group or department manager, but from the stated goals of the company. This helps everyone to see the problem more accurately. It also causes clearer communication regarding priorities and economics. Look at the example below for a difference between a single point of view and one that looks at the overall goals of the organization. The Primary Effect or “What” of this example was stated as a “Pump Loss.”

Significance: (Single point of view) Minor leakage, no reportable spills, $2,500 pump replacement Significance: (Expanded to reflect all company goals) Safety: Near miss of flying pump impeller parts Environmental: Minor Leakage, no Reportable spills Production: 20% reduction in pulp processing line #2, cut back in throughput of 15,000 pounds per hour at $0.30/lbs. for 6 hrs. equals $27,000 Maintenance: Materials = $1,250 pump replacement Labor = $1,250 Frequency: Second time this year, three times in 1998, five similar pumps in unit 3 In the above example the significance has been expanded into five areas that reflect the overall goals of the company. Even though the maintenance group only saw the pump loss as the “problem,” it is important for everyone in the organization to define the problem within the context of the overall company goals. Properly stating the significance generates the primary effect for the cause and effect chart. For example, the primary effect could be changed to “Production Loss,” which is eventually caused by the “Pump Loss.” It is a subtle but important The Pump Handbook Series

distinction required to improve communications, because the focus changes from one of operational or maintenance perspective to one of the company goals—produce product safely. Problem definition is not only the first step, it is critical to effective problem solving because it forces us to understand the problem from all perspectives, thus improving the communications of the organization.

Create an Apollo Cause and Effect Chart For each Primary Effect (the “What” in your problem definition) begin asking why. Use yellow stickies on a board to document your progress. As answers come, connect the causes with the words “Caused By.” This will go a long way to ensuring a good chart, because the words “caused by” force the mind to go from the present to the past, and thus minimize storytelling. When you reach the end of a line of causes, go back to the beginning and start over. Each time you ask why, try to find two or more reasons. Look for condition and action causes. If you have identified an action cause like an operator-pushed button, identify the conditional causes that had to be present at that same point in time and space to allow the effect to occur, like “button existed,” “operator at panel,” etc. After going through the cause chart four or five times, begin to write down the evidence that explains how you know each cause exists. Once you have done this several times it gets very easy and natural, but it may be difficult at first if you are used to thinking and communicating by storytelling. Storytelling is a significant barrier to effective problem solving.

The Problem with Storytelling Storytelling is a linear understanding of an event in a time sequence from past to present, and it significantly violates the cause and effect principle. While stories are our primary form of communication, they conflict with the cause and effect principle in three ways: 1. Stories start in the past, while causal relationships start with the present.

2. Stories are linear, while causal relationships follow the branches of the infinite set. 3. Stories use inference to communicate meaning, while problems are known by sensed causal relationships. Let’s examine a simple little story to see how detrimental these conflicts are: The little crippled boy lost control of the run-down wagon and it took off down the hill on a wild ride until it hit the little blind girl next to the drinking fountain by Mrs. Goodwin. The little boy was in the wagon the whole way, but was not injured. The boy’s mother should never have left him unsupervised. The root cause of the girl’s injury was human error. Stories Start in the Past This story starts in the past at the top of the hill and progresses through time from the past to the present; from the beginning of the ride to the end; from the safe condition to the stated problem of injury. The conflict this creates is that by going from past to present, we do not see the branched causal relationships of actions and conditions. If we could know every cause of this injury example, we would see a diagram of cause and effect relationships similar to Figure 3. That is, we would see a set of ever-expanding causes starting with the injury and proceeding into the past. To express what we know causally in story format, we would first need to express all the causes on the right side of the diagram, i.e., starting from the past. Our language and the rules of storytelling simply do not allow for this. We can not express 16 causes and then tell what they caused and so on. No one would sit still for a story told this way, because stories are about people, places and things as a linear function of time. Stories Are Linear As we look at this simple story (or any story), we find that our language and our mind restrict us to a linear path through time and space. Stories go from A to B to C, without regard for the order of causal relationships. We are told of

the little boy losing control of the wagon as it goes down the hill and strikes the little blind girl. There is no ever-expanding set of branched causes expressed like those in Figure 5. We have the ability to escape this linearity and express branches if we use the words “and” and “or,” but the rules of grammar tell us not to use these connecting words excessively. The best we can accomplish is one or two branches for each sentence. The conflict arises because the cause and effect principle theoretically dictates an infinite set of causes for everything that happens while stories are created and expressed linearly. Stories Use Inference to Communicate Causes Since good stories seem to provide us with a valid perception of what happened, we need to question how this can occur in light of the above two conflicts. The key word here is perception. When we read or hear a story our mind provides most of the information. As we read the words, we are busy creating images in our mind’s eye. These images are created from past experience and assembled into a sequence of events. We don’t necessarily need causal information to create the image; our mind fills in with its own causes. Because stories or the pictures we create do not express the branched causes of the infinite set, we must make up for it somehow, and we do this by inference. We infer causes within the story that are not stated. For example, we read that the little crippled boy lost control of the wagon. Since no cause is stated for how he lost control, we can infer anything our mind will provide, and we do just that if questioned about it. Furthermore, stories imply cause by the use of prepositions such as in, on, with, etc. Prepositions and conjunctions by definition imply a relationship between words, and the relationship is left to the reader. The word “and” is often used to mean “caused.” In this story we read that the boy lost control of the wagon and it took off down the hill, meaning the loss of control caused The Pump Handbook Series

the wagon to take off down the hill. Within this “and,” is the potential for many causal relationships and they are left for the reader to interpret. For me, the “and” between “lost control” and “took-off down the hill” is obviously a broken steering mechanism. You may have inferred that crippled means a paraplegic and this condition was the cause. The next person sees the wagon wheel strike a rock, which causes the wagon to veer sharply, while another person is so shocked by the politically incorrect usage of “crippled” and “blind children,” that she has lost the ability to think about the problem altogether. Because we do not express what is happening causally, each word in the story provides the reader with the opportunity to think they know more about the event than is stated. We interpret the situation from our own biased mind, which is not necessarily what happened or what the storyteller meant. In the end, each one of us thinks we know what happened but we really don’t because stories can not express the full set of causal relationships. Our linear language, and the linear thinking behind it, prevents us from knowing and expressing what really happens in any given situation. And when we get together to discuss our problems, we usually end up arguing and making presumptuous statements like: “It’s obvious why this happened,” or “the solution is clear.” By breaking away from storytelling and creating an Apollo Cause and Effect Chart, we are able to include all possible causes without the usual arguing and politics.

Identify Effective Solutions Once we have an Apollo Cause and Effect Chart, we can look at the causes and propose solutions to remove or control them in a way that will prevent recurrence of the primary effect. This is a two-step process. First the creation stage, where we challenge each cause and offer solutions. Next, we evaluate each proposed solution against the three solution criteria: prevents recurrence, is within our control, and meets our goals and objectives. Looking at

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Figure 5, we can see several opportunities for prevention, such as not parking on the shoulder of the road, or not listening to annoying radio personalities while driving. Each cause provides the potential for effective solutions and the more causes we can put on the chart, the greater the chance of finding a creative solution. Every solution is a function of our goals and objectives. It matters not what someone outside our organization thinks of our solutions, because we are the ones who are going to be responsible for the success or failure. Accepting this responsibility is what makes any solution “right” for our organization. If our team or organization can show that our solutions are supported by clear, evidence-based cause and effect relationships and meet the three solution criteria, then our solutions will be effective. If, along the way to our solution, another stakeholder does not agree with our analysis or solutions, they are enthusiastically asked to share their evidenced-based causes. If they are able to improve our understanding, then we are obliged to add them to the chart and modify our conclusions as necessary. A common occurrence for people who use Apollo Root Cause Analysis is to take the problem analysis with its causes and recommended solution all the way to the president of the corporation and rarely encounter opposition. Why? Because there are no opinionated stories, just evidence-based causes. The chart provides a clear picture for anyone to see how effective the solution will be. Because the causes are clearly stated and supported by evidence, there is no need to test the solutions, as the other methodologies would have us do. During a visit to a manufacturing plant where Apollo Root Cause Analysis has been in use for almost a year, the plant manager confided in me, “It is a little scary letting go, but if I can successfully apply the solution criteria to the corrective actions, then I let it go—regardless of my gut feelings or opinions. I have not been disappointed yet, and this process has truly empowered our workforce. They love it because it really works.”

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Competitive Advantage In the week following a training session on this new approach, a reliability engineer at a prominent chemical company in Europe put these methods to use. After clearly defining the problem and its causes, he and his team fixed an old problem for the last time. After 35 lifetime failures of a main process compressor, their solution completely eliminated the problem. To date, they have realized savings in maintenance costs in excess of $750,000. This company, like many others who have learned to create a common reality, is now using this new approach for solving event-based problems throughout the world and is gaining a competitive advantage by reducing maintenance costs while improving productivity and safety. In today’s global marketplace, everyone can work harder, longer, or faster, but only a few will choose to work smarter. Working smarter means better communications and better problem solving, which translates to a more productive and safer work place. Using the ineffective methods of storytelling and categorization may have been good enough last year, but they won’t take you into the future. Effective problem solving is a definite competitive advantage both from a production standpoint and a safety perspective. We have seen how using the Apollo Cause and Effect Chart can help us communicate better and thus become more effective problem-solvers and decision-makers. By first defining the problem with the What, When, Where and Significance, and then preparing an Apollo Cause and Effect Chart, a common reality is created such that everyone in the organization can provide creative solutions that meet the three solution criteria. As more personnel within an organization learn this simple structured approach to event-based problem solving and effective communications, a questioning attitude begins to permeate the work culture. With this new attitude comes the understanding that things do not just happen, and if we look more closely at causes and effects, we can prevent problems from occurring The Pump Handbook Series

in the first place. With these simple communication tools, employees are able to communicate their ideas like never before. Likewise, the time and money saved not having to resolve repeat events and emerging issues can be used for positive business activities such as increasing productivity or upgrading equipment. You can’t afford not to improve your communications and problem-solving skills. ■

References 1. Churchland, Paul M., 1996, The Engine of Reason, The Seat of the Soul, MIT, Cambridge, MA 2. Covey, Stephen, 1990, The Seven Habits of Highly Effective People, Simon & Schuster Fireside Books, New York, NY 3. Damasio, Antonio, 1994, Descartes’ Error, Grosset/Putman, New York, NY 4. Gano, Dean, 1999 Apollo Root Cause Analysis - A New Way Of Thinking, Apollonian Publications, Yakima, WA 5. Goleman, Daniel, 1995, Emotional Intelligence, Bantam Books, New York, NY 6. National Safety Council, 1983, Accident Investigation: A New Approach 7. Senge, Peter M., 1990, The Fifth Discipline, Currency Doubleday, New York, NY 8. Van Doren, Charles, 1991, A History of Knowledge, Ballantine Books, New York, NY 9. Wilson, Paul, et. al., 1993, Root Cause Analysis, ASQ, Quality Press, Milwaukee, WI Dean Gano is President of Apollo Associated Services, an international business that provides Root Cause Analysis training. Mr. Gano has more than 25 years of experience in process industries, including nuclear, petroleum and chemical, and has been teaching Root Cause Analysis for the past 11 years. He is a member of the American Society of Mechanical Engineers, the American Society of Safety Engineers and the American Society for Quality.

Coupling Alignment Misalignment can lead to failure throughout your system. Mark F. McCullough, Lovejoy, Inc. flexible coupling should never be the basis for allowable misalignment just because it can accommodate more misalignment than other connected pieces of equipment. The coupling may have ten times the capability for misalignment than the rotating equipment. System alignment should be based first on the minimum requirements of the driven equipment, or the driver, and then the coupling. Misalignment is a leading cause of bearing and seal failures, vibration, oil leakage from bearing frames, broken shafts and coupling failures. So, equipment always should be aligned as close as possible, keeping within the economics and sophistication of the system

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Planes of Flexibility Users should consider the plane of flexibility as pivot points within the coupling. A full flex coupling has two pivot points, with one attached to each of the connected shafts. A coupling with a pivot point on one side and a rigid shaft attachment on the other is called a single flex, flex rigid, or half coupling. The pivot point can be in the loose fit between separate parts, such as the hub tooth-to -sleeve tooth interface in a gear coupling, or in the bending of a continuous flexing element as used in disc or diaphragm, or link couplings. The flex plane of an elastomeric coupling is within the elastomer itself. There are three variations to shaft misalignment: • parallel offset (radial) misalignment; • angular misalignment; and • the combination of angular and parallel offset.

Axial displacement is considered a bearing systems and on one end of form of misalignment with which the floating shaft systems (Figure 2). coupling may need to deal.

Axial Displacement Parallel or Radial Misalignment Parallel (radial) misalignment occurs when the driving and driven shafts are parallel but with some offset between their axial centers. Accommodating such offset requires either a full flex coupling (with two flex planes) or two single flex couplings in series. In either case, the greater the axial distance between the two flex planes, the greater the coupling's parallel (radial) capability. Typical full flex couplings include gear, grid, and dual element disc or diaphragm types. Spacer couplings are good for extra radial displacement and closecoupled couplings such as a standard flanged sleeve gear coupling provide the minimum. Spindle couplings and floating shaft couplings provide the maximum capability by further spreading the flex planes. Although the elastomeric type has only one flex plane, the elastomer can distort enough in some cases to provide significant parallel offset capability if it has sufficient resilience. Elastomeric couplings can also be made as spacer or floating shaft types to a limited extent (Figure 1).

Angular Misalignment Angular misalignment occurs when driving and driven shafts are not parallel but their axial center lines intersect. Flex-rigid or half couplings provide only angular misalignment, because there is only one flex plane. Single element disc or diaphragm couplings provide for angular misalignment only. Single element couplings are used on three The Pump Handbook Series

Axial misalignment is an in-out movement along the shaft axis, often associated with thermal shaft growth and floating rotors. Thermal growth is the result of high temperature in the rotating equipment causing an unconfined growth along the length of its shaft. Sometimes a thrust bearing at the coupling end of the shaft will direct axial movement or growth away from the coupling, but sometimes the thrust bearing is at the other end of the shaft. Another well-known source of axial movement is the rotor that seeks its magnetic centers. The coupling must either accommodate axial movement or contain it by transferring the thrust to the bearing system of the rotor. Those that contain it are called limited end float couplings. Sometimes axial thrust is deliberately transferred to another machine through the coupling. Limited end float may or may not be invoked in such a case. Gear couplings exhibit the best capability to handle axial misalignment or movement. In the gear coupling the hub teeth are free to slide axially within the sleeve while enmeshed. Other types such as the diaphragm coupling can flex or stretch to allow some axial displacement. The disc coupling can also, but to a lesser degree than the diaphragm coupling. In both the disc and the diaphragm coupling, axial movement is met with resistance that increases as the displacement increases. The elastomeric coupling is not a good unit for axial float or axial displacement. Sometimes the unit can slide in one direc-

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tion, but there are no limiters that stop the slide and the coupling disengages. In other types of elastomeric couplings axial distortion would overload the elastomer when combined with normal misalignment and nominal torque transmission.

Comparison of Misalignment for Various Coupling Styles The coupling manufacturer’s full line catalog is a good starting point for a comparison of misalignment capabilities. Coupling misalignment is usually given in terms of angular misalignment that can be converted to radial or parallel when two flex planes are used. The conversion of angular to radial misalignment capability is a matter of plane geometry. The radial offset distance is the product of the tangent of the angle of misalignment and the distance between flex points. It is true that misalignment, torque capabilities and coupling life are intertwined. The torque capability of a coupling is reduced when the coupling is misaligned. The reduction in life can come from higher wear and the reduction in torque from high fatigue forces. Misalignment will cause fatigue forces. Some manufacturers publish their torque ratings as a maximum value and require that the user de-rate by some factor to determine usable nominal torque. Others publish their torque ratings at rated full misalignment. When the coupling is selected it should be able to carry the nominal torque while misaligned per the application. Gear couplings are capable of 11⁄2% of misalignment per gear mesh, although the gear coupling needs some misalignment to push the lubricant to the friction surfaces of the teeth. With tooth modification that misalignment can be increase to as much as 6°. At that much misalignment the pressure angle of the tooth may be changed to strengthen the tooth. That would be the method for spindle couplings. Highspeed gear coupling units might be reduced to 1⁄4° per mesh. Other metallic element flexible couplings have a wide range of angular

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Figure 1. Parallel or radial misalignment

Figure 2. Angular misalignment

capabilities. The flexing link coupling is suitable for 6° while the torsional spring coupling can be used up to 41⁄2% of angular. Published data for the disc and diaphragm coupling types range from 1⁄4° to 11⁄2°. The elastomer couplings such as the jaw types and the unclamped donut elastomer in shear type couplings are typically limited to 1°. Donut and Tire type elastomer couplings are capable of 3° of angular misalignment.

Acceptable Misalignment While a high-speed, high-power system requires close alignment, generalpurpose equipment is aligned to a looser specification. For example high-speed equipment (running at 3000 rpm or more), needs alignment to .0005 inch (or better) per inch of flex point separation as defined here. General-purpose equipment can be acceptable at .001 inches per inch of separation. Smaller values will improve the operation but should be consistent with the equipment manufacturer recommendation. The preceding recommendations are typically used with "close coupled" equipment. When spacers and floating shafts are chosen for the ability to allow radial displacements of two pieces of equipment, these rules of thumb would not necessarily apply. The Pump Handbook Series

It is sometimes necessary to have large amounts of parallel displacement built into the equipment installation. Special design alterations to the couplings and the connected equipment can also be required in those cases. The drivers and driven equipment must have sufficient design strength to deal with the increased reactionary load. Equipment with spacer couplings or even floating shafts may be designed for ease of maintenance rather than for the increased misalignment capability that would be possible. Exceeding acceptable misalignment contributes to vibration problems. Severe misalignment will impose a heavy vibratory force on the equipment. Large amounts of parallel misalignment are acceptable only on machines operating at slow speed. High-speed machines, even those with spacers and floating shafts, must be aligned closely to limit the vibration and the reactionary loads.

Reactionary Loads from Misalignment Shaft-to-shaft misalignment causes couplings to impose reactionary loads on the connected equipment. The greater the misalignment, the greater the reactionary load. Different types of couplings produce different reactionary loads.

In gear couplings, reactionary forces result from the sliding friction of the tooth-to-tooth movement. The sliding friction is considerable when dealing with metal-to-metal contact. There is some lubricant but it is a very thin film. When a gear coupling is misaligned, the friction or drag forces become a bending moment on the system. The bending moment reactionary load can be 10% of the value of the torque transmitted. At that value it is many times the reaction load of a disc coupling. Disc and diaphragm couplings utilize the flexing of thin metal elements to handle the misalignment, both angular and axial. The thinner the element, the less force it takes to bend the metal. The force to bend becomes the force of the reaction in the classic physics truth about equal and opposite reaction forces. Link couplings and spring couplings also provide a reactionary force in proportion to the loading force. Elastomeric couplings exhibit different kinds of forces depending on the type of coupling. Elastomeric couplings in shear will impose a thrust (axial) load on the adjacent machine parts. Distortion of an elastomeric donut or misalignment in a jaw coupling will impose a bending moment. Tire-type couplings exert a thrust load also as the centrifugal force acts on the tire. Reaction forces in the elastomer coupling are a function of the resiliency and may also have some friction components. The friction is not unlike the sliding of gear teeth and occurs on couplings like the jaw designs. The reactionary loads include the weight of the coupling. Because weight usually reflects size, the more power intensive couplings (i.e. those with higher torque capacity for smaller size, such as gear couplings) will have the advantage of lower reactionary forces at comparable torque loads. Large size couplings (i.e. less power intensive types) may also be made from light materials to reduce their weight and their inertia. The loads imposed by the coupling are related to the point of loading. The pivot point or flex plane is the location of the loading for the coupling. Those

couplings that move the pivot point closer to the next available bearing are termed "reduced moment" couplings, because they reduce the reactionary load imposed on the bearing.

The Alignment Occurs at Installation Alignment is accomplished at the final installation point of the equipment. Alignment is done when the coupling is installed. Equipment can be aligned when the driver and the driven equipment are assembled at the manufacturer, however the alignment still must be checked and likely adjusted at the final installation. When aligning the coupling, the installer must take into account the conditions affecting the rotating machinery. Not only is it important to understand where the equipment is at standstill, but also it is important to know where the equipment will move when in operation. This is the most difficult part of alignment and installation. For example, hot operating equipment grows when it is brought up to operating temperature. The coupling is aligned cold at one location with the expectation that movement will bring it to closer alignment. In addition to temperature considerations the rotating equipment alignment can be affected by tolerance stackup, pipe loading, the foundation and conditions such as bent shafts or soft foot. Before aligning the equipment it is best to check for those occurrences. Alignment can be measured by use of a straight edge and feeler gage (or calipers or taper gage). That would be the simplest method, but the least accurate. Dial indicators are used with the reverse indicator and the face and rim method. The most modern method uses a laser alignment system. Each method has its strengths and weakness and all can be satisfactory depending on the skill of the installer. There are many books and papers written on the "how to" of alignment. Here are general descriptions of the different processes. A straight edge is used to determine the shaft offset by eye. It is used for both The Pump Handbook Series

vertical and horizontal planes. The taper gage (or calipers or feeler gage) is used for angular misalignment. The shaft separation or "BSE" dimension is measured with a ruler. It is a trial and error process. Dial indicators are used with the reverse indicator or the face and rim method. The dial indicators are mounted on the shaft opposite to the reading to be taken. These indicators are accurate to ± 1 mil. In the reverse indicator method, readings are taken from coupling hub on shaft "A" to the rim of the coupling hub on shaft "B". A second set of readings is taken from the coupling hub on shaft "B" to the rim of the coupling hub on shaft "A". Both sets of readings are plotted on graph paper or become the input to a computer program. With the proper calculations in plane geometry, the misalignment of both parallel offset, and angularity of the shafts can be determined. The face and rim method uses the dial indicator mounted on one coupling hub to take readings on the face and the rim of the second coupling hub. Again with graphical plotting or a computer, and plane geometry the misalignment of both types can be determined. This method can be very accurate if done with graphical assistance or computer assistance. Other commercial mechanical and electrical devices can obtain the results by measuring the positions of two shafts. Laser beam alignment uses the laser to replace the dial indicator. It is a little more accurate, but is much more costly. Included with the laser package is the means of direct input to a computer program that calculates the moves necessary to align the equipment. Lasers are accurate to ± 3 micron or better. The laser is a light beam that is very narrow and focused. The beam generating equipment is mounted on the equipment shaft and aimed at a device on the opposite shaft. The device can be a reflector or can be the photodiode target cell that will generate a voltage. The amount of voltage that is generated will depend on the position of the

157

light beam as it hits the cell. A reflector will cause the beam to return to a target cell that is mounted with the laser generator. The generated voltage becomes the input to a system that calculates the misalignment and needed corrections.

Conclusions

often is ordered with the services of installation start-up supervision from the manufacturer. That service is well worth its extra cost. Furthermore, there are many companies within specific industries that provide alignment service. Always keep in mind that equipment should be aligned to the rotating equipment manufacturer's standards and requirements, not the coupling manufacturer’s. When operating misaligned, the flexible coupling can transmit reactionary loads and vibrations that are within the coupling’s capabilities, but not within the capabilities of the equipment. ■

Proper installation and alignment procedures are included with the installation instructions of rotating equipment and couplings. Often the coupling manufacturer can provide further guidelines for installing the coupling and aligning the rotating equipment. In addition, there are many published papers and pamphlets on the subject. Mark F. McCullough is Director of High-speed, high- powered equipment Marketing & Application Engineering for

158

The Pump Handbook Series

Lovejoy, Inc. He has been with the company for 10 years, starting in the Sales Department and serving as Product Manager for elastomeric couplings for six years before assuming his current position in 1998. McCullough is Lovejoy's representative to the marketing committee in the Hydraulic Institute, and the author of several technical articles on flexible coupling selection and chain/belt tensioner selection. He received his B.S. degree in Marketing from Arizona State University and his MBA from DePaul University. (Editors Note: This article was adapted from The Coupling Handbook published by Lovejoy, Inc. For information, e-mail Tracy Laskowski at [email protected].)

Subject Index - 2002 abrasion combating ……………………………………………………………………Vol abrasives gear pumps …………………………………………………………………Vol mechanical seals …………………………………………………………Vol peristaltic pumps …………………………………………………………Vol positive displacement pumps …………………………………………Vol progressing cavity pumps ………………………………………………Vol thermoplastic pumps ……………………………………………………Vol accumulator tanks …………………………………………………………………Vol acoustic emissions reducing ………………………………………………………………………Vol additives pumps for ……………………………………………………………………Vol adjustable speed drives ……………………………………………………………Vol affinity laws rotary pumps ………………………………………………………………Vol variable frequency drives ………………………………………………Vol air injection ………………………………………………………………………Vol leakage ………………………………………………………………………Vol low flow pump designs……………………………………………………Vol mechanical seals …………………………………………………………Vol pumping with ………………………………………………………………Vol suction line …………………………………………………………………Vol circulation ……………………………………………………………………Vol air chambers …………………………………………………………………………Vol air operated diaphragm pumps overview ………………………………………………………………………Vol pulsation control …………………………………………………………Vol specifying ……………………………………………………………………Vol alignment approaches …………………………………………………………………Vol baseplates and ……………………………………………………………Vol c frame adapter and………………………………………………………Vol canned motor vs. magnetic drive ………………………………………Vol centrifugal pump installation …………………………………………Vol couplings ……………………………………………………………………Vol final ……………………………………………………………………………Vol laser……………………………………………………………………………Vol machinery vs. coupling……………………………………………………Vol maintaining …………………………………………………………………Vol methods ………………………………………………………………………Vol methods (summary) ………………………………………………………Vol misaligned shafts …………………………………………………………Vol monitoring report …………………………………………………………Vol movement methods ………………………………………………………Vol

1, 111-114 2, 3, 2, 2, 2, 1, 2,

174 62-65 183-184 19-21 13-14, 181 282-287 9-10

4, 22-24 2, 31-34 4, 132-134, 192; Vol 2, 213-218 2, 119-120 4, 451 4, 1, 2, 3, 5, 5, 1, 2,

520 208 194 63-65 75-76 25 299 9-14

2, 35-38 2, 27-28 2, 42 4, 4, 4, 1, 1, 4, 4, 4, 4, 5, 4, 4, 4, 4, 4,

108; Vol 5, 155-158 87; Vol 1, 336-341 107 69 47; Vol 1, 336-341 147-148; Vol 5, 155-158 397 323-324 149-150, 433 77 294-295 379-385 47 399-402 105-106

MTBR and …………………………………………………………………Vol Mounting ……………………………………………………………………Vol piping …………………………………………………………………………Vol piping-to-pump ……………………………………………………………Vol rim and face…………………………………………………………………Vol rotary pump start-ups ……………………………………………………Vol shaft-to-shaft ………………………………………………………………Vol spacer couplings ……………………………………………………………Vol tools ……………………………………………………………………………Vol vertical pumps………………………………………………………………Vol alkylation process ……………………………………………………………………Vol angular contact bearings …………………………………………………………Vol angular misalignment spacer couplings ……………………………………………………………Vol ANSI B73.1 ……………………………………………………………………………Vol ANSI impellers low flow/high head ………………………………………………………Vol ANSI pumps design choices ………………………………………………………………Vol criteria for repair …………………………………………………………Vol CPI service …………………………………………………………………Vol DuPont Corporate Design ………………………………………………Vol material choices ……………………………………………………………Vol reliability improvements …………………………………………………Vol repairing ……………………………………………………………………Vol selecting a pump …………………………………………………………Vol troubleshooting with limited information …………………………Vol upgrades ……………………………………………………………………Vol ANSI/ASME B73 ……………………………………………………………………Vol revisions ………………………………………………………………………Vol anti-friction bearings centrifugal pumps …………………………………………………………Vol API 610 pumps ………………………………………………………………………Vol criteria for repair …………………………………………………………Vol hot fluids ……………………………………………………………………Vol noise investigation …………………………………………………………Vol pressure ratings ……………………………………………………………Vol rebuilding, at Avon ………………………………………………………Vol upgrading techniques ……………………………………………………Vol API 682 standard ……………………………………………………………………Vol application …………………………………………………………………Vol reference guide ……………………………………………………………Vol VOC control …………………………………………………………………Vol API 682 seals …………………………………………………………………………Vol API applications hazardous fluids……………………………………………………………Vol API plans circulating pump arrangements …………………………………………Vol plan 53 ………………………………………………………………………Vol high temperature …………………………………………………………Vol API process pumps retrofit…………………………………………………………………………Vol API pumps casing design ………………………………………………………………Vol noise, reducing ……………………………………………………………Vol seals for ………………………………………………………………………Vol API seal flush plans ………………………………………………………………Vol API vs. ANSI procurement …………………………………………………………………Vol ASME pumps chemical applications ……………………………………………………Vol

4, 1, 1, 4, 4, 2, 4, 4, 4, 4, 1, 4,

358, 526-531 290 11-12; Vol 5, 25 393-398 224-226 78-79 379-385 143-146 204-205, 434 58-59, 190, 239-240 260-264 38-39

4, 143-146 3, 1 1, 234 1, 4, 1, 4, 1, 1, 4, 1, 5, 1, 1, 1,

198-204, 324-335 78-79 101-105 483 198-204 101-102 79 148-152, 324-335 134-136 153-159 153-154 151

1, 3, 4, 1, 5, 1, 1, 1, 3, 3, 3, 2, 3,

237-247 1; Vol 4, 30, 33 79-80 190-192 82-84 131-134 213-221 54 1, 130-134 57, 130-134 29-34 114-117 72-73

1, 260-264, 324-335 1, 324-335 1, 264 3, 78 1, 54 4, 5, 3, 3,

80 82-84 96 1-8

4, 361-362 3, 96

assembly internal ………………………………………………………………………Vol pumps …………………………………………………………………………Vol ATSM specifications …………………………………………………………………Vol axial spacing couplings ……………………………………………………………………Vol axial thrust damage ………………………………………………………………………Vol vertical pumps………………………………………………………………Vol backward running …………………………………………………………………Vol balancing boiler feed pumps …………………………………………………………Vol devices…………………………………………………………………………Vol impellers ……………………………………………………………………Vol pumps …………………………………………………………………………Vol requirements ………………………………………………………………Vol ball bearings inspection procedures ……………………………………………………Vol types …………………………………………………………………………Vol barrier systems mechanical seals …………………………………………………………Vol baseplate ANSI pump …………………………………………………………………Vol bouncing ……………………………………………………………………Vol design …………………………………………………………………………Vol installation, at BASF ……………………………………………………Vol preparation …………………………………………………………………Vol progressing cavity pumps ………………………………………………Vol rebuilding, at Avon ………………………………………………………Vol replacements ………………………………………………………………Vol specifying ……………………………………………………………………Vol surface preparation ………………………………………………………Vol bearing defects vibration………………………………………………………………………Vol bearing failures vibration………………………………………………………………………Vol bearing housing ANSI pump …………………………………………………………………Vol protection ……………………………………………………………………Vol sealing ………………………………………………………………………Vol bearing problems progressing cavity pumps ………………………………………………Vol bearing rating life ……………………………………………………………………Vol bearings anti-friction …………………………………………………………………Vol basics …………………………………………………………………………Vol canned motor pumps ……………………………………………………Vol condition ……………………………………………………………………Vol configurations ………………………………………………………………Vol extending life ………………………………………………………………Vol failure, sealless pumps……………………………………………………Vol failures ………………………………………………………………………Vol gear pumps …………………………………………………………………Vol grouting/installation ………………………………………………………Vol hydrofluoric acid service ………………………………………………Vol installation …………………………………………………………………Vol integral thrust design ……………………………………………………Vol leakage ………………………………………………………………………Vol liquid ring pumps …………………………………………………………Vol

1, 289 4, 185 4, 386-387 4, 146 1, 309 4, 236 5, 129 5, 1, 4, 1, 4,

131 309 325-328 24-25 432

4, 426 1, 265-266 3, 5 1, 4, 4, 4, 4, 2, 1, 4, 4, 4,

150 461 87-89, 166, 271-272 18-21 168-169; Vol 1, 336-341 180 213-221 438 363 440; Vol 1, 336-341

4, 307-308 4, 119-120 1, 149-150 1, 246-247 4, 509 2, 158 1, 270 1, 1, 1, 1, 1, 4, 1, 1, 2, 4, 1, 1, 5, 4, 4,

237-247 26-29; Vol 4, 38-41 172, 278 322 26 95-98, 548-552 81 28; Vol 4, 40-41 174 270 75-76 27 45-47 293 297

lubrication …………………………………………………………………Vol lubrication option …………………………………………………………Vol magnetic drive pumps ……………………………………………………Vol maintenance ………………………………………………………………Vol materials ……………………………………………………………………Vol modifications ………………………………………………………………Vol monitoring, sealless pumps ……………………………………………Vol motor, failures ………………………………………………………………Vol nomenclature ………………………………………………………………Vol preload ………………………………………………………………………Vol radial load …………………………………………………………………Vol reliability, centrifugal pumps …………………………………………Vol repairs…………………………………………………………………………Vol replacement …………………………………………………………………Vol seal life, effect on …………………………………………………………Vol sealless pumps ……………………………………………………………Vol sleeve …………………………………………………………………………Vol specifying ……………………………………………………………………Vol temperature factors, sealless pumps …………………………………Vol thrust damage………………………………………………………………Vol types …………………………………………………………………………Vol vibration analysis …………………………………………………………Vol water-lubricated, fluid-film ……………………………………………Vol beat vibration …………………………………………………………………………Vol bellows seal repair …………………………………………………………………………Vol belt vibration …………………………………………………………………………Vol best efficiency point (BEP) capacity characteristics …………………………………………………Vol off-BEP operation …………………………………………………………Vol point on curve ………………………………………………………………Vol bid evaluation…………………………………………………………………………Vol blower dry gas seals and …………………………………………………………Vol boiler feed pumps case study ……………………………………………………………………Vol replacement case study …………………………………………………Vol start-up ………………………………………………………………………Vol troubleshooting ……………………………………………………………Vol upgrading ……………………………………………………………………Vol boiler water circulation ……………………………………………………………Vol bowls vertical process pumps ……………………………………………………Vol buffer system API Plan 53 …………………………………………………………………Vol buffer tanks……………………………………………………………………………Vol bushings close clearance throat ……………………………………………………Vol materials ……………………………………………………………………Vol buying new pump ……………………………………………………………………Vol calculating MTBF …………………………………………………………………………Vol canned motor pumps ………………………………………………………………Vol applications …………………………………………………………………Vol basics …………………………………………………………………………Vol design for unique applications …………………………………………Vol design/performance evaluation…………………………………………Vol design/vertical motor under pumps …………………………………Vol hazardous fluids……………………………………………………………Vol

4, 4, 1, 1, 4, 1, 1, 5, 4, 1, 1, 1, 4, 1, 3, 1, 5, 4, 4, 1, 1, 1, 4, 4,

49-52, 532-537, 545-547, 548-552 90-94, 545-547, 548-552 161, 172, 274 29 187-188; Vol 1, 316 147; Vol 4, 357 147 103-105 97 27 166-167 265-270 181-182 313 14-15 315 128-129 363 93-94 309 238-239 316 244-247 308

3, 41-42 4, 308-309 1, 3, 1, 4,

141 22-24; Vol 4, 557-561 305 14

2, 134 4, 5, 4, 4, 1, 1,

298-304 130-133 278-281 42 51-53 229-230

4, 367-368 1, 264 4, 213 3, 101 4, 187-188 1, 107-110, 136-139 4, 1, 4, 1, 1, 1, 1, 1,

282-290; Vol 3, 130-134 106, 170-171, 324-335 206-208 278-281 263-264 222-230 234 260-264

hydrofluoric acid service ………………………………………………Vol NEC regulatory impact …………………………………………………Vol position monitoring ………………………………………………………Vol reliability ……………………………………………………………………Vol vertical in-line ………………………………………………………………Vol vs. magnetic drive …………………………………………………………Vol canned rotary pump ………………………………………………………………Vol capacity cavitation ……………………………………………………………………Vol head …………………………………………………………………………Vol pump, problems ……………………………………………………………Vol troubleshooting ……………………………………………………………Vol cartridge seals ………………………………………………………………………Vol high temperature …………………………………………………………Vol installation …………………………………………………………………Vol casing wear rings mounting, clearances ……………………………………………………Vol casings ANSI pump …………………………………………………………………Vol design, inspection, repair ………………………………………………Vol dual pressure ………………………………………………………………Vol screw pumps…………………………………………………………………Vol specifying ……………………………………………………………………Vol cavitation ………………………………………………………………………………Vol basics …………………………………………………………………………Vol bearing wear and …………………………………………………………Vol centrifugal pumps …………………………………………………………Vol control valves ………………………………………………………………Vol elements of minimum flow ……………………………………………Vol erosion ………………………………………………………………………Vol gear pumps …………………………………………………………………Vol hydraulic instabilities and ………………………………………………Vol liquid ring pumps …………………………………………………………Vol low suction pressure ………………………………………………………Vol noise …………………………………………………………………………Vol NPSH and……………………………………………………………………Vol power plant …………………………………………………………………Vol pressure ………………………………………………………………………Vol progressing cavity pump …………………………………………………Vol pulsation ……………………………………………………………………Vol pump failure ………………………………………………………………Vol reciprocating pumps ………………………………………………………Vol recirculation, high axial vibration ……………………………………Vol recirculation zones…………………………………………………………Vol troubleshooting ……………………………………………………………Vol turbulence……………………………………………………………………Vol vibration………………………………………………………………………Vol vs. entrainment ……………………………………………………………Vol centrifugal pumps ……………………………………………………………………Vol ANSI design and material options ……………………………………Vol ANSI process pumps, selecting …………………………………………Vol ANSI upgrades ……………………………………………………………Vol API 610 noise investigation ……………………………………………Vol balancing ……………………………………………………………………Vol bearing life …………………………………………………………………Vol bearing lubrication ………………………………………………………Vol bearing reliability …………………………………………………………Vol buying …………………………………………………………………………Vol canned motor pumps ……………………………………………………Vol capacity ………………………………………………………………………Vol

1, 1, 1, 1, 1, 1, 2,

75-78 70-74 322 106; Vol 4, 318 263-264 68-69 108-110

1, 1, 1, 5, 3, 3, 3,

308 306 42-43; Vol 4, 520 78 43-45 77 67

4, 99-101 1, 4, 1, 2, 4, 1, 5, 3, 1, 1, 1, 1, 2, 1, 4, 1, 1, 1, 4, 1, 2, 2, 1, 2, 4, 1, 5, 1, 1, 1, 1, 1, 1, 1, 5, 1, 4, 1, 1, 1, 1, 5,

149, 199 80-85 133 5-6 363 1, 233 106-108 15 6-9 312 15-16 120-121 172-173 118-124 296-297 320; Vol 4, 522 1 6-8; Vol 4, 6-9; Vol 5, 3-6 3-5 318 104 188-189 322 170-171; Vol 4, 28 118 14 77 7 1; Vol 4, 480 87 1 198-204 148-152 153-159 82-84 24-25 95 98-100 265-270 107-110 68-69, 106, 324-335 78

capacity, problems …………………………………………………………Vol 1, 42-43 cavitation ……………………………………………………………………Vol 1, 6-9, 118-124 cavitation and NPSH ……………………………………………………Vol 5, 3-6 chemical plant rerate case history ……………………………………Vol 5, 85-92 design/sealless ………………………………………………………………Vol 1, 222-230 design/vertical motor under pumps …………………………………Vol 1, 234 downhill………………………………………………………………………Vol 5, 50-51 efficiency ……………………………………………………………………Vol 1, 30-31, 39-41; Vol 5, 60-61 gasket repair ………………………………………………………………Vol 4, 175-179 head …………………………………………………………………………Vol 1, 1, 4; Vol 5, 16-17 horsepower …………………………………………………………………Vol 1, 308 hot fluids ……………………………………………………………………Vol 1, 190-192 hydraulic instabilities ……………………………………………………Vol 1, 118-124 hydrofluoric acid service ………………………………………………Vol 1, 75-78 impeller repair………………………………………………………………Vol 4, 175-179 impeller,volute design factors …………………………………………Vol 1, 139-141 impellers ……………………………………………………………………Vol 1, 139-141 installation and start-up …………………………………………………Vol 1, 46-49, 336-341 low flow options ……………………………………………………………Vol 1, 94-97 magnetic couplings ………………………………………………………Vol 1, 83-86 magnetic drive pumps ……………………………………………………Vol 1, 68-80, 103-104, 166-169, 170-174, 324-335 mechanical pumps…………………………………………………………Vol 3, 108 minimum flow………………………………………………………………Vol 1, 13-16, 34-36 motor size selection ………………………………………………………Vol 1, 32-33 multi-stage horizontal/repairs …………………………………………Vol 4, 182-185 multi-stage pump operations …………………………………………Vol 4, 263-269 multi-stage vertical pumps ……………………………………………Vol 4, 53-59 nomenclature ………………………………………………………………Vol 1, 1-3 operating systems …………………………………………………………Vol 1, 4-5 optimizing ……………………………………………………………………Vol 4, 232-235 oversizing ……………………………………………………………………Vol 1, 17-19 parallel ………………………………………………………………………Vol 5, 41-42 performance curve …………………………………………………………Vol 1, 4 performance in operating system ………………………………………Vol 2, 3-4 pressure monitoring ………………………………………………………Vol 1, 209-212 rating pumps ………………………………………………………………Vol 1, 131-135 rebuilding ……………………………………………………………………Vol 1, 213-221 revamp ………………………………………………………………………Vol 1, 184-189 sanitary ………………………………………………………………………Vol 4, 162 sealless ………………………………………………………………………Vol 1, 70-74 sealless, design features …………………………………………………Vol 1, 103-104, 146-147, 170-174, 324-335 sealless, failures ……………………………………………………………Vol 1, 81-82 selection ………………………………………………………………………Vol 1, 261-263, 324-335; Vol 4, 557-561 self-priming …………………………………………………………………Vol 1, 60-63, 205-208; Vol 5, 137-139 series pumping ……………………………………………………………Vol 5, 43-44 single-stage horizontal/repairs …………………………………………Vol 4, 180-181 slurry pump wear factors ………………………………………………Vol 4, 194-203 spare parts requirements ………………………………………………Vol 4, 77 speed and staging combinations ………………………………………Vol 1, 193-197 suction ………………………………………………………………………Vol 1, 231-233 suction lift……………………………………………………………………Vol 1, 2 suction specific speed ……………………………………………………Vol 1, 8-9 switching services …………………………………………………………Vol 1, 42 system considerations ……………………………………………………Vol 1, 205-208 system curve …………………………………………………………………Vol 1, 4 temperature monitoring …………………………………………………Vol 1, 209-212 testing …………………………………………………………………………Vol 1, 64-67 troubleshooting ……………………………………………………………Vol 5, 77-79 turbulence……………………………………………………………………Vol 1, 7

variable speed ………………………………………………………………Vol venting ………………………………………………………………………Vol vertical pumps/repairs ……………………………………………………Vol vertical pumps/service changes …………………………………………Vol vertical turbine ……………………………………………………………Vol viscosity effects ……………………………………………………………Vol volutes ………………………………………………………………………Vol vs. twin screw pumps ……………………………………………………Vol wear rings ……………………………………………………………………Vol charge pumps 13-stage ………………………………………………………………………Vol energy savings ………………………………………………………………Vol chemical feed pumps positive displacement pumps …………………………………………Vol chemical injection……………………………………………………………………Vol chemical processing corrosion and abrasion …………………………………………………Vol FCCU bottoms applications ……………………………………………Vol maintenance strategies …………………………………………………Vol multi-stage pumps …………………………………………………………Vol pump selection ……………………………………………………………Vol upgrading ……………………………………………………………………Vol chemical pumps corrosion and abrasion …………………………………………………Vol selection ………………………………………………………………………Vol chemicals classifying ……………………………………………………………………Vol hazardous ……………………………………………………………………Vol injection ………………………………………………………………………Vol chopper pumps solids-handling ……………………………………………………………Vol circulating water pumps boiler …………………………………………………………………………Vol classification coupling ………………………………………………………………………Vol viscosity ………………………………………………………………………Vol clearances high temperatures …………………………………………………………Vol wear ring ……………………………………………………………………Vol CMMS …………………………………………………………………………………Vol developing ……………………………………………………………………Vol coatings …………………………………………………………………………………Vol aluminum frames …………………………………………………………Vol material ………………………………………………………………………Vol complex systems sizing pumps for ……………………………………………………………Vol compressed air air operated diaphragm pumps ………………………………………Vol computer-based reliability …………………………………………………………Vol condition monitoring bearings ………………………………………………………………………Vol machinery ……………………………………………………………………Vol consolidation pump upgrades and ………………………………………………………Vol seal upgrades ………………………………………………………………Vol construction boiler feed pumps …………………………………………………………Vol impacts on pumps …………………………………………………………Vol magnetic drive pumps ……………………………………………………Vol

1, 1, 4, 4, 1, 1, 1, 2, 4,

55-59; Vol 5, 13-15 44-45 186-189 236-240 175-179 20-23 139-141 205 99-101

4, 68-75 4, 312-313 2, 31-34, 213-218 2, 96-99; Vol 1, 299 1, 1, 4, 4, 1, 4,

111-114 142-145 76-86 263-269 101-105 68-75

1, 112-113 1, 150-152, 324-335 4, 344-346 1, 295; Vol 4, 444 1, 299 1, 180-182 1, 54 4, 431 4, 33 1, 4, 4, 5, 4, 1, 4,

308 99-101 164-175, 334-336 112-116 220-223 289 391-392

5, 71-74 2, 37 4, 164-175 1, 268-269 4, 471 1, 154 1, 156 4, 278-281 4, 403-411 1, 271-276

containment canned motor and magnetic drive pumps …………………………Vol double, in sealless pumps ………………………………………………Vol containment shell canned motor pumps ……………………………………………………Vol magnetic drive pumps ……………………………………………………Vol control valves cavitation ……………………………………………………………………Vol high speed/low flow ………………………………………………………Vol system head curves ………………………………………………………Vol vs. VFD ………………………………………………………………………Vol controllers fire pump systems …………………………………………………………Vol controls automatic ……………………………………………………………………Vol flush piping …………………………………………………………………Vol motor …………………………………………………………………………Vol operational checks …………………………………………………………Vol safety …………………………………………………………………………Vol visible parts on pump and in station …………………………………Vol conversion chart (pumping units) ………………………………………………Vol coolers lubrication …………………………………………………………………Vol seal ……………………………………………………………………………Vol cooling canned motor and magnetic drive pumps …………………………Vol requirements, sealless pumps …………………………………………Vol cooling tower bearing failures and ………………………………………………………Vol noise …………………………………………………………………………Vol corrosion combating ……………………………………………………………………Vol erosion/corrosion phenomenon …………………………………………Vol gear pumps …………………………………………………………………Vol injection valve ………………………………………………………………Vol materials to combat ………………………………………………………Vol thermal spray coatings……………………………………………………Vol corrosive fluids peristaltic pumps …………………………………………………………Vol positive displacement pumps …………………………………………Vol thermoplastic pumps ……………………………………………………Vol cost analysis …………………………………………………………………………Vol costs energy …………………………………………………………………………Vol higher heads or flows ……………………………………………………Vol maintenance ………………………………………………………………Vol metering………………………………………………………………………Vol couplings alignment ……………………………………………………………………Vol balancing requirements …………………………………………………Vol characteristics ………………………………………………………………Vol classifications ………………………………………………………………Vol design …………………………………………………………………………Vol dry ……………………………………………………………………………Vol magnetic ……………………………………………………………………Vol modification …………………………………………………………………Vol MTBR and …………………………………………………………………Vol piping problems ……………………………………………………………Vol spacer coupling alignment ……………………………………………Vol specifying ……………………………………………………………………Vol

1, 171, 322, 324-335 1, 146 1, 278 1, 272-274 1, 1, 1, 4,

312 130 306 452

1, 253 1, 3, 4, 1, 1, 1, 1,

292 99 135-137 291 291 298 3

4, 32 3, 101 1, 171 1, 146, 315 5, 103-105 5, 117-119 1, 4, 2, 1, 1, 4,

111-114 83 174 300 111 220-223

2, 2, 1, 4,

184 19-21 282-287 174

1, 1, 4, 1,

307 312 60-62; Vol 1, 320 303

4, 4, 1, 4, 1, 4, 1, 1, 4, 5, 4, 4,

147-150, 164-166; Vol 5, 155-158 432 83 431 86 457 83-86 51 358 109 143-145 363

strategies ……………………………………………………………………Vol tolerances ……………………………………………………………………Vol torsional analysis …………………………………………………………Vol vertical pumps………………………………………………………………Vol CPI pumping …………………………………………………………………………Vol cryogenic pumping dry gas seals and …………………………………………………………Vol custom fluids peristaltic pumps and ……………………………………………………Vol cylinder pressure reciprocating pumps ………………………………………………………Vol dampeners ……………………………………………………………………………Vol pulsation ……………………………………………………………………Vol data failure, developing …………………………………………………………Vol published ……………………………………………………………………Vol test results……………………………………………………………………Vol defects bearings ………………………………………………………………………Vol deflection shaft, calculating …………………………………………………………Vol demand side management ………………………………………………………Vol design backpressure valves ………………………………………………………Vol boiler feed pumps …………………………………………………………Vol changes ………………………………………………………………………Vol discharge system ……………………………………………………………Vol enclosure ……………………………………………………………………Vol how to examine a system ………………………………………………Vol long bore ……………………………………………………………………Vol magnetic drive pumps ……………………………………………………Vol mechanical seals …………………………………………………………Vol metering applications ……………………………………………………Vol metering pumps ……………………………………………………………Vol pressure relief valve ………………………………………………………Vol progressing cavity pumps ………………………………………………Vol pump system ………………………………………………………………Vol pump, selection ……………………………………………………………Vol pumps for hydrocarbon service…………………………………………Vol sealless centrifugal pumps ……………………………………………Vol seals……………………………………………………………………………Vol short bore ……………………………………………………………………Vol slurry pump …………………………………………………………………Vol vertical pumps………………………………………………………………Vol dial indicator alignment ……………………………………………………………………Vol measurement system………………………………………………………Vol diaphragm pumps……………………………………………………………………Vol air operated …………………………………………………………………Vol high pressure ………………………………………………………………Vol metal diaphragm designs ………………………………………………Vol metering applications ……………………………………………………Vol sealing ………………………………………………………………………Vol disassembly pumps …………………………………………………………………………Vol discharge cavitation ……………………………………………………………………Vol internal assembly …………………………………………………………Vol pulsation control …………………………………………………………Vol pulsation dampeners ……………………………………………………Vol

4, 4, 4, 4, 1,

127-131 146-147 43-46 56-57 101-105

2, 134-135 2, 67-70 2, 1-2 2, 28-29 2, 190-193; Vol 1, 301 5, 112-116 1, 307 1, 318 1, 243 5, 67-68 4, 311 1, 4, 4, 2, 1, 1, 2, 1, 3, 2, 2, 1, 2, 5, 5, 4, 1, 3, 2, 4, 1,

300 300-301 6-7 201 290 310 210 160, 273-274 14-15 97, 111-114 87-88 300 11-14 29-36 53 212-213 222-226; Vol 4, 320 20 210 194-195 285

4, 4, 2, 2, 2, 2, 2, 2,

380-381 466 18 35-38, 42 130 93 61-62, 86-89, 90-95, 100-102 142

4, 188-189 5, 1, 2, 1,

106 289 189-190 301

pump curve …………………………………………………………………Vol pump station installation ………………………………………………Vol discharge piping ………………………………………………………………………Vol discharge pressure gear pumps …………………………………………………………………Vol low, rotary pumps …………………………………………………………Vol double and tandem seals …………………………………………………………Vol double end face seals ………………………………………………………………Vol double seals hazardous fluids……………………………………………………………Vol downhill pumping ……………………………………………………………………Vol drive problems progressing cavity pumps ………………………………………………Vol drives variable frequency …………………………………………………………Vol variable speed ………………………………………………………………Vol evaluating ……………………………………………………………………Vol driver system vertical pumps………………………………………………………………Vol drivers fire pump systems …………………………………………………………Vol lubrication …………………………………………………………………Vol ranges in power ……………………………………………………………Vol dry gas seals …………………………………………………………………………Vol dry products …………………………………………………………………………Vol dry running ……………………………………………………………………………Vol high frequency tracking …………………………………………………Vol magnetic drive pumps ……………………………………………………Vol progressing cavity pumps ………………………………………………Vol pump failure ………………………………………………………………Vol dry seals hazardous fluids and ……………………………………………………Vol dual gas seals …………………………………………………………………………Vol dual pressure casings ………………………………………………………………Vol dual seals ………………………………………………………………………………Vol eccentricity rotor …………………………………………………………………………Vol eddy current losses magnetic drive………………………………………………………………Vol efficiency boiler feed pumps …………………………………………………………Vol calculation, pump …………………………………………………………Vol canned motor ………………………………………………………………Vol centrifugal pumps …………………………………………………………Vol energy …………………………………………………………………………Vol losses …………………………………………………………………………Vol magnetic drive pumps ……………………………………………………Vol motor …………………………………………………………………………Vol pump selection ……………………………………………………………Vol screw pumps…………………………………………………………………Vol volumetric ……………………………………………………………………Vol elastomers mechanical seals …………………………………………………………Vol progressing cavity pumps ………………………………………………Vol electric motors ………………………………………………………………………Vol electric submersible pumps mine dewatering ……………………………………………………………Vol electrical-induced vibration ………………………………………………………Vol emission control regulations …………………………………………………………………Vol

1, 304 1, 288 1, 206-207 2, 2, 3, 2,

173 83-84 50-54 143

1, 261-263 5, 50-51 2, 158-160 4, 448 4, 132-134; Vol 5, 13-15 4, 191-193 4, 54-55 1, 4, 2, 2, 5, 4, 1, 1, 2, 4,

252-253 227-229 199 116-117; Vol 3, 79-82; Vol 4, 342 128 319-320; Vol 5, 128 320 168 156, 179 524

2, 2, 1, 2,

133-136 117; Vol 3, 81, 94 133-134 115-116; Vol 3, 9-10, 50-54, 74-75

1, 24 1, 167 4, 1, 1, 1, 1, 1, 1, 1, 5, 2, 2,

302-303 31 172 30-31, 39-41 193-197 30 172 30-31 60-61 43-47 207

3, 116 2, 181 4, 219 1, 258; Vol 2, 139; Vol 4, 339 4, 309-310 3, 51, 90-96

energy consumption …………………………………………………………………Vol energy costs screw pumps…………………………………………………………………Vol energy savings reliability improvements and …………………………………………Vol engineered pumps ……………………………………………………………………Vol engineered systems mechanical seals …………………………………………………………Vol entrainment suction side problems ……………………………………………………Vol environmental considerations variable speed pumping …………………………………………………Vol environmental controls ……………………………………………………………Vol erosion cavitation ……………………………………………………………………Vol evaluation bids ……………………………………………………………………………Vol expansion joints ………………………………………………………………………Vol piping problems ……………………………………………………………Vol face and rim alignment ……………………………………………………………………Vol failure analysis mechanical seals …………………………………………………………Vol sealless pumps ……………………………………………………………Vol failure data developing ……………………………………………………………………Vol FCCU Bottoms ………………………………………………………………………Vol feed pumps cavitation (case history) …………………………………………………Vol cost analysis …………………………………………………………………Vol hidden dangers ……………………………………………………………Vol thermal plant reliability (case history) ………………………………Vol troubleshooting ……………………………………………………………Vol upgrading techniques ……………………………………………………Vol filters ……………………………………………………………………………………Vol fire pump systems ……………………………………………………………………Vol flexible tubing pumps ………………………………………………………………Vol flow friction loss …………………………………………………………………Vol minimum ……………………………………………………………………Vol none, troubleshooting ……………………………………………………Vol pulsating ……………………………………………………………………Vol restriction devices …………………………………………………………Vol shutoff ………………………………………………………………………Vol system resistance …………………………………………………………Vol turbulent ……………………………………………………………………Vol flow characteristics metering applications ……………………………………………………Vol flow control metering applications ……………………………………………………Vol motors and …………………………………………………………………Vol flow sensing ……………………………………………………………………………Vol flowback ………………………………………………………………………………Vol fluid characteristics motors and …………………………………………………………………Vol fluid properties metering applications ……………………………………………………Vol reduction of seal life ………………………………………………………Vol fluid-film bearings……………………………………………………………………Vol

1, 193-197 2, 46 4, 311-314 3, 19 3, 19-21 1, 87-90 1, 58-59 3, 21 1, 120-121 4, 14 2, 9-14 5, 109 4, 381-382 3, 70-71, 114 1, 81-82 5, 112-116 1, 142-145 1, 4, 5, 4, 4, 1, 4, 1, 2,

121 60-62 93-94 3-5 42 50-54 252 251-255 67-70

1, 5, 5, 1, 1, 1, 1, 1,

306 18-21 48-49 301 95-96 308 304 307

2, 111-112, 213-218 2, 1, 4, 5,

146-149, 213-218 33 154-156 129

1, 33 2, 22-24 4, 458 4, 244-247

fluids basics …………………………………………………………………………Vol 4, 85 classifying ……………………………………………………………………Vol 4, 344-346 shear-sensitive ………………………………………………………………Vol 2, 12-13 special cases …………………………………………………………………Vol 2, 24-25, 219-223 viscous fluids ………………………………………………………………Vol 2, 13 flush plans seals……………………………………………………………………………Vol 3, 1-8 API ……………………………………………………………………………Vol 3, 102 flushing packing ………………………………………………………………………Vol 4, 292-293; Vol 3, 124-129 food processing chopper pumps ……………………………………………………………Vol 1, 182 foundations ……………………………………………………………………………Vol 4, 109; Vol 1, 336-341 baseplate design ……………………………………………………………Vol 4, 87-89 building ………………………………………………………………………Vol 4, 151-153 centrifugal pumps …………………………………………………………Vol 1, 46-49 grouting ………………………………………………………………………Vol 4, 47 installation …………………………………………………………………Vol 4, 358 rotary pump start-ups ……………………………………………………Vol 2, 78-79 standardized installation…………………………………………………Vol 4, 270 friction losses …………………………………………………………………………Vol 1, 10; Vol 5, 8, 24, 32-33, 96 fully lined slurry pumps ……………………………………………………………Vol 4, 421-424 gas barrier seals ……………………………………………………………………Vol 3, 35-40 gas purge seal…………………………………………………………………………Vol 3, 73-74 gas/liquid containment ……………………………………………………………Vol 3, 92-94 gaskets compression …………………………………………………………………Vol 4, 182 leaks …………………………………………………………………………Vol 5, 127 mechanical seal basics……………………………………………………Vol 3, 114 process pump ………………………………………………………………Vol 4, 177 replacements ………………………………………………………………Vol 4, 505 gases multiphase pumping ………………………………………………………Vol 2, 219-223 trapped, venting ……………………………………………………………Vol 1, 44-45 gear drives rotary pumps ………………………………………………………………Vol 2, 75-76 gear pumps additive pumps, as…………………………………………………………Vol 2, 33 do’s and don’ts ……………………………………………………………Vol 2, 16-17 extending the life of ………………………………………………………Vol 2, 172-175 maintaining and operating ……………………………………………Vol 2, 64-66 progressing cavity pumps and …………………………………………Vol 2, 122 user’s guide …………………………………………………………………Vol 2, 118-123 good engineering practices…………………………………………………………Vol 1, 153 GPA mechanical seal ………………………………………………………………Vol 3, 74 grease bearing lubrication ………………………………………………………Vol 1, 244; Vol 4, 92-93 grout ……………………………………………………………………………………Vol 4, 47-48; Vol 1, 335-341 guidelines ……………………………………………………………………Vol 1, 46 installation …………………………………………………………………Vol 4, 396 standardized practices and ……………………………………………Vol 4, 168, 270-273 standards ……………………………………………………………………Vol 4, 168 vs. stilt mounting …………………………………………………………Vol 4, 87 haz/rad wastewater service ………………………………………………………Vol 4, 138-142 hazardous fluids API applications ……………………………………………………………Vol 1, 260-264 dry seals………………………………………………………………………Vol 2, 133-136 high energy pumps ………………………………………………………Vol 1, 37-38; Vol 5, 16-17

optimizing energy …………………………………………………………Vol 4, 232-235 sealing ………………………………………………………………………Vol 4, 341-343 head centrifugal pumps …………………………………………………………Vol circulation ……………………………………………………………………Vol differential …………………………………………………………………Vol discharge ……………………………………………………………………Vol gear pumps …………………………………………………………………Vol maximum, estimating ……………………………………………………Vol NPSHR for reciprocating pumps ………………………………………Vol peristaltic pumps …………………………………………………………Vol static …………………………………………………………………………Vol user’s guide …………………………………………………………………Vol heat exchangers………………………………………………………………………Vol heat generation face temperature and flush rates………………………………………Vol magnetic drive pumps ……………………………………………………Vol heat transfer services ………………………………………………………………Vol hidden dangers inspecting for ………………………………………………………………Vol high axial vibration …………………………………………………………………Vol high efficiency motors ………………………………………………………………Vol high energy pumps optimizing ……………………………………………………………………Vol high pressure applications canned motor vs. magnetic drive ………………………………………Vol high pressure pumps reciprocating…………………………………………………………………Vol high speed/low flow top 10 issues…………………………………………………………………Vol high temperature sealing …………………………………………………………Vol high temperature service canned motor vs. magnetic drive ………………………………………Vol high temperature slurry applications best practices ………………………………………………………………Vol horizontal pump bearings ………………………………………………………………………Vol horsepower availability ……………………………………………………………Vol hot fluids pumping………………………………………………………………………Vol housing seals …………………………………………………………………………Vol housings bearings ………………………………………………………………………Vol H-Q curve/H-Q range operating system ……………………………………………………………Vol performance vs. system curve …………………………………………Vol hydraulic capabilities sealless centrifugal pumps ……………………………………………Vol hydraulic diaphragm designs metering applications ……………………………………………………Vol hydraulic fit evaluating ……………………………………………………………………Vol hydraulic instabilities cavitation and ………………………………………………………………Vol hydraulic losses magnetic drive pumps ……………………………………………………Vol hydraulic pressure reactions in pump piping systems ……………………………………Vol hydraulics modifications ………………………………………………………………Vol

1, 3, 1, 1, 2, 5, 5, 2, 5, 2, 4,

1-4, 37-38 104 304 304 16 16-17 22-23 68-69 96 118-123 253-254

3, 108 1, 161 1, 190-192 5, 93-94 4, 118 4, 121-126 4, 232-235 1, 68 2, 129-132 1, 125-130 3, 76-78 1, 68 4, 421-424 1, 242 4, 218-219 1, 190-192 1, 267-268 4, 49 1, 4-5 5, 64-66 1, 222-223 2, 88, 90-95 1, 96-97 1, 118-124 1, 167 5, 109-111 4, 357

power plant …………………………………………………………………Vol specifying ……………………………………………………………………Vol vibration, effects ……………………………………………………………Vol hydrocarbon service …………………………………………………………………Vol FCCU bottoms application ……………………………………………Vol multi-stage vertical pumps ……………………………………………Vol nitrogen quench ……………………………………………………………Vol hydrofluoric acid service……………………………………………………………Vol impeller life evaluation ……………………………………………………………………Vol impeller wear rings …………………………………………………………………Vol impellers ………………………………………………………………………………Vol ANSI pump …………………………………………………………………Vol attachment …………………………………………………………………Vol balancing ……………………………………………………………………Vol gas bound ……………………………………………………………………Vol inlet/outlet recirculation …………………………………………………Vol major servicing ……………………………………………………………Vol modifications ………………………………………………………………Vol pump, trimming ……………………………………………………………Vol vertical lineshaft turbines ………………………………………………Vol vertical process pumps ……………………………………………………Vol inertia head ……………………………………………………………………………Vol injectable packing ……………………………………………………………………Vol inlet piping configuration and motor efficiency ……………………………………Vol inlet pressure rotary pumps ………………………………………………………………Vol inspection API pumps …………………………………………………………………Vol bearings ………………………………………………………………………Vol boiler feed pumps …………………………………………………………Vol MTBR and …………………………………………………………………Vol shaft …………………………………………………………………………Vol visual …………………………………………………………………………Vol installation baseplate design ……………………………………………………………Vol bearings ………………………………………………………………………Vol centrifugal pumps …………………………………………………………Vol equipment ……………………………………………………………………Vol foundations …………………………………………………………………Vol foundations, building ……………………………………………………Vol foundations, dimensions ………………………………………………Vol grout …………………………………………………………………………Vol high pressure charge pump ……………………………………………Vol magnetic drive pumps ……………………………………………………Vol mechanical seals …………………………………………………………Vol multi-stage pumps …………………………………………………………Vol packing ………………………………………………………………………Vol piping …………………………………………………………………………Vol pump …………………………………………………………………………Vol pump/motor/baseplate ……………………………………………………Vol reciprocating pumps ………………………………………………………Vol rotary pumps ………………………………………………………………Vol troubleshooting ……………………………………………………………Vol vertical pumps………………………………………………………………Vol instrumentation ………………………………………………………………………Vol fire pump systems …………………………………………………………Vol reciprocating pumps ………………………………………………………Vol screw pumps…………………………………………………………………Vol sealless centrifugal pumps ……………………………………………Vol

4, 4, 4, 4, 1, 4, 3, 1,

3-5 363, 557-561 27-28 211-216 142-145 53-59 106 75-78

4, 4, 1, 1, 4, 4, 1, 1, 1, 4, 1, 4, 4, 5, 4,

8 180 139-141 148-149 178 325-328 87-88 118-124 298 6-9 39-41 331 237, 367-368 39-40 260-262, 293

5, 1-2 2, 124-126 4, 1, 5, 4, 4, 1,

177-178 266 93-94 356-359 189 290

4, 1, 1, 1, 4, 4, 4, 4, 4, 1, 3, 4, 4, 4, 4, 4, 2, 2, 1, 1, 3, 1, 2, 2, 1,

87-89 266-267; Vol 4, 97 46-49, 336-341 288 358; Vol 1, 336-341 150-152 109; Vol 1, 336-341 47, 271, 396; Vol 1, 336-341 71-75 104-105 66-69 266 291-292; Vol 3, 124-129 167, 393-394 18-19 396 165-171 77-81 46-47 175-179; Vol 4, 53-54, 239 19 253 53 50-51 225-226

intake noise modeling ……………………………………………………………………Vol 5, 100-102 integral thrust bearings vertical pumps………………………………………………………………Vol 5, 45-47 internal feed system magnetic drive pumps ……………………………………………………Vol 1, 160-161 internal gear pumps…………………………………………………………………Vol 2, 17 inventory ………………………………………………………………………………Vol 4, 210 managing ……………………………………………………………………Vol 1, 156-157 isolation valves ………………………………………………………………………Vol 4, 255 jacketing canned motor ………………………………………………………………Vol 1, 171, 173 magnetic drive pumps ……………………………………………………Vol 1, 171, 173 journal bearings ……………………………………………………………………Vol 4, 246 key indicators reliability ……………………………………………………………………Vol 4, 217-219 labyrinth seals ………………………………………………………………………Vol 1, 246-247 vs. lip seal ……………………………………………………………………Vol 1, 99-100 laser alignment ………………………………………………………………………Vol 4, 295, 323-324, 383-384 leakage air………………………………………………………………………………Vol 1, 208 bearings ………………………………………………………………………Vol 4, 293 dry gas seals…………………………………………………………………Vol 3, 80-81 flush rate ……………………………………………………………………Vol 3, 99 managing, seals ……………………………………………………………Vol 3, 3 visible …………………………………………………………………………Vol 3, 115 leaks ……………………………………………………………………………………Vol 5, 127 life cycle cost magnetic drive pumps ……………………………………………………Vol 1, 79 line shaft pumps mine dewatering ……………………………………………………………Vol 1, 258; Vol 2, 139; Vol 4, 339 lined pumps slurries ………………………………………………………………………Vol 1, 142-145 lip seals …………………………………………………………………………………Vol 1, 246; Vol 4, 50 liquid characteristics ………………………………………………………………Vol 1, 232-233 liquid ring pumps troubleshooting ……………………………………………………………Vol 4, 296-297 liquid seal designs low emission…………………………………………………………………Vol 3, 91 lobe pumps ……………………………………………………………………………Vol 2, 17 progressing cavity pumps and …………………………………………Vol 2, 123 lock-up start-up ………………………………………………………………………Vol 5, 128 low emission seals……………………………………………………………………Vol 2, 114-115 low flow. boiler feed pumps …………………………………………………………Vol 4, 42 options ………………………………………………………………………Vol 1, 94-97 precision solutions …………………………………………………………Vol 1, 299 pump designs ………………………………………………………………Vol 1, 299 recirculation …………………………………………………………………Vol 1, 309 rotary pumps ………………………………………………………………Vol 2, 83 low flow/high head options ………………………………………………………………………Vol 1, 234-236 low flow/high speed pumping top 10 issues…………………………………………………………………Vol 1, 125-130 lube oil piping…………………………………………………………………………Vol 4, 255 lube oil pumps ………………………………………………………………………Vol 4, 250-252 lubrication ……………………………………………………………………………Vol 5, 127; Vol 4, 526-531, 532-537 barrier fluid for reservoir ………………………………………………Vol 3, 105 bearings ………………………………………………………………………Vol 1, 269-270; Vol 4, 49-52. 90-84, 97, 545-547, 548-552

effects of pump design ……………………………………………………Vol gear pumps …………………………………………………………………Vol mechanical seals …………………………………………………………Vol oil ………………………………………………………………………………Vol packing ………………………………………………………………………Vol progressing cavity pumps ………………………………………………Vol pumps and drivers …………………………………………………………Vol reciprocating pumps ………………………………………………………Vol rotary pump start-ups ……………………………………………………Vol systems ………………………………………………………………………Vol viscosity classifications ……………………………………………………Vol lubrication paths canned motor pumps ……………………………………………………Vol magnetic drive pumps ……………………………………………………Vol lubrication systems bearings ………………………………………………………………………Vol machinery alignment ………………………………………………………………Vol magnetic coupling……………………………………………………………………Vol magnetic drive pumps ……………………………………………………Vol magnetic drive pumps ………………………………………………………………Vol basics …………………………………………………………………………Vol couplings ……………………………………………………………………Vol failures, interpreting ………………………………………………………Vol hazardous fluids……………………………………………………………Vol hazardous fluids……………………………………………………………Vol installation …………………………………………………………………Vol NEC impact …………………………………………………………………Vol non-metallic …………………………………………………………………Vol operation protection ………………………………………………………Vol options ………………………………………………………………………Vol performance evaluation …………………………………………………Vol reliability improvements …………………………………………………Vol selecting a pump …………………………………………………………Vol temperature monitoring …………………………………………………Vol vs. canned motor …………………………………………………………Vol when to apply ………………………………………………………………Vol magnetic liquid seals ………………………………………………………………Vol VOC containment …………………………………………………………Vol magnetic seals bearing housing ……………………………………………………………Vol maintenance access covers…………………………………………………………………Vol computer-based (Tennessee Eastman) ………………………………Vol cost ……………………………………………………………………………Vol key indicators ………………………………………………………………Vol low flow metering application …………………………………………Vol outsourcing …………………………………………………………………Vol predictive programs ………………………………………………………Vol program (Champion International) …………………………………Vol pump rebuilding (Avon) …………………………………………………Vol record keeping ………………………………………………………………Vol regular inspection …………………………………………………………Vol reliability-driven ……………………………………………………………Vol stand pipe ……………………………………………………………………Vol strategies (Thorn Creek Sanitation) …………………………………Vol team approach ……………………………………………………………Vol maintenance staff ……………………………………………………………………Vol makedown pumps positive displacement pumps …………………………………………Vol

1, 2, 1, 4, 4, 2, 4, 2, 2, 4, 4,

98-100 174 38; Vol 3, 2-3, 14; Vol 5, 17 248-259, 545-547, 548-552 291 179 227-229 171 79 29-33 33

1, 279-280 1, 274-275 1, 4, 1, 1, 1, 1, 1, 1, 1, 1, 1, 1, 4, 1, 1, 1, 1, 1, 1, 1, 1, 3, 3,

244-246; Vol 4, 545-547, 548-552 149-150 83-86 271-272 170, 324-335 271-276 83-86 81-82 260-264 261-263 104-105 70-74 319 166-169 170-174 222-230 103-105, 277 160-165, 324-335 209-212 68-69, 324-335 79-80 74 16-18

1, 247 1, 4, 4, 4, 2, 4, 4, 4, 1, 4, 1, 4, 2, 4, 4, 4,

289 163-174 60-62 217-219, 538-544 198 553-556 112-115, 538-544 156-158 213-221 63-67 295 76-86, 526-531, 538-544 204 209-210 12-21 10-11

2, 32

management fluid sealing …………………………………………………………………Vol materials abrasives ……………………………………………………………………Vol basics …………………………………………………………………………Vol bearings, bushings …………………………………………………………Vol characteristics ………………………………………………………………Vol construction and …………………………………………………………Vol corrosion ……………………………………………………………………Vol failure, reciprocating pumps ……………………………………………Vol heat transfer ………………………………………………………………Vol mechanical seals …………………………………………………………Vol metals …………………………………………………………………………Vol non-metallics ………………………………………………………………Vol reliability and ………………………………………………………………Vol seal faces ……………………………………………………………………Vol sealless pumps ……………………………………………………………Vol seals……………………………………………………………………………Vol selection ………………………………………………………………………Vol specifying ……………………………………………………………………Vol thermoplastics ………………………………………………………………Vol thermosets……………………………………………………………………Vol maximum allowable working pressure (MAWP) …………………………Vol maximum head estimating ……………………………………………………………………Vol mean time between failure data ……………………………………………………………………………Vol mechanical barrier seals canned motor pumps ……………………………………………………Vol mechanical losses screw pumps…………………………………………………………………Vol mechanical seals ……………………………………………………………………Vol accessories, API 682 ………………………………………………………Vol air and abrasives, protection …………………………………………Vol ANSI pump …………………………………………………………………Vol API plans 11, 12, 13, 21, 22, 31, 32, 41, 62 ………………………Vol API plan 52 …………………………………………………………………Vol API plan 53 …………………………………………………………………Vol API plan 54 …………………………………………………………………Vol API 682 reference guide …………………………………………………Vol barrier systems ……………………………………………………………Vol bearings and ………………………………………………………………Vol bellows ………………………………………………………………………Vol cartridge seals ………………………………………………………………Vol chamber design ……………………………………………………………Vol chemical injection designs ………………………………………………Vol damage, hidden ……………………………………………………………Vol design, API 682 ……………………………………………………………Vol double end face ……………………………………………………………Vol dry gas ………………………………………………………………………Vol dual arrangements ………………………………………………………Vol elastomers ……………………………………………………………………Vol environment, controlling ………………………………………………Vol extending the life of ………………………………………………………Vol failure analysis ……………………………………………………………Vol gas barrier …………………………………………………………………Vol heat generation ……………………………………………………………Vol high speed/low flow ………………………………………………………Vol high temperature …………………………………………………………Vol hp losses ……………………………………………………………………Vol

3, 86-90 2, 4, 4, 1, 4, 1, 2, 3, 3, 1, 1, 4, 3, 4, 3, 1, 4, 1, 1, 1,

19-21 85-86 187-188, 246 201 404-405 111-114 170 109 1-2, 25-28 201 203 386-392 46-49 320 84, 120-123 202-204, 285; Vol 5, 52 363 201-202, 282-287 201 131-135

1, 37-38 5, 115-116 1, 263-264 2, 4, 3, 3, 1, 3, 3, 3, 3, 3, 3, 3, 3, 3, 3, 2, 1, 3, 2, 3, 3, 3, 3, 3, 3, 3, 3, 1, 3, 1,

45 219 32-33 62-65 200 4 5 6 7 29-34 5 14-15 41-42 43-45 10, 58-61 96-99 45 29-30 143 79-82 50-54 116 1-8; Vol 4, 353-355 11-13 70-71 35-40 108 127 76-78 32

inspection and testing ……………………………………………………Vol installation …………………………………………………………………Vol instrumentation, API 682 ………………………………………………Vol low emission…………………………………………………………………Vol lubrication …………………………………………………………………Vol magnetic liquid seals ……………………………………………………Vol management ………………………………………………………………Vol materials of construction ………………………………………………Vol materials, API 682 ………………………………………………………Vol new developments …………………………………………………………Vol non-contacting………………………………………………………………Vol off-design operation ………………………………………………………Vol reliability case study ………………………………………………………Vol seal face application ………………………………………………………Vol shaft sealing…………………………………………………………………Vol single end face………………………………………………………………Vol slurries ………………………………………………………………………Vol steam turbines ……………………………………………………………Vol testing in adverse situations ……………………………………………Vol totally engineered sealing system………………………………………Vol troubleshooting ……………………………………………………………Vol upgrades ……………………………………………………………………Vol VOC control …………………………………………………………………Vol zero-leak ……………………………………………………………………Vol mechanically actuated diaphragm pumps metering applications ……………………………………………………Vol metal bellows seals high temperature applications …………………………………………Vol metallic magnetic pumps …………………………………………………………Vol metals …………………………………………………………………………………Vol metering applications chemical metering …………………………………………………………Vol design considerations ……………………………………………………Vol diaphragm pumps and……………………………………………………Vol flow control …………………………………………………………………Vol options ………………………………………………………………………Vol peristaltic pumps and ……………………………………………………Vol positive displacement pumps and ……………………………………Vol pulsation control …………………………………………………………Vol metering pumps low flow pump designs……………………………………………………Vol reciprocating…………………………………………………………………Vol selection guide………………………………………………………………Vol updates ………………………………………………………………………Vol mine dewatering ……………………………………………………………………Vol minimum flow ………………………………………………………………………Vol boiler feed pumps …………………………………………………………Vol centrifugal pumps …………………………………………………………Vol magnetic drive pumps ……………………………………………………Vol prevent low flow recirculation …………………………………………Vol recirculation …………………………………………………………………Vol vertical pumps………………………………………………………………Vol vibration………………………………………………………………………Vol misalignment …………………………………………………………………………Vol vibration and ………………………………………………………………Vol mixers …………………………………………………………………………………Vol low emission seals for ……………………………………………………Vol mixing pumps positive displacement pumps …………………………………………Vol

3, 3, 3, 3, 1, 3, 3, 3, 3, 3, 3, 3, 3, 3, 3, 2, 1, 3, 3, 3, 5, 1, 2, 3,

33 66-69 33 91 38; Vol 3, 2-3; Vol 5, 17 16-18 86-90 9, 25-28, 32, 46-49, 120-123 32 9-10 39-40 22-25 55-57, 130-134 46-49 83 143 142-145 95 25-28 19-21 128 157 114-117 72-75

2, 88, 90-95 3, 76-78 4, 319 1, 201 2, 2, 2, 2, 2, 2, 2, 2,

33-34, 96-99 111-114 61-62 146-149 23-26, 213-218 184 18, 213-218 28, 60-63, 161-164, 213-218

1, 2, 2, 2, 1, 5, 5, 1, 1, 1, 1, 4, 1, 4, 4, 2, 3,

299 90-95 86-87, 213-218 194 256-259 18-21 131 13-16, 34-36 166-169 309 34 238 35 118-119 306 26 95-96

2, 32-33

mixing tanks/blenders controlling surge……………………………………………………………Vol modifications bearings ………………………………………………………………………Vol controls ………………………………………………………………………Vol pump hydraulics……………………………………………………………Vol pumps …………………………………………………………………………Vol seals……………………………………………………………………………Vol turbines ………………………………………………………………………Vol monitoring alignment ……………………………………………………………………Vol magnetic drive pumps ……………………………………………………Vol multi-stage pumps …………………………………………………………Vol pumps …………………………………………………………………………Vol rotor position ………………………………………………………………Vol sealless pumps ……………………………………………………………Vol vertical lineshaft turbines ………………………………………………Vol motors analysis ………………………………………………………………………Vol bearings ………………………………………………………………………Vol controls ………………………………………………………………………Vol current analysis ……………………………………………………………Vol efficiency ……………………………………………………………………Vol evaluating ……………………………………………………………………Vol high-efficiency ………………………………………………………………Vol lubrication …………………………………………………………………Vol oil filled and non-oil filled ……………………………………………Vol optimizing ……………………………………………………………………Vol overload problems …………………………………………………………Vol performance curve limitations …………………………………………Vol sealless pumps ……………………………………………………………Vol selection ………………………………………………………………………Vol selection, size ………………………………………………………………Vol variable speed ………………………………………………………………Vol vertical process pumps ……………………………………………………Vol mounting guidelines ……………………………………………………………………Vol screw pump options ………………………………………………………Vol sensor …………………………………………………………………………Vol MTBF……………………………………………………………………………………Vol calculating …………………………………………………………………Vol strategies ……………………………………………………………………Vol multi-phase operations ……………………………………………………………Vol multiple screw pumps extending the life of ………………………………………………………Vol multi-stage pumps chemical processing ………………………………………………………Vol estimating head ……………………………………………………………Vol head …………………………………………………………………………Vol hydrocarbon service ………………………………………………………Vol improvements (13-stage pump) …………………………………………Vol synchronous vibration ……………………………………………………Vol National Electric Code sealless centrifugal pumps ……………………………………………Vol National Fire Protection Association …………………………………………Vol Newtonian fluids ……………………………………………………………………Vol no flow shutoff ………………………………………………………………………Vol troubleshooting ……………………………………………………………Vol

5, 123 4, 5, 4, 5, 4, 5,

357 85 357 85 357 85

4, 1, 4, 4, 1, 4, 4,

399-402 162, 166-169 268 110-111, 213 322 321; Vol 1, 315 330

4, 5, 4, 4, 5, 4, 4, 4, 1, 4, 5, 1, 4, 1, 1, 1, 4,

415 103-105 135-137 115 1-2 191-193 121-126 32 298 371-378 79 32 319 57 32-33 55-59; Vol 4, 132-134; Vol 5, 13-15 368

1, 2, 1, 4, 4, 4, 4,

47 6 320 217-218; Vol 3, 130-134 282-290 356-359 211-216; Vol 2, 219-223

2, 176-178 4, 1, 5, 4, 4, 4,

263-269 37-38 16-17 4, 211-216 68-75 117-118

1, 70-74 1, 254 1, 21 1, 308 5, 48-49

noise API pumps …………………………………………………………………Vol cooling tower ………………………………………………………………Vol intake, modeling……………………………………………………………Vol pulsation and ………………………………………………………………Vol reducing ………………………………………………………………………Vol rotary pumps ………………………………………………………………Vol nomenclature bearings ………………………………………………………………………Vol centrifugal pumps …………………………………………………………Vol detecting problems…………………………………………………………Vol symbols ………………………………………………………………………Vol non-API seals zero-leak ……………………………………………………………………Vol non-contacting seals design …………………………………………………………………………Vol non-metallic, lined magnetic drive pumps ……………………………………Vol non-metallics …………………………………………………………………………Vol non-Newtonian fluids ………………………………………………………………Vol nozzle loading…………………………………………………………………………Vol NPSH ……………………………………………………………………………………Vol boiler feed pumps …………………………………………………………Vol cavitation ……………………………………………………………………Vol centrifugal pumps …………………………………………………………Vol changes in pump flow ……………………………………………………Vol high speed/low flow ………………………………………………………Vol how much is enough? ……………………………………………………Vol low, correcting ………………………………………………………………Vol minimum flow………………………………………………………………Vol no flow ………………………………………………………………………Vol priming ………………………………………………………………………Vol pump selection ……………………………………………………………Vol reciprocating pumps ………………………………………………………Vol required, reciprocating PD pumps ……………………………………Vol suction conditions …………………………………………………………Vol sufficient for suction ………………………………………………………Vol testing …………………………………………………………………………Vol unit revamp strategies ……………………………………………………Vol verifying ………………………………………………………………………Vol viscous fluid …………………………………………………………………Vol NPSHA suction lifts …………………………………………………………………Vol NPSHR progressing cavity pumps and …………………………………………Vol OEM training …………………………………………………………………………Vol off-BEP operation boiler feed pumps …………………………………………………………Vol off-design operation mechanical seal performance …………………………………………Vol sealless pumps ……………………………………………………………Vol offshore oil platforms reciprocating pumps on …………………………………………………Vol oil analysis ……………………………………………………………………………Vol oil bath lubrication …………………………………………………………………Vol oil flooded bearing lubrication……………………………………………………Vol oil lubrication …………………………………………………………………………Vol oil mist lubrication bearing lubrication ………………………………………………………Vol oil ring lubrication …………………………………………………………………Vol

5, 5, 5, 2, 4, 2,

82-84 117-119 100-102 188 22-24 84

4, 1, 4, 1,

97-98 1-3 524 309

3, 73-74 3, 4, 1, 1, 1, 1, 5, 1, 1, 1, 1, 5, 5, 1, 5, 5, 5, 2, 5, 1, 5, 1, 1, 5, 5,

39-40 319 203 21 91-93, 199 1, 232 130-133 6-8; Vol 4, 6-9; Vol 5, 3-6 6-8, 43 308 127-128 11-12 95-97 14; Vol 5, 19-20 48-49 7-10 54; Vol 4, 557-561 1-2 22-23 10 24 65 184-189 106 9-10

1, 207-208 2, 103-107 1, 157-158 5, 131 3, 22-25 4, 320 2, 4, 4, 1, 4,

165-171 415 91 244 248-259, 532-537, 545-547, 548-552

1, 245; Vol 4, 51, 92 4, 91

operation protection magnetic drive pumps ……………………………………………………Vol overhaul rotary pumps, guide ………………………………………………………Vol shop ……………………………………………………………………………Vol vertical process pumps ……………………………………………………Vol overload motor, troubleshooting ………………………………………………Vol oversizing effects of ……………………………………………………………………Vol packed plunger pumps metering applications ……………………………………………………Vol packing …………………………………………………………………………………Vol injectable ……………………………………………………………………Vol positive displacement pumps and ……………………………………Vol progressing cavity pumps ………………………………………………Vol paper mills chemical additive pumps for …………………………………………Vol parallel pumping ……………………………………………………………………Vol centrifugal pumps …………………………………………………………Vol variable speed ………………………………………………………………Vol partial emission pumps ……………………………………………………………Vol performance characteristics, screw pumps …………………………………………Vol degradation, troubleshooting …………………………………………Vol flush rates ……………………………………………………………………Vol heat transfer ………………………………………………………………Vol improvement, new pump ………………………………………………Vol sealless centrifugal pumps ……………………………………………Vol performance curve motor sizing …………………………………………………………………Vol operating point ……………………………………………………………Vol pump oversizing factor ……………………………………………………Vol revamp strategies and ……………………………………………………Vol system selection ……………………………………………………………Vol vs. system curve ……………………………………………………………Vol periodic maintenance ………………………………………………………………Vol peristaltic pumps ……………………………………………………………………Vol custom fluids and …………………………………………………………Vol pulsation control …………………………………………………………Vol pharmaceutical pumps ……………………………………………………………Vol PIP standards…………………………………………………………………………Vol pipeline pressure, metering applications ………………………………………Vol pumps, screw ………………………………………………………………Vol valving ………………………………………………………………………Vol piping additive system design ……………………………………………………Vol alignment ……………………………………………………………………Vol design …………………………………………………………………………Vol discharge ……………………………………………………………………Vol external pipe stress ………………………………………………………Vol fittings…………………………………………………………………………Vol flush …………………………………………………………………………Vol hydraulic pressure …………………………………………………………Vol inlet, configuration ………………………………………………………Vol installation …………………………………………………………………Vol lubrication systems ………………………………………………………Vol progressing cavity pumps ………………………………………………Vol reducing friction losses……………………………………………………Vol replacements ………………………………………………………………Vol

1, 166-169 2, 4, 4, 5,

127-128 77 367-370 79

1, 17-19 2, 4, 4, 2, 2,

88 292-293, Vol 3, 124-129 260-262 143 179

2, 1, 1, 5, 1,

31-34 188; Vol 5, 41-42 248-250 69-70 94-95

2, 5, 3, 3, 1, 1,

6-7 55-57 110 109 138 222-226

1, 1, 1, 1, 5, 5, 4, 2, 2, 2, 4, 1,

32-33 308 17-19 184-189 62-63 64-66 397-398 17-18, 182-186 67-70 28 161 151

2, 162-163 2, 43-47, 48-51 2, 48-51 2, 1, 1, 1, 4, 4, 3, 5, 5, 1, 4, 2, 5, 4,

31-32 11-12; Vol 5, 25 92; Vol 4, 167 206-207 73 104 99 109-111 1-2 47 32 180 32-33 447

rotary pumps and …………………………………………………………Vol size ……………………………………………………………………………Vol slurry pump …………………………………………………………………Vol standards ……………………………………………………………………Vol strain …………………………………………………………………………Vol suction ………………………………………………………………………Vol suction, layout………………………………………………………………Vol supports ………………………………………………………………………Vol system design ………………………………………………………………Vol temperature …………………………………………………………………Vol piping losses priming and …………………………………………………………………Vol piping-to-pump alignment …………………………………………………………Vol piston pumps high pressure ………………………………………………………………Vol pulsation control …………………………………………………………Vol reciprocating pumps ………………………………………………………Vol sealing ………………………………………………………………………Vol pitot tube pumps low flow/high head ………………………………………………………Vol plant turnaround ……………………………………………………………………Vol plastic pumps …………………………………………………………………………Vol plunger pumps high pressure ………………………………………………………………Vol positive displacement pumps abrasives and corrosives …………………………………………………Vol built-in relief valves ………………………………………………………Vol canned rotary pumps ……………………………………………………Vol chemical additive pumps ………………………………………………Vol chemical injection applications ………………………………………Vol gear pumps …………………………………………………………………Vol inlet pressure requirements ……………………………………………Vol makedown pumps …………………………………………………………Vol mixing pumps ………………………………………………………………Vol multiple screw ………………………………………………………………Vol NPSH …………………………………………………………………………Vol operating principles ………………………………………………………Vol peristaltic design……………………………………………………………Vol pipeline screw pump efficiency …………………………………………Vol progressing cavity pump reliability ……………………………………Vol progressing cavity, selecting ……………………………………………Vol pulsation control …………………………………………………………Vol reciprocating pumps ………………………………………………………Vol rotary gear pumps …………………………………………………………Vol rotary pumps ………………………………………………………………Vol screw …………………………………………………………………………Vol sealing ………………………………………………………………………Vol speed control ………………………………………………………………Vol types, review …………………………………………………………………Vol unloading pumps …………………………………………………………Vol vibration………………………………………………………………………Vol positive sealing elements positive displacement pumps and ……………………………………Vol power consumption couplings and ………………………………………………………………Vol loss, bearings and …………………………………………………………Vol pump reliability ……………………………………………………………Vol when oversizing ……………………………………………………………Vol power plants troubleshooting, boiler feed ……………………………………………Vol

2, 4, 4, 4, 4, 1, 5, 1, 5, 3,

77-78 103, 394-395 194 167 102-103 207, 231-232; Vol 4, 102-104 59 93 37-38 103

5, 8 4, 393-398 2, 2, 2, 2,

129 27-28 206 141-142

1, 234-235 4, 276-277 1, 114 2 129-130 2, 2, 2, 2, 2, 2, 2, 2, 2, 2, 5, 2, 2, 2, 4, 2, 2, 2, 2, 2, 2, 2, 2, 2, 2, 2,

19-21 100-102 108-110 31-34, 213-218 96-99 172-175 124 32 32-33 176-178 22-23 3-4, 213-218 67-70 43-47 241-243 11-14 27-30 1-4 64-65 118-123 5-8, 219-223 141-145 150-152 15-18 32 39-41

2, 144-145 4, 4, 4, 1,

127-128 247 3-5 17-19

4, 42

upgrading, utility …………………………………………………………Vol usage …………………………………………………………………………Vol precision tools…………………………………………………………………………Vol predictive maintenance ……………………………………………………………Vol building a program ………………………………………………………Vol pressure close clearance throat bushings ………………………………………Vol control valves, API 610 …………………………………………………Vol cylinder, reciprocating pumps …………………………………………Vol discharge ……………………………………………………………………Vol gauges, screw pumps………………………………………………………Vol maximum allowable ………………………………………………………Vol mechanical seals …………………………………………………………Vol pipeline, metering applications …………………………………………Vol relief valves, API 610 ……………………………………………………Vol seal coolers …………………………………………………………………Vol series pumping ……………………………………………………………Vol stuffing box …………………………………………………………………Vol switches ………………………………………………………………………Vol preventive maintenance ……………………………………………………………Vol priming centrifugal pumps …………………………………………………………Vol suction life and NPSH ……………………………………………………Vol troubleshooting ……………………………………………………………Vol proactive maintenance ……………………………………………………………Vol probes installation tips ……………………………………………………………Vol process control ………………………………………………………………………Vol procurement …………………………………………………………………………Vol benchmarking ………………………………………………………………Vol engineered systems ………………………………………………………Vol maintenance’s role ………………………………………………………Vol progressing cavity pumps …………………………………………………………Vol additive pumps, as…………………………………………………………Vol design basics ………………………………………………………………Vol extending the life of ………………………………………………………Vol gear pumps and ……………………………………………………………Vol lobe pumps and ……………………………………………………………Vol maximizing performance ………………………………………………Vol NPSHR ………………………………………………………………………Vol progressing cavity pumps and …………………………………………Vol reliability ……………………………………………………………………Vol selecting ………………………………………………………………………Vol three screw pumps and …………………………………………………Vol vane pumps and ……………………………………………………………Vol pulp and paper ………………………………………………………………………Vol shaft sealing…………………………………………………………………Vol pulsation ………………………………………………………………………………Vol controlling……………………………………………………………………Vol dampeners …………………………………………………………………Vol metering pumps ……………………………………………………………Vol minimizing, metering applications ……………………………………Vol positive displacement pumps and ……………………………………Vol reciprocating pumps ………………………………………………………Vol waterhammer ………………………………………………………………Vol pump materials of construction ………………………………………………Vol monitoring …………………………………………………………………Vol rerate problems ……………………………………………………………Vol selection ………………………………………………………………………Vol

1, 2, 4, 4, 4,

50-54 84 78 112-115, 358-359 412-420

3, 4, 2, 1, 2, 1, 3, 2, 4, 3, 5, 3, 4, 4,

101 254-255 1-2 132 49-50; Vol 1, 301 131-135 58-59 162-163 254; Vol 1, 300 101 43-44 98 32 358-359, 526-531, 532-537

1, 5, 5, 1,

60-63 7-10 27-28 215-216

4, 4, 4, 4, 3, 4, 2, 2, 2, 2, 2, 2, 2, 2, 2, 4, 2, 2, 2, 2, 3, 2, 2, 2, 2, 2, 2, 2, 5,

35-37 214 14 115 19 361-366 17-18, 123 33 11 179-181 122 123 153-160 103-107 123 241-243 11-14 122 123 31-34 83-85 28-29 27-30; Vol 5, 120-125 190-193; Vol 1, 301 60-63 161-164 39-41 187-193 80-81

1, 4, 5, 5,

114, 198-204; Vol 4, 84 110-111 85-92 52-53, 62-63, 324-335

selection (efficiency-based) ……………………………………………Vol selection, suction pressure ………………………………………………Vol pump controls flow sensing …………………………………………………………………Vol pump curve centrifugal……………………………………………………………………Vol conflicting ……………………………………………………………………Vol consists of ……………………………………………………………………Vol performance vs. system …………………………………………………Vol rotary pumps ………………………………………………………………Vol troubleshooting with ………………………………………………………Vol pump discharge heads vertical process pumps ……………………………………………………Vol pump dynamics mechanical seals and ……………………………………………………Vol pump noise API 610 noise investigation ……………………………………………Vol modeling intake noise ……………………………………………………Vol reducing ………………………………………………………………………Vol pump record keeping ………………………………………………………………Vol pump reliability systems approach …………………………………………………………Vol pump sizing ……………………………………………………………………………Vol pump speed valves and ……………………………………………………………………Vol pumping terms ………………………………………………………………………Vol pumps centrifugal sealless ………………………………………………………Vol handling ……………………………………………………………………Vol low flow designs ……………………………………………………………Vol lubrication …………………………………………………………………Vol major servicing ……………………………………………………………Vol sizing for complex systems ……………………………………………Vol specifications ………………………………………………………………Vol start up and testing ………………………………………………………Vol purchasing communicating needs ……………………………………………………Vol specifications ………………………………………………………………Vol strategies ……………………………………………………………………Vol pure mist bearing lubrication ………………………………………………………Vol purge mist bearing lubrication ………………………………………………………Vol radial bearings ………………………………………………………………………Vol radial load bearings and shaft…………………………………………………………Vol radially split pumps …………………………………………………………………Vol rating centrifugal pumps …………………………………………………………Vol vs. required …………………………………………………………………Vol rebowling vertical lineshaft turbines ………………………………………………Vol rebuilding centrifugal pumps …………………………………………………………Vol reciprocating pumps…………………………………………………………………Vol cavitation ……………………………………………………………………Vol high pressure ………………………………………………………………Vol metering applications ……………………………………………………Vol NPSH …………………………………………………………………………Vol performance in operating system ………………………………………Vol

5, 60-61 5, 54 4, 153-155 1, 1, 1, 5, 2, 5,

4; Vol 2, 3 129 304 64-66 118-119 134-136

4, 369 3, 14-15 5, 5, 4, 4,

82-84 100-102 22-24 63-67

4, 1-2 4, 171-172 2, 58 1, 3 1, 1, 1, 4, 1, 5, 4, 1,

315, 324-335 288 299 227-229 298 71-74 363-366 291

1, 136-138 1, 216-217; Vol 4, 14 1, 107-110 1, 245-246 1, 246 4, 39 1, 166-167 4, 176-177 1, 131-135 1, 133 4, 331 1, 2, 4, 2, 2, 2, 2,

213-221 18 28 129-132 90-95 1-2; Vol 5, 22-23 3-4

pulsation control …………………………………………………………Vol selection, installation, start-up and operation ……………………Vol valve dynamics and reliability …………………………………………Vol vibration………………………………………………………………………Vol recirculation calculating …………………………………………………………………Vol cavitation and ………………………………………………………………Vol minimum flow………………………………………………………………Vol seal coolers …………………………………………………………………Vol vibration………………………………………………………………………Vol record-keeping maintenance ………………………………………………………………Vol MTBR and …………………………………………………………………Vol reliability key indicators ………………………………………………………………Vol materials ……………………………………………………………………Vol systems approach …………………………………………………………Vol thermal power plants ……………………………………………………Vol reliability improvements ANSI pumps…………………………………………………………………Vol magnetic drives ……………………………………………………………Vol reliability profile Chevron ………………………………………………………………………Vol reliability programs …………………………………………………………………Vol strategies ……………………………………………………………………Vol reliability tips magnetic drive pumps ……………………………………………………Vol reliability-driven maintenance …………………………………………………Vol relief valves ……………………………………………………………………………Vol built-in ………………………………………………………………………Vol repairs proper pump …………………………………………………………………………Vol repeller pump dry gas seals and …………………………………………………………Vol rerates energy savings and…………………………………………………………Vol troubleshooting problems ………………………………………………Vol reservoirs cleaning ………………………………………………………………………Vol lubrication …………………………………………………………………Vol residual unbalance …………………………………………………………………Vol resonance vibration and ………………………………………………………………Vol retrofit API process pumps ………………………………………………………Vol sealless pumps ……………………………………………………………Vol utility and process pumps ………………………………………………Vol revamp centrifugal pumps …………………………………………………………Vol reverse indicator alignment ……………………………………………………………………Vol rigid spacer couplings vertical process pumps ……………………………………………………Vol rim and face alignment procedures ……………………………………………………Vol ring oiled bearing lubrication ………………………………………………………Vol root cause analysis …………………………………………………………………Vol root valves ……………………………………………………………………………Vol

2, 2, 2, 2,

187-193 165-171 52-59 39-41

1, 1, 1, 3, 4,

15 118-124 15; Vol 5, 20-21 101 25-26, 118

4, 63-67, 538-544 4, 357 4, 4, 4, 4,

217-219, 557-561; Vol 3, 130-134 386-392 1-2, 526-531 3-5

1, 101-102 1, 103 4, 157-159 4, 334-336, 360 4, 347-352, 526-531, 538-544 1, 4, 4, 2,

277 76-86 254 100-102

4, 176-190 2, 134 4, 313 5, 85-92 3, 105 4, 30 1, 24 4, 306 1, 54 1, 322; Vol 4, 524 1, 50-55 1, 184-189 4, 295, 382-383 4, 368-369 4, 224-226, 294-295 1, 244-245 4, 76; Vol 4, 499 4, 255

rotary gear pumps maintaining and operating ……………………………………………Vol selection guide………………………………………………………………Vol rotary pumps …………………………………………………………………………Vol abrasives ……………………………………………………………………Vol built-in relief valves ………………………………………………………Vol canned ………………………………………………………………………Vol gear pumps …………………………………………………………………Vol operating principles ………………………………………………………Vol peristaltic pumps …………………………………………………………Vol pipeline screw pump efficiency …………………………………………Vol progressing cavity pumps, selecting……………………………………Vol two screw design ……………………………………………………………Vol gear drive options …………………………………………………………Vol inlet pressure ………………………………………………………………Vol overhauling …………………………………………………………………Vol sanitary ………………………………………………………………………Vol start-ups………………………………………………………………………Vol troubleshooting ……………………………………………………………Vol user’s guide …………………………………………………………………Vol rotation reverse, progressing cavity pumps ……………………………………Vol rotary pump start-ups ……………………………………………………Vol shaft …………………………………………………………………………Vol rotor assembly, balancing ………………………………………………………Vol canned motor pumps ……………………………………………………Vol dynamics ……………………………………………………………………Vol flexible, balancing …………………………………………………………Vol magnetic drive pumps ……………………………………………………Vol position ………………………………………………………………………Vol rigid, balancing ……………………………………………………………Vol stacking ………………………………………………………………………Vol safety equipment ……………………………………………………………………Vol sanitary pumping ……………………………………………………………………Vol peristaltic pumps and ……………………………………………………Vol screw pumps canned rotary ………………………………………………………………Vol pipeline efficiency …………………………………………………………Vol two screw ……………………………………………………………………Vol valving ………………………………………………………………………Vol seal chamber ANSI pump …………………………………………………………………Vol design advances ……………………………………………………………Vol reliability ……………………………………………………………………Vol seal environment controlling……………………………………………………………………Vol seal faces application …………………………………………………………………Vol dry gas ………………………………………………………………………Vol materials ……………………………………………………………………Vol support ………………………………………………………………………Vol seal flush ………………………………………………………………………………Vol seal pot piping ………………………………………………………………Vol sealed pumps CPI applications ……………………………………………………………Vol sealing arrangements, high temperature ………………………………………Vol chemicals ……………………………………………………………………Vol hazardous fluids……………………………………………………………Vol

2, 2, 2, 2, 2, 2, 2, 2, 2, 2, 2, 2, 2, 2, 2, 4, 2, 2, 2,

64-66 71-74 15-16 19-21 100-102 108-110 32, 64-66 3-4 67-70 43-47 11-14 5-8, 219-223 75-76 124-126 127-128 163 77-81 81-85 118-123

2, 106 2, 78-79 3, 115 4, 1, 4, 1, 1, 1, 1, 4, 4, 4, 2,

184-185 278 3-5 25 273 315 25 183-184 78 160-163 185

2, 2, 2, 2,

108-110 43-47 5-8, 219-223 48-51

1, 149 3, 10 3, 58-61 4, 353-355; Vol 3, 118 3, 3, 3, 3, 4, 3,

46-49 79-82 9, 120-123 117 353-355; Vol 3, 98 68

1, 101 3, 77-78 4, 344-346 4, 341-343

problems, progressing cavity pumps …………………………………Vol totally engineered systems ………………………………………………Vol sealing technology dry seals………………………………………………………………………Vol low emission…………………………………………………………………Vol positive displacement pumps …………………………………………Vol VOC control …………………………………………………………………Vol sealless pumps applying canned motor …………………………………………………Vol canned motor vs. magnetic drive ………………………………………Vol containment barrier ………………………………………………………Vol continuous monitoring ……………………………………………………Vol CPI applications ……………………………………………………………Vol design and performance …………………………………………………Vol failures ………………………………………………………………………Vol hazardous fluids……………………………………………………………Vol hydrofluoric acid applications …………………………………………Vol interpreting failures ………………………………………………………Vol magnetic couplings ………………………………………………………Vol magnetic drive operation protection …………………………………Vol magnetic drive pumps ……………………………………………………Vol National Electric Code……………………………………………………Vol NEC impact …………………………………………………………………Vol options ………………………………………………………………………Vol reliability ……………………………………………………………………Vol selecting magnetic drive pumps ………………………………………Vol seals dry gas ………………………………………………………………………Vol dual ……………………………………………………………………………Vol gear pumps …………………………………………………………………Vol modifications ………………………………………………………………Vol secondary seals ………………………………………………………………………Vol selection additive pumps ……………………………………………………………Vol ANSI pump …………………………………………………………………Vol boiler feed pumps …………………………………………………………Vol centrifugal pumps …………………………………………………………Vol couplings ……………………………………………………………………Vol haz/rad wastewater service ……………………………………………Vol high pressure reciprocating pumps ……………………………………Vol materials ……………………………………………………………………Vol mechanical seals …………………………………………………………Vol motor …………………………………………………………………………Vol motors and drives …………………………………………………………Vol multi-stage pumps …………………………………………………………Vol progressing cavity pumps ………………………………………………Vol pump …………………………………………………………………………Vol pump (efficiency-based) …………………………………………………Vol pump, parameters …………………………………………………………Vol pump, suction pressure …………………………………………………Vol pump, variable speed applications ……………………………………Vol pumps …………………………………………………………………………Vol pumps for abrasive or corrosive service ………………………………Vol reciprocating pumps ………………………………………………………Vol rotary gear pumps …………………………………………………………Vol sanitary pumps ……………………………………………………………Vol screw pumps…………………………………………………………………Vol seals, sanitary pumps ……………………………………………………Vol self-priming pumps ………………………………………………………Vol slurry pump …………………………………………………………………Vol

2, 157-158 3, 19-21 2, 2, 2, 2,

133-136 114-115 141-145 114-117

1, 1, 1, 1, 1, 1, 1, 1, 1, 1, 1, 1, 1, 1, 1, 1, 4, 1,

68-69 68-69, 324-335 85-86 315 103-104 222-226 81-82 260-264 75-78 81-82 83-86 166-169 79-81 70-74 70-74 146-147, 170-174, 324-335 318-322 160-165

2, 2, 2, 4, 3,

116-117 115-117 174 357 67

2, 1, 4, 1, 4, 4, 2, 1, 3, 1, 4, 4, 2, 5, 5, 4, 5, 1, 1, 1, 2, 2, 4, 2, 4, 1, 4,

31, 213-218 148-152 299-300 184-189, 261-263, 324-335 130 138-142 129-132 202-204 22-23 32-33, 57 371-378 264-266 11-14; Vol 4, 241-242 52-52, 62-63; Vol 4, 557-561 60-61 215 54 56-57 160-165 114 166-168 71-74 160-161 6 161-162 63 196-197

thermoplastic pumps ……………………………………………………Vol variable frequency drive …………………………………………………Vol self-priming pumps …………………………………………………………………Vol troubleshooting ……………………………………………………………Vol sensing flow ……………………………………………………………………………Vol series pumping ………………………………………………………………………Vol shafts ……………………………………………………………………………………Vol alignment ……………………………………………………………………Vol alignment tools ……………………………………………………………Vol ANSI pump …………………………………………………………………Vol broken ………………………………………………………………………Vol canned motor pumps ……………………………………………………Vol deflection, calculating ……………………………………………………Vol inspection ……………………………………………………………………Vol machining ……………………………………………………………………Vol maintenance ………………………………………………………………Vol radial load …………………………………………………………………Vol runout ………………………………………………………………………Vol sealing, pulp and paper …………………………………………………Vol vertical process pumps ……………………………………………………Vol shaft to coupling spool method alignment ……………………………………………………………………Vol shear-sensitive fluids progressing cavity pumps and …………………………………………Vol shipping and handling guidelines ……………………………………………………………………Vol shutdown MTBR and …………………………………………………………………Vol pulsation and ………………………………………………………………Vol reciprocating pumps ………………………………………………………Vol side channel pumps low flow/high head ………………………………………………………Vol signature analysis vibration………………………………………………………………………Vol single end face seals…………………………………………………………………Vol single seals design …………………………………………………………………………Vol single-stage pumps head …………………………………………………………………………Vol size piping………………………………………………………………………………Vol sizing pumps …………………………………………………………………………Vol pumps for complex systems ……………………………………………Vol sleeve bearings ………………………………………………………………………Vol slip flow screw pumps…………………………………………………………………Vol sludge transfer ………………………………………………………………………Vol slurry sealing …………………………………………………………………………Vol slurry pumps …………………………………………………………………………Vol corrosion and abrasion …………………………………………………Vol fully lined ……………………………………………………………………Vol sealless ………………………………………………………………………Vol soft foot …………………………………………………………………………………Vol solids-handling canned motor and magnetic drive pumps …………………………Vol canned motor vs. magnetic drive ………………………………………Vol seal chamber and …………………………………………………………Vol spacer couplings alignment ……………………………………………………………………Vol

1, 1, 1, 5,

284-285 57 60-63, 205-208 137-139

4, 1, 1, 4, 4, 1, 5, 1, 5, 4, 4, 4, 1, 4, 3, 4,

154-156 188; Vol 5, 43-44 200 47 204-205 199 39-40 278 67-68 189 190 230-231 166-167 183 83-85 368

4, 383 2, 12-13 1, 46 4, 359 5, 123 2, 170 1, 235-236 4, 305-310 2, 143 3, 9 5, 16-17 4, 394-395 4, 171-172 5, 71-74 5, 128-129; Vol 2, 207 2, 1, 4, 4, 1, 1, 1, 4,

44, 219-223 181 333, 345-346 194-203 113 142-145 146 151

1, 171 1, 68 3, 59-60 4, 143-146

spare parts major servicing ……………………………………………………………Vol minimum requirements …………………………………………………Vol planning ……………………………………………………………………Vol specialty pumps reliability ……………………………………………………………………Vol specific gravity high, lifting …………………………………………………………………Vol specific speed efficiency and ………………………………………………………………Vol specifications air operated diaphragm pumps ………………………………………Vol baseplates ……………………………………………………………………Vol bearings ………………………………………………………………………Vol casings ………………………………………………………………………Vol centrifugal pumps …………………………………………………………Vol couplings ……………………………………………………………………Vol developing ……………………………………………………………………Vol hydraulics ……………………………………………………………………Vol materials ……………………………………………………………………Vol MTBR and …………………………………………………………………Vol pumps …………………………………………………………………………Vol pumps …………………………………………………………………………Vol pumps for hydrocarbon service…………………………………………Vol purchasing …………………………………………………………………Vol purchasing …………………………………………………………………Vol writing ………………………………………………………………………Vol speed and efficiency characteristics ……………………………………………Vol speed and staging combinations ………………………………………………Vol speed changes progressing cavity pumps ………………………………………………Vol speed control positive displacement pumps …………………………………………Vol staff maintenance, involvement ……………………………………………Vol standards ANSI/ASME B73 …………………………………………………………Vol nozzle loading ………………………………………………………………Vol PIP ……………………………………………………………………………Vol starters motor …………………………………………………………………………Vol start-up . boiler feed pumps …………………………………………………………Vol checklist ………………………………………………………………………Vol continuous processing metering ………………………………………Vol lined slurry pumps ………………………………………………………Vol MTBR…………………………………………………………………………Vol packing ………………………………………………………………………Vol procedures……………………………………………………………………Vol pulsation ……………………………………………………………………Vol reciprocating pumps ………………………………………………………Vol rotary pumps ………………………………………………………………Vol troubleshooting ……………………………………………………………Vol static suction head …………………………………………………………………Vol stators canned motor pumps ……………………………………………………Vol progressing cavity pumps ………………………………………………Vol steam quench …………………………………………………………………………Vol steam turbine dry gas seals and …………………………………………………………Vol seals for ………………………………………………………………………Vol

1, 298 4, 77 4, 410 1, 102-103 5, 8-9 1, 3, 193-197 2, 4, 4, 4, 1, 4, 1, 4, 4, 4, 3, 4, 4, 1, 4, 4, 1, 1,

42 363 363 363 184-189 363 108 363 363 356-359 20-21 363-366 211-212 216-217 14 403 30-31; Vol 5, 60-61 193-197

2, 180 2, 150-152 4, 10-11 1, 151 1, 91-93 1, 151 4, 135-137 4, 1, 2, 1, 4, 4, 1, 5, 2, 2, 1, 1,

278-281 47 24-25 145 359 292-293 46-49 123 165-171 77-81 46-47; Vol 5, 126-128 2

1, 278 2, 155-156; Vol 4, 242 3, 68 2, 134 3, 95

storage pump …………………………………………………………………………Vol straight edge method alignment ……………………………………………………………………Vol suction …………………………………………………………………………………Vol cavitation ……………………………………………………………………Vol centrifugal pumps …………………………………………………………Vol conditions ……………………………………………………………………Vol conditions, centrifugal ……………………………………………………Vol control, pulsation …………………………………………………………Vol dynamic lift …………………………………………………………………Vol entrainment problems ……………………………………………………Vol flooded ………………………………………………………………………Vol lift, priming and ……………………………………………………………Vol loss, rotary pumps …………………………………………………………Vol magnetic drive pumps ……………………………………………………Vol noise in, cooling tower ……………………………………………………Vol pressure, canned motor …………………………………………………Vol pressure induced cavitation ……………………………………………Vol pump selection and ………………………………………………………Vol vapor in ………………………………………………………………………Vol suction piping …………………………………………………………………………Vol centrifugal pump revamp ………………………………………………Vol layout …………………………………………………………………………Vol recommended configurations …………………………………………Vol rule of thumb ………………………………………………………………Vol suction flange ………………………………………………………………Vol suction specific speed boiler feed pumps …………………………………………………………Vol cavitation ……………………………………………………………………Vol centrifugal pumps …………………………………………………………Vol high speed/low flow ………………………………………………………Vol suction volume ………………………………………………………………………Vol sump pump thermoplastics ………………………………………………………………Vol surface pressure………………………………………………………………………Vol surge control, positive displacement pumps ………………………………Vol controlling……………………………………………………………………Vol pressure, calculating ………………………………………………………Vol suppressors, pulsation ……………………………………………………Vol synchronous vibration ………………………………………………………………Vol system design/analysis/selection …………………………………………………Vol chemical injection …………………………………………………………Vol control, metering applications …………………………………………Vol curve vs. performance curve ……………………………………………Vol curve, centrifugal …………………………………………………………Vol design …………………………………………………………………………Vol design, piping ………………………………………………………………Vol efficiency, defining …………………………………………………………Vol failure, torsional vibration and…………………………………………Vol head calculations …………………………………………………………Vol interaction……………………………………………………………………Vol metering system design …………………………………………………Vol no flow ………………………………………………………………………Vol parallel operation …………………………………………………………Vol parallel with variable speed drive ……………………………………Vol performance degradation ………………………………………………Vol pump reliability ……………………………………………………………Vol pumping downhill …………………………………………………………Vol

1, 46 4, 1, 5, 1, 5, 1, 2, 1, 1, 1, 5, 2, 1, 5, 1, 1, 5, 1, 1, 1, 5, 4, 5, 1,

294-295 2, 207; Vol 4, 102-104 106 231-233 24-26 10-12 189 2 87-90 44-45 7-10 83 172 119 172 320 54 11 231-232 184-189 59 102-104 59 10-11

5, 5, 1, 1, 1,

131 5-6 8-9 126 207

1, 282-287 5, 96 2, 5, 5, 5, 4, 2, 2, 2, 5, 1, 5, 5, 5, 5, 1, 1, 2, 5, 5, 5, 5, 4, 5,

27-30 120-125 81 125 117-118 3-4 96-99 113 64-66 4 29-36 37-38 60-61 98-99 1-3 184 111-113 48-49 41-42 69-70 55-57 1-2 50-51

pumping with air …………………………………………………………Vol pumps in series ……………………………………………………………Vol reliability approach ………………………………………………………Vol revamp strategies …………………………………………………………Vol self-priming pumps ………………………………………………………Vol sizing for complex systems ……………………………………………Vol system selection requirements …………………………………………Vol tandem seals …………………………………………………………………………Vol tank farm transfer pumps controlling surge……………………………………………………………Vol team approach maintenance ………………………………………………………………Vol temperature bearings ………………………………………………………………………Vol minimum flow………………………………………………………………Vol monitoring, magnetic drive pumps ……………………………………Vol testing centrifugal pumps …………………………………………………………Vol simulation……………………………………………………………………Vol start up ………………………………………………………………………Vol thermal movement spacer couplings ……………………………………………………………Vol thermography …………………………………………………………………………Vol thermoplastic pumps ………………………………………………………………Vol thermosets ……………………………………………………………………………Vol three screw pumps …………………………………………………………………Vol progressing cavity pumps and …………………………………………Vol thrust vertical pumps………………………………………………………………Vol thrust bearing…………………………………………………………………………Vol assemblies ……………………………………………………………………Vol integral, vertical pumps …………………………………………………Vol thrust loading …………………………………………………………………………Vol minimum flow………………………………………………………………Vol tie rod troubleshooting ……………………………………………………………Vol tools alignment ……………………………………………………………………Vol torque couplings ……………………………………………………………………Vol measuring ……………………………………………………………………Vol motor …………………………………………………………………………Vol torsional analysis couplings ……………………………………………………………………Vol torsional vibration …………………………………………………………………Vol training …………………………………………………………………………………Vol troubleshooting cavitation ……………………………………………………………………Vol centrifugal pumps …………………………………………………………Vol high pressure reciprocating pumps ……………………………………Vol installation …………………………………………………………………Vol no flow ………………………………………………………………………Vol performance degradation ………………………………………………Vol prime …………………………………………………………………………Vol reciprocating pumps ………………………………………………………Vol rotary pumps ………………………………………………………………Vol self-priming pumps ………………………………………………………Vol start-up ………………………………………………………………………Vol vertical lineshaft turbines ………………………………………………Vol vibration………………………………………………………………………Vol

4, 5, 4, 1, 1, 5, 5, 3,

75-76 43-44 1-2 184-189 205-208 71-74 52-54 50-54

5, 121-123 4, 12-21 4, 93-94, 247 1, 13-14; Vol 5, 18-19 1, 209-212 1, 64-67 1, 292 1, 291 4, 4, 1, 1, 2, 2,

145 115, 416 201-202, 282-287 201 176 122

4, 4, 3, 5, 5, 1,

236 245-246 25-28 45-47 19 14

5, 110 4, 204-205 4, 128 4, 37 4, 371 4, 43-46 5, 98-99 1, 157-158 5, 5, 2, 1, 5, 5, 5, 2, 2, 5, 1, 4, 5,

77 77-79 131-132 46-47 48-49 55-57 27-28 170 81-85 137-139 46-47 331 58-59

tubing selecting for peristaltic pumps …………………………………………Vol 2, 69-70 turbine rerate problems ……………………………………………………………Vol turbulence centrifugal pumps …………………………………………………………Vol turnaround plant, managing ……………………………………………………………Vol two screw pumps ……………………………………………………………………Vol ultrapure services thermoplastic pumps ……………………………………………………Vol ultrasonics ……………………………………………………………………………Vol unbalance vibration and ………………………………………………………………Vol Underwriters Laboratories ………………………………………………………Vol universal drives maintenance ………………………………………………………………Vol universal joints progressing cavity pumps ………………………………………………Vol unloading pumps ……………………………………………………………………Vol upgrades ANSI …………………………………………………………………………Vol mechanical seal ……………………………………………………………Vol pumps, at Chevron ………………………………………………………Vol utility and process pumps ………………………………………………Vol utility and process pumps upgrading ……………………………………………………………………Vol utility service …………………………………………………………………………Vol vacuum pumps troubleshooting ……………………………………………………………Vol variable speed drives………………………………………………………Vol valves ……………………………………………………………………………………Vol air operated diaphragm pumps ………………………………………Vol backpressure ………………………………………………………………Vol dynamics, reciprocating pumps ………………………………………Vol injection ………………………………………………………………………Vol material ………………………………………………………………………Vol multi-function ………………………………………………………………Vol performance …………………………………………………………………Vol pipeline screw pumps ……………………………………………………Vol pressure relief ………………………………………………………………Vol pulsation and ………………………………………………………………Vol pump speed and ……………………………………………………………Vol relief, built-in ………………………………………………………………Vol rotary pumps ………………………………………………………………Vol vibration, positive displacement pumps………………………………Vol vane pass vibration …………………………………………………………………Vol vane pumps ……………………………………………………………………………Vol progressing cavity pumps and …………………………………………Vol vapor in suction line ………………………………………………………………Vol removal ………………………………………………………………………Vol vapor pressure ………………………………………………………………………Vol fluid metering pumps ……………………………………………………Vol magnetic drive pumps ……………………………………………………Vol suction lift and ……………………………………………………………Vol variable speed actuators positive displacement pumps …………………………………………Vol variable speed controls ……………………………………………………………Vol

5, 85-92 1, 7 4, 276-277 2, 176-177, 219-223 1, 282-287 4, 416-417 4, 305 1, 254-255; Vol 1, 298 4, 230-231 2, 156-157 2, 32 1, 1, 4, 1,

153-159 157 315-317 50-55

1, 50-55 1, 227-229 4, 4, 4, 2, 1, 2, 1, 2, 1, 2, 2, 1, 5, 2, 2, 2, 2, 4, 2, 2,

297 274-275 254-256 42 300 52-59 300 57-58 301 40-41 48-51 300 124 58 100-102 77-78 39-41 117 17 123

1, 11; Vol 5, 25 3, 60 1,1; Vol 5, 96 2, 22-23 1, 163 5, 7-8 2, 150-151 4, 122

variable speed drives ………………………………………………………………Vol advantages and pitfalls …………………………………………………Vol energy savings and…………………………………………………………Vol inverters ………………………………………………………………………Vol overview ………………………………………………………………………Vol parallel pumping …………………………………………………………Vol positive displacement pumps …………………………………………Vol vacuum pumps ……………………………………………………………Vol vertical pumps………………………………………………………………Vol variable speed pumping ……………………………………………………………Vol environmental considerations …………………………………………Vol velocity head …………………………………………………………………………Vol venting centrifugal pumping ………………………………………………………Vol vertical lineshaft turbines revitalizing……………………………………………………………………Vol vertical motor-under pumps ……………………………………………………Vol vertical pumps bearings ………………………………………………………………………Vol design …………………………………………………………………………Vol hydrocarbon service ………………………………………………………Vol intakes ………………………………………………………………………Vol integral thrust bearings in ……………………………………………Vol managing ……………………………………………………………………Vol multi-stage …………………………………………………………………Vol overhauling …………………………………………………………………Vol petrochemical applications ……………………………………………Vol repair techniques …………………………………………………………Vol rerates…………………………………………………………………………Vol service changes ……………………………………………………………Vol troubleshooting ……………………………………………………………Vol vibration ………………………………………………………………………………Vol amplification ………………………………………………………………Vol analysis ………………………………………………………………………Vol couplings ……………………………………………………………………Vol flow phenomena……………………………………………………………Vol high axial ……………………………………………………………………Vol minimum flow………………………………………………………………Vol motor bearing failure studies …………………………………………Vol positive displacement pumps and ……………………………………Vol probe installation tips ……………………………………………………Vol programs ……………………………………………………………………Vol pulsation and ………………………………………………………………Vol pump, reducing ……………………………………………………………Vol radial …………………………………………………………………………Vol reciprocating pumps ………………………………………………………Vol rotary pumps ………………………………………………………………Vol signature analysis …………………………………………………………Vol start-up ………………………………………………………………………Vol studies…………………………………………………………………………Vol system approach ……………………………………………………………Vol torsional………………………………………………………………………Vol vane passing frequencies ………………………………………………Vol vertical lineshaft turbines ………………………………………………Vol vertical turbine pumps……………………………………………………Vol vibration-related pump problems top ten ………………………………………………………………………Vol viscosity classifications ………………………………………………………………Vol effects on centrifugal pumps …………………………………………Vol

4, 5, 4, 4, 1, 5, 2, 4, 4, 1, 1, 1,

374-375; Vol 5, 13-15 13-15 313 122 55-59 69-70 151-152 274-275 239 55-59 58-59 2-3

1, 44-45 4, 329-322 1, 227-229 1, 242 1, 234, 285 4, 53-55 1, 115-117 5, 45-47 4, 236-240 4, 53-59 4, 367-370 1, 175-179 4, 186-189 5, 85-92 4, 236-240 4, 187 4, 219, 526-531 5, 58-59 4, 112-113, 116-120, 413-415 4, 44, 131 1, 123 43, 118-119 1, 35 5, 103-105 2, 39-41 4, 54-55 4, 112-113, 349-350 2, 188 4, 25-28 1, 40 2, 171 2, 84 4, 305-310 5, 128 4, 131 4, 25-28 5, 98-99 1, 41 4, 330 1, 175-179 4, 26 4, 33 1, 20-23

fluid metering pumps ……………………………………………………Vol 2, 23-24, 213-218 NPSH considerations ……………………………………………………Vol 5, 9-10 rotary pumps ………………………………………………………………Vol 2, 118-123 specific gravity for common liquids……………………………………Vol 1, 20-23 viscous fluids lifting …………………………………………………………………………Vol 5, 9-10 progressing cavity pumps and …………………………………………Vol 2, 13 VOC control …………………………………………………………………………Vol 2, 114-117 containment …………………………………………………………………Vol 3, 16-18 volutes …………………………………………………………………………………Vol 1, 139-141 vortex breakers ………………………………………………………………………Vol 1, 89 vortexing ………………………………………………………………………………Vol 1, 207 cooling towers ………………………………………………………………Vol 5, 118-119 wastewater chopper pumps and ………………………………………………………Vol 1, 181 haz/rad ………………………………………………………………………Vol 4, 138-142 strategies ……………………………………………………………………Vol 4, 209-210 water pumps circulating ……………………………………………………………………Vol 1, 53-54 water transport pump reliability ……………………………………………………………………Vol 4, 6-9 waterhammer …………………………………………………………………………Vol 5, 80-81; Vol 1, 301 water-lubricated bearings …………………………………………………………Vol 4, 244-247 wear pump …………………………………………………………………………Vol 1, 33 rapid, rotary pumps ………………………………………………………Vol 2, 84-85 vertical lineshaft turbines ………………………………………………Vol 4, 329-330 wear particle analysis ………………………………………………………………Vol 4, 113-114 wear rings impeller ………………………………………………………………………Vol 4, 180 mechanical seals …………………………………………………………Vol 3, 14-15 mounting, clearances ……………………………………………………Vol 4, 99-101 wear technology slurry pump …………………………………………………………………Vol 4, 198-203 well pump mine dewatering ……………………………………………………………Vol 1, 256-259; Vol 2, 137-140; Vol 4, 337-340 zero-leak seals ………………………………………………………………………Vol 3, 72-75

Author Index - 2002 Able, Stephen D. … … … … Pump Suction Lift, NPSH and Priming ……………………………………Vol 5, 7-10 Abramovitz, Stanley … … … Water-Lubricated Fluid Film Bearings Can Be Trouble-Free ……………Vol 4, 244-247 Adams, William V. … … … Applying Dry Gas Sealing Technology to Pumps …………………………Vol 3, 79-82 Adams, William V. … … … Gas-Barrier Seals Establish Beachhead ……………………………………Vol 3, 35-39 Adams, William V … … … Off-Design Operation and Seal Performance………………………………Vol 3, 22-24 Addie, Graeme … … … … Slurry Pump Wear Factors …………………………………………………Vol 4, 194-203 Aliasso, Joe … … … … … Troubleshooting Liquid Ring Pumps ………………………………………Vol 4, 296-297 Anderson, Larry … … … … How to Extend Pump Bearing Life ………………………………………Vol 4, 95-98 Atkins, Ken … … … … … Positive Displacement Pump Vibration ……………………………………Vol 2, 39-41 Bailey, Craig J. … … … … Canned Motor Pumps: Back to Basics ……………………………………Vol 1, 278-281 Bailey, Craig J. … … … … Magnetic Drive Pumps without the Hype …………………………………Vol 1, 271-276 Bailey, Derek … … … … … Optimizing High Energy Pump Operation …………………………………Vol 4, 232-235 Barnes, Brian … … … … … Applying Predictive Maintenance Measures ………………………………Vol 4, 112-115 Barnhart, Pete … … … … … Sealing Hazardous Fluids with Dry Seal Technology ……………………Vol 2, 133-136 Beahm, John A. … … … … Peristaltic Pump Technology for Industrial Applications …………………Vol 2, 182-186 Bechtler, Terry W.… … … … Self-Priming Centrifugal Pumps……………………………………………Vol 1, 60-63 Berezowskyj, Oleh … … … Developing Meaningful Pump Failure Data ………………………………Vol 5, 112-116 Bertucci, John … … … … … Communicating Your Pump Needs ………………………………………Vol 1, 136-138 Besic, Dan … … … … … … A Guide to ANSI Centrifugal Pump Design and Material Choices ………Vol 1, 198-204 Beynart, V. Larry … … … … Pulsation Control for Reciprocating Pumps ………………………………Vol 2, 187-193 Bloch, Heinz P. … … … … Bearing Protection Devices and Equipment Reliability Part I – Constant-Level Lubricators ………………………………………Vol 4, 545-547 Bloch, Heinz P. … … … … Bearing Protection Devices and Equipment Reliability Part II – What is Really Justified? …………………………………………Vol 4, 548-552 Bloch, Heinz P. … … … … Best-of-Class Lubrication for Pumps and Drivers …………………………Vol 4, 227-229 Bloch, Heinz P. … … … … Shaft Sealing for Pulp & Paper ……………………………………………Vol 3, 83-85 Blong, Richard … … … … CPI Pumping ………………………………………………………………Vol 1, 101-105 Blong, Richard … … … … Tips for Selecting ANSI Process Pumps …………………………………Vol 1, 148-152 Bolam, Jack … … … … … Coupling Alignment—Let’s Talk Tolerances ………………………………Vol 4, 147-148 Bolam, Jack … … … … … Tools That Do Our Work for Us—Or Do They? …………………………Vol 4, 204-205 Bolam, Jack … … … … … Which Alignment Method Is Right for You? ………………………………Vol 4, 294-295 Bolleter, Ulrich … … … … Improving Pump Reliability at Thermal Power Plants ……………………Vol 4, 3-5 Bowan, Gary … … … … … Specifying Air Operated Double Diaphragm Pumps ……………………Vol 2, 42 Brennan, James … … … … Built-In Relief Valves: The Case for and Against …………………………Vol 2, 100-102 Brennan, James … … … … Extending the Life of the Positive Displacement Pumps: Screw Pumps …Vol 2, 176-178 Brennan, James … … … … In the Pipeline: PD Screw Pump Valving …………………………………Vol 2, 48-51 Brennan, James … … … … Pipeline Screw Pump Efficiency …………………………………………Vol 2, 43-47 Brennan, James … … … … Rotary Pump Inlet Pressure Requirements…………………………………Vol 2, 124-126 Brennan, James … … … … Rotary Pump Overhaul Guide ……………………………………………Vol 2, 127 Brennan, James … … … … Rotary Pump Startups ………………………………………………………Vol 2, 77-81 Brennan, James … … … … Rotary Pump Troubleshooting ……………………………………………Vol 2, 82-85 Brennan, James … … … … The Canned Rotary Pump—Circa 1997……………………………………Vol 2, 108-110 Brown, Ivan … … … … … Applying Predictive Maintenance Measures ………………………………Vol 4, 112-115 Buck, Gordon … … … … … Estimating Heat Generation, Face Temperature and Flush Rate for Mechanical Seals …………………………………………Vol 3, 108-113

Buck, Gordon S. … … … … Materials for Seal Faces ……………………………………………………Vol 3, 120-123 Budris, Allan R. … … … … Pump Reliability – Hydraulic Selection to Minimize the Unscheduled Maintenance Portion of Life-Cycle Cost ………………Vol 4, 557-561 Bullentini, Scott … … … … The Role of Grouting in Standardized Installation Practices………………Vol 4, 270-273 Bullentini, Scott … … … … Building A Better Foundation………………………………………………Vol 4, 436-443 Burgner, John H. … … … … Overhauling 36 Process Pumps and Motors under Adverse Conditions …Vol 4, 367-370 Burgner, John H. … … … … The Critical First Hour Start-up for a Turbine-Driven Boiler Feed Pump Vol 4, 278-281 Burke, Peter Y.… … … … … Pumps for Haz/Rad Wastewater Service …………………………………Vol 4, 138-142 Burke, William … … … … Packing & Rotating Equipment ……………………………………………Vol 4, 291-293 Burke, William … … … … The Importance of Seal Failure Analysis …………………………………Vol 3, 70-71 Calistrat, Michael M. … … Coupling Strategies…………………………………………………………Vol 4, 127-131 Calistrat, Michael M. … … Machinery Alignment vs. Coupling Misalignment ………………………Vol 4, 149-150 Campanelli, Joe … … … … Vertical Motor-Under Pumps Expand Their Range ………………………Vol 1, 227-229 Campenelli, Joe … … … … Upgrading Boiler Water Circulation Process ………………………………Vol 1, 229-230 Cappellino, Charles … … … Tips for Selecting ANSI Process Pumps …………………………………Vol 1, 148-152 Carr, Dave … … … … … … Evaluating Sealless Centrifugal Pump Design & Performance ……………Vol 1, 222-226 Carr, Dave … … … … … … High Speed/Low Flow Pumps: Top 10 Issues ……………………………Vol 1, 125-130 Carr, Dave … … … … … … The Power of Speed and Staging …………………………………………Vol 1, 193-197 Cary, John B. … … … … … Pump Rebuilding at Avon …………………………………………………Vol 1, 213-221 Casucci, David P. … … … … A Quick Reference Guide to the API 682 Standard ………………………Vol 3, 29-35 Cayro, Julio R.… … … … … Fully Lined Slurry Pump for FCCU Bottoms Use…………………………Vol 1, 142-145 Ciotola, Al … … … … … … Coupling and Alignment Strategies ………………………………………Vol 4, 431-435 Clark, Dan … … … … … … Fully Lined Slurry Pump for FCCU Bottoms Use…………………………Vol 1, 142-145 Clarke, Kenneth … … … … Factors in Pump Suction Piping ……………………………………………Vol 4, 102-104 Claxton, Jack … … … … … Hydraulic Pressure Reactions in Pump Piping Systems …………………Vol 5, 109-111 Cleary, Joe … … … … … … Canned Motor Pumps ………………………………………………………Vol 1, 106 Clother, Alan … … … … … Improving Performance of a Water Transport Pump ……………………Vol 4, 6-9 Cole, Richard D. … … … … Getting the Full Benefit from High-Efficiency Motors ……………………Vol 4, 121-126 Coppins, Donald G. … … … Oil Lubrication for Process Pumps and Related Equipment ………………Vol 4, 248-259 Cornell, Gary … … … … … Controlling Surge and Pulsation Problems…………………………………Vol 5, 120-125 Cornell, Gary … … … … … Pulsation and Surge Control ………………………………………………Vol 2, 27-30 Covino, Larry … … … … … Probe Installation Tips ……………………………………………………Vol 4, 35-37 Crowley, John … … … … … Chemical Injection: Simplex or Complex …………………………………Vol 2, 96-99 Cummings, David… … … … Centrifugal Pump Efficiency ………………………………………………Vol 1, 30-31 Damato, David … … … … Chemical Additive Pumps for Paper Mills…………………………………Vol 2, 31-34 de Marolles, Charles … … … Innovation in Multi-phase Hydrocarbon Operations ………………………Vol 4, 211-216 Dillon, Michael L. … … … Applying the NPSHR Standard to Progressing Cavity Pumps ……………Vol 2, 103-107 Dillon, Michael L. … … … Maximize the Performance of Progressive Cavity Pumps ………………Vol 2, 153-160 Dillon, Michael L. … … … Using PC Pumps and Other Rotary PD Pumps for Metering Applications Vol 2, 231-218 Dingman, Randy R. … … … Gas-Barrier Seals Establish Beachhead ……………………………………Vol 3, 35-39 Dingman, Randy R. … … … Off-Design Operation and Seal Performance………………………………Vol 3, 22-24 Djuric, David … … … … … Sealing Lime Slurry at Alberta-Pacific ……………………………………Vol 4, 333 Dodd, V. Ray … … … … … Foundation Tips ……………………………………………………………Vol 4, 109 Dodd, V. Ray … … … … … Seal Reliability at Chevron…………………………………………………Vol 3, 55-57 Dolniak, Joseph … … … … ANSI Upgrades Require More Than Technology …………………………Vol 1, 153-159 Dolniak, Joseph F. … … … Installation: The Foundation of Equipment Reliability ……………………Vol 1,326-341 Dow, Gaylan … … … … … Vertical Motor-Under Pumps Expand Their Range ………………………Vol 1, 227-229 Downing, Jim … … … … … Extending Mechanical Seal Life……………………………………………Vol 3, 11-13 Dufour, John … … … … … Installation and Start-up Troubleshooting …………………………………Vol 1, 46-49 Dunford, Joe … … … … … Seal Protection: Guarding Against Air and Abrasives ……………………Vol 3, 62-65 Dunn, Sandy … … … … … Maintenance Outsourcing – Critical Issues ………………………………Vol 4, 553-556 Ehlert, Don … … … … … … Bearing Lubrication Trends and Tips ………………………………………Vol 4, 49-52 Enlow, Richard L.… … … … Just What is a Totally Engineered Sealing System? ………………………Vol 3, 19-21 Erickson, Dr. R. Barry … … Pump Reliability – Hydraulic Selection to Minimize the Unscheduled Maintenance Portion of Life-Cycle Cost ………………Vol 4, 557-561 Essinger, Jack … … … … … Alignment Tolerances for Spacer Couplings ………………………………Vol 4, 143-146 Fedoronko, John … … … … In-Plant Perspective: KoSa …………………………………………………Vol 4, 496-502 Feese, Troy … … … … … … Torsional Vibration Linked to Water Pumping System Failure ……………Vol 5, 98-99 Flach, Patrick M. … … … … A Quick Reference Guide to the API 682 Standard ………………………Vol 3, 29-35 Flach, Patrick M. … … … … Dual Seals: a.k.a. Double and Tandem ……………………………………Vol 3, 51-54

Flach, Patrick M. … … … … High Temperature Sealing Systems ………………………………………Vol 3, 76-78 Flach, Patrick M. … … … … Nomenclature and Definitions ……………………………………………Vol 1, 1-3 Florjancic, Stephan … … … Improving Performance of a Water Transport Pump ………………………Vol 4, 6-9 Florjanic, Dusan … … … … Improving Pump Reliability at Thermal Power Plants ……………………Vol 4, 3-5 Forsberg, Ronald P. … … … Design Practices for Safe Handling of Hazardous Fluids in API Applications …………………………………………………………Vol 1, 260-264 Forsberg, Ron … … … … … Selecting Sealless Pumps and Circulation Systems for Difficult Pumping Applications ………………………………………Vol 1, 324-335 Fortier, Kimberly … … … … Nozzle Loading — Who Sets the Standards? ……………………………Vol 1, 91-93 Fulton, Lynn C. … … … … Installation and Start-up Troubleshooting …………………………………Vol 1, 46-49 Gano, Dean… … … … … … Root Cause Failure Analysis ………………………………………………Vol 4, 510-516; Vol 5, 148-154 Ganzon, Nick … … … … … Seal Chamber Design Affects Reliability, Emissions ……………………Vol 3, 58-61 Geeslin, James W. (Tray) … Installation, Start-up and Operation of a Reciprocating Pump ……………Vol 2, 165-171 Gerber, Dean R. … … … … Testing Seals in Adverse Situations ………………………………………Vol 3, 25-28 Gibson, Ken. … … … … … Precision Solutions for Low-Flow Handling ………………………………Vol 1, 299-303; Vol 2, 194-198 Glidden, Gary E. … … … … Looking for Hidden Dangers ………………………………………………Vol 5, 93-94 Glidden, Gary E. … … … … Troubleshooting Boiler Feed Pumps ………………………………………Vol 4, 42 Glidden, Gary E. … … … … What Does It Cost?…………………………………………………………Vol 4, 60-62 Golden, Coker … … … … … Seal Environmental Controls ………………………………………………Vol 3, 1-8 Goodenberger, Bob … … … Effect of Bearing Performance on Seal Life ………………………………Vol 3, 14-15 Gourley, Dan … … … … … Teamwork Key to Pump Reliability Upgrades at Conoco ………………Vol 4, 315-317 Greutink, Herman… … … … Recommendations for Vertical Pump Intake ………………………………Vol 1, 115-117 Greutink, Herman… … … … Vertical Turbine Pumps Power Petrochemicals ……………………………Vol 1, 175-179 Grimmenstein, Larry F. … … New Coatings Make Their Mark …………………………………………Vol 4, 220-223 Guelick, J.F. … … … … … Hydraulic Instabilities and Cavitation ……………………………………Vol 1, 118-124 Gulich, Johann… … … … … Improving Pump Reliability at Thermal Power Plants ……………………Vol 4, 3-5 Guthrie, Daniel … … … … Metering System Design Requirements ……………………………………Vol 2, 111-113 Hammock, Cliff … … … … Motor Bearing Failures in Cooling Tower Water Pump …………………Vol 5, 103-105 Hammock, Cliff … … … … Reliability: Big Opportunity or Big Distraction……………………………Vol 4, 347-352 Hammock, Cliff … … … … The Nuts and Bolts of Vibration Signature Analysis ……………………Vol 4, 305-310 Hansen, Scott … … … … … Lubrication Systems for Rotating Equipment ……………………………Vol 4, 29-33 Harmon, Alan … … … … … Reliability Engineers’ Advice: Make the First Steps the Right Steps ……Vol 4, 360 Hart, Robert J. … … … … … Motor Size Selection for Centrifugal Pumps ………………………………Vol 1, 32-33 Harting, Gerald … … … … Reciprocating Metering Pumps in Leak-Free Design ……………………Vol 2, 90-95 Hawkins, Kenneth … … … Rim and Face Alignment Procedures for Direct Motor Driven Equipment Vol 4, 224-226 Hawkins, Kenneth … … … Vibration Amplification ……………………………………………………Vol 5, 58-59 Hawks, Jeff… … … … … … Cavitation in a Nutshell ……………………………………………………Vol 5, 106-108 Hayes, John … … … … … Chopper Pumps Digest the Solids …………………………………………Vol 1, 180-183 Hernandez, Terry … … … … Reliability-Driven Pump Maintenance ……………………………………Vol 4, 76-85 Henson, Gregory. … … … … In-Plant Perspective: Eli Lilly and Company………………………………Vol 4, 477-482 Heronema, John V. … … … Pumping Hydrofluoric Acid ………………………………………………Vol 1, 75-78 Hines, Gordon … … … … … Balancing Pump Impellers …………………………………………………Vol 4, 325-328 Hole, Gunnar … … … … … Fluid Viscosity Effects on Centrifugal Pumps ……………………………Vol 1, 20-23 Hole, Gunnar … … … … … Pump Balancing Criteria……………………………………………………Vol 1, 24-25 Holmes, Tim … … … … … In-Plant Perspective: Dupont ………………………………………………Vol 4, 483-488 Hopkins, Sam … … … … … Lubrication Systems for Rotating Equipment ……………………………Vol 4, 29-33 Horwath, John H. … … … … Centrifugal Pump Suction …………………………………………………Vol 1, 231-233 Hrivnak, Steven J. … … … Adhering to Specifications and Standards Key for MTBR Improvements at Eastman Chemical ………………………………Vol 4, 356-359 Hrivnak, Steven J. … … … Computer-based Pump Reliability …………………………………………Vol 4, 164-175 Hrivnak, Steven J. … … … Why Doesn’t Your Self-Priming Pump Work? ……………………………Vol 5, 137-139 Hrivnak, Steven J. … … … In-Plant Perspective: Eastman Chemical Company ………………………Vol 5, 140-147 Ingram, James H. … … … … Suction Side Problems—Gas Entrainment…………………………………Vol 1, 87-90 Ives, Jeff … … … … … … Metering System Design Requirements ……………………………………Vol 2, 111-113 Jackson, Maurice G. … … … User Perspective: When to Apply Mag Drive Pumps ……………………Vol 1, 79-80 Jacoby, Rodger … … … … Handling Abrasives and Corrosives with Positive Displacement Pumps …Vol 2, 19-21 Johns, David A. … … … … Reliability Improvements to a 13-Stage Charge Pump ……………………Vol 4, 68-70 Johns, David A, … … … … Reliability Improvements to a High Pressure Charge Pump ……………Vol 4, 71-75 Johns, Will E. … … … … … Universal Drive Shaft Maintenance ………………………………………Vol 4, 230-231 Jones, Randall … … … … … AlliedSignal Geismar Works ………………………………………………Vol 4, 334-336

Jones, Merwin W. … … … … Experience with Replacement of Boiler Feed Pumps for Reliability Enhancement ………………………………………Vol 5, 130-133 Kahler, Thomas G. … … … Construction Impacts on Pumps and Systems ……………………………Vol 4, 403-411 Kamaraju, Satish … … … … Using VFD Technology in a Water Distribution System …………………Vol 4, 448-454 Karassik, Igor J. … … … … Effects of Oversizing ………………………………………………………Vol 1, 17-19 Karassik, Igor J. … … … … Setting the Minimum Flows for Centrifugal Pumps ………………………Vol 1, 34-36 Key, William (Bill) … … … Sealing Technology for VOC Control ……………………………………Vol 2, 114-117 Khan, Tanveer … … … … … Coupling and Alignment Strategies ………………………………………Vol 4, 431-435 Klein, Manfred … … … … Operation Protection for Mag Drive Pumps ………………………………Vol 1, 166-169 Koelbl, Gary C. … … … … Energy Savings Pay for Reliability Improvements at Chevron ……………Vol 4, 311-314 Kracke, George … … … … Winning Maintenance Strategies at Thorn Creek …………………………Vol 4, 209-210 Kracke, George … … … … In-Plant Perspective: Thorn Creek Basin Sanitary District ………………Vol 4, 503-506 Krebs, Robert … … … … … Calculating Shaft Deflection ………………………………………………Vol 5, 67-68 Krebs, Robert … … … … … Have you Broken Any Shafts Lately? ……………………………………Vol 5, 39-40 Krebs, Robert … … … … … Pump System Design—Part 1 ……………………………………………Vol 5, 29-31 Krebs, Robert … … … … … Pump System Design—Part 2 ……………………………………………Vol 5, 32-34 Krebs, Robert … … … … … Pump System Design—Part 3 ……………………………………………Vol 5, 35-36 Krebs, Robert … … … … … Pumping Downhill …………………………………………………………Vol 5, 50-51 Krebs, Robert … … … … … Pumping System Piping ……………………………………………………Vol 5, 37-38 Krebs, Robert … … … … … Pumping Terms ……………………………………………………………Vol 1, 3 Krebs, Robert … … … … … Pumping with Air …………………………………………………………Vol 5, 75-76 Krebs, Robert … … … … … Pumps in Parallel - When One is Not Enough ……………………………Vol 5, 41-42 Krebs, Robert … … … … … Pumps in Parallel with Variable Speed Drive ……………………………Vol 5, 69-70 Krebs, Robert … … … … … Pumps in Series - For More Pressure ………………………………………Vol 5, 43-44 Krebs, Robert … … … … … Selecting a Pump: Will it Operate Where You Want it to? ………………Vol 5, 62-63 Krebs, Robert … … … … … Selecting a Pump? Define Efficiency First ………………………………Vol 5, 60-61 Krebs, Robert … … … … … Selecting a Pump-Suction Pressure ………………………………………Vol 5, 54 Krebs, Robert … … … … … Selecting a Pump-The Right Start Leads to the Right Finish ……………Vol 5, 52 Krebs, Robert … … … … … Selecting a Pump-What Type Should it Be? ………………………………Vol 5, 53 Krebs, Robert … … … … … Sizing Pumps for Complex Systems—Part 1………………………………Vol 5, 71 Krebs, Robert … … … … … Sizing Pumps for Complex Systems—Part 2………………………………Vol 5, 72-73 Krebs, Robert … … … … … Sizing Pumps for Complex Systems—Part 3………………………………Vol 5, 74 Krebs, Robert … … … … … Waterhammer - Containing the Surge ……………………………………Vol 5, 80-81 Krebs, Robert … … … … … Where’s the Prime?…………………………………………………………Vol 5, 27-28 Krebs, Robert … … … … … Why Does the Pump Pump? ………………………………………………Vol 5, 48-49 Kriebel, Doug … … … … … Key Centrifugal Pump Parameters and How They Impact Your Applications – Part 1 ……………………………Vol 1, 304-309 Kriebel, Doug … … … … … Key Centrifugal Pump Parameters and How They Impact Your Applications – Part 2 ……………………………Vol 1, 310-314 LaCombe, David … … … … Revitalizing Vertical Lineshaft Turbines …………………………………Vol 4, 329-332 Lahr, Paul T. … … … … … Cavitation and NPSH in Centrifugal Pumps ………………………………Vol 1, 6-9; Vol 5, 3-6 Laplant, Kenneth … … … … A Multi-Stage Vertical Pump in Light Hydrocarbon Service—Part 1 ………………………………………………Vol 4, 53-55 Laplant, Kenneth … … … … A Multi-Stage Vertical Pump in Light Hydrocarbon Service—Part 2 ………………………………………………Vol 4, 56-59 Laux, C.H. … … … … … … Replacement of the Boiler Feed Pumps at Janschwalde Power Station …………………………………………………Vol 4, 298-304 Lavelle, Ken … … … … … Choices in Emission Control for Rotating Equipment ……………………Vol 3, 90-97 Lavelle, Ken … … … … … Sealing Technology for VOC Control ……………………………………Vol 2, 114-117 Lawler, John … … … … … Multistage Pumps in the CPI ………………………………………………Vol 4, 263-269 Le Bleu, Julien … … … … Continuous Monitoring of Sealless Pumps ………………………………Vol 1, 315-323; Vol 4, 517-525 Le Bleu, Julien … … … … Keys to Improve Pump Record-Keeping …………………………………Vol 4, 62-67 LeFrapper, Claude … … … Flow Control with Metering Pumps ………………………………………Vol 2, 146-149 Liddle, Bob… … … … … … Evaluating Pump Monitoring Options ……………………………………Vol 4, 110-111 Linden, Robert… … … … … High Speed/Low Flow Pumps: Top 10 Issues ……………………………Vol 1, 125-130 Lingenfelder, George … … … Fire Pump Systems—Design and Specification……………………………Vol 1, 251-255 List, Mark … … … … … … Well Pump Applications for Mine Dewatering ……………………………Vol 1, 256-259; Vol 2 137-140; Vol 4, 337-340 Lobach, James … … … … … Continuous Monitoring of Sealless Pumps ………………………………Vol 1, 315-323; Vol 4, 517-525 Lubcyik, Mike … … … … … Energy Savings Pay for Reliability Improvements at Chevron ……………Vol 4, 311-314

Mabe, William … … … … … Reducing Acoustic Emissions………………………………………………Vol 4, 22-24 Mabe, William … … … … … The Power of Speed and Staging …………………………………………Vol 1, 193-197 Mackay, Ross C. … … … … Centrifugal and Positive Displacement Pumps in the Operating System …Vol 1, 4-5; Vol 2, 3-4 Mackay, Ross C. … … … … Inlet Piping Configuration and Motor Efficiency …………………………Vol 5, 1-2 Mackay, Ross C. … … … … Pump Suction Conditions …………………………………………………Vol 1, 10-12; Vol 5, 24-26 Mackay, Ross C. … … … … The Correction and Prevention of Low NPSH Conditions ………………Vol 5, 95-97 Mahan, Jim … … … … … … Torsional Analysis in Couplings Selection…………………………………Vol 4, 43-46 Malinowski, John … … … … Evaluating Motor and Drive Solutions ……………………………………Vol 4, 191-193 Manion, Bob … … … … … CPI Pumping ………………………………………………………………Vol 1, 101-105 Mares, Art … … … … … … Energy Savings Pay for Reliability Improvements at Chevron ……………Vol 4, 311-314 Martelli, Robert … … … … National Electric Code Impact on Sealless Centrifugal Pumps …………Vol 1, 70-74 Martin, Ana M. … … … … Multiphase – The Final Pumping Frontier …………………………………Vol 2, 219-223 Matthews, Robert … … … … Avoiding Gremlins and Alligators at Pump Start-Up………………………Vol 5, 126-129 Matthews, Bob … … … … Packing in the 21st Century ………………………………………………Vol 3, 124-129 Matthews, Bob … … … … Paper Mill Stock Improvements ……………………………………………Vol 4, 507-509 Matthews, Robert … … … … Reliability Profile #1: Champion International Corporation ………………Vol 4, 157-159 Mayleben, Phil … … … … Performance Curve vs. System Curve ……………………………………Vol 5, 64-66 McCaul, Colin … … … … … OEM Pump Materials and Their Relationship to Quality and Reliability …Vol 4, 386-392 McCauley, Michael L. … … Reliability Through Fluid Sealing Management …………………………Vol 3, 86-89 McCloskey, T.H. … … … … Hydraulic Instabilities and Cavitation ……………………………………Vol 1, 118-124 McCloskey, Tom … … … … Improving Pump Reliability at Thermal Power Plants ……………………Vol 4, 3-5 McCullough. Mark F. … … Coupling Alignment ………………………………………………………Vol 5, 155-158 McGuire, J.T. … … … … … Pump Buying Strategies ……………………………………………………Vol 1, 107-110 McGuire, J.T. … … … … … Pump Ratings Vital when Pressure’s On …………………………………Vol 1, 131-135 McGuire, J.T. … … … … … So You Need Pumps for a Revamp ………………………………………Vol 1, 184-189 McKay, Scott … … … … … Selection Guide to Metering Pumps ………………………………………Vol 2, 86-89 McNally, William … … … … Classifying Chemicals to Ensure Effective Sealing ………………………Vol 4, 344-346 Mechelay, Eddie … … … … Bellows Seal Repair: The Plant Perspective ………………………………Vol 3, 41-42 Mechelay, Eddie … … … … Cartridge Seals: Pros, Cons and the Human Factor ………………………Vol 3, 43-44 Mechelay, Eddie … … … … Increased Reliability at a Small Refinery …………………………………Vol 4, 526-531 Mechelay, Eddie … … … … Key Indicators—The Measure of Reliability ………………………………Vol 4, 217-219 Mechelay, Eddie … … … … Mechanical Seal Installation ………………………………………………Vol 3, 66-69 Mechelay, Eddie … … … … Turnaround Time! …………………………………………………………Vol 4, 276-277 Mefford, Jody … … … … … Reliability Engineers’ Advice: Make the First Steps the Right Steps ……Vol 4, 360 Merullo, Ralph … … … … Mechanical Seal Failure Analysis …………………………………………Vol 3, 114-119 Messina, J.P. … … … … … Modeling Pump Intake Noise ………………………………………………Vol 5, 100-102 Meyer, Erich A. … … … … Selecting the Right Thermoplastic Sump Pump for Corrosive, Abrasive and Ultrapure Services ………………………………Vol 1, 282-287 Micheletti, Wayne C. … … Examining Pump Capacity Problems ………………………………………Vol 1, 42-43 Mikalonis, Dave … … … … Bearing Reliability in Centrifugal Pumps …………………………………Vol 1, 265-270 Miller, Jim … … … … … … NPSH Required for Reciprocating PD Pump………………………………Vol 2, 1-2; Vol 5, 22-23 Moore, Stan … … … … … The Role of Grouting in Standardized Installation Practices………………Vol 4, 270-273 Mraz, Will … … … … … … Magnetic Liquid Seals Contain Voices ……………………………………Vol 3, 16-18 Muehlbauer, Jerry… … … … Selecting Motors & Drives for Optimum System Reliability ……………Vol 4, 371-378 Mulholland, Robert … … … Maximizing Reliability in Sealless Pump Operation ……………………Vol 4, 318-322 Muller, Fredrich … … … … Fluid Metering System Options ……………………………………………Vol 2, 22-26 Murphy, Stephen … … … … Variable Speed Pumping……………………………………………………Vol 1, 55-59 Murray, Malcolm … … … … Alignment Movement Methods ……………………………………………Vol 4, 105-106 Myers, Charles A.… … … … Interpreting Sealless Pump Failures ………………………………………Vol 1, 81-82 Myers, Richard D. … … … Build a Better Foundation to Reduce Costly Downtime …………………Vol 4, 151-153 Myers, Richard D. … … … Not All Grouts are Created Equal …………………………………………Vol 4, 47-48 Neis, William … … … … … Fluid Metering System Options ……………………………………………Vol 2, 22-26 Nelik, Lev … … … … … … Extending the Life of the Positive Displacement Pumps: Gear Pumps……Vol 2, 172-175 Nelik, Lev … … … … … … Piping-to-Pump Alignment: Getting it Right! ……………………………Vol 4, 393-398 Nelik, Lev … … … … … … Pump Design Changes Improve Lubrication ………………………………Vol 1, 98-100 Nelson, William E. (Ed) … … Addressing Pump Vibrations—Part 1………………………………………Vol 4, 25-26 Nelson, William E. (Ed) … … Addressing Pump Vibrations—Part 2………………………………………Vol 4, 27-28 Nelson, William E. (Ed) … … Anti-Friction Bearings in Centrifugal Pumps………………………………Vol 1, 237-247 Nelson, William E. (Ed) … … Ed’s Rule of Thumb ………………………………………………………Vol 5, 59 Nelson, William E. (Ed) … … Mounting and Clearance for Casing Wear Rings …………………………Vol 4, 99-101

Nelson, William E. (Ed) … … Proper Repairs Avert Failures—Part 1 ……………………………………Vol 4, 176-180 Nelson, William E. (Ed) … … Proper Repairs Avert Failures—Part 2 ……………………………………Vol 4, 181-182 Nelson, William E. (Ed) … … Proper Repairs Avert Failures—Part 3 ……………………………………Vol 4, 183-186 Nelson, William E. (Ed) … … Proper Repairs Avert Failures—Part 4 ……………………………………Vol 4, 187-190 Nelson, William E. (Ed) … … Tips on Pump Efficiency …………………………………………………Vol 1, 39-41 Nelson, William E. (Ed) … … Troubleshooting Centrifugal Pumps ………………………………………Vol 5, 77-79 Netzel, James P. … … … … Built to Last: Working with a New Mechanical Seal Standard ……………Vol 3, 130-134 Netzel, James … … … … … Estimating Maximum Head in Single and Multi-Stage Pump Systems …………………………………………………Vol 1, 37-38; Vol 5, 16-17 Noble, Ken … … … … … … Methodologies for Calculating MTBF—Part 1 ……………………………Vol 4, 282-286 Noble, Ken … … … … … … Methodologies for Calculating MTBF—Part 2 ……………………………Vol 4, 287-290 Noble, Ken … … … … … … Strategies to Reduce Pump Repairs at Petro-Canada………………………Vol 4, 455-460 Ochs, Greg … … … … … … Flow Sensing for Pump Protection…………………………………………Vol 4, 154-156 Ooka, Kaz … … … … … … Operation Protection for Mag Drive Pumps ………………………………Vol 1, 166-169 Pagalthivarthi, Krishnan V … Slurry Pump Wear Factors …………………………………………………Vol 4, 194-203 Parker, Joseph C. … … … … Off-Design Operation and Seal Performance………………………………Vol 3, 22-24 Patel, Vinod … … … … … Oil Lubrication for Process Pumps and Related Equipment ………………Vol 4, 248-259 Perry, Steven … … … … … Alignment Monitoring Report ……………………………………………Vol 4, 399-402 Petersen, Ray … … … … … Self-Priming Pumps: It’s in the System ……………………………………Vol 1, 205-208 Peterson, John … … … … … Handling Abrasives and Corrosives with Positive Displacement Pumps …Vol 2, 19-21 Pfaff, David A. … … … … Testing Seals in Adverse Situations ………………………………………Vol 3, 25-28 Piotrowski, John … … … … Understanding and Using Shaft-to-Shaft Alignment ……………………… Measurement Systems………………………………………………………Vol 4, 379-385 Piotrowski, John … … … … Finding and Solving Vibration Problems …………………………………Vol 4, 461-466 Piotrowski, John … … … … Pro-Active Maintenance for Pumps ………………………………………Vol 4, 538-544 Platt, Robert A. … … … … A User’s Guide to Rotary Pumps …………………………………………Vol 2, 118-123 Platt, Robert A. … … … … Maintaining and Operating Positive Displacement Rotary Gear Pumps …Vol 2, 64-66 Platt, Robert A. … … … … Positive Displacement Two Screw Pumps …………………………………Vol 2, 5-8 Polk, Mike … … … … … … Maximize Seal Flush Performance for Longer Seal Life …………………Vol 3, 98-107 Powers, Gary … … … … … Vibration Analysis Yields Good Vibes ……………………………………Vol 4, 116-120 Price, Stephen … … … … … Valve Dynamics Affect Recip Pump Reliability—Part 1 ………………Vol 2, 52-54 Price, Stephen … … … … … Valve Dynamics Affect Recip Pump Reliability—Part 2 ………………Vol 2, 55-59 Pumps and Systems Staff … Alignment Using the C Frame Adapter ……………………………………Vol 4, 107-108 Pumps and Systems Staff … Canned Motor Pump Application Profiles …………………………………Vol 4, 206-208 Pumps and Systems Staff … Early Maintenance Staff Involvement ……………………………………Vol 4, 10-11 Pumps and Systems Staff … Hazardous Fluid Pump and Sealing Systems: Reliability Driven Improvements …………………………………………Vol 4, 341-343 Pumps and Systems Staff … Laser Shaft Alignment of a Bark Hog ……………………………………Vol 4, 323-324 Pumps and Systems Staff … Low Flow Options …………………………………………………………Vol 1, 94-97 Pumps and Systems Staff … Mechanical Seals Keep Pace with User Demands ………………………Vol 3, 9-10 Pumps and Systems Staff … More on Non-Contacting Seal Design ……………………………………Vol 3, 39 Pumps and Systems Staff … Options for Sealless Centrifugals …………………………………………Vol 1, 146-147 Pumps and Systems Staff … Pumping Hot Stuff—Another Perspective …………………………………Vol 1, 190-192 Pumps and Systems Staff … Vertical Motor-Under Pumps Expand Their Range ………………………Vol 1, 227-229 Pumps and Systems Staff … Winning MTBR Strategies …………………………………………………Vol 4, 356-359 Purcell, John … … … … … Gear Drive Options for Rotary Pumps ……………………………………Vol 2, 75-76 Quarfordt, Katherine … … … Want to Go from 0 to 100 in 3 seconds? …………………………………Vol 2, 150-152 Rankin, Kenneth … … … … Predictive Maintenance Programs: Building a Complete Package ………………………………………………………Vol 4, 412-420 Rake, Brad … … … … … … Water Contamination of Equipment-Lubricating Oil………………………Vol 4, 533-537 Redpath, David … … … … Built to Last: Working with a New Mechanical Seal Standard ……………Vol 3, 130-134 Rhoe, Ray … … … … … … Bearing Basics ……………………………………………………………Vol 1, 26-29; Vol 4, 38-41 Richard, Leo … … … … … Centrifugal Pump Testing …………………………………………………Vol 1, 64-67 Rienks, Patrick … … … … Pumping Options for Low Flow/High Head Applications ………………Vol 1, 234-236 Rinard, John … … … … … A Common Sense Approach to Combating Corrosion and Abrasion ……Vol 1, 111-114 Rizo, Luis F. … … … … … Piping-to-Pump Alignment: Getting it Right! ……………………………Vol 4, 393-398 Rizo, Luis F. … … … … … Troubleshooting Pump Performance Degradation …………………………Vol 5, 55-57 Robinson, Jeihri … … … … Solid State Controls Tune Pump Operation ………………………………Vol 4, 135-137 Roland, Brent … … … … … A Review of Positive Displacement Pumps ………………………………Vol 2, 15-18 Ross, Robert R. … … … … Impellers and Volutes: Power with Control ………………………………Vol 1, 139-141

Rossi, Stephen C. … … … … Pump Rebuilding at Avon …………………………………………………Vol 1, 213-221 Rutan, Charlie … … … … … A Multi-Stage Vertical Pump in Light Hydrocarbon Service—Part 1 ……Vol 4, 53-55 Rutan, Charlie … … … … … A Multi-Stage Vertical Pump in Light Hydrocarbon Service—Part 2 ……Vol 4, 56-59 Sabini, Eugene P. … … … … Pump Reliability – Hydraulic Selection to Minimize the Unscheduled Maintenance Portion of Life-Cycle Cost ………………Vol 4, 557-561 Schindler, Keith … … … … Controlling the Seal Environment—A Key to Seal Reliability ……………Vol 4, 353-355 Schmidt, George C. … … … Best Practice: High Temperature Slurry Applications ……………………Vol 4, 421-424 Schmitz, Steve … … … … … The Mystery of Cooling Tower Pump Noise ………………………………Vol 5, 117-119 Schommer, Harry … … … … Know the Inside Story of Your Mag Drive Pumps ………………………Vol 1, 209-212 Schultz, John … … … … … Reliability: Big Opportunity or Big Distraction……………………………Vol 4, 347-352 Schumann, Kurt … … … … Upgrading Utility and Process Pumps ……………………………………Vol 1, 50-54 Sciascia, Robert … … … … Chemical Additive Pumps for Paper Mills…………………………………Vol 2, 31-34 Scott, Dr. Stuart L. … … … Multiphase – The Final Pumping Frontier …………………………………Vol 2, 219-223 Self, Karen … … … … … … Team Approach Produces Winning O&M Formula—Part 1 ……………Vol 4, 12-14 Self, Karen … … … … … … Team Approach Produces Winning O&M Formula—Part 2 ……………Vol 4, 15-17 Self, Karen … … … … … … Team Approach Produces Winning O&M Formula—Part 3 ……………Vol 4, 18-21 Semotink, Dave … … … … A Guide to High Pressure Reciprocating Pumps …………………………Vol 2, 129-132 Shank, Paul… … … … … … Fire Pump Systems—Design and Specification……………………………Vol 1, 251-255 Shanley, Larry … … … … … Improving Progressing Cavity Pump Reliability …………………………Vol 4, 241-243 Shelby, Jerry T. … … … … Lubrication Oil Viscosity Classifications…Don’t Get Confused …………Vol 4, 34 Sidelko, John … … … … … How Much NPSH is Enough? ……………………………………………Vol 5, 11-12 Simon, Albert … … … … … Improving Pump Reliability at Thermal Power Plants ……………………Vol 4, 3-5 Simonette, Dallas … … … … A Guide to High Pressure Reciprocating Pumps …………………………Vol 2, 129-132 Sinclair, Jerry L. … … … … The Importance of a Bearings Inspection Program ………………………Vol 4, 425-430 Skelton, Kevin … … … … … Variable Frequency Drives for a Vacuum Pump System …………………Vol 4, 274-275 SKF USA, Inc. … … … … Exploring Bearing Lubrication Options ……………………………………Vol 4, 90-94 Smith, Ronald P. … … … … Magnetic Couplings for Sealless Pumps …………………………………Vol 1, 83-86 Smith, Donald … … … … … Valve Dynamics Affect Recip Pump Reliability—Part 1 ………………Vol 2, 52-54 Smith, Donald … … … … … Valve Dynamics Affect Recip Pump Reliability—Part 2 ………………Vol 2, 55-59 Smith, Michael D. … … … Venting Pump Systems ……………………………………………………Vol 1, 44-45 Smith, Stephen … … … … Twin-Screw Pumps vs. Centrifugal and Reciprocating Pumps ……………Vol 2, 205-212 Spitzer, David William … … Variable Speed Drives: Advantages and Pitfalls …………………………Vol 5, 13-15 Stanmore, Leon K. … … … Pump and Turbine Rerate Programs ………………………………………Vol 5, 85-92 Stanmore, Leon K. … … … Reducing Noise in API Process Pumps ……………………………………Vol 5, 82-84 Stefl, Timothy S. … … … … Reliability Starting with Procurement: Maintenance’s Role in the Purchase Cycle …………………………………………………Vol 4, 361-366 Stover, Bob… … … … … … Expansion Joints and Air Chambers ………………………………………Vol 2, 9-10 Submersible Wastewater… … Successful Submersible Operation Pump Assocation Part 1: Pump Installation and Start-Up ……………………………………Vol 1, 288-294 Submersible Wastewater… … Successful Submersible Operation Pump Association Part 2: Inspection and Maintenance ………………………………………Vol 1, 295-298; Vol 4,444-447 Thome, Terry E. … … … … Getting the Full Benefit from High-Efficiency Motors ……………………Vol 4, 121-126 Thorwart, Lawrence J. … … Seal Face Application: Little Room for Error ……………………………Vol 3, 46-49 Tison, Jim … … … … … … Valve Dynamics Affect Recip Pump Reliability—Part 1 ………………Vol 2, 52-54 Tison, Jim … … … … … … Valve Dynamics Affect Recip Pump Reliability—Part 2 ………………Vol 2, 55-59 Treichler, Stephen R … … … Team Approach Produces Winning O&M Formula—Part 1 ……………Vol 4, 12-14 Treichler, Stephen R … … … Team Approach Produces Winning O&M Formula—Part 2 ……………Vol 4, 15-17 Treichler, Stephen R … … … Team Approach Produces Winning O&M Formula—Part 3 ……………Vol 4, 18-21 Tuck, Alan Jr. … … … … … Air Operated Diaphragm Pumps—Nineties Style …………………………Vol 2, 35-38 Uwe, Burchhardt … … … … Replacement of the Boiler Feed Pumps at Janschwalde Power Station …Vol 4, 298-304 Uwe, Kirstein … … … … … Replacement of the Boiler Feed Pumps at Janschwalde Power Station …Vol 4, 298-304 Valente, Nick … … … … … Reliability Tips for Operating Magnetic Drive Pumps ……………………Vol 1, 277 Van Bogaert, Larry … … … Peristaltic Pumps Offer Custom Fluid Solutions …………………………Vol 2, 67-70 Visintainer, Robert J. … … … Slurry Pump Wear Factors …………………………………………………Vol 4, 194-203 Vullings, Klaus … … … … Applying the NPSHR Standard to Progressing Cavity Pumps ……………Vol 2, 103-107 Wahren, Uno … … … … … Centrifugal Pumps Operating in Parallel …………………………………Vol 1, 248-250 Waling, Lee A.… … … … … Off-Design Operation and Seal Performance………………………………Vol 3, 22-24 Wallace, Bob … … … … … Baseplate Design Affects Reliability ………………………………………Vol 4, 87-89 Wallace, Ed… … … … … … Extending the Life of the Positive Displacement Pumps: Progressing Cavity …………………………………………………………Vol 2, 179-181 Wallace, Neil M. … … … … Built to Last: Working with a New Mechanical Seal Standard ……………Vol 3, 130-134

Warwick, Ed … … … … … Minimizing Pressure & Flow Pulsations from Piston/Diaphragm Metering Pumps ………………………………………Vol 2, 161-164 Warwick, Ed … … … … … Reducing Pulsations in Metering Pumps …………………………………Vol 2, 60-63 Warwick, Ed … … … … … Metering Pump System Design for Dependable Performance ……………Vol 2, 199-204 Waterbury, Robert C. … … Adjustable Speed Drive Offers Pump Flexibility …………………………Vol 4, 132-134 Waterbury, Robert C. … … Sealless Options Optimize Solutions ………………………………………Vol 1, 170-174 Waterbury, Robert C. … … Selecting Mag Drive Pumps ………………………………………………Vol 1, 160-165 Waterbury, Robert C. … … Tips and Trends in Sanitary Pumping ……………………………………Vol 4, 160-163 Waterbury, Robert C. … … Zero-Leak Seals Cut Emissions ……………………………………………Vol 3, 72-75 Watkins, Robert C. … … … Seal Reliability at Chevron…………………………………………………Vol 3, 55-57 Weehunt, Dennis L. … … … Injectable Packing Compounds—An Alternative Sealing Technology …………………………………………………………Vol 4, 260-262 Weehut, Dennis … … … … Building Reliability and Safety into Rotating Equipment ………………Vol 4, 489-495 Weidner, David … … … … Troubleshooting ANSI Pumps Using Limited Information ………………Vol 5, 134-136 White, Scott … … … … … In-Plant Perspective: Dupont ………………………………………………Vol 4, 483-488 Whitmire, C.M. … … … … Kent Selection Guide, Rotary Gear Pumps ………………………………Vol 2, 71-74 Wild, Alan G. … … … … … Selecting a Progressing Cavity Pump………………………………………Vol 2, 11-14 Wold, Terry M. … … … … Elements of Minimum Flow ………………………………………………Vol 1, 13-16; Vol 5, 18-21 Wold, Terry… … … … … … Managing a Vertical Pumps in a Changing World ………………………Vol 4, 236-240 Wold, Terry M … … … … … Vertical Pumps with Integral Thrust Bearings ……………………………Vol 5, 45-47 Wood, John W. … … … … Sealing Positive Displacement Pumps: What Are Your Choices? ………Vol 2, 141-145 Xu, Ming … … … … … … Spike Energy™ Measurement ……………………………………………Vol 4, 467-476 Zamin, Mohammad … … … Team Approach Produces Winning O&M Formula—Part 1 ……………Vol 4, 12-14 Zamin, Mohammad … … … Team Approach Produces Winning O&M Formula—Part 2 ……………Vol 4, 15-17 Zamin, Mohammad … … … Team Approach Produces Winning O&M Formula—Part 3 ……………Vol 4, 18-21 Zimmerman, Gregory … … A Review of Positive Displacement Pumps ………………………………Vol 2, 15-18 Zimmerman, Gregory … … A Systems Approach to Pump Reliability …………………………………Vol 4, 1-2 Zimmerman, Gregory … … The Canned Motor vs. Magnetic Drive Debate ……………………………Vol 1, 68-69

200 Diagnostic Engineering

Why are mechanical seals used on centrifugal pumps

M

echanical Seals has always been synonymus to Centrifugal Pump of all types. Typically, when someone discusses pumps of this description, mechanical seals appears to jive-in audaciously into every deliberations that takes place on this subject. I have talked to numerous production and maintenance technicians in Petronas Refinery on their understanding of this subject.

As the process operator in a petrochemical complex (its good for true-blue maintenance boys to appreciates this too), your dominant task may possibly involves routine assurances that process liquids are being efficiently and precisely transported through a 'leak-proofed' centrifugal pump. Irrespective from a point to another or from plant to plant via dedicated piping systems.

Interestingly, it appears that each and every person of us seems to have only one conception for its existence, that the mechanical seals are being stuffed somewhere in the pump for none other than to prevent 'lethal' fluids from escaping into the atmosphere. Of course, I totally agree with them as its name implies, this must be true, otherwise why bother put one in the first place? But, are we sure that they are being put there exclusively to prevent leakage to the outside world or was it the other way around?

In this case, it is your duty too to ensure that only uncontaminated fluids are being moved around throughout the processing systems. Taking into consideration that the interior of a centrifugal pump (casing) is under a suction pressure of constant (*vacuum) strength whilst running, it must becomes apparent that any available outside influences such as moisture and cool air are more than likely to be introduced into the process system.

Before we could make factual conclusions on this subject, lets us take a closer look at what basically or practically goes-on within a working (running) centrifugal pump. All this while, we typically presumed pumps to be more or less of a high-pressure generating mechanism that will elevate pressure of liquids but infrequently addresses them as a vacuum creating medium. Let me ask you this simple question: How could there conceivably be flow or fluid pressure at the pump's discharge flange if there's non-existence of an 'energy' to somehow or other transmit or convey the said liquid or gaseous mass through to the suction flange? Hence, without hesitation I would fittingly say that this machine in question must surely be a device exclusively designed to exert 'positive suction' forces upon any liquids or gaseous body by the logic of reduced air pressure (vacuum) over parts of its internal surfaces where in this instances, literally is the pump's casing. After all is said and done, if it turns out that the casing of a running pump is without exception being covered or primed by a suction influence of high potencies, how could there possibly be leakage of fluids from the casing to the atmosphere? In order to get the concept proper, I believed it is more appropriate for us to immediately change the way we speculate of the way a pump works before we comprehend the rationality of why a mechanical seal is being packed into its stuffing box.

Subsequently, this inducement of contaminants will indefinitely provoke (efficiency) degradation of the pump. The performance reductions customarily resulted in excessive cavitation to the pump's behaviour. if it turns out that hot fluids is the pumped media, the reactivity of introduced moisture and chilly atmosphere into the pump's vacuity will beyond doubt bring about or effectuate (an abrupt) severe implosions on the spinning impeller's wheel, vanes and casing surfaces respectively. The aftermath of these clattering bursts ordinarily encourages pinhole cavities to materialise on impeller surfaces thus if serious, will in the fullness of time affect it to go out of balance etc. (*A vacuum described here is hypothetical statement addressing the region of space around or near the pump impeller's eye where the pressure is less than the normal atmospheric pressure). It is for this reason that for an 'unambiguous' operational requirement, a pump purchaser (the project engineer) must usually furnish the available positive suction pressure (NPSHA*) to the reputable pump's manufacturer in order for it to accordingly design a suitable pump that will meet his requirement. The manufacturer later in the purchasing course typically returns an 'obligatory' pump performance curves of the demanded specifications to the intended buyer. This include data of the Positive Suction requirements (NPSHR*) and of the manufacturer's capability to realistically meet the determined Net Positive Suction Head (NPSH*) in order to meet the owner's explicit discharge requirements. (*The

Institution of Diagnostic Engineers

Diagnostic Engineering 201

N.P.S.H available can be calculated for a specific situation and depends on the barometric pressure, the friction loss between the system inlet and the pump suction flange, and other factors.

opening near the rotating shaft called a packing or sea ‘stuffing box’. The design of this stuffing box has not change inasmuch as packing was first introduced as sealing media to centrifugal pumps.

The (*N.P.S.H.R.) required is given by the pump manufacturer and depends on the head, flow and type of pump. The N.P.S.H. available must always be greater than the N.P.S.H. required for the pump to operate properly whereas the Net Positive Suction Head (N.P.S.H.) is the head at the suction flange of the pump less the vapour pressure converted to fluid column height of the fluid. The N.P.S.H. is always positive since it is expressed in terms of absolute fluid column height.

However, taking into consideration on the availability of modern, contemporary and compact mechanical seals designs, It is now possible to position a tandem or double seal architecture typically of the cartridge type into the same sized stuffing box thus tolerating or subjecting the pump in question to a more securer and strenuous operation. Seal configurations includes two or more extremely flat and precision lapped rotating rings or faces to prevent internal intrusions during operation and a set of stationary gaskets, 'O'-rings and 'chevron' seal rings at the mechanical seal interior's furnishings.

The term "Net" refers to the actual head at the pump suction flange and not the static head. The N.P.S.H. is independent of the fluid density as are all head terms.). Accordingly, to prevent a deterioration of the required suction pressure within its casing due to an opening around the pump's shaft connected to the impeller, pump manufactures (by convention of various applicable standards) provides a small

This is to avert leakages to the external environment when the pump is in a primed and or motionless situation. M.Y.Gazee FIDiag.E (9582) PETRONAS REFINERY, MALAYSIA.

Institution of Diagnostic Engineers

Centrifugal Pump Fundamentals

-Goulds Pumps-

t 12/2003

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12/2003

Section TECH-A Centrifugal Pump Fundamentals TECH-A-1 Head The pressure at any point in a liquid can be thought of as being caused by a vertical column of the liquid which, due to its weight, exerts a pressure equal to the pressure at the point in question. The height of this column is called the static head and is expressed in terms of feet of liquid. The static head corresponding to any specific pressure is dependent upon the weight of the liquid according to the following formula.

Where H = Total head developed in feet. v = Velocity at periphery of impeller in feet per sec. g = 32.2 Feet/Sec.2 We can predict the approximate head of any centrifugal pump by calculating the peripheral velocity of the impeller and substituting into the above formula. A handy formula for peripheral velocity is: v = RPM x D 229

Head in Feet = Pressure in psi x 2.31 Specific Gravity A Centrifugal pump imparts velocity to a liquid. This velocity energy is then transformed largely into pressure energy as the liquid leaves the pump. Therefore, the head developed is approximately equal to the velocity energy at the periphery of the impeller This relationship is expressed by the following well-known formula:

Where D = Impeller diameter in inches

The above demonstrates why we must always think in terms of feet of liquid rather than pressure when working with centrifugal pumps. A given pump with a given impeller diameter and speed will raise a liquid to a certain height regardless of the weight of the liquid, as shown in Fig. 1.

2 H= v 2g

100 Ft.

100 Ft.

32.5 psi

100 Ft.

52 psi

43 psi

Gasoline, Sp. Gr. = 0.75

Water, Sp. Gr. = 1.0

Brine, Sp. Gr. = 1.2

Discharge 100' X 0.75 = = 32.5 PSI Pressure 2.31

Discharge 100' X 1.0 = = 43 PSI Pressure 2.31

Discharge 100' X 1.2 = = 52 PSI Pressure 2.31

Fig. 1 Identical Pumps Handling Liquids of Different Specific Gravities. All of the forms of energy involved in a liquid flow system can be expressed in terms of feet of liquid. The total of these various heads determines the total system head or the work which a pump must perform in the system. The various forms of head are defined as follows. SUCTION LIFT exists when the source of supply is below the center line of the pump. Thus the STATIC SUCTION LIFT is the vertical distance in feet from the centerline of the pump to the free level of the liquid to be pumped. SUCTION HEAD exists when the source of supply is above the centerline of the pump. Thus the STATIC SUCTION HEAD is the vertical distance in feet from the centerline of the pump to the free level of the liquid to be pumped.

STATIC DISCHARGE HEAD is the vertical distance in feet between the pump centerline and the point of free discharge or the surface of the liquid in the discharge tank. TOTAL STATIC HEAD is the vertical distance in feet between the free level of the source of supply and the point of free discharge or the free surface of the discharge liquid. The above forms of static head are shown graphically in Fig. 2a & b FRICTION HEAD (hf) is the head required to overcome the resistance to flow in the pipe and fittings. It is dependent upon the size and type of pipe, flow rate, and nature of the liquid. Frictional tables are included in section TECH-C.

TECH-A

VELOCITY HEAD (hv) is the energy of a liquid as a result of its motion at some velocity V. It is the equivalent head in feet through which the water would have to fall to acquire the same velocity, or in other words, the head necessary to accelerate the water. Velocity head can be calculated from the following formula: 2 hv = V 2g

2 where g = 32.2 ft/sec. V = liquid velocity in feet per second

The velocity head is usually insignificant and can be ignored in most high head systems. However, it can be a large factor and must be considered in low head systems. PRESSURE HEAD must be considered when a pumping system either begins or terminates in a tank which is under some pressure other than atmospheric. The pressure in such a tank must first be converted to feet of liquid. A vacuum in the suction tank or a positive pressure in the discharge tank must be added to the system head, whereas a positive pressure in the suction tank or vacuum in the discharge tank would be subtracted. The following is a handy formula for converting inches of mercury vacuum into feet of liquid. Vacuum, ft. of liquid = Vacuum, in. of Hg x 1.13 Sp. Gr. The above forms of head, namely static, friction, velocity, and pressure, are combined to make up the total system head at any particular flow rate. Following are definitions of these combined or “Dynamic” head terms as they apply to the pump.

TOTAL DYNAMIC SUCTION LIFT (hs) is the static suction lift minus the velocity head at the pump suction flange plus the total friction head in the suction line. The total dynamic suction lift, as determined on pump test, is the reading of a gauge on the suction flange, converted to feet of liquid and corrected to the pump centerline*, minus the velocity head at the point of gauge attachment. TOTAL DYNAMIC SUCTION HEAD (hs) is the static suction head plus the velocity head at the pump suction flange minus the total friction head in the suction line. The total dynamic suction head, as determined on pump test, is the reading of the gauge on the suction flange, converted to feet of liquid and corrected to the pump centerline*, plus the velocity head at the point of gauge attachment. TOTAL DYNAMIC DISCHARGE HEAD (hd) is the static discharge head plus the velocity head at the pump discharge flange plus the total friction head in the discharge line. The total dynamic discharge head, as determined on pump test, is the reading of a gauge at the discharge flange, converted to feet of liquid and corrected to the pump centerline*, plus the velocity head at the point of gauge attachment. TOTAL HEAD (H) or TOTAL Dynamic HEAD (TDH) is the total dynamic discharge head minus the total dynamic suction head or plus the total dynamic suction lift.

TDH = hd + hs (with a suction lift) TDH = hd – hs (with a suction head)

STATIC DISCHG HEAD TOTAL STATIC HEAD

STATIC SUCTION LIFT

Fig. 2-a Suction Lift – Showing Static Heads in a Pumping System Where the Pump is Located Above the Suction Tank. (Static Suction Head)

TECH-A

TOTAL STATIC HEAD

STATIC DISCHARGE HEAD

STATIC SUCTION HEAD

Fig. 2-b Suction Head – Showing Static Heads in a Pumping System Where the Pump is Located Below the Suction Tank. (Static Suction Head)

TECH-A-2 Capacity Capacity (Q) is normally expressed in gallons per minute (gpm). Since liquids are essentially incompressible, there is a direct relationship between the capacity in a pipe and the velocity of flow. This relationship is as follows: Q = A x V or V = Q A

Where A = Area of pipe or conduit in square feet. V = Velocity of flow in feet per second. *On vertical pumps the correction should be made to the eye of the suction or lowest impeller.

TECH-A-3 Power and Efficiency The work performed by a pump is a function of the total head and the weight of the liquid pumped in a given time period. The pump capacity in gpm and the liquid specific gravity are normally used in the formulas rather than the actual weight of the liquid pumped. Pump input or brake horsepower (bhp) is the actual horsepower delivered to the pump shaft. Pump output or hydraulic horsepower (whp) is the liquid horsepower delivered by the pump. These two terms are defined by the following formulas. whp = Q x TDH x Sp. Gr. 3960

bhp =

Q x TDH x Sp. Gr. 3960 x Pump Efficiency

The constant 3960 is obtained by dividing the number or foot pounds for one horsepower (33,000) by the weight of one gallon of water (8.33 pounds.) The brake horsepower or input to a pump is greater than the hydraulic horsepower or output due to the mechanical and hydraulic losses incurred in the pump. Therefore the pump efficiency is the ratio of these two values. Pump Eff = whp = Q x TDH x Sp. Gr. bhp 3960 x bhp

TECH-A

TECH-A-4 Specific Speed and Pump Type Specific speed (Ns) is a non-dimensional design index used to classify pump impellers as to their type and proportions. It is defined as the speed in revolutions per minute at which a geometrically similar impeller would operate if it were of such a size as to deliver one gallon per minute against one foot head. The understanding of this definition is of design engineering significance only, however, and specific speed should be thought of only as an index used to predict certain pump characteristics. The following formula is used to determine specific speed:

 Ns = N Q H3/4 Where N = Pump speed in RPM Q = Capacity in gpm at the best efficiency point H = Total head per stage at the best efficiency point

The specific speed determines the general shape or class of the impeller as depicted in Fig. 3. As the specific speed increases, the ratio of the impeller outlet diameter, D2, to the inlet or eye diameter, D1, decreases. This ratio becomes 1.0 for a true axial flow impeller. Radial flow impellers develop head principally through centrifugal force. Pumps of higher specific speeds develop head partly by centrifugal force and partly by axial force. A higher specific speed indicates a pump design with head generation more by axial forces and less by centrifugal forces. An axial flow or propeller pump with a specific speed of 10,000 or greater generates it's head exclusively through axial forces. Radial impellers are generally low flow high head designs whereas axial flow impellers are high flow low head designs.

Values of Specific Speed, Ns

Fig. 3 Impeller Design vs Specific Speed

TECH-A-5 Net Positive Suction Head (NPSH) and Cavitation The Hydraulic Institute defines NPSH as the total suction head in feet absolute, determined at the suction nozzle and corrected to datum, less the vapor pressure of the liquid in feet absolute. Simply stated, it is an analysis of energy conditions on the suction side of a pump to determine if the liquid will vaporize at the lowest pressure point in the pump. The pressure which a liquid exerts on its surroundings is dependent upon its temperature. This pressure, called vapor pressure, is a unique characteristic of every fluid and increases with increasing temperature. When the vapor pressure within the fluid reaches the pressure of the surrounding medium, the fluid begins to vaporize or boil. The temperature at which this vaporization occurs will decrease as the pressure of the surrounding medium decreases. A liquid increases greatly in volume when it vaporizes. One cubic foot of water at room temperature becomes 1700 cu. ft. of vapor at the same temperature. It is obvious from the above that if we are to pump a fluid effectively, we must keep it in liquid form. NPSH is simply a measure of the amount of suction head present to prevent this excess vaporization at the lowest pressure point in the pump.

TECH-A

NPSH Required is a function of the pump design. As the liquid passes from the pump suction to the eye of the impeller, the velocity increases and the pressure decreases. There are also pressure losses due to shock and turbulence as the liquid strikes the impeller. The centrifugal force of the impeller vanes further increases the velocity and decreases the pressure of the liquid. The NPSH Required is the positive head in feet absolute required at the pump suction to overcome these pressure drops in the pump and maintain enough of the liquid above its vapor pressure to limit the head loss, due to the blockage of the cavitation vapor bubble, to 3 percent. The 3% head drop criteria for NPSH Required is used worldwide and is based on the ease of determining the exact head drop off point. Most standard low suction energy pumps can operate with little or no margin above the NPSH Required, without seriously affecting the service life of the pump. The NPSH Required varies with speed and capacity within any particular pump. Pump manufacturer’s curves normally provide this information.

NPSH Available is a function of the system in which the pump operates. It is the excess pressure of the liquid in feet absolute over its vapor pressure as it arrives at the pump suction. Fig. 4 shows four typical suction systems with the NPSH Available formulas applicable to each. It is important to correct for the specific gravity of the liquid and to convert all terms to units of “feet absolute” in using the formulas.

4a SUCTION SUPPLY OPEN TO ATMOSPHERE - with Suction Lift

4b SUCTION SUPPLY OPEN TO ATMOSPHERE - with Suction Head

4c CLOSED SUCTION SUPPLY - with Suction Lift

4d CLOSED SUCTION SUPPLY - with Suction Head

PB = Barometric pressure, in feet absolute.

Ls

VP = Vapor pressure of the liquid at maximum pumping temperature, in feet absolute.

LH = Minimum static suction head in feet.

p

= Pressure on surface of liquid in closed suction tank, in feet absolute.

hf

= Maximum static suction lift in feet.

= Friction loss in feet in suction pipe at required capacity

Fig. 4 Calculation of system Net Positive Suction Head Available for typical suction conditions.

TECH-A

In an existing system, the NPSH Available can be determined by a gauge on the pump suction. The following formula applies: NPSHA= PB – Vp ± Gr + hV Where Gr = Gauge reading at the pump suction expressed in feet (plus if above atmospheric, minus if below atmospheric) corrected to the pump centerline. hv = Velocity head in the suction pipe at the gauge connection, expressed in feet. Cavitation is a term used to describe the phenomenon, which occurs in a pump when there is insufficient NPSH Available. The pressure of the liquid is reduced to a value equal to or below its vapor pressure and small vapor bubbles or pockets begin to form. As these vapor bubbles move along the impeller vanes to a higher pressure area, they rapidly collapse. The collapse, or “implosion” is so rapid that it may be heard as a rumbling noise, as if you were pumping gravel. In high suction energy pumps, the collapses are generally high enough to cause minute

pockets of fatigue failure on the impeller vane surfaces. This action may be progressive, and under severe (very high suction energy) conditions can cause serious pitting damage to the impeller. The accompanying noise is the easiest way to recognize cavitation. Besides possible impeller damage, excessive cavitation results in reduced capacity due to the vapor present in the pump. Also, the head may be reduced and/or be unstable and the power consumption may be erratic. Vibration and mechanical damage such as bearing failure can also occur as a result of operating in excessive cavitation, with high and very high suction energy pumps. The way to prevent the undesirable effects of cavitation in standard low suction energy pumps is to insure that the NPSH Available in the system is greater than the NPSH Required by the pump. High suction energy pumps require an additional NPSH margin, above the NPSH Required. Hydraulic Institute Standard (ANSI/HI 9.6.1) suggests NPSH margin ratios of from 1.2 to 2.5 times the NPSH Required, for high and very high suction energy pumps, when operating in the allowable operating range.

TECH-A-6 NPSH Suction Specific Speed and Suction Energy 1/2 S = N (GPM) (NPSH) 3/4

In designing a pumping system, it is essential to provide adequate NPSH available for proper pump operation. Insufficient NPSH available may seriously restrict pump selection, or even force an expensive system redesign. On the other hand, providing excessive NPSH available may needlessly increase system cost.

1/2 9000 = N (2000) 30 3/4

N = 2580 RPM

Suction specific speed may provide help in this situation. Suction specific speed (S) is defined as: 1/2 S = N (GPM) (NPSHR ) 3/4

Where

N GPM

NPSH

= Pump speed RPM

Running a pump at this speed would require a gear and at this speed, the pump might not develop the required head. At a minimum, existing NPSHA is constraining pump selection. Same system as 1. Is a double suction pump practical? For a double suction pump, flow is divided by two. 1/2 S = N (GPM) (NPSH) 3/4

= Pump flow at best efficiency point at impeller inlet (for double suction impellers divide total pump flow by two).

1/2 9000 = N (1000) (30 )3/4

= Pump NPSH required at best efficiency point.

N = 3700 RPM For a given pump, the suction specific speed is generally a constant - it does not change when the pump speed is changed. Experience has shown that 9000 is a reasonable value of suction specific speed. Pumps with a minimum suction specific speed of 9000 are readily available, and are not normally subject to severe operating restrictions. An example: Flow 2,000 GPM; head 600 ft. What NPSH will be required? Assume: at 600 ft., 3550 RPM operation will be required. 1/2 S = N (GPM) (NPSHR ) 3/4 1/2 9000 = 3550 (2000) (NPSHR ) 3/4

NPSH R 3/4 = 17.7 NPSH R = 46 ft. A related problem is in selecting a new pump, especially at higher flow, for an existing system. Suction specific speed will highlight applications where NPSHA may restrict pump selection. An example: Existing system: Flow 2000 GPM; head 600 ft.: NPSHA 30 ft. What is the maximum speed at which a pump can be run without exceeding NPSH available?

Using a double suction pump is one way of meeting system NPSH. The amount of energy in a pumped fluid, that flashes into vapor and then collapses back to a liquid in the higher pressure area of the impeller inlet, determines the extent of the noise and/or damage from cavitation. Suction Energy is defined as: Suction Energy = De x N x S x Sg Where

De

= Impeller eye diameter (inches)

Sg

= Specific gravity of liquid (Sg - 1.0 for cold water)

High Suction Energy starts at 160 x 106 for end suction pumps and 120 x 106 for horizontal split case pumps. Very high suction energy starts at 1.5 times the High Suction Energy values. For estimating purposes you can normally assume that the impeller eye diameter is approximately 90% of the suction nozzle size, for an end suction pump, and 75% of the suction size for a double suction split case pump. An example: Suction specific speed 9,000, pump speed 3550 RPM, suction nozzle size 6 inch, specific gravity 1.0, and the pump type is end suction. De  .9 x 6" = 5.4" Suction Energy = De x N x S x Sg = 5.4 x 3550 x 9,000 x 1.0 = 173 x 106 Since 173 x 106 > 160 x 106, this is a High Suction Energy pump.

TECH-A

TECH-A-7 Pump Characteristic Curves The performance of a centrifugal pump can be shown graphically on a characteristic curve. A typical characteristic curve shows the total dynamic head, brake horsepower, efficiency, and net positive suction head all plotted over the capacity range of the pump.

pump. The shut-off head is usually 150% to 200% of the design head. The brake horsepower remains fairly constant over the flow range. For a typical axial flow pump, the head and brake horsepower both increase drastically near shutoff as shown in Fig. 7.

Figures 5, 6, & 7 are non-dimensional curves which indicate the general shape of the characteristic curves for the various types of pumps. They show the head, brake horsepower, and efficiency plotted as a percent of their values at the design or best efficiency point of the pump.

The distinction between the above three classes is not absolute, and there are many pumps with characteristics falling somewhere between the three. For instance, the Francis vane impeller would have a characteristic between the radial and mixed flow classes. Most turbine pumps are also in this same range depending upon their specific speeds.

Fig. 5 shows that the head curve for a radial flow pump is relatively flat and that the head decreases gradually as the flow increases. Note that the brake horsepower increases gradually over the flow range with the maximum normally at the point of maximum flow. Mixed flow centrifugal pumps and axial flow or propeller pumps have considerably different characteristics as shown in Figs. 6 and 7. The head curve for a mixed flow pump is steeper than for a radial flow

Fig. 8 shows a typical pump curve as furnished by a manufacturer. It is a composite curve which tells at a glance what the pump will do at a given speed with various impeller diameters from maximum to minimum. Constant horsepower, efficiency, and NPSHR lines are superimposed over the various head curves. It is made up from individual test curves at various diameters.

Fig. 5 Radial Flow Pump

Fig. 6 Mixed Flow Pump

TECH-A

Fig. 7 Axial Flow Pump

Fig. 8 Composite Performance Curve

TECH-A

TECH-A-8 Affinity Laws The affinity laws express the mathematical relationship between the several variables involved in pump performance. They apply to all types of centrifugal and axial flow pumps. They are as follows: 1. With impeller diameter, D, held constant: Where: Q H BHP N

A.

Q1 N = 1 Q2 N2

B.

H1 N1 = H2 N2

C.

BHP1 N1 = BHP2 N2

= = = =

Capacity, GPM Total Head, Feet Brake Horsepower Pump Speed, RPM

EXAMPLE: To illustrate the use of these laws, refer to Fig. 8. It shows the performance of a particular pump at 1750 RPM with various impeller diameters. This performance data has been determined by actual tests by the manufacturer. Now assume that you have a 13" maximum diameter impeller, but you want to belt drive the pump at 2000 RPM. The affinity laws listed under 1 above will be used to determine the new performance, with N1 = 1750 RPM and N2 = 2000 RPM. The first step is to read the capacity, head, and horsepower at several points on the 13” dia. curve in Fig. 9. For example, one point may be near the best efficiency point where the capacity is 300 GPM, the head is 160 ft, and the BHP is approx. 20 hp.

2

( )

3

( )

300 1750 = Q2 2000

Q2 = 343 gpm

2. With speed, N, held constant:

A.

Q1 D = 1 Q2 D2

B.

H1 D1 = H2 D2

C.

160 = H2 20 = BHP2

2

( )

BHP1 D1 = BHP2 D2

1750

2

(2000) 1750

3

( 2000)

H2 = 209 ft.

BHP2 – 30 hp

This will then be the best efficiency point on the new 2000 RPM curve. By performing the same calculations for several other points on the 1750 RPM curve, a new curve can be drawn which will approximate the pump's performance at 2000 RPM, Fig. 9.

3

( )

When the performance (Q1, H1, & BHP1) is known at some particular speed (N1) or diameter (D1), the formulas can be used to estimate the performance (Q2, H2, & BHP2) at some other speed (N2) or diameter (D2). The efficiency remains nearly constant for speed changes and for small changes in impeller diameter.

Trial and error would be required to solve this problem in reverse. In other words, assume you want to determine the speed required to make a rating of 343 GPM at a head of 209 ft. You would begin by selecting a trial speed and applying the affinity laws to convert the desired rating to the corresponding rating at 1750 RPM. When you arrive at the correct speed, 2000 RPM in this case, the corresponding 1750 RPM rating will fall on the 13" diameter curve.

Fig. 9

TECH-A

TECH-A-9 System Curves For a specified impeller diameter and speed, a centrifugal pump has a fixed and predictable performance curve. The point where the pump operates on its curve is dependent upon the characteristics of the system in which it is operating, commonly called the System Head Curve...or, the relationship between flow and hydraulic losses* in a system. This representation is in a graphic form and, since friction losses vary as a square of the flow rate, the system curve is parabolic in shape.

POSITIVE STATIC HEAD The parabolic shape of the system curve is again determined by the friction losses through the system including all bends and valves. But in this case there is a positive static head involved. This static head does not affect the shape of the system curve or its “steepness”, but it does dictate the head of the system curve at zero flow rate. The operating point is at the intersection of the system curve and pump curve. Again, the flow rate can be reduced by throttling the discharge valve.

HEAD

PUMP CURVE

THROTTLED SYSTEM CURVE PUMP CURVE

0

By plotting the system head curve and pump curve together, it can be determined: 1. Where the pump will operate on its curve.

HEAD

FLOW RATE

THROTTLED

2. What changes will occur if the system head curve or the pump performance curve changes.

SYSTEM CURVE

NO STATIC HEAD – ALL FRICTION As the levels in the suction and discharge are the same (Fig. 1), there is no static head and, therefore, the system curve starts at zero flow and zero head and its shape is determined solely from pipeline losses. The point of operation is at the intersection of the system head curve and the pump curve. The flow rate may be reduced by throttling valve.

H 0

FLOW RATE Fig. 2 Positive Suction Head

HEAD

PUMP CURVE

THROTTLED SYSTEM CURVE

0

FLOW RATE Fig. 1 No Static Head - All Friction

TECH-A

* Hydraulic losses in piping systems are composed of pipe friction losses, valves, elbows and other fittings, entrance and exit losses (these to the entrance and exit to and from the pipeline normally at the beginning and end – not the pump) and losses from changes in pipe size by enlargement or reduction in diameter.

NEGATIVE (GRAVITY) HEAD

MOSTLY LIFT- LITTLE FRICTION HEAD

In this illustration, a certain flow rate will occur by gravity head alone. But to obtain higher flows, a pump is required to overcome the pipe friction losses in excess of “H” – the head of the suction above the level of the discharge. In other words, the system curve is plotted exactly as for any other case involving a static head and friction head, except the static head is now negative. The system curve begins at a negative value and shows the limited flow rate obtained by gravity alone. More capacity requires extra work.

The system head curve in this illustration starts at the static head “H” and zero flow. Since the friction losses are relatively small (possibly due to the large diameter pipe), the system curve is “flat”. In this case, the pump is required to overcome the comparatively large static head before it will deliver any flow at all.

PUMP CURVE HEAD

H (NEGATIVE)

H

“FLAT” SYSTEM H PUMP CURVE

FLOW RATE HEAD

Fig. 4 Mostly Lift - Little Friction Head

SYSTEM CURVE

0 FLOW RATE -H

Fig. 3 Negative (Gravity) Head

TECH-A

TECH-A-10 Basic Formulas and Symbols Symbols

Formulas GPM = 0.002 x Lb./Hr. Sp. Gr. GPM =

Lbs./Hr. 500 x Sp. Gr.

GPM = 449 x CFS GPM = 0.7 x BBL /Hr.

GPM = gallons per minute CFS = cubic feet per second Lb. = pounds Hr. = hour BBL = barrel (42 gallons) Sp. Gr. = specific gravity

H = 2.31 x psi Sp. Gr. H = 1.134 x In. Hg. Sp. Gr. 2 hv = V = .0155 V2 2g

H = head in feet psi = pounds per square inch In. Hg. = inches of mercury hv = velocity head in feet V = velocity in feet per second g = 32.16 ft/sec2 (acceleration of gravity)

V = GPM x 0.321 = GPM x 0.409 A (I.D.) 2 BHP = GPM x H x Sp. Gr. = GPM x psi 3960 x Eff. 1715 x Eff. Eff. = GPM x H x Sp. Gr. 3960 x BHP Sp. Gr. =

141.5 131.5 x degrees A.P.I.

A = area in square inches I.D. = inside diameter in inches BHP = brake horsepower Eff. = pump efficiency expressed as a decimal Ns = specific speed N = speed in revolutions per minute

NC = 187.7 f 3 f = PL mEI

Ns = N  GPM H 3/4 2 H = v 2g

v =NxD 229 DEG. C

= (DEG. F - 32) x 5 / 9

DEG. F

= (DEG. C x 5 / 9) + 32

*SEE SECTION TECH-D-8C FOR SLURRY FORMULAS

TECH-A

v = peripheral velocity of an impeller in feet per second D = Impeller in inches Nc = critical speed f = shaft deflection in inches P = total force in lbs. L = bearing span in inches m = constant usually between 48 and 75 for pump shafts E = modules of elasticity, psi – 27 to 30 million for steel

Section TECH-B Pump Application Data TECH-B-1 Corrosion & Materials of Construction Selecting the right pump type and sizing it correctly are critical to the success of any pump application. Equally important is the selection of materials of construction. Choices must be made between metals and/or non-metals for pump components that come into contact with the pumpage. In addition, gaskets and O-ring material selections must be made to assure long leak-free operation of the pump's dynamic and static sealing joints. To assist in proper selection, included in this section is a brief discussion of specific types of corrosion and a general material selection guide.

Corrosion Corrosion is the destructive attack of a metal by chemical or electrachemical reaction with its environment. It is important to understand the various types of corrosion and factors affecting corrosion rate to properly select materials. TYPES OF CORROSION (1) Galvanic corrosion is the electro-chemical action produced when one metal is in electrical contact with another more noble metal, with both being immersed in the same corroding medium called the electrolyte. A galvanic cell is formed and current flows between the two materials. The least noble material called the anode will corrode while the more noble cathode will be protected. It is important that the smaller wearing parts in a pump be of a more noble material than the larger more massive parts, as in an iron pump with bronze or stainless steel trim. Following is a galvanic series listing the more common metals and alloys. Corroded End (Anodic, or least noble) Magnesium Magnesium Alloys Zinc Aluminum 2S Cadmium Aluminum 175T Steel or Iron Cast Iron Stainless Steel, 400 Series (Active) Stainless Steel, Type 304 (Active) Stainless Steel, Type 316 (Active) Lead-tin Solders Lead Tin Nickel (Active)

Nickel base alloy (active) Brasses Copper Bronzes Copper-Nickel Alloy Monel Silver Solder Nickel (Passive) Nickel Base Alloy (Passive) Stainless Steel, 400 Series (Passive) Stainless Steel, Type 304 (Passive) Stainless Steel, Type 316 (Passive) Silver Graphite Gold Platinum Protected End (Cathodic, or most noble)

(2) Uniform Corrosion is the overall attack on a metal by a corroding liquid resulting in a relatively uniform metal loss over the exposed surface. This is the most common type of corrosion and it can be minimized by the selection of a material which offers resistance to the corroding liquid. (3) Intergranular corrosion is the precipitation of chromium carbides at the grain boundaries of stainless steels. It results in the complete destruction of the mechanical properties of the steel for the depth of the attack. Solution annealing or the use of extra low carbon stainless steels will eliminate intergranular corrosion. (4) Pitting Corrosion is a localized rather than uniform type of attack. It is caused by a breakdown of the protective film and results in rapid pit formation at random locations on the surface. (5) Crevice or Concentration Cell Corrosion occurs in joints or small surface imperfections. Portions of the liquid become trapped and a difference in potential is established due to the oxygen concentration difference in these cells. The resulting corrosion may progress rapidly leaving the surrounding area unaffected. (6) Stress Corrosion is the failure of a material due to a combination of stress and corrosive environment, whereas the material would not be affected by the environment alone. (7) Erosion-Corrosion is the corrosion resulting when a metal’s protective film is destroyed by high velocity fluids. It is distinguished from abrasion which is destruction by fluids containing abrasive solid particles. pH VALUES The pH of a liquid is an indication of its corrosive qualities, either acidic or alkaline. It is a measure of the hydrogen or hydroxide ion concentration in gram equivalents per liter. pH value is expressed as the logarithm to the base 10 of the reciprocal of the hydrogen ion concentration. The scale of pH values is from zero to 14, with 7 as a neutral point. From 6 to zero denotes increasing hydrogen ion concentration and thus increasing acidity, and from 8 to 14 denotes increasing hydroxide ion concentration and thus increasing alkalinity. The table below outlines materials of construction usually recommended for pumps handling liquids of known pH value

pH Value

Material of Construction

10 to 14

Corrosion Resistant Alloys

8 to 10 6 to 8 4 to 6

Iron, Stainless Steel, Bronze, Carbon Steel

0 to 4

Corrosion Resistant Alloys

The pH value should only be used as a guide with weak aqueous solutions. For more corrosive solutions, temperature and chemical composition should be carefully evaluated in the selection of materials of construction.

TECH-B

TECH-B-2 Material Selection Chart This chart is intended as a guide in the selection of economical materials. It must be kept in mind that corrosion rates may vary widely with temperature, concentration, and the presence of trace elements or abrasive solids. Blank spaces in the chart indicate a lack of accurate corrosion data for those specific conditions. In general, the chart is limited to metals and non-metals regularly furnished by ITT-Goulds. Note: Maximum temperature limits are shown where data is available. Contact a Goulds representative for temperature limits of all materials before final material selection.

Corrosive

Code: A B X Steel

Recommended Useful resistance Unsuitable Carbon steel, cast iron and ductile iron Brz Bronze 316 Stainless steel A-20 Carpenter stainless CD4MCu CD4MCu stainless steel Alloy 2205 Alloy 2205 stainless steel C-276 Wrought Hastelloy® C-276 alloy Ti Titanium unalloyed Zi Zirconium ETFE Ethylenetetrafluoroethylene (Tefzel ®) FP Fluoropolymers (e.g.,

Steel Brz

316 A-20 CD4MCu

Teflon®) including perfluoroalkoxy (PFA), polytetrafluoroethylene (PTFE) and fluorinated ethylene propylene (FEP) FRP Fiber-reinforced plastic (vinylester resin) EPDM Ethylenepropylene rubber (Nordel ®) FKM1 Standard grades; dipolymers of hexafluoropropylene (HFP) and vinylidene fluoride (VF2) (Viton®) FKM2 Specialty grades; terpolymers comprising at least three of the following: HFP, VF2, tetrafluorethylene (TFE), perfluoromethylvinyl ether

ALLOY 2205 C-276

FFKM PVDF

(PMVE) or ethylene (E). Specialty grades may have significantly improved chemical compatibility compared to standard grades in many harsh chemical environments (Viton®). Copolymer of TFE and PMVE (Kalrez®) Polyvinylidene fluoride (Kynar ®, Solef ®)

1Compatibility

is dependent on specific freon. Contact elastomer manufacturer.

Ti

Zi

ETFE

FP

Acetaldehyde, 70°F

B

A

A

A

A

A

A

A

A

A

A

FRP EPDM FKM1 FKM2 FFKM PVDF X

A

X

X

A

X

Acetic acid, 70°F

X

A

A

A

A

A

A

A

A

A

A

X

A

X

B

A

A

Acetic acid, <50%, to boiling

X

B

A

A

B

A

A

A

A

A

X

B

A

B

Acetic acid, >50%, to boiling

X

X

B

A

X

A

A

A

A

104°C

A

X

B

X

B

A

X

A

X

Acetone, to boiling

A

A

A

A

A

A

A

A

104°C

A

X

A

X

X

A

Aluminum chloride, <10%, 70°F

X

B

X

B

X

B

B

A

A

A

A

A

A

A

A

Aluminum chloride, >10%, 70°F

X

X

X

B

X

B

B

A

A

A

A

A

A

A

A

Aluminum chloride, <10%, to boiling

X

X

X

X

X

X

X

A

104°C

A

X

A

A

A

A

A

Aluminum chloride, >10%, to boiling

X

X

X

X

X

X

X

A

104°C

A

X

A

A

A

A

A (to 40°C)

A

X

Aluminum sulphate, 70°F

X

B

A

A

A

A

B

A

A

A

A

Aluminum sulphate, <10%, to boiling

X

B

B

A

B

A

A

A

A

104°C

A

Aluminum sulphate, >10%, to boiling

X

X

X

B

X

B

B

X

B

104°C

A

Ammonium chloride, 70°F

X

X

B

B

B

B

A

A

A

A

A

Ammonium chloride, <10%, to boiling

X

X

B

B

X

B

A

A

A

104°C

Ammonium chloride, >10%, to boiling

X

X

X

X

X

X

X

X

X

104°C

Ammonium fluosilicate, 70°F

X

X

X

B

X

B

X

X

X

Ammonium sulphate, <40%, to boiling

X

X

B

B

X

B

B

A

A

Arsenic acid, to 225°F

X

X

X

B

X

B

Barium chloride, 70°F <30%

X

B

X

B

X

B

B

B

Barium chloride, <5%, to boiling

X

B

X

B

X

B

B

Barium chloride, >5%, to boiling

X

X

X

X

X

X

X

Barium hydroxide, 70°F

B

X

A

A

A

A

B

Barium nitrate, to boiling

X

X

B

B

B

B

Barium sulphide, 70°F

X

X

B

B

B

B

Benzoic acid

X

X

B

B

B

B

A

A

A

A

A

A

A

A

A

A

A

A

A

A

A

A

A

A

A

A

A

A

A

A

A

A

A

A

A

A

A

A

A

A 104°C

A

A

X

B

A

A

A

A

A

A

A

A

B

A

A

A

A

A

A

A

A

A

104°C

A

A

A

A

A

A

X

X

104°C

A

A

A

A

A

A

A

A

A

A

A

A

A

A

A

B

B

104°C

A

A

A

A

A

A

A

A

A

A

B

A

A

A

A

A

A

A

X

A

A

A

110°C

B

B

104°C

A

A

A

A

X

X

B

B

B

B

A

Boron trichloride, 70°F dry

B

B

B

B

B

B

B

Boron trifluoride, 70°F 10%, dry

B

B

B

A

B

A

A

Brine (acid), 70°F

X

X

X

X

X

X

B

B

Bromine (dry), 70°F

X

X

X

X

X

X

B

X

Bromine (wet), 70°F

X

X

X

X

X

X

B

Calcium bisulphite, 70°F

X

X

B

B

B

B

B

A

A

A

Boric acid, to boiling

A

A

A

A

X

X

X

B

A

A

A

A

A

A

A

A

A

X

A

A

X

X

A

A

A

A

X

X

A

A

X

X

B

A

A

A

A

A

A

A

X

A

A

A

A 95°C

Calcium bisulphite

X

X

X

B

X

B

X

A

A

A

A

Calcium chloride, 70°F

B

X

B

B

B

B

A

A

A

A

A

Calcium chloride <5%, to boiling

X

X

B

B

B

B

A

A

A

104°C

A

Calcium chloride >5%, to boiling

X

X

X

B

X

B

A

B

B

104°C

A

Calcium hydroxide, 70°F

B

B

B

B

B

B

A

A

A

A

Calcium hydroxide, <30%, to boiling

X

B

B

B

B

B

A

A

104°C

Calcium hydroxide, >30%, to boiling

X

X

X

X

X

X

B

A

104°C

TECH-B

A (to 40°C)

X

A

A

A

A

A

A

A

A

A

A

A

A

A

A

B

A

A

A

A

A

A

A

A

A

A

B

A

A

A

A

A

B

A

A

A

A

A

Corrosive

Steel Brz

316 A-20 CD4MCu

ALLOY 2205 C-276

Ti

Zi

ETFE

FP

FRP EPDM FKM1 FKM2 FFKM PVDF

Calcium hypochlorite, <2%, 70°F

X

X

X

X

X

X

A

A

A

A

A

X

B

A

A

A

Calcium hypochlorite, >2%, 70°F

X

X

X

X

X

X

B

A

B

A

A

X

B

A

A

A

A

Carbolic acid, 70°F (phenol)

X

B

A

A

A

A

A

A

A

A

A

B

B

A

A

50°C

Carbon bisulphide, 70°F

B

B

A

A

A

A

A

A

X

A

A

A

Carbonic acid, 70°F

B

X

A

A

A

A

A

A

A

A

A

A

A

A

A

A

Carbon tetrachloride, dry to boiling

B

B

A

A

A

A

B

A

A

X

B

A

A

A

A

A

A

A

A

A

A

A

Chloric acid, 70°F

X

X

X

B

X

B

X

Chlorinated water, 70°F

X

X

B

B

B

B

A

A

X

104°C 149°C A

A

A

A

A

Chloroacetic acid, 70°F

X

X

X

A

B

A

A

A

A

X

B

A

X

Chlorosulphonic acid, 70°F

X

X

X

X

X

X

A

B

X

A

A

A

X

X

X

A

X

Chromic acid, <30%

X

X

X

B

X

B

B

A

A

65°C

A

X

X

A

A

A

80°C

Citric acid

X

X

A

A

A

A

A

A

A

A

A

A

A

A

A

A

A

Copper nitrate, to 175°F

X

X

B

B

B

B

X

B

A

A

A

A

A

Copper sulphate, to boiling

X

X

X

B

X

B

A

A

104°C

A

A

A

A

A

A

Cresylic acid

X

X

B

B

B

B

B

A

A

X

A

A

A

65°C

Cupric chloride

X

X

X

A

Cyanohydrin, 70°F

X

X

X

X

X

B

B

B

B

B

A X

A

A

A

A

A

Dichloroethane

X

B

B

B

B

B

B

A

B

65°C

A

Diethylene glycol, 70°F

A

B

A

A

A

A

B

A

A

A

A

Dinitrochlorobenzene, 70°F (dry)

X

B

A

A

A

A

A

A

A

A

A

Ethanolamine, 70°F

B

X

B

B

B

B

A

A

A

A

B

Ethers, 70°F

B

B

B

A

A

A

B

A

A

A

A

Ethyl alcohol, to boiling

A

A

A

A

A

A

A

A

A

104°C

A

Ethyl cellulose, 70°F

A

B

B

B

B

B

B

A

A

A

A

Ethyl chloride, 70°F

X

B

B

A

B

A

B

A

A

A

A

X

Ethyl mercaptan, 70°F

X

X

B

A

B

A

B

A

A

X

Ethyl sulphate, 70°F

X

B

B

A

B

A

A

A

X

Ethylene chlorohydrin, 70°F

X

B

B

B

B

B

B

A

A

A

A

X

Ethylene dichloride, 70°F

X

B

B

B

B

B

X

A

A

A

A

X

Ethylene glycol, 70°F

B

B

B

B

B

B

A

A

A

A

A

A

Ethylene oxide, 70°F

X

X

B

B

B

B

A

A

A

A

A

Ferric chloride, <5%, 70°F

X

X

X

X

X

X

A

A

B

A

A

Ferric chloride, >5%, 70°F

X

X

X

X

X

X

B

B

X

A

Ferric nitrate, 70°F

X

X

B

A

B

A

B

Ferric sulphate, 70°F

X

X

X

B

X

B

B

B

B

A

A

A

X

X

A

B

A

A

A

B

A

A

A

X

B

A

A

X

X

A

X

X

X

X

A

B

A

B

A

A

A

B

X

X

A

X

B

A

A

A

X

B

B

A

A

X

X

A

B

A

A

A

A

X

A

A

A

A

A

A

A

A

A

X

X

X

A

A

A

A

A

A

A

A

A

X

A

A

A

A

A

A

A

A

A

A

A

A

A

A

A

A

A

A

A

A

A

A

A

A

A

A

A

X

B

A

X

A

Ferrous sulphate, 70°F

X

X

X

B

X

B

B

A

A

A

A

Formaldehyde, to boiling

B

B

A

A

A

A

B

A

A

104°C

A

Formic acid, to 212°F

X

X

X

A

B

A

A

X

A

A

A

A

X

X

A

A

Freon, 70°F

A

A

A

A

A

A

A

A

A

A

A

A/X1

A/X1

A/X1

A/B1

A

Hydrochloride acid, <1%, 70°F

X

X

X

B

X

B

A

B

A

A

A

A

A

A

A

A

A

Hydrochloric acid, 1% to 20%, 70°F

X

X

X

X

X

X

A

X

A

A

A

A

A

A

A

A

A

Hydrochloric acid, >20%, 70°F

X

X

X

X

X

A

X

B

A

A

X

A

B

A

A

A

Hydrochloric acid, <1/2%, 175°F

X

X

X

X

X

A

X

A

A

A

X

X

B

A

A

A

Hydrochloric acid, 1/2% to 2%, 175°F

X

X

X

A

X

A

A

A

X

X

B

A

A

A

A

X X

Hydrocyanic acid, 70°F

X

X

X

B

X

B

X

A

A

A

A

A

A

A

Hydrogen peroxide, <30%, <150°F

X

X

B

B

B

B

B

A

A

A

A

B

B

A

A

A

Hydrofluoric acid, <20%, 70°F

X

B

X

B

X

B

A

X

X

A

A

X

B

A

A

A

Hydrofluoric acid, >20%, 50°F

X

X

X

X

X

X

B

X

X

A

A

X

B

A

A

A

Hydrofluoric acid, to boiling

X

X

X

X

X

X

X

X

X

X

X

B

A

B

Hydrofluorsilicic acid, 70°F

X

X

B

X

B

B

Lactic acid, <50%, 70°F

X

B

A

A

A

A

B

A

Lactic acid, >50%, 70°F

X

B

B

B

B

B

B

A

Lactic acid, <5%, to boiling

X

X

X

B

X

B

B

A

A

A

B

A

A

A

A

A

A

A

A

A

A

A

A

A

A

A

A

A

A

A

A

A

104°C

A

X

B

A

A

50°C

Lime slurries, 70°F

B

B

B

B

A

B

B

B

B

Magnesium chloride, 70°F

X

X

B

A

B

A

A

A

A

A

A

A

A

A

A

A

A

A

A

Magnesium chloride, <5%, to boiling

X

X

X

B

X

B

A

A

A

104°C

A

A

A

A

A

A

140°C

Magnesium chloride, >5%, to boiling

X

X

X

X

X

X

B

B

B

104°C

A

A

A

A

A

140°C

A

TECH-B

Corrosive

Steel Brz

316 A-20 CD4MCu

ALLOY 2205 C-276

Ti

Magnesium hydroxide, 70°F

B

A

B

B

A

B

B

A

Magnesium sulphate

X

X

B

A

B

A

X

B

Maleic acid

X

X

B

B

B

B

B

A

Mercaptans

A

X

A

A

A

A

Mercuric chloride, <2%, 70°F

X

X

X

X

X

X

B

Zi

B

ETFE

FP

FRP EPDM FKM1 FKM2 FFKM PVDF

A

A

A

A

A

A

A

A

A

A

A

A

A

A

135°C

A

A

B

A

A

A

120°C

A

A

X

B

A

A

A

A

A

A

A

A

A

A

A

A

A

X

A

A

A

A

A

X

A

A

A

A

A

A

A

A

A

A

A

A

A

A

X

X

B

A

70%, 50°C A

A

A

A A

A

A

A

A

A

Mercurous nitrate, 70°F

X

X

B

B

B

B

C

Methyl alcohol, 70°F

A

A

A

A

A

A

A

Naphthalene sulphonic acid, 70°F

X

X

B

B

B

B

B

A

A

Napthalenic acid

X

X

B

B

B

B

B

A

A

Nickel chloride, 70°F

X

X

X

B

X

B

B

A

A

Nickel sulphate

X

X

B

B

B

B

A

A

A

Nitric acid

X

X

B

B

B

B

Nitrobenzene, 70°F

A

X

A

A

A

A

B

A

A

A

X

A

B

A

A

Nitroethane, 70°F

A

A

A

A

A

A

A

A

A

A

A

X

B

X

X

A

A

Nitropropane, 70°F

A

A

A

A

A

A

A

A

A

A

A

X

X

X

X

A

B

A

B B B

Nitrous acid, 70°F

X

X

X

X

X

X

Nitrous oxide, 70°F

X

X

X

X

X

X

X

Oleic acid

X

X

B

B

B

B

X

X

Oleum acid, 70°F

B

X

B

B

B

B

B

B

Oxalic acid

X

X

X

B

X

B

B

X

Palmitic acid

B

B

B

A

B

A

B

X A

Phenol (see carbolic acid) Phosgene, 70°F

X

X

B

B

B

B

B

Phosphoric acid, <10%, 70°F

X

X

A

A

A

A

A

A

A

A

A

A

A

A

X

X

A

B

B

A

A

A

X

X

B

B

B

A

120°C

A

A

A

A

X

X

B

A

A

X

X

A

A

A

A

50°C

A A

A

B

A

A

A

120°C

A

B

B

A

A

A

A

50°C

X

X

A

A

A

A

A

A

A

A

A

A

Phosphoric acid, >10% to 70%, 70°F

X

X

A

A

A

A

X

B

B

A

A

X

A

A

A

A

A

Phosphoric acid, <20%, 175°F

X

X

B

B

B

B

A

X

B

A

A

X

A

A

A

A

A

A

A

X

A

A

A

A

A

A

A

A

A

A A

Phosphoric acid, >20%, 175°F, <85%

X

X

X

B

X

B

X

X

X

Phosphoric acid, >10%, boil, <85%

X

X

X

X

X

X

X

X

X

Phthalic acid, 70°F

X

B

B

A

B

A

B

A

A

Phthalic anhydride, 70°F

B

X

A

A

A

A

A

Picric acid, 70°F

X

X

X

B

X

B

B

Potassium carbonate

B

B

A

A

A

A

B

A

A

A

A

B

B

A

A

A

B

B

A

A

A

A

A

B A

A

A

A

A

A

A

A

140°C

Potassium chlorate

B

X

A

A

A

A

B

A

A

A

A

A

A

A

A

95°C

Potassium chloride, 70°F

X

X

B

A

B

A

B

A

A

A

A

A

A

A

A

A

A

Potassium cyanide, 70°F

B

X

B

B

B

B

B

A

A

A

A

A

A

A

A

Potassium dichromate

B

B

A

A

A

A

B

A

A

A

A

A

A

A

140°C

A

Potassium ferricyanide

X

B

B

B

B

B

B

Potassium ferrocyanide, 70°F

X

B

B

B

B

B

B

Potassium hydroxide, 70°F

X

X

B

A

B

A

X

B

Potassium hypochlorite

X

X

X

B

X

B

B

A

Potassium iodide, 70°F

X

B

B

B

B

B

B

A

Potassium permanganate

B

B

B

B

B

B

B

Potassium phosphate

X

X

B

B

B

B

B

Seawater, 70°F

X

B

B

A

B

A

A

A

Sodium bisulphate, 70°F

X

X

X

B

X

B

B

Sodium bromide, 70°F

B

X

B

B

B

B

B

Sodium carbonate

B

B

B

A

B

A

Sodium chloride, 70°F

X

B

B

B

B

B

Sodium cyanide

B

X

B

B

B

B

B B

A A

A

A

B

A

A

A

A

A

A

A

A

A

A

A

A

A

A

B

A

A

A

A

A

A

A

B

A

A

A

A

B

A

A

A

B

A

A

A

B

B

A

140°C

B

B

A

A

X

B

A

X

X

X

A

95°C

A

A

A

A

B

B

A

120°C

A

A

A

A

A

A

A

A

A

A

A

A

A

A

A

A

A

A

A

A

A

A

A

140°C

A

A

A

A

A

A

A

A

A

A

A

A

A

A

135°C

100°C

A

A

A

A

95°C

A

A

Sodium dichromate

B

X

B

B

B

B

Sodium ethylate

B

A

A

A

A

A

Sodium fluoride

X

X

B

B

B

B

X

B

B

A

A

Sodium hydroxide, 70°F

B

B

B

A

B

A

A

A

A

A

A

A

Sodium hypochlorite

X

X

X

X

X

X

B

A

B

A

A

X

Sodium lactate, 70°F

B

X

X

X

X

X

X

A

A

A

TECH-B

A

A

A

A 140°C

A

A

A

A

B

A

A

X

B

B

A

A

40%, 95°C

A

A

Corrosive

Steel Brz

316 A-20 CD4MCu

ALLOY 2205 C-276

Ti

Zi

ETFE

FP

Stannic chloride, <5%, 70°F

X

X

X

X

X

X

B

A

A

A

A

Stannic chloride, >5%, 70°F

X

X

X

X

X

X

X

B

B

A

A

Sulphite liquors, to 175°F

X

X

B

B

B

B

B

Sulphur (molten)

B

X

A

A

A

A

FRP EPDM FKM1 FKM2 FFKM PVDF A

A

A

A

A

A

A

A

A

A

A

A

B

A

A

A

A

A

A

A

A

A

A

120°C A

Sulphur dioxide (spray), 70°F

X

X

B

B

B

B

B

X

A

A

A

A

A

A

Sulphuric acid, <2%, 70°F

X

X

B

A

B

A

A

B

A

A

A

A

A

A

A

A

A

Sulphuric acid, 2%t o 40%, 70°F

X

X

X

B

X

B

A

X

A

A

A

A

B

A

A

A

A

Sulphuric acid, 40%, <90%, 70°F

X

X

X

B

X

B

A

X

X

A

A

X

B

B

A

A

A

Sulphuric acid, 93% to 98%, 70°F

B

X

B

B

B

B

A

X

X

A

A

X

X

B

A

A

A

Sulphuric acid, <10%, 175°F

X

X

X

B

X

B

A

X

B

A

A

A

X

A

A

A

A

Sulphuric acid, 10% to 60% & >80%, 175°F

X

X

X

B

X

B

X

X

X

A

A

X

B

A

A

A

A

Sulphuric acid, 60% to 80%, 175°F

X

X

X

X

X

X

B

X

X

A

A

X

X

B

A

A

A

Sulphuric acid, <3/4%, boiling

X

X

X

B

X

B

A

X

B

X

B

A

A

120°C

Sulphuric acid, 3/4% to 40%, boiling

X

X

X

X

X

X

X

X

B

X

B

A

A

120°C

Sulphuric acid, 40% to 65% & >85%, boiling

X

X

X

X

X

X

X

X

X

X

X

B

A

B

Sulphuric acid, 65% to 85%, boiling

X

X

X

X

X

X

X

X

X

X

X

B

A

95°C

Sulphurous acid, 70°F

X

X

X

B

X

B

B

A

B

X

X

B

A

A

Titanium tetrachloride, 70°F

X

X

B

X

B

X

X

B

B

B

A

Tirchlorethylene, to boiling

B

B

B

B

B

B

A

A

X

B

A

A

A

B

B

X

Urea, 70°F

X

X

B

B

B

B

X

Vinyl acetate

B

B

B

B

B

B

B

Vinyl chloride

B

X

B

B

B

B

B

A

Water, to boiling

B

A

A

A

A

A

A

A

A

Zinc chloride

X

X

B

A

B

A

A

A B

Zinc cyanide, 70°F

X

B

B

B

B

B

B

B

Zinc sulphate

X

X

A

A

A

A

X

A

X

A

A

A

A

A

A

A

B

A

A

A

A

A

B

B

A

120°C

A

A

X

A

A

A

95°C

A

A

A

A

A

A

A

A

A

A

140°C

A

A

A

A

A

A

A

140°C

A

A

A

A

A

A

Elastomer Selection Guide Please use the following chart as a general guide only. Refer to detailed selection tables or the factory for specific elastomer recommendations.

Elastomer

Shore (A) Hardness

Max Temp Limit

pH Range

Abrasion

Resistance to Moderate Chemicals

Oils Hydrocarbons

Natural Rubber

40

154 F

5 - 12

E

G (1)

P

Polyurethane

81

149 F

3 - 11

E (2)

G (1)

E

Neoprene

60

212 F

3 - 12

G

G (1)

G

Nitrile

60

220 F

4 - 12

G

G (1)

E

Hypalon

55

230 F

1 - 14

G

E

G

Chlorobutyl

50

300 F

3 - 12

G

E

P

(1) Poor for oxidizing chemicals and strong acids. (2) Fine particles only (200 mesh or less). E = Excellent G = Good P = Poor

TECH-B

CHECK VALVE

ECCENTRIC REDUCER

GATE VALVE

LONG RADIUS ELBOW

(1a) CORRECT

FOOT VALVE (IF USED) STRAINER

CHECK VALVE

ECCENTRIC REDUCER

LONG RADIUS ELBOW GATE VALVE

SUCTION PIPE SLOPES UPWARDS FROM SOURCE OF SUPPLY

(1b) CORRECT

FOOT VALVE (IF USED) STRAINER

AIR POCKET BECAUSE ECCENTRIC REDUCER IS NOT USED AND BECAUSE SUCTION PIPE DOES NOT SLOPE GRADUALLY UPWARD FROM SUPPLY

GATE VALVE

GATE VALVE SHOULD NOT BE BETWEEN CHECK VALVE AND PUMP

(1c) WRONG

Fig. 1 Air Pockets in Suction Piping

TECH-B

CHECK VALVE

TECH-B-3 Piping Design The design of a piping system can have an important effect on the successful operation of a centrifugal pump. Such items as sump design, suction piping design, suction and discharge pipe size, and pipe supports must all be carefully considered.

liquid from evenly filling the impeller. This upsets hydraulic balance leading to noise vibration, possible cavitation, and excessive shaft deflection. Cavitation erosion damage, shaft breakage or premature bearing failure may result.

Selection of the discharge pipe size is primarily a matter of economics. The cost of the various pipe sizes must be compared to the pump size and power cost required to overcome the resulting friction head.

On pump installations involving suction lift, air pockets in the suction line can be a source of trouble. The Suction pipe should be exactly horizontal, or with a uniform slope upward from the sump to the pump as shown in Fig. 1. There should be no high spots where air can collect and cause the pump to lose its prime. Eccentric rather than concentric reducers should always be used.

The suction piping size and design is far more important. Many centrifugal pump troubles are caused by poor suction conditions. The Suction pipe should never be smaller than the suction connection of the pump, and in most cases should be at least one size larger. Suction pipes should be as short and as straight as possible. Suction pipe velocities should be in the 5 to 8 feet per second range unless suction conditions are unusually good.

LEAST 5D

ECCENTRIC REDUCER-WITH TOP HORIZONTAL

MUST BE AT

Higher velocities will increase the friction loss and can result in troublesome air or vapor separation. This is further complicated when elbows or tees are located adjacent to the pump suction nozzle, in that uneven flow patterns or vapor separation keeps the

If an elbow is required at the suction of a double suction pump, it should be in a vertical position if at all possible. Where it is necessary for some reason to use a horizontal elbow, it should be a long radius elbow and there should be a minimum of five diameters of straight pipe between the elbow and the pump as shown in Fig. 2. Fig. 3 shows the effect of an elbow directly on the suction. The liquid will flow toward the outside of the elbow and result in an uneven flow distribution into the two inlets of the double suction impeller. Noise and excessive axial thrust will result.

ELBOW MUST BE VERTICAL WHEN NEXT TO PUMP

(2a) PERMISSABLE

(2b) WRONG

Fig. 2 Elbows At Pump Suction

Fig. 3 Effect of Elbow Directly on Suction There are several important considerations in the design of a suction supply tank or sump. It is imperative that the amount of turbulence and entrained air be kept to a minimum. Entrained air may cause reduced capacity and efficiency as well as vibration, noise, shaft breakage, loss of prime, and/or accelerated corrosion.

The free discharge of liquid above the surface of the supply tank at or near the pump suction can cause entrained air to enter the pump. All lines should be submerged in the tank, and baffles should be used in extreme cases as shown in Fig. 4.

TECH-B

PUMP SUCTION

RECOMMENDED

RECOMMENDED

PUMP SUCTION

BAFFLE

RECOMMENDED

PUMP SUCTION

Fig. 4 Keeping Air Out of Pump

For horizontal pumps, Fig. 5 can be used as a guide for minimum submergence and sump dimensions for flows up to approximately 5000 gpm. Baffles can be used to help prevent vortexing in cases where it is impractical or impossible to maintain the required submergence. Fig. 6 shows three such baffling arrangements. On horizontal pumps, a bell should be used on the end of the suction pipe to limit the entrance velocity to 5.5 feet per second. Also, a reducer at the pump suction flange to smoothly accelerate and stabilize the flow into the pump is desirable.

16

H-SUBMERGENCE IN FEET (MIN.)

14

The submergence of the suction pipe must also be carefully considered. The amount of submergence required depends upon the size and capacity of the individual pumps as well as on the sump design. Past experience is the best guide for determining the submergence. The pump manufacturer should be consulted for recommendations in the absence of other reliable data. 16

14 H-SUBMERGENCE IN FEET (MIN.)

Improper submergence of the pump suction line can cause a vortex, which is a swirling funnel of air from the surface directly into the pump suction pipe. In addition to submergence, the location of the pipe in the sump and the actual dimensions of the sump are also important in preventing vortexing and/or excess turbulence.

12

10

8

6

4

5,000 GPM 3,000 GPM

12

1,000 GPM 2

10

200 GPM 2

4

6

8

10

12

14

16

8 VELOCITY IN FEET PER SEC. =

6

4

QUAN. (G.P.M.) x .321

5,000 GPM 3,000 GPM 1,000 GPM

2 200 GPM 2

4

6

VELOCITY IN FEET PER SEC. =

8

10

QUAN. (G.P.M.) x .321 AREA (inches)2

12

OR

14

16

G.P.M. x .4085 D2

Fig. 5 Minimum Suction Pipe Submergence and Sump Dimensions

TECH-B

AREA (inches)2

OR

G.P.M. x .4085 D2

FLAT BAFFLE

SIDE VIEW

BAFFLE SMOOTHS OUT VORTEX SUCTION PIPE

SUCTION PIPE

TOP VIEW

(6a)

(6b)

(6c)

Fig. 6 Baffle Arrangements for Vortex Prevention For larger units (over 5000 GPM) taking their suction supply for an intake sump (especially vertically submerged pumps), requires special attention. The following section (intake System Design) addresses these larger pumps. INTAKE SYSTEM DESIGN The function of the intake structure, whether it be an open channel, a fully wetted tunnel, a sump, or a tank, is to supply an evenly distributed flow to the pump suction. An uneven distribution of flow, characterized by strong local currents, can result in formation of surface or submerged vortices and with certain low values of submergence, may introduce air into the pump, causing a reduction of capacity, an increase in vibration and additional noise. Uneven flow distribution can also increase or decrease the power consumption with a change in total developed head. The ideal approach is a straight channel coming directly to the pump or suction pipe. Turns and obstructions are detrimental, since they may cause eddy currents and tend to initiate deep-cored vortices. The amount of submergence available is only one factor affecting vortex-free operation. It is possible to have adequate submergence and still have submerged vortices that may have an adverse effect on pump operation. Successful, vortex-free operation will depend greatly on the approach upstream of the sump. Complete analysis of intake structures can only be accurately accomplished by scale model tests. Model testing is especially recommended for larger pumping units.

All of the dimensions In Figures 7 through 10 are based on the rated capacity of the pump. If operation at an increased capacity is to be undertaken for extended periods of time, the maximum capacity should be used for obtaining sump dimensions. If the position of the back wall is determined structurally, dimension B in Figures 7 to 10 may become excessive and a false back wall should be installed. Dimension S in Figures 7 and 9 is a minimum value based on the normal low water level at the pump or suction pipe bell, taking into consideration friction losses through the inlet screen and approach channel. Note that this dimension represents submergence at the intake, or the physical height of the water level above the intake relating to the prevention of eddy formations and vortexing. The channel floor should be level for at least a distance Y (see Figures 7 through 10) upstream before any slope begins. The screen or gate widths should not be substantially less than W, and heights should not be less than the maximum anticipated water level to avoid overflow. Depending on the approach conditions before the sump, it may be necessary to construct straightening vanes in the approach channel, increase dimension A and/or conduct an intake model test to work out some other combination of these factors. Dimension W is the width of an individual pump cell or the center-tocenter distance of two pumps if no dividing wall is used. On multiple intake installations, the recommended dimensions in Figures 7 and 8 apply as noted above, and the following additional factors should be considered.

GENERAL DATA INFORMATION Subject to the qualifications of the foregoing statements, Figures 7 through 10 have been constructed for single and multiple intake arrangements to provide guidelines for basic sump dimensions. Since these values are composite averages for many pump types and cover the entire range of specific speeds, they are not absolute values but typical values subject to variations.

Reprinted from Hydraulic Institute Standard

TECH-B

As shown in Fig. 10 (A), low velocity and straight in-line flow to all units simultaneously is a primary recommendation. Velocities in the sump should be approximately one foot per second, but velocities of two feet per second may prove satisfactory. This is particularly true when the design is based on a model study. Not recommended would be an abrupt change in the size of the inlet pipe to the sump or the inlet from one side introducing eddying. In many cases, as shown in Fig. 10 (B), pumps operate satisfactorily without separating walls below 5,000 GPM. If walls must be used for structural purposes or some pumps operate intermittently, then the walls should extend from the rear wall approximately five times the D dimension given in Fig. 7. If walls are used, increase dimension W by the thickness of the wall for correct centerline spacing and use round or ogive ends of walls. Not recommended is the placement of a number of pumps or suction pipes around the sides of a sump with or without dividing walls. Abrupt changes in size, as shown in Fig. 10 (C), from inlet pipe or channel to the sump are not desirable. Connection of a pipe to a sump is best accomplished using a gradually increasing taper section. The angle should be as small as possible, preferably not more than 10 degrees. With this arrangement, sump velocities less than one foot per second are desirable. Specifically not recommended is a pipe directly connected to a sump with suction intakes close to the sump inlet, since this results in an abrupt change in the flow direction. Centering pumps or suction

pipes in the sump leaves large vortex areas behind the intake which will cause operational trouble. If the sump velocity, as shown in Fig. 10 (D), can be kept low (approximately one foot per second), an abrupt change from inlet pipe to sump can be accommodated if the sump length equals or exceeds the values shown. As ratio Z/P increases, the inlet velocity at P may be increased up to an allowed maximum of eight feet per second at Z/P 10. Intakes “in line” are not recommended unless a trench-type of intake is provided (per ANSI/HI 9.8), or the ratio of sump to intake size is quite large and intakes are separated by a substantial margin longitudinally. A sump can generally be constructed at less cost by using a recommended design. As shown in Fig. 10 (E), it is sometimes desirable to install pumps in tunnels or pipe lines. A drop pipe or false well to house the unit with a vaned inlet elbow facing upstream is satisfactory in flows up to eight feet per second. Without inlet elbow, the suction bell should be positioned at least two pipe (vertical) diameters above the top of the tunnel. The unit should not be suspended in the tunnel flow, unless the tunnel velocity Is less than two feet per second. There must be no air along the top of the tunnel, and the minimum submergence must be provided. In general: Keep inlet velocity to the sump below two feet per second. Keep velocity in sump below 1.5 foot per second. Avoid changing direction of flow from inlet to pump or suction pipe, or change direction gradually and smoothly, guiding flow.

D =

(.0744Q)0.5 Recommended

W =

2D

S =

Y  5D

Where:

A  5D

S -

inches

C =

.3D to .5D

Q -

Flow (GPM)

B =

.75D

D -

inches

Fig. 7 Sump Dimensions

TECH-B

D + 0.574 Q / D1.5

Pump

W/2

Single pump W

W/2

Flow

Trash Rack

W

Multiple sump

Screen

W Optional partial dividers (increase dimension “W” by the divider thickness) required above 5,000 GPM B

Flow

Y A

Fig. 8 Sump dimensions, plan view, wet pit type pumps

A B

Screen

Trash Rack

Y Min. Water Level

D

Note: 10° or less preferred with 1 ft./sec velocity max. at screen location shown. 15° max. with velocity reduced to 0.5 ft./sec

Fig. 9 Sump dimensions, elevation view, wet pit type pumps

TECH-B

Fig. 10 Multiple pump installations Reprinted from Hydraulic Institute Standard

TECH-B

TECH-B-4A Sealing The proper selection of a seal is critical to the success of every pump application. For maximum pump reliability, choices must be made between the type of seal and the seal environment. In addition, a sealless pump is an alternative which would eliminate the need for a dynamic type seal entirely.

Sealing Area

Sealing Basics There are two basic kinds of seals: static and dynamic. Static seals are employed where no movement occurs at the juncture to be sealed. Gaskets and O-rings are typical static seals. Dynamic seals are used where surfaces move relative to one another. Dynamic seals are used, for example, where a rotating shaft transmits power through the wall of a tank (Fig. 1), through the casing of a pump (Fig. 2), or through the housing of other rotating equipment such as a filter or screen.

Rotating Shaft

A common application of sealing devices is to seal the rotating shaft of a centrifugal pump. To best understand how such a seal functions, a quick review of pump fundamentals is in order.

Fig. 1 Cross Section of Tank and Mixer

In a centrifugal pump, the liquid enters the suction of the pump at the center (eye) of the rotating impeller (Figures 3 and 4).

Sealing Area

Fig. 2 Typical Centrifugal Pump Discharge

Throat Stuffing Box or Seal Chamber

Rotary Impeller Suction Eye

Gland Shaft

Fig. 3 Centrifugal Pump, Liquid End Casing

TECH-B

As the impeller vanes rotate, they transmit motion to the incoming product, which then leaves the impeller, collects in the pump casing, and leaves the pump under pressure through the pump discharge.

Discharge

Discharge pressure will force some product down behind the impeller to the drive shaft, where it attempts to escape along the rotating drive shaft. Pump manufacturers use various design techniques to reduce the pressure of the product trying to escape. Such techniques include: 1) the addition of balance holes through the impeller to permit most of the pressure to escape into the suction side of the impeller, or 2) the addition of back pump-out vanes on the back side of the impeller.

Casing

Impeller Vanes

However, as there is no way to eliminate this pressure completely, sealing devices are necessary to limit the escape of the product to the atmosphere. Such sealing devices are typically either compression packing or end-face mechanical seals.

Fig. 4 Fluid Flow in a Centrifugal Pump

Impeller

Suction Eye

Stuffing Box Packing A typical packed stuffing box arrangement is shown in Fig. 5. It consists of: A) Five rings of packing, B) A lantern ring used for the injection of a lubricating and/or flushing liquid, and C) A gland to hold the packing and maintain the desired compression for a proper seal. The function of packing is to control leakage and not to eliminate it completely. The packing must be lubricated, and a flow from 40 to 60 drops per minute out of the stuffing box must be maintained for proper lubrication.

(Fig. 8). A flow of from .2 to .5 gpm is desirable and a valve and flowmeter should be used for accurate control. The seal water pressure should be from 10 to 15 psi above the stuffing box pressure, and anything above this will only add to packing wear. The lantern ring is normally located in the center of the stuffing box. However, for extremely thick slurries like paper stock, it is recommended that the lantern ring be located at the stuffing box throat to prevent stock from contaminating the packing.

The method of lubricating the packing depends on the nature of the liquid being pumped as well as on the pressure in the stuffing box. When the pump stuffing box pressure is above atmospheric pressure and the liquid is clean and nonabrasive, the pumped liquid itself will lubricate the packing (Fig. 6). When the stuffing box pressure is below atmospheric pressure, a lantern ring is employed and lubrication is injected into the stuffing box (Fig. 7). A bypass line from the pump discharge to the lantern ring connection is normally used providing the pumped liquid is clean.

The gland shown in Figures 5 through 8 is a quench type gland. Water, oil, or other fluids can be injected into the gland to remove heat from the shaft, thus limiting heat transfer to the bearing frame. This permits the operating temperature of the pump to be higher than the limits of the bearing and lubricant design. The same quench gland can be used to prevent the escape of a toxic or volatile liquid into the air around the pump. This is called a smothering gland, with an external liquid simply flushing away the undesirable leakage to a sewer or waste receiver.

When pumping slurries or abrasive liquids, it is necessary to inject a clean lubricating liquid from an external source into the lantern ring

Today, however, stringent emission standards limit use of packing to non-hazardous water based liquids. This, plus a desire to reduce maintenance costs, has increased preference for mechanical seals.

TECH-B

Lantern Ring

Sealing Liquid Connection

Packing Gland (Quench Type)

Stuffing Box Bushing

Positive Fluid Pressure Above Atmospheric Pressure

Atmospheric Pressure

Stuffing Box Throat

Leakage Mechanical Packing

Fig. 5 Typical Stuffing Box Arrangement (Description of Parts)

Injected Fluid Atmospheric Pressure

Leakage Into Pump

Fig. 6 Typical Stuffing Box Arrangement When Stuffing Box Pressure is Above Atmospheric Pressure

Lantern Ring Location For Thick Slurries Including Paper Stock

Injected Fluid From External Source Atmospheric Pressure

Leakage Into Pump

Normal Lantern Ring Connection

Fig. 7 Typical Stuffing Box Arrangement When Stuffing Box Pressure is Below Atmospheric Pressure

Fig. 8 Typical Stuffing Box Arrangement When Pumping Slurries

Mechanical Seals A mechanical seal is a sealing device which forms a running seal between rotating and stationary parts. They were developed to overcome the disadvantages of compression packing. Leakage can be reduced to a level meeting environmental standards of government regulating agencies and maintenance costs can be lower. Advantages of mechanical seals over conventional packing are as follows:

1. Zero or limited leakage of product (meet emission regulations.) 2. Reduced friction and power loss. 3. Elimination of shaft or sleeve wear. 4. Reduced maintenance costs. 5. Ability to seal higher pressures and more corrosive environments. 6. The wide variety of designs allows use of mechanical seals in almost all pump applications.

TECH-B

The Basic Mechanical Seal All mechanical seals are constructed of three basic sets of parts as shown in Fig. 9:

2. A set of secondary seals known as shaft packings and insert mountings such as O-rings, wedges and V-rings.

1. A set of primary seal faces: one rotary and one stationary...shown in Fig. 9 as seal ring and insert.

3. Mechanical seal hardware including gland rings, collars, compression rings, pins, springs and bellows.

Coil Spring

Insert

Insert Mounting

Gland Ring Shaft Packing

Seal Ring

Gland Gasket

Fig. 9 A Simple Mechanical Seal

How A Mechanical Seal Works The primary seal is achieved by two very flat, lapped faces which create a difficult leakage path perpendicular to the shaft. Rubbing contact between these two flat mating surfaces minimizes leakage. As in all seals, one face is held stationary in a housing and the other face is fixed to, and rotates with, the shaft. One of the faces is usually a non-galling material such as carbon-graphite. The other is usually a relatively hard material like silicon-carbide. Dissimilar materials are usually used for the stationary Insert and the rotating seal ring face in order to prevent adhesion of the two faces. The softer face usually has the smaller mating surface and is commonly called the wear nose.

POINT C Gland Gasket

POINT D Insert Mounting

POINT A Face

There are four main sealing points within an end face mechanical seal (Fig. 10). The primary seal is at the seal face, Point A. The leakage path at Point B is blocked by either an O-ring, a V-ring or a wedge. Leakage paths at Points C and D are blocked by gaskets or O-rings. The faces in a typical mechanical seal are lubricated with a boundary layer of gas or liquid between the faces. In designing seals for the desired leakage, seal life, and energy consumption, the designer must consider how the faces are to be lubricated and select from a number of modes of seal face lubrication. To select the best seal design, it’s necessary to know as much as possible about the operating conditions and the product to be sealed. Complete information about the product and environment will allow selection of the best seal for the application.

POINT B Shaft Packing Fig. 10 Sealing Points for Mechanical Seal

TECH-B

Mechanical Seal Types Mechanical seals can be classified into several types and arrangements:

PUSHER:

NON-PUSHER:

Incorporate secondary seals that move axially along a shaft or sleeve to maintain contact at the seal faces. This feature compensates for seal face wear and wobble due to misalignment. The pusher seals advantage is that it’s inexpensive and commercially available in a wide range of sizes and configurations. Its disadvantage is that it's prone to secondary seal hang-up and fretting of the shaft or sleeve. Examples are Dura RO and Crane Type 9T.

The non-pusher or bellows seal does not have to move along the shaft or sleeve to maintain seal face contact. The main advantages are its ability to handle high and low temperature applications, and does not require a secondary seal (not prone to secondary seal hang-up). A disadvantage of this style seal is that its thin bellows cross sections must be upgraded for use in corrosive environments. Examples are Dura CBR and Crane 215, and Sealol 680.

UNBALANCED:

BALANCED:

They are inexpensive, leak less, and are more stable when subjected to vibration, misalignment, and cavitation. The disadvantage is their relative low pressure limit. If the closing force exerted on the seal faces exceeds the pressure limit, the lubricating film between the faces is squeezed out and the highly loaded dry running seal fails. Examples are the Dura RO and Crane 9T.

Balancing a mechanical seal involves a simple design change which reduces the hydraulic forces acting to close the seal faces. Balanced seals have higher pressure limits, lower seal face loading, and generate less heat. This makes them well suited to handle liquids with poor lubricity and high vapor pressures such as light hydrocarbons. Examples are Dura CBR and PBR and Crane 98T and 215.

CONVENTIONAL:

CARTRIDGE:

Examples are the Dura RO and Crane Type 1 which require setting and alignment of the seal (single, double, tandem) on the shaft or sleeve of the pump. Although setting a mechanical seal is relatively simple, today's emphasis on reducing maintenance costs has increased preference for cartridge seals.

Examples are Dura P-50 and Crane 1100 which have the mechanical seal premounted on a sleeve including the gland and fit directly over the Model 3196 shaft or shaft sleeve (available single, double, tandem). The major benefit, of course is no requirement for the usual seal setting measurements for their installation. Cartridge seals lower maintenance costs and reduce seal setting errors.

TECH-B

Mechanical Seal Arrangements SINGLE INSIDE: This is the most common type of mechanical seal. These seals are easily modified to accommodate seal flush plans and can be balanced to withstand high seal environment pressures. Recommended for relatively clear non-corrosive and corrosive liquids with satisfactory lubricating properties where cost of operation does not exceed that of a double seal. Examples are Dura RO and CBR and Crane 9T and 215. Reference Conventional Seal. SINGLE OUTSIDE: If an extremely corrosive liquid has good lubricating properties, an outside seal offers an economical alternative to the expensive metal required for an inside seal to resist corrosion. The disadvantage is that it is exposed outside of the pump which makes it vulnerable to damage from impact and hydraulic pressure works to open the seal faces so they have low pressure limits (balanced or unbalanced).

DOUBLE GAS BARRIER (PRESSURIZED DUAL GAS): Very similar to cartridge double seals...sealing involves an inert gas, like nitrogen, to act as a surface lubricant and coolant in place of a liquid barrier system or external flush required with conventional or cartridge double seals. This concept was developed because many barrier fluids commonly used with double seals can no longer be used due to new emission regulations. The gas barrier seal uses nitrogen or air as a harmless and inexpensive barrier fluid that helps prevent product emissions to the atmosphere and fully complies with emission regulations. The double gas barrier seal should be considered for use on toxic or hazardous liquids that are regulated or in situations where increased reliability is the required on an application. Examples are Dura GB200, GF200, and Crane 2800.

DOUBLE (DUAL PRESSURIZED): This arrangement is recommended for liquids that are not compatible with a single mechanical seal (i.e. liquids that are toxic, hazardous [regulated by the EPA], have suspended abrasives, or corrosives which require costly materials). The advantages of the double seal are that it can have five times the life of a single seal in severe environments. Also, the metal inner seal parts are never exposed to the liquid product being pumped, so viscous, abrasive, or thermosetting liquids are easily sealed without a need for expensive metallurgy. In addition, recent testing has shown that double seal life is virtually unaffected by process upset conditions during pump operation. A significant advantage of using a double seal over a single seal. The final decision between choosing a double or single seal comes down to the initial cost to purchase the seal, cost of operation of the seal, and environmental and user plant emission standards for leakage from seals. Examples are Dura double RO and X-200 and Crane double 811T.

TANDEM (DUAL UNPRESSURIZED): Due to health, safety, and environmental considerations, tandem seals have been used for products such as vinyl chloride, carbon monoxide, light hydrocarbons, and a wide range of other volatile, toxic, carcinogenic, or hazardous liquids. Tandem seals eliminate icing and freezing of light hydrocarbons and other liquids which could fall below the atmospheric freezing point of water in air (32°F or 0°C). (Typical buffer liquids in these applications are ethylene glycol, methanol, and propanol.) A tandem also increases online reliability. If the primary seal fails, the outboard seal can take over and function until maintenance of the equipment can be scheduled. Examples are Dura TMB-73 and tandem PTO.

TECH-B

Mechanical Seal Selection The proper selection of a mechanical seal can be made only if the full operating conditions are known:

2. Pressure. The proper type of seal, balanced or unbalanced, is based on the pressure on the seal and on the seal size. 3. Temperature. In part, determines the use of the sealing members. Materials must be selected to handle liquid temperature.

1. Liquid 2. Pressure 3. Temperature 4. Characteristics of Liquid 5. Reliability and Emission Concerns

1. Liquid. Identification of the exact liquid to be handled is the first step in seal selection. The metal parts must be corrosion resistant, usually steel, bronze, stainless steel, or Hastelloy. The mating faces must also resist corrosion and wear. Carbon, ceramic, silicon carbide or tungsten carbide may be considered. Stationary sealing members of Buna, EPR, Viton and Teflon are common.

4. Characteristics of Liquid. Abrasive liquids create excessive wear and short seal life. Double seals or clear liquid flushing from an external source allow the use of mechanical seals on these difficult liquids. On light hydrocarbons balanced seals are often used for longer seal life even though pressures are low. 5. Reliability and Emission Concerns. The seal type and arrangement selected must meet the desired reliability and emission standards for the pump application. Double seals and double gas barrier seals are becoming the seals of choice.

Seal Environment The number one cause of pump downtime is failure of the shaft seal. These failures are normally the result of an unfavorable seal environment such as improper heat dissipation (cooling), poor lubrication of seal faces, or seals operating in liquids containing solids, air or vapors. To achieve maximum reliability of a seal application, proper choices of seal housings (standard bore stuffing box, large bore, or large tapered bore seal chamber) and seal environmental controls (CPI and API seal flush plans) must be made. STANDARD BORE STUFFING BOX COVER Designed thirty years ago specifically for packing. Also accommodates mechanical seals (clamped seat outside seals and conventional double seals.)

CONVENTIONAL LARGE BORE SEAL CHAMBER Designed specifically for mechanical seals. Large bore provides increased life of seals through improved lubrication and cooling of faces. Seal environment should be controlled through use of CPI or API flush plans. Often available with internal bypass t o provide circulation of liquid to faces without using external flush. Ideal for conventional or cartridge single mechanical seals in conjunction with a flush and throat bushing in bottom of chamber. Also excellent for conventional or cartridge double or tandem seals.

LARGE BORE SEAL CHAMBERS Introduced in the mid-80’s, enlarged bore seal chambers with increased radial clearance between the mechanical seal and seal chamber wall, provide better circulation of liquid to and from seal faces. Improved lubrication and heat removal (cooling) of seal faces extend seal life and lower maintenance costs.

BigBoreTM Seal Chamber

TaperBoreTM Seal Chamber

TECH-B

Large Tapered Bore Seal Chambers Provide increased circulation of liquid at seal faces without use of external flush. Offers advantages of lower maintenance costs, elimination of tubing/piping, lower utility costs (associated with seal flushing) and extended seal reliability. The tapered bore seal chamber is commonly available with ANSI chemical pumps. API process pumps use conventional large bore seal chambers. Paper stock pumps use both conventional large bore and large tapered bore seal chambers. Only tapered bore seal chambers with flow modifiers provide expected reliability on services with or without solids, air or vapors. Conventional Tapered Bore Seal Chamber: Mechanical Seals Fail When Solids or Vapors Are Present in Liquid Many users have applied the conventional tapered bore seal chamber to improve seal life on services containing solids or vapors. Seals in this environment failed prematurely due to entrapped solids and vapors. Severe erosion of seal and pump parts, damaged seal faces and dry running were the result.

Modified Tapered Bore Seal Chamber with Axial Ribs: Good for Services Containing Air, Minimum Solids This type of seal chamber will provide better seal life when air or vapors are present in the liquid. The axial ribs prevent entrapment of vapors through improved flow in the chamber. Dry running failures are eliminated. In addition, solids less than 1% are not a problem. The new flow pattern, however, still places the seal in the path of solids/liquid flow. The consequence on services with significant solids (greater than 1%) is solids packing the seal spring or bellows, solids impingement on seal faces and ultimate seal failure.

Goulds Standard TaperBoreTM PLUS Seal Chamber: The Best Solution for Services Containing Solids and Air or Vapors To eliminate seal failures on services containing vapors as well as solids, the flow pattern must direct solids away from the mechanical seal, and purge air and vapors. Goulds Standard TaperBoreTM PLUS completely reconfigures the flow in the seal chamber with the result that seal failures due to solids are eliminated. Air and vapors are efficiently removed eliminating dry run failures. Extended seal and pump life with lower maintenance costs are the results.

Goulds TaperBoreTM Plus: How It Works The unique flow path created by the Vane Particle Ejector directs solids away from the mechanical seal, not at the seal as with other tapered bore designs. And the amount of solids entering the bore is minimized. Air and vapors are also efficiently removed. On services with or without solids, air or vapors, Goulds TaperBoreTM PLUS is the effective solution for extended seal and pump life and lower maintenance costs.

1

Solids/liquid mixture flows toward mechanical seal/seal chamber.

2

Turbulent zone. Some solids continue to flow toward shaft. Other solids are forced back out by centrifugal force (generated by back pump-out vanes).

3

Clean liquid continues to move toward mechanical seal faces. Solids, air, vapors flow away from seal.

4

Low pressure zone create by Vane Particle Ejector. Solids, air, vapor liquid mixture exit seal chamber bore.

5

Flow in TaperBoreTM PLUS seal chamber assures efficient heat removal (cooling) and lubrication. Seal face heat is dissipated. Seal faces are continuously flushed with clean liquid.

TECH-B

1

2

4 5

3

JACKETED STUFFING BOX COVER

JACKETED LARGE BORE SEAL CHAMBER

Designed to maintain proper temperature control (heating or cooling) of seal environment. (Jacketed covers do not help lower seal face temperatures to any significant degree). Good for high temperature services that require use of a conventional double seal or single seal with a flush and API or CPI plan 21.

Maintains proper temperature control (heating or cooling) of seal environment with improved lubrication of seal faces. Ideal for controlling temperature for services such as molten sulfur and polymerizing liquids. Excellent for high temperature services that require use of conventional or cartridge single mechanical seals with flush and throat bushing in bottom of seal chamber. Also, great for conventional or cartridge double or tandem seals.

Stuffing Box Cover and Seal Chamber Guides The following two selection guides are designed to assist selection of the proper seal housing for a pump application.

Stuffing Box and Seal Chamber Application Guide Stuffing Box Cover Seal Chamber

Application

Standard Bore Stuffing Box Cover

Use for soft packing. Outside mechanical seals. Double seals. Also, accommodates other Mechanical seals.

Jacketed Stuffing Box Cover

Same as but also need to control temperatures of liquid in seal area.

Conventional Large Bore

Use for all mechanical seal applications where the seal environment requires use of CPI or API seal flush pans. Cannot be used with outside type mechanical seals

Jacked Large Bore

Same as Large Bore but also need to control temperature of liquid in seal area.

Tapered Large Bore with Axial Ribs

Clean services that require use of single mechanical seals. Can also be used with cartridge double seals. Also, effective on services with light solids up to 1% by weight. Paper stock to 1% by weight.

Tapered Large Bore with Patented Vane Particle Ejector (Alloy Construction)

Services with light to moderate solids up to 10% by weight. Paper stock to 5% by Weight. Ideal for single mechanical seals. No flush required. Also, accommodates cartridge double seals. Cannot be used with outside mechanical seals.

TECH-B

Selection Guide Goulds Engineered Seal Chambers Provide Best Seal Environment For Selected Sealing Arrangements/Services

TYPE 1

TYPE 2

TYPE 3

TYPE 4

TYPE 5

Standard Bore Stuffing Box Cover

Conventional Large Bore

Tapered Bore

Jacketed Stuffing Box

Jacketed Large Bore

Maintains proper temperature control (heating or cooling) of seal environment.

Maintains proper temperature control (heating or cooling) of seal environment with improved lubrication of seal faces. Ideal for controlling temperatures on services such as molten sulfur and polymerizing liquids.

Designed for packing. Also accommodates mechanical seals.

A

Ideally Suited

B

Acceptable

C

Not Recommended

Enlarged chamber for increased seal life through improved lubrication and cooling. Seal environment should be controlled through use of CPI flush plans.

Lower seal face temperatures, self-venting and draining. Solids and vapors circulated away from seal faces. Often no flush required. Superior patented design maximizes seal life with or without solids and vapor in liquid.

Service Acceptable Ideally Suited Ambient Water With Flush

A

A

A

-

-

Entrained Air or Vapor

C

B

A

C

B

Solids 0-10%, No Flush

C

C

A

C

C

Solids up to and greater than 10% With Flush Paper Stock 0-5%, With No Flush Paper Stock 0-5%, With Flush

B

A

A

B

A

C

C

A

-

-

B

A

A

-

-

Slurries 0-5%, No Flush

C

C

A

C

C

High Boiling Point Liquids, no flush

C

C

A

C

C

Temperature Control

C

C

C

B

A

C

C

A

C

C

C

A

A

C

A

C

C

B

C

C

C

B

B

C

A

Self-Venting and Draining Seal Face Heat Removal Molten or Polymerizing Liquid, No Flush Molten or Polymerizing Liquid With Flush

TECH-B

Environmental Controls Environmental controls are necessary for reliable performance of a mechanical seal on many applications. Goulds Pumps and the seal vendors offer a variety of arrangements to combat these problems. 1. Corrosion 2. Temperature Control 3. Dirty or incompatible environments CORROSION Corrosion can be controlled by selecting seal materials that are not attacked by the pumpage. When this is difficult, external fluid injection of a non-corrosive chemical to lubricate the seal is possible. Single or double seals could be used, depending on if the customer can stand delusion of his product. TEMPERATURE CONTROL As the seal rotates, the faces are in contact. This generates heat and if this heat is not removed, the temperature in the stuffing box or seal chamber can increase and cause sealing problems. A simple by-pass of product over the seal faces will remove the heat generated by the seal (Fig. 25). For higher temperature services, by-pass of product through a cooler may be required to cool the seal sufficiently (Fig. 26). External cooling fluid injection can also be used.

Fig. 25

DIRTY or INCOMPATIBLE ENVIRONMENTS Mechanical seals do not normally function well on liquids which contain solids or can solidify on contact with the atmosphere. Here, by-pass flush through a filter, a cyclone separator or a strainer are methods of providing a clean fluid to lubricate seal faces. Strainers are effective for particles larger than the openings on a 40 mesh screen. Cyclone separators are effective on solids 10 micron or more in diameter, if they have a specific gravity of 2.7 and the pump develops a differential pressure of 30-40 psi. Filters are available to remove solids 2 microns and larger. If external flush with clean liquid is available, this is the most fail proof system. Lip seal or restricting bushings are available to control flow of injected fluid to flows as low as 1⁄8 GPM.

Fig. 26

Quench type glands are used on fluids which tend to crystallize on exposure to air. Water or steam is put through this gland to wash away any build up. Other systems are available as required by the service.

TECH-B

API and CPI Plans API and CPI mechanical seal flush plans are commonly used with API and CPI process pumps. The general arrangement of the plans are similar regardless of the designation whether API or CPI. The difference between the flush plans is the construction which provides applicable pressure-temperature capability for each type of pump. API plans have higher pressure and temperature capability than CPI plans. Each plan helps provide critical lubrication and cooling of seal faces to maximize seal reliability. Plan No. Recommended Applications 01

Single mechanical seals and TDH less then 125 feet.

02

Used with some outside seals. In most cases not recommended.

11

Single and tandem seals. Always consider a plan 11 with balanced seals. Apply when TDH is greater than 125 ft.

12

Same application as 11. Additionally, a 12 will strain particles from the flush liquid. This helps prevent solid impingement on seal faces.

13

Single and tandem seals. Use when difference in pressure between the seal chamber or stuffing box and pump suction exceed 35 psi.

21

Single and tandem seals. Required when the flush needs to be cooled before flushing at the seal faces. (ex. water above 200°F, light hydrocarbons or any other liquids with poor lubricating qualities and high vapor pressures.)

22

Same application as 21. Additionally, a plan 22 will strain particles from the flush liquid. This helps prevent solid impingement on seal faces.

23

Single and tandem seals. Use when difference in pressure between the seal chamber or stuffing box and pump suction exceed 35 psi. 3600 RPM only.

31

Single and tandem seals. Apply when strainers are inadequate to clean flushing liquid.

32

Single and tandem seals. Required when pumpage is not suitable to lubricate seal faces. Use of bushing or lip seal is also recommended.

33

Used with double seals when external system is available from user.

41

Apply with liquids that require simultaneous cyclone separation and cooling. (Single and tandem seals).

51

Single seals. Required when sealed liquid will crystallize, coke, solidify, etc. at seal faces if contact with air. Common blankets are isopropyl alcohol, glycol, and water. Normally used with FVD gland and bushing or packed auxiliary box.

52

Tandem seals. Plan provides buffer liquid for outside seal. A plan 01 or plan 11 is also recommended with tandem seals to properly flush inboard seal. Pumping rings recommended.

53

Double seals. Plan provides flushing and cooling to both sets of seal faces. Pumping ring recommended.

54

Double seals or packed auxiliary stuffing box.

Dynamic Seal - an Alternative to the Mechanical Seal On some tough pumping services like paper stock and slurries, mechanical seals require outside flush and constant, costly attention. Even then, seal failures are common, resulting in downtime.

Repeller Plate

Goulds offers a Dynamic Seal which, simply by fitting a repeller between the stuffing box and impeller, eliminates the need for a mechanical seal.

Repeller

BENEFITS OF GOULDS DYNAMIC SEAL: • External seal water not required. • Elimination of pumpage contamination and product dilution • Reduces utility cost • No need to treat seal water • Eliminates problems associated with piping from a remote source HOW IT WORKS At start-up, the repeller functions like an impeller, and pumps liquid and solids from the stuffing box. When pump is shut down, packing (illustrated) or other type of secondary seal prevents pumpage from leaking.

TECH-B

Stuffing Box Cover

Impeller

TECH-B-4B Magnetic Drive Pumps INTRODUCTION

PRINCIPLES OF OPERATION

Environmental concerns and recurring mechanical seal problems have created a need for sealless pumps in the chemical and petrochemical industries. In some cases, more stringent regulations by the EPA, OSHA and local agencies are mandating the use of sealless pumps. One type of sealless pump is the magnetic drive pump which uses a permanent magnetic coupling to transmit torque to the impeller without the need for a mechanical seal for packing.

Magnetic drive pumps use a standard electric motor to drive a set of permanent magnets that are mounted on a carrier or drive assembly located outside of the containment shell. The drive magnet assembly is mounted on a second shaft which is driven by a standard motor. The external rotating magnetic field drives the inner rotor. The coaxial synchronous torque coupling consists of two rings of permanent magnets as shown in Fig. 1. A magnetic force field is established between the north and south pole magnets in the drive and driven assemblies. This provides the no slip or synchronous capability of the torque coupling. The magnetic field is shown as dashed lines and shaded areas in Fig. 3.

Fig. 1 Typical Magnetic Drive Pump Driven Magnet Drive Magnet Carrier Assembly Carrier Assembly Containment Shell

Bearing Frame Assembly

Bearings

A

MOTOR (DRIVE)

PUMP (DRIVEN)

Driven Magnet Assembly

A

Drive Magnet Assembly

Fig. 2. Coaxial Synchronous Magnetic Torque Coupling Fig. 3

TECH-B

Two Types of Magnetic Drive Pump Designs A. Rotating Driven Shaft This type of design typically uses metal components and is best suited for heavy duty applications. The metallic construction offers the best strength, temperature and pressure capability required for heavy duty applications. Corrosion resistant high alloy materials such as 316SS, Hastelloy, and Alloy 20 are offered. The rotating shaft does, however, increase the number of parts required and thus increases the complexity and cost of the pump. This type of design typically uses a pressurized recirculation circuit, which helps prevent vaporization of liquid required for process lubricated bearings. (Refer to Model 3296, Section CHEM-3A).

B. Stationary Shaft This type of design typically uses non-metallic components such as ceramics and plastics. It is best suited for light to medium duty applications. The stationary shaft design significantly reduces the number of parts required, simplifying maintenance and reducing cost. Corrosion resistant materials such as silicon carbide ceramics and fluoropolymer plastics (Teflon, Tefzel, etc.) provide excellent range of application. The use of plastics materials does, however, limit the temperature range of these designs to 200° F to 250° F. (Refer to Model 3298, Section CHEM-3C).

Containment Shell Designs The containment shell is the pressure containing barrier which is fitted between the drive and the driven magnet assembly. It must contain full working pressure of the pump, since it isolates the pumped liquid from the atmosphere. One-piece formed shells offer the best reliability, eliminating welds used for two-piece shells. Since the torque coupling magnetic force field must pass through the shell, it must be made of a non-magnetic material. Non-magnetic metals such as Hastelloy and 316SS are typical choices for the containment shell. The motion of the magnets past an electrically conductive containment shell produces eddy currents, which generate heat and must be removed by a process fluid recirculation circuit.

The eddy currents also create a horsepower loss, which reduces the efficiency of the pump. Metals with low electrical conductivity have lower eddy current losses, providing superior pump efficiency. Hastelloy has a relatively low electrical conductivity and good corrosion resistance, thus is an excellent choice for metal containment shells. Electrically non-conductive materials such as plastic and ceramics are also good choices for containment shells, since the eddy current losses are totally eliminated. This results in pump efficiencies equal to conventionally sealed pumps. Plastic containment shells are generally limited to lower pressures and temperatures due to the limited strength of plastics.

Sleeve and Thrust Bearings Magnetic drive pumps utilize process lubricated bearings to support the inner drive rotor. These bearings are subject to the corrosive nature of the liquids being pumped, thus need to be made from corrosion resistant materials. Two commonly used materials are hard carbon and silicon carbide (SIC). Pure sintered SIC is superior to reaction bonded SIC, since reaction bonded SIC has free silicon left in the matrix, resulting in lower chemical resistance and lower strength.

TECH-B

Hard carbon against silicon carbide offers excellent service life for many chemical applications and also offers the advantage of short term operation in marginal lubrication conditions. Silicon carbide against silicon carbide offers excellent service life for nearly all chemical applications. Its hardness, high thermal conductivity, and strength make it an excellent bearing material. Silicon carbide must be handled carefully to prevent chipping. Silicon carbide against silicon carbide has very limited capability in marginal lubrication conditions.

Recirculation Circuit All magnetic drive pumps circulate some of the process fluid to lubricate and cool the bearings supporting the inner rotor. Magnetic drive pumps with metal containment shells, also require a circulation of some process fluid through the containment shell to remove heat generated by eddy currents. For pumps with metal containment shells, the fluid recirculation path must be carefully engineered to prevent vaporization of the process liquid necessary to lubricate the bearings. A pressurized circuit as shown in Fig. 4 offers excellent reliability for pumps with metal containment shells. Magnetic drive pumps with electrically non-conductive containment shells, such as plastic or ceramic have no heat generated by eddy currents. Since no heat is required to be removed from the containment shell, a much simpler recirculation circuit can be used. For liquids near vaporization, a calculation must be made to ensure the process fluid does not vaporize at the bearings. This calculation includes the effects of process fluid specific heat, vapor pressure, drive losses, recirculation flow, etc. This calculation procedure can be found in the GOULDS PUMPS HANDBOOK FOR MAGNETIC DRIVE PUMPS. An external cooling system can be added to the recirculation circuit to prevent vaporization. Fig. 4 Recirculation Circuit

Fail Safe Devices DESCRIPTION

These malfunctions can contribute to:

Condition monitoring of the pump is a "key objective" and provides the user with an assurance of safety and reliability.

•

Overheating of the drive and driven magnet assemblies

•

Overload of drive motor and drive magnetic assembly

System and pump malfunctions can result from the following:

•

Extreme pump bearing load conditions

•

Damage to pump due to extremes in temperatures and pressures due to transients that exceed normal design.

•

No-flow condition through the pump

•

Dry running as a result of plugged liquid circulation paths in the pump bearing and magnets assembly section

•

Cavitation due to insufficient NPSHA

•

Uncoupling of the magnetic drive due to overload

•

thermocouple / controller

•

Temperature and pressure transients in the system

•

low amp relay

•

"Flashing" in the pump liquid circulation paths due to pressure and temperature transients.

•

liquid leak detector

•

power monitor

Various fail safe devices are available with the pump to control malfunctions and provide safety and reliability including:

TECH-B

TECH-B-5 Field Testing Methods A. Determination of total head The total head of a pump can be determined by gauge readings as illustrated in Fig. 1. hd WATER

hs MERCURY

hd

h

h

hd

Vacuum

Datum

Pressure hs

hs

B. Measurement of capacity

A calibrated magnetic flow meter is an accurate means of measuring flow in a pumping system. However, due to the expense involved, magnetic flow meters are only practical in small factory test loops and in certain process pumping systems where flow is critical. b.) Volumetric measurement

Negative Suction Pressure: TDH = Discharge gauge reading converted to feet of liquid + vacuum gauge reading converted to feet of liquid + distance between point of attachment of vacuum gauge and the centerline of the discharge 2 2 gauge, h, in feet + Vd – Vs 2g 2g

(

)

Positive Suction Pressure: or TDH = Discharge gauge reading converted to feet of liquidpressure gauge reading in suction line converted to ft. of liquid + distance between center of discharge and suction gauges, h, in feet 2 2 + Vd – Vs 2g 2g

)

In using gauges when the pressure is positive or above atmospheric pressure, any air in the gauge line should be vented off by loosening the gauge until liquid appears. This assures that the entire gauge line is filled with liquid and thus the gauge will read the pressure at the elevation of the centerline of the gauge. However, the gauge line will be empty of liquid when measuring vacuum and the gauge will read the vacuum at the elevation of the point of attachment of the gauge line to the pipe line. These assumptions are reflected in the above definitions. The final term in the above definitions accounts for a difference in size between the suction and discharge lines. The discharge line is normally smaller than the suction line and thus the discharge velocity is higher. A higher velocity results in a lower pressure since the sum of the pressure head and velocity head in any flowing liquid remains constant. Thus, when the suction and discharge line sizes at the gauge attachment points are different, the resulting difference in velocity head must be included in the total head calculation. Manometers can also be used to measure pressure. The liquid used in a manometer is normally water or mercury, but any liquid of known specific gravity can be used. Manometers are extremely accurate for determining low pressures or vacuums and no calibration is needed. They are also easily fabricated in the field to suit any particular application. Figs. 2 & 3 illustrate typical manometer set ups.

TECH-B

Fig. 3 Manometer Indicating Pressure

a.) Magnetic Flow Meter

Fig. 1 Determination of Total Head From Gauge Readings

(

Fig. 2 Manometer Indicating Vacuum

Pump capacity can be determined by weighing the liquid pumped or measuring its volume in a calibrated vessel. This is often practical when pumping into an accurately measured reservoir or tank, or when it is possible to use small containers which can be accurately weighed. These methods, however, are normally suited only to relatively small capacity systems. c.) Venturi meter A venturi meter consists of a converging section, a short constricting throat section and then a diverging section. The object is to accelerate the fluid and temporarily lower its static pressure. The flow is then a function of the pressure differential between the full diameter line and the throat. Fig. 4 shows the general shape and flow equation. The meter coefficient is determined by actual calibration by the manufacturer and when properly installed the Venturi meter is accurate to within plus or minus 1%.

Q(GPM) = 5.67 CD22

H

1 – R4

C = Instrument Coefficient D1 = Entrance Diameter in Inches D2 = Throat Diameter in Inches R = D2/ D1 H = Differential Head in Inches = h1 – h2 h1

h2

D2

D1

D1

Fig. 5 Venturi Meter d.) Nozzle A nozzle is simply the converging portion of a venturi tube with the liquid exiting to the atmosphere. Therefore, the same formula can be used with the differential head equal to the gauge reading ahead of the nozzle. Fig. 5 lists theoretical nozzle discharge flows.

Theoretical Discharge of Nozzles in U.S. GPM Head Lbs.

Feet

10 15 20 25 30 35 40 45 50 55 60 65 70 75 80 85 90 95 100 105 110 115 120 125 130 135 140 145 150 175 200 250 300

23.1 34.6 46.2 57.7 69.3 80.8 92.4 103.9 115.5 127.0 138.6 150.1 161.7 173.2 184.8 196.3 207.9 219.4 230.9 242.4 254.0 265.5 277.1 288.6 300.2 311.7 323.3 334.8 346.4 404.1 461.9 577.4 692.8

10 15 20 25 30 35 40 45 50 55 60 65 70 75 80 85 90 95 100 105 110 115 120 125 130 135 140 145 150 175 200 250 300

23.1 34.6 46.2 57.7 69.3 80.8 92.4 103.9 115.5 127.0 138.6 150.1 161.7 173.2 184.8 196.3 207.9 219.4 230.9 242.4 254.0 265.5 277.1 288.6 300.2 311.7 323.3 334.8 346.4 404.1 461.9 577.4 692.8

Veloc’y of Disch. Feet per Sec. 38.6 47.25 54.55 61.0 66.85 72.2 77.2 81.8 86.25 90.4 94.5 98.3 102.1 105.7 109.1 112.5 115.8 119.0 122.0 125.0 128.0 130.9 133.7 136.4 139.1 141.8 144.3 146.9 149.5 161.4 172.6 193.0 211.2

38.6 47.25 54.55 61.0 66.85 72.2 77.2 81.8 86.25 90.4 94.5 98.3 102.1 105.7 109.1 112.5 115.8 119.0 122.0 125.0 128.0 130.9 133.7 136.4 139.1 141.8 144.3 146.9 149.5 161.4 172.6 193.0 211.2

Diameter of Nozzle in Inches 1⁄16

1⁄8

0.37 0.45 0.52 0.58 0.64 0.69 0.74 0.78 0.83 0.87 0.90 0.94 0.98 1.01 1.05 1.08 1.11 1.14 1.17 1.20 1.23 1.25 1.28 1.31 1.33 1.36 1.38 1.41 1.43 1.55 1.65 1.85 2.02

1.48 1.81 2.09 2.34 2.56 2.77 2.96 3.13 3.30 3.46 3.62 3.77 3.91 4.05 4.18 4.31 4.43 4.56 4.67 4.79 4.90 5.01 5.12 5.22 5.33 5.43 5.53 5.62 5.72 6.18 6.61 7.39 8.08

11⁄2 213 260 301 336 368 398 425 451 475 498 521 542 563 582 602 620 638 656 672 689 705 720 736 751 767 780 795 809 824 890 950 1063 1163

3⁄16

1⁄4

3⁄8

1⁄2

5⁄8

3⁄4

7⁄8

53.1 65.0 75.1 84.0 92.0 99.5 106. 113. 119. 125. 130. 136. 141. 146. 150. 155. 160. 164. 168. 172. 176. 180. 184. 188. 192. 195. 199. 202. 206. 222. 238. 266. 291.

72.4 88.5 102. 114. 125. 135. 145. 153. 162. 169. 177. 184. 191. 198. 205. 211. 217. 223. 229. 234 240 245. 251. 256. 261. 266. 271. 275. 280. 302. 323. 362. 396.

3.32 4.06 4.69 5.25 5.75 6.21 6.64 7.03 7.41 7.77 8.12 8.45 8.78 9.08 9.39 9.67 9.95 10.2 10.5 10.8 11.0 11.2 11.5 11.7 12.0 12.2 12.4 12.6 12.9 13.9 14.8 16.6 18.2

5.91 7.24 8.35 9.34 10.2 11.1 11.8 12.5 13.2 13.8 14.5 15.1 15.7 16.2 16.7 17.3 17.7 18.2 18.7 19.2 19.6 20.0 20.5 20.9 21.3 21.7 22.1 22.5 22.9 24.7 26.4 29.6 32.4

13.3 16.3 18.8 21.0 23.0 24.8 26.6 28.2 29.7 31.1 32.5 33.8 35.2 36.4 37.6 38.8 39.9 41.0 42.1 43.1 44.1 45.1 46.0 47.0 48.0 48.9 49.8 50.6 51.5 55.6 59.5 66.5 72.8

23.6 28.9 33.4 37.3 40.9 44.2 47.3 50.1 52.8 55.3 57.8 60.2 62.5 64.7 66.8 68.9 70.8 72.8 74.7 76.5 78.4 80.1 81.8 83.5 85.2 86.7 88.4 89.9 91.5 98.8 106. 118. 129.

36.9 45.2 52.2 58.3 63.9 69.0 73.8 78.2 82.5 86.4 90.4 94.0 97.7 101. 104. 108. 111. 114. 117. 120. 122. 125. 128. 130. 133. 136. 138. 140. 143. 154. 165. 185. 202.

13⁄4

2

21⁄4

21⁄2

23⁄4

3

289 354 409 458 501 541 578 613 647 678 708 737 765 792 818 844 868 892 915 937 960 980 1002 1022 1043 1063 1082 1100 1120 1210 1294 1447 1582

378 463 535 598 655 708 756 801 845 886 926 964 1001 1037 1070 1103 1136 1168 1196 1226 1255 1282 1310 1338 1365 1390 1415 1440 1466 1582 1691 1891 2070

479 585 676 756 828 895 957 1015 1070 1121 1172 1220 1267 1310 1354 1395 1436 1476 1512 1550 1588 1621 1659 1690 1726 1759 1790 1820 1853 2000 2140 2392 2615

591 723 835 934 1023 1106 1182 1252 1320 1385 1447 1506 1565 1619 1672 1723 1773 1824 1870 1916 1961 2005 2050 2090 2132 2173 2212 2250 2290 2473 2645 2955 3235

714 874 1009 1128 1236 1335 1428 1512 1595 1671 1748 1819 1888 1955 2020 2080 2140 2200 2255 2312 2366 2420 2470 2520 2575 2620 2670 2715 2760 2985 3190 3570 3900

851 1041 1203 1345 1473 1591 1701 1802 1900 1991 2085 2165 2250 2330 2405 2480 2550 2625 2690 2755 2820 2885 2945 3005 3070 3125 3180 3235 3295 3560 3800 4250 4650

31⁄2 1158 1418 1638 1830 2005 2168 2315 2455 2590 2710 2835 2950 3065 3170 3280 3375 3475 3570 3660 3750 3840 3930 4015 4090 4175 4250 4330 4410 4485 4840 5175 5795 6330

11⁄8

11⁄4

13⁄8

94.5 116. 134. 149 164. 177. 189. 200. 211. 221. 231. 241. 250. 259. 267. 276. 284. 292. 299. 306. 314. 320. 327. 334. 341. 347. 354 360 366. 395. 423 473. 517.

120 147 169 189 207 224 239 253 267 280 293 305 317 327 338 349 359 369 378 388 397 406 414 423 432 439 448 455 463 500 535 598 655

148 181 209 234 256 277 296 313 330 346 362 376 391 404 418 431 443 456 467 479 490 501 512 522 533 543 553 562 572 618 660 739 808

179 219 253 283 309 334 357 379 339 418 438 455 473 489 505 521 536 551 565 579 593 606 619 632 645 656 668 680 692 747 799 894 977

4

41⁄2

5

51⁄2

6

1510 1850 2135 2385 2615 2825 3020 3200 3375 3540 3700 3850 4000 4135 4270 4440 4530 4655 4775 4655 5010 5120 5225 5340 5450 5550 5650 5740 5850 6310 6750 7550 8260

1915 2345 2710 3025 3315 3580 3830 4055 4275 4480 4685 4875 5060 5240 5410 5575 5740 5900 6050 5900 6350 6490 6630 6760 6900 7030 7160 7280 7410 8000 8550 9570 10480

2365 2890 3340 3730 4090 4415 4725 5000 5280 5530 5790 6020 6250 6475 6690 6890 7090 7290 7470 7290 7840 8010 8180 8350 8530 8680 8850 8990 9150 9890 10580 11820 12940

1

2855 3490 4040 4510 4940 5340 5280 6050 6380 6690 6980 7270 7560 7820 8080 8320 8560 8800 9030 8800 9470 9680 9900 10100 10300 10490 10690 10880 11070 11940 12770 14290 15620

3405 4165 4810 5380 5895 6370 6380 7210 7600 7970 8330 8670 9000 9320 9630 9920 10210 10500 10770 10500 11300 11500 11800 12030 12290 12510 12730 12960 13200 14250 15220 17020 18610

NOTE: – The actual quantities will vary from these figures, the amount of variation depending upon the shape of nozzle and size of pipe at the point where the pressure is determined. With smooth taper nozzles the actual discharge is about 94% of the figures given in the tables.

Fig. 5

TECH-B

e.) Orifice

f.) Weir

An orifice is a thin plate containing an opening of specific shape and dimensions. The plate is installed in a pipe and the flow is a function of the pressure upstream of the orifice. There are numerous types of orifices available and their descriptions and applications are covered in the Hydraulic Institute Standards and the ASME Fluid Meters Report. Orifices are not recommended for permanent installations due to the inherent high head loss across the plate.

A weir is particularly well suited to measuring flows in open conduits and can be adapted to extremely large capacity systems. For best accuracy, a weir should be calibrated in place. However, when this is impractical, there are formulas which can be used for the various weir configurations. The most common types are the rectangular contracted weir and the 90 V-notch weir. These are shown in Fig. 6 with the applicable flow formulas.

(6a) - Rectangular Weir With Complete End Contractions Q(G.P.M.) = 1495 H 3/2 (B-O.2H) H = Head in Feet Above Weir B = Crest Width in Feet

(6b) - 90° V-Notch Weir Q(G.P.M.) = 1140 H 5/2 H = Head in Feet Above Weir

Fig. 6 Weirs

g.) Pilot tube A pilot tube measures fluid velocity. A small tube placed in the flow stream gives two pressure readings: one receiving the full impact of the flowing stream reads static head + velocity head, and the other reads the static head only (Fig. 7). The difference between the two readings is the velocity head. The velocity and the flow are then determined from the following well known formulas.

Total head

Small holes on both sides of outer tube

V= C 2ghv where C is a coefficient for the meter determined by calibration, and hv = velocity head, Capacity = Area x Average Velocity Since the velocity varies across the pipe, it is necessary to obtain a velocity profile to determine the average velocity. This involves some error, but when properly applied a calibrated pilot tube is within plus or minus 2% accuracy.

TECH-B

Fig. 7 Pilot Tube

Static head

TECH-B-6 Vibration Analysis Vibration analysis equipment enables you to tell when "normal" vibration becomes "problem" vibration or exceeds acceptable levels. It may also allow you to determine the source and cause of the vibration, thus becoming an effective preventive maintenance and troubleshooting aid. A vibration analyser measures the amplitude, frequency and phase of vibration. Also when vibration occurs at several frequencies, it separates one frequency from another so that each individual vibration characteristic can be measured. The vibration pickup senses the velocity of the vibration and converts it into an electrical signal. The analyzer receives this signal, converting it to the corresponding amplitude and frequency. The amplitude is measured in terms of peak-to-peak displacement in mils (1 mil = .001") and is indicated on the amplitude meter. Some instruments are equipped with a frequency meter which gives a direct readout of the predominant frequency of the vibration. Other instruments have tunable filters which allow scanning the frequency scale and reading amplitude at any particular frequency, all others being filtered out.

By analyzing the tabulated vibration data one or several causes may be found. Each must be checked, starting with the most likely cause or easiest to check. For example, assume the axial vibration is 50% or more of the radial vibration and the predominant frequency is the same as the RPM of the pump. The chart indicates probable misalignment or bent shaft. Coupling misalignment is probably the most common single cause of pump vibration and is one of the easiest to check. If after checking, the alignment proves to be good, then inspect for flange loading. Finally, check for a bent shaft. Cavitation in a pump can cause serious vibration. Vibration at random frequencies can also be caused by hydraulic disturbances in poorly designed suction or discharge systems. The use of vibration equipment in preventative maintenance involves keeping a vibration history on individual pieces of equipment in a plant. A form similar to that shown in Fig 3 can be used to record the vibration data on a periodic routine basis. Abrupt changes are a sign of impending failure. A gradual increase in vibration can also be detected and corrective measures can be taken before it reaches a dangerous level.

A strob light is used to determine the phase of vibration. It can be made to flash at the frequency of the vibration present or at any arbitrary frequency set on an internal oscillator. A reference mark on a rotating part viewed under the strob light flashing at the vibration frequency may appear as a single frozen (or rotating) mark, or as several frozen (or rotating) marks. The number of marks viewed is useful in determining the source of the vibration. The location of the mark or marks is used in balancing rotating parts. The first step in vibration analysis is to determine the severity of the vibration, then, if the vibration is serious, a complete set of vibration readings should be taken before attempting to analyze the cause. Fig. 1 is the general guide for horizontal centrifugal pumps as published by the Hydraulic Institute. The amplitudes shown are the overall maximum obtained without filtering to specific frequencies. Amplitudes at specific frequencies, such as vane pass frequency with multi-vane impellers, should be less than 75% of the unfiltered amplitudes allowed in Fig. 1 at the operating RPM. For horizontal non-clog and vertical submerged pumps, refer to Hydraulic Institute standards or pump manufacturer. Severity of vibration is a function of amplitude and pump speed; however, it should be noted that a change in severity over a period of time is usually a warning of impending failure. This change is often more important than vibration in the "slightly rough" or "rough" ranges which does not change with time. Complete pump vibration analysis requires taking vibration readings at each bearing in three planes (horizontal, vertical and axial). Readings at the pump suction and discharge flanges may also be useful in some cases.

Fig. 1 Acceptable field vibration limits for horizontal or vertical in-line pumps (Figures 1.107 to 1.109) - clear liquids Reprinted from HYDRAULIC INSTITUTE STANDARDS. 1994 Edition, Copyright by Hydraulic Institute.

After all data has been tabulated, it can be analyzed to determine the most likely cause or causes of vibration and the identifying characteristics of each.

TECH-B

Vibration Analysis – Continued Cause

Amplitude

Frequency

Phase

Remarks

Unbalance

Largest in radial direction. Proportional to unbalance

1 x RPM

Single reference mark

Unbalance

Misalignment of coupling or bearings and bent shaft

Axial direction vibration 50% or more of radial

1 x RPM normally

single, double, or triple

Easily recognized by large axial vibration. Excessive flange loading can contribute to misalignment

Bad Anti-friction bearings

Unsteady

Very high. Erratic Several time RPM

Largest high-frequency vibration near the bad bearing.

2 x RPM

Two reference marks. Slightly erratic.

Check grouting and bed plate bolting.

Mechanical looseness Bad drive belts

Erratic or pulsing

1, 2, 3 & 4 x RPM of belts

Unsteady

Use strobe light to freeze faulty belt.

Electrical

Disappears when power is turned off.

1 or 2 x synchronous frequency

Single or rotating double mark

3600 or 7200 cps for 60 cycle current.

Hydraulic forces

No. of impeller vanes x RPM

Rarely a cause of serious vibration

Fig. 3 Vibration Identification Chart

Fig. 4 Vibration Data Sheet

TECH-B-7 Vertical Turbine Pumps

DISCHARGE LINE FRICTION LOSSES

Turbine Nomenclature 1. DATUM OR GRADE - The elevation of the surface from which the pump is supported. 2. STATIC LIQUID LEVEL - The vertical distance from grade to the liquid level when no liquid is being drawn from the well or source.

HEAD ABOVE DISCHARGE

3. DRAWDOWN - The distance between the static liquid level and the liquid level when pumping at required capacity. 4. PUMPING LIQUID LEVEL - The vertical distance from grade to liquid level when pumping at rated capacity. Pumping liquid level equals static water level plus drawdown. 5. SETTING - The distance from grade to the top of the pump bowl assembly. 6. TPL (TOTAL PUMP LENGTH) - The distance from grade to lowest point of pump.

GRADE

PUMP SETTING

STATIC LEVEL

PUMPING LEVEL

7. RATED PUMP HEAD - Lift below discharge plus head above discharge plus friction losses in discharge line. This is the head for which the customer is responsible and does not include any losses within the pump.

TOTAL PUMP 8. COLUMN AND DISCHARGE HEAD FRICTION LOSS - Head LENGTH (TPL) loss in the pump due to friction in the column assembly and discharge head. Friction loss is measured in feet and is dependent upon column size, shaft size, setting, and discharge head size. Values given in appropriate charts in Data Section. 9. BOWL HEAD - Total head which the pump bowl assembly will deliver at the rated capacity. This is curve performance. 10. BOWL EFFICIENCY- The efficiency of the bowl unit only. This value is read directly from the performance curve. 11. BOWL HORSEPOWER- The horsepower - required by the bowls only to deliver a specified capacity against bowl head. BOWL HP = Bowl Head x Capacity 3960 x Bowl Efficiency 12. TOTAL PUMP HEAD - Rated pump head plus column and discharge head loss. Note: This is new or final bowl head. 13. SHAFT FRICTION LOSS - The horsepower required to turn the lineshaft in the bearings. These values are given in appropriate table in Data Section.

TECH-B

SPECIFIED PUMP HEAD HEAD BELOW DISCHARGE

DRAWDOWN

SUBM.

14. PUMP BRAKE HORSEPOWER - Sum of bowl horsepower plus shaft loss (and the driver thrust bearing loss under certain conditions). 15. TOTAL PUMP EFFICIENCY (WATER TO WATER) -The efficiency of the complete pump less.the driver, with all pump losses taken into account. Efficiency = Specified Pump Head x Capacity 3960 x Brake Horsepower 16. OVERALL EFFICIENCY (WIRE TO WATER) - The efficiency of the pump and motor complete. Overall efficiency = total pump efficiency x motor efficiency. 17. SUBMERGENCE - Distance from liquid level to suction bell.

Vertical Turbine Pumps - Calculating Axial Thrust Under normal circumstances Vertical Turbine Pumps have a thrust load acting parallel to the pump shaft. This load is due to unbalanced pressure, dead weight and liquid direction change. Optimum selection of the motor bearing and correct determination of required bowl lateral for deep setting pumps require accurate knowledge of both the magnitude and direction (usually down) of the resultant of these forces. In addition, but with a less significant role, thrust influences shaft H.P. rating and shaft critical speeds.

DEAD WEIGHT In addition to the impeller force, dead weight (shaft plus impeller weight less the weight of the liquid displaced) acts downward. On pumps with settings less than 50 feet, dead weight may be neglected on all but the most critical applications as it represents only a small part of the total force. On deeper setting pumps, dead weight becomes significant and must be taken into account. NOTE:

IMPELLER THRUST Impeller Thrust in the downward direction is due to the unbalanced discharge pressure across the eye area of the impeller. See diagram A. Counteracting this load is an upward force primarily due to the change in direction of the liquid passing through the impeller. The resultant of these two forces constitutes impeller thrust. Calculating this thrust using a thrust constant (K) will often produce only an approximate thrust value because a single constant cannot express the upthrust component which varies with capacity. To accurately determine impeller thrust, thrust-capacity curves based on actual tests are required. Such curves now exist for the "A" Line. To determine thrust, the thrust factor "K" is read from the thrust-capacity curve at the required capacity and given RPM. "K" is then multiplied by the Total Pump Head (Final Lab Head) times Specific Gravity of the pumped liquid. If impeller thrust is excessively high, the impeller can usually be hydraulically balanced. This reduces the value of "K". Balancing is achieved by reducing the discharge pressure above the impeller eye by use of balancing holes and rings. See diagram B.

We normally only take shaft weight into consideration as dead weight, the reason being that impeller weight less its liquid displacement weight is usually a small part of the total. SHAFT SLEEVES Finally, there can be an upward force across a head shaft sleeve or mechanical seal sleeve. In the case of can pumps with suction pressure, there can be an additional upward force across the impeller shaft area. Again for most applications these forces are small and can be neglected; however, when there is a danger of upthrusts or when there is high discharge pressure (above 600 psi) or high suction pressure (above 400 psi) these forces should be considered. MOTOR BEARING SIZING Generally speaking a motor for a normal thrust application has as standard, a bearing adequate for shutoff thrust. When practical, motor bearings rated for shutoff conditions are preferred. For high thrust applications (when shutoff thrust exceeds the standard motor bearing rating) the motor bearing may be sized for the maximum anticipated operating range of the pump. Should the pump operate to the left of this range for a short period of time, anti-fraction bearings such as angular contact or spherical roller can handle the overload. It should be remembered, however, that bearing life is approximately inversely proportional to the cube of the load. Should the load double, motor bearing life will be cut to 1⁄8 of its original value. Although down thrust overloading is possible, the pump must never be allowed to operate in a continuous up thrust condition even for a short interval without a special motor bearing equipped to handle it. Such upthrust will tail the motor bearing. CALCULATING MOTOR BEARING LOAD

(A)

(B)

Suction Pressure Discharge Pressure

As previously stated, for short setting non-hydraulic balanced pumps below 50 feet with discharge pressures below 600 psi and can pumps with Suction pressures below 100 psi only impeller thrust need be considered. Under these conditions:

Where:

Motor Bearing Load (lbs.) Timp = KHL x SG

Impeller Thrust (lbs.) K=Thrust factors (lbs./ft.) HL, = Lab Head (ft.) SG = Specific Gravity

NOTE: Although hydraulic balancing reduces impeller thrust, it also decreases efficiency by one to five points by providing an additional path for liquid recirculation. Of even greater concern is that should the hydraulic balancing holes become clogged, (unclean fluids, fluids with solid content, intermittent services, etc.), the impeller thrust will increase and possibly cause the driver to fail. Hydraulically balanced impellers cannot be used in applications requiring rubber bowl bearings because the flutes on the inside diameter of the bearings provide an additional path to the top side of the impeller, thus creating an additional down thrust.

For more demanding applications, the forces which should be considered are impeller thrust plus dead weight minus any sleeve or shaft area force. In equation form: Motor Bearing Load = Timp + Wt(1) – sleeve force(2) – shaft area force(3) =Tt

Hydraulically balanced impellers should be used as a ''last resort" for those situations where the pump thrust exceeds the motor thrust bearing capabilities.

TECH-B

CALCULATING AXIAL THRUST - CONTINUED

Shaft Dia (in)

Shaft Dead Wt. (lbs/ft.) Open Closed Lineshaft Lineshaft

THRUST BEARING LOSS

1.1

1.1

6.0

1.8

1.1

6.7

7.6

2.2

1.5

Thrust bearing loss is the loss of horsepower delivered to the pump at the thrust bearings due to thrust. In equation form: Tt LTB = .0075 BHP 100 1000

8.8

10.0

2.9

1.8

11.2

12.8

3.7

2.0

2.3

2.6

3

1 ⁄16

3.3

3.8

11⁄2

5.3

11

1 ⁄16 115⁄16 2 ⁄16

Sleeve Area (in) 1.0

1

3

Shaft Area (in2)

(1) Wt.= Shaft Dead Wt. x Setting In Ft. (2) Sleeve Force=Sleeve area x Discharge pressure (3) Shaft Area Force = Shaft area x Suction pressure *Oil Lube shaft does not displace liquid above the pumping water level and therefore has a greater net weight.

.78

( )(

)

where: LTB BHP Tt

= = = =

Thrust bearing loss (HP) Brake horsepower Motor Bearing Load (Lbs.) Timp+ Wt(1) – sleeve force(2) – shaft area force(3)

Vertical Turbine Bearing Material Data Material Description

Temp. and S.G. Limits

Remarks

1. Bronze-SAE 660 (Standard) #1104 ASTM-B-584-932

-50 to 250°F. Min S.G. of 0.6

General purpose material for non- abrasive, neutral pH service. 7% Tin/7% Lead/3% Zinc/83% Cu.

2 Bronze-SAE 64 (Zincless) #1107 ASTM-B-584-937

-50 to 180°F. Min. S.G. of 0.6

Similar to std. Bronze. Used for salt water services. 10% Tin/ 10% Lead/80% Cu.

3 Carbon Graphite Impregnated with Babbit2

-450 to 300o F. All Gravities

Corrosion resistant material not suitable for abrasive services. Special materials available for severe acid services and for temp. as high as 650°. Good for low specific gravity fluids because the carbon is self-lubricating.

4. Teflon 25% Graphite with 75% Teflon

-50 to 250° F. All Gravities

Corrosion resistant except for highly oxidizing solutions. Not suitable for abrasive services. Glass filled Teflon also available.

5. Cast Iron3 ASTM-A-48 CL30 Flash Chrome Coated

32 to 180° F. Min. S.G of 0.6

Used on non abrasive caustic services and some oil products. Avoid water services as bearings can rust to shaft when idle. Test with bronze bearings.

6. Lead Babbit

32 to 300° F.

Excellent corrosion resistance to pH of 2. Good in mildly abrasive sevices. 80% Lead/3% Tin/17% Antimony.

7. Rubber w/Phenolic backing (Nitrite Butadiene or Neoprene)

32 to 150° F.

Use in abrasive water services. Bearings must be wet prior to start-up for TPL 50’. Do not use: For oily services, for stuffing box bushing, or with hydraulically balanced impellers. For services that are corrosive, backing material other than Phenolic must be specified.

8. Hardened Metals: Sprayed on stainless steel shell (Tungsten Carbide)

All temperatures All Gravities

Expensive alternate for abrasive services. Hardfaced surfaces typically in the range of Rc72. Other coatings are chromium oxide, tungsten carbide, colmonoy, etc. Consult factory for pricing and specific recommendation.

TECH-B-8 Self Priming Pump System Guidelines Self-priming pumps are inherently designed to allow the pump to re-prime itself typically under lift conditions. These pumps are very effective to the end user in that they will eliminate the need for foot valves, vacuum and ejector pumps which can become clogged or be impractical to use for prolonged or remote operation. Although the pump itself is designed to accomplish this task, it is important to understand the principle of how self-priming is achieved so that the piping system can be designed so as not to conflict with this function. A self-priming pump, by definition, is a pump which will clear its passages of air if it becomes air bound and resume delivery of the pumpage without outside attention. To accomplish this, a charge of

TECH-B

liquid sufficient to prime the pump must be retained in the casing (See Fig. A) or in an accessory priming chamber. When the pump starts, the rotating impeller creates a partial vacuum; air from the suction piping is then drawn into this vacuum and is entrained in the liquid drawn from the priming chamber. This air-liquid mixture is then pumped into the air separation chamber (within the casing) where the air is separated from the liquid with the air being expelled out the discharge piping (Fig. B) and the liquid returning to the priming chamber. This cycle is repeated until all of the air from the suction piping has been expelled and replaced by pumpage and the prime has been established (Fig. C).

Fig. A

Fig. B

Fig. C

The following considerations should be made when designing a piping system for which a self-priming pump is to be used: •

Care should be exercised to insure that adequate liquid is retained in the priming chamber. For outdoor/remote installations a heating element may be required to prevent freezing. For dirty services a strainer may be required to keep solids from accumulating in the priming chamber, thus displacing priming liquid.

•

The static lift and suction piping should be minimized to keep priming time to a minimum. Excessive priming time can cause liquid in the priming chamber to vaporize before prime is achieved.

•

All connections in the suction piping should be leak-free as air could be sucked in, thus extending/compromising priming of the pump. (Pumps sealed with packing should be flushed to prevent air from being introduced.)

•

A priming bypass line (See Fig. D) should be installed so that back pressure is not created in the discharge piping during priming which would prevent the pump from priming Itself. (Self-priming pumps are not good air compressors!)

•

The suction piping should be designed such that no high points are created where air can be trapped/accumulate which can prevent priming. Historically this has been problematic on top unloading of rail cars. (See Fig. E)

Fig. D

NOTE: Goulds Model 3796 self-priming process pump is outlined in Section 1F.

NOT RECOMMENDED

RECOMMENDED

Fig. E Tank Car Unloading

TECH-B

TECH-B-9 Priming Time Calculations Priming time data for each Model 3796 pump size and speed is displayed on the individual performance curves where priming time is plotted versus effective static lift for maximum, minimum and intermediate impeller diameters. This data is for suction piping of the same nominal diameter as the pump suction, i.e. 3" piping and 3" pump suction, and must be corrected for suction pipe diameters different from the pump suction and for suction pipe lengths greater than the effective static lift. To calculate the total priming time for a given system: 1. Select the correct size and speed pump from the performance curve for the given rating.

2. Calculate the NPSH Available for the system. The available NPSH must be equal to or greater than the NPSH Required by the selected pump at the rating point.

4. Enter the priming time curve at the effective static lift calculated in Step 3. Proceed across to the impeller diameter selected for the specified rating and then downward to the bottom coordinate to determine the priming time (PTLes) to achieve the given lift. 5. Insert the priming time from Step 4 into the following formula to calculate the total system priming time: Priming Time - Seconds 2 PTT = PTLes x SPL x Dp Les Ds

()

NPSHA = P - (Ls + Vp + hf) where: where:

P = Pressure on surface of liquid in feet absolute Ls = Maximum static lift in feet from free surface of the liquid to the centerline of the impeller. Vp = Vapor pressure of the liquid at maximum pumping temperature in feet absolute. hf

= Suction pipe friction loss in feet at the required capacity.

3. Determine the effective static lift. Les = Ls x Sp. Gr. where:

Les = Effective static lift in feet. Ls = Maximum static lift in feet from free surface of the liquid to the centerline of the pump suction, or the highest point in the suction piping, whichever is greater.

Sp. Gr. = Specific gravity of the liquid.

TECH-B

PTT

= Total system priming time.

PTLes = Priming time in seconds for the effective static lift (Step 4.) SPL = Total suction pipe length above the free surface of the liquid in feet. Les

= Effective static lift.

Dp

= Nominal pipe diameter.

Ds

= Nominal pump suction diameter.

Section TECH-C Water Data TECH-C-1 Friction Loss for Water – Sched 40 Steel Pipe 1 U.S. ⁄ 8 In. (0.269" I.D.) Gallons per V V2 hf Minute (Ft./Sec.) 2g (Ft./100 ft.) 0.2 1.13 0.020 2.72 0.4 2.26 0.079 16.2 0.6 3.39 0.178 33.8 0.8 4.52 0.317 57.4 1.0 5.65 0.495 87.0 1.5 8.48 1.12 188 2.0 11.3 1.98 324 2.5 3.0 3.5 4.0 4.5 5 6 7 8 9 10 12 14

U.S. Gallons per Minute 4 5 6 7 8 9 10 12 14 16 18 20 25 30 35 40 45 50 60 70 80 90 100 120 140

1

3

⁄4 In. (0.364" I.D.) V2

V

hf

2g 1.23 1.85 2.47 3.08 4.62 6.17 7.17 9.25 10.79 12.33 13.87 15.42

3

0.024 0.053 0.095 0.148 0.332 0.591 0.923 1.33 1.81 2.36 2.99 3.69

3.7 7.6 12.7 19.1 40.1 69.0 105 148 200 259 326 398

⁄ 4 In. (0.824" I.D.)

1 In. (1.049" I.D.)

V

V2

2.41 3.01 3.61 4.21 4.81 5.42 6.02 7.22 8.42 9.63 10.8 12.0 15.1 18.1

2g 0.090 0.141 0.203 0.276 0.360 0.456 0.563 0.810 1.10 1.44 1.82 2.25 3.54 5.06

hf

V

V2

4.21 6.32 8.87 11.8 15.0 18.8 23.0 32.6 43.5 56.3 70.3 86.1 134 187

1.48 1.86 2.23 2.60 2.97 3.34 3.71 4.45 5.20 5.94 6.68 7.42 9.29 11.1 13.0 14.8 16.7 18.6 22.3 26.0

2g 0.034 0.053 0.077 0.105 0.137 0.173 0.214 0.308 0.420 0.548 0.694 0.857 1.34 1.93 2.62 3.43 4.33 5.35 7.71 10.5

⁄ 8 In. (0.493" I.D.)

V

V2 2g

1.01 1.34 1.68 2.52 3.36 4.20 5.04 5.88 6.72 7.56 8.40 10.1 11.8 13.4 15.1 16.8

0.016 0.028 0.044 0.099 0.176 0.274 0.395 0.538 0.702 0.889 1.10 1.58 2.15 2.81 3.56 4.39

hf

1.74 2.89 4.30 8.93 15.0 22.6 31.8 42.6 54.9 68.4 83.5 118 158 205 258 316

1 1⁄ 4 In. (1.3880" I.D.) hf

1.29 1.93 2.68 3.56 4.54 5.65 6.86 9.62 12.8 16.5 20.6 25.1 37.4 54.6 73.3 95.0 119 146 209 283

V

V2 2g

1.29 1.50 1.72 1.93 2.15 2.57 3.00 3.43 3.86 4.29 5.37 6.44 7.52 8.58 9.66 10.7 12.9 15.0 17.2 19.3 21.5 25.7

0.026 0.035 0.046 0.058 0.071 0.103 0.140 0.183 0.232 0.286 0.448 0.644 0.879 1.14 1.45 1.79 2.57 3.50 4.58 5.79 7.15 10.3

hf

0.70 0.93 1.18 1.46 1.77 2.48 3.28 4.20 5.22 6.34 9.66 13.6 18.5 23.5 29.5 36.0 51.0 68.8 89.2 112 138 197

1

⁄ 2 In. (0.622" I.D.)

V

V2

hf

2g

1.06 1.58 2.11 2.64 3.17 3.70 4.22 4.75 5.28 6.34 7.39 8.45 9.50 10.6 12.7 14.8

0.017 0.039 0.069 0.108 0.156 0.212 0.277 0.351 0.433 0.624 0.849 1.11 1.40 1.73 2.49 3.40

1.86 2.85 4.78 7.16 10.0 13.3 17.1 21.3 25.8 36.5 48.7 62.7 78.3 95.9 136 183

1 1⁄ 2 In. (1.610" I.D.) V

V2 2g

1.26 1.42 1.58 1.89 2.21 2.52 2.84 3.15 3.94 4.73 5.52 6.30 7.10 7.88 9.46 11.0 12.6 14.2 15.8 18.9 22.1

0.025 0.031 0.039 0.056 0.076 0.99 0.125 0.154 0.241 0.347 0.473 0.618 0.783 0.965 1.39 1.89 2.47 3.13 3.86 5.56 7.56

hf

0.56 0.69 0.83 1.16 1.53 1.96 2.42 2.94 4.50 6.26 8.38 10.8 13.5 16.4 23.2 31.3 40.5 51.0 62.2 88.3 119

U.S. Gallons per Minute 0.2 0.4 0.6 0.8 1.0 1.5 2.0 2.5 3.0 3.5 4.0 4.5 5 6 7 8 9 10 12 14

U.S. Gallons per Minute 4 5 6 7 8 9 10 12 14 16 18 20 25 30 35 40 45 50 60 70 80 90 100 120 140

TECH-C

U.S. Gallons per Minute 30 35 40 50 60 80 100 120 140 160 180 200 220 240 260 280 300 350 400 500 600 700 800 1000

U.S. Gallons per Minute 140 160 180 200 240 280 320 360 400 450 500 600 700 800 900 1000 1200 1400 1600 1800 2000 2400 2800 3200 3600 4000

21⁄ 2 In. (2.469" I.D.)

2 In. (2.067" I.D.) V 2.87 3.35 3.82 4.78 5.74 7.65 9.56 11.5 13.4 15.3 17.2 19.1 21.0 22.9 24.9 26.8 28.7

V2 2g 0.128 0.174 0.227 0.355 0.511 0.909 1.42 2.05 2.78 3.64 4.60 5.68 6.88 8.18 9.60 11.1 12.8

hf

V

1.82 2.42 3.10 4.67 6.59 11.4 17.4 24.7 33.2 43.0 54.1 66.3 80.0 95.0 111 128 146

2.01 2.35 2.68 3.35 4.02 5.36 6.70 8.04 9.38 10.7 12.1 13.4 14.7 16.1 17.4 18.8 20.1 23.5 26.8 33.5

V2 2g 0.063 0.085 0.112 0.174 0.251 0.447 0.698 1.00 1.37 1.79 2.26 2.79 3.38 4.02 4.72 5.47 6.28 8.55 11.2 17.4

hf 0.75 1.00 1.28 1.94 2.72 4.66 7.11 10.0 13.5 17.4 21.9 26.7 32.2 38.1 44.5 51.3 58.5 79.2 103 160

3 In. (3.068" I.D.) V

V2 2g

hf

2.17 2.60 3.47 4.34 5.21 6.08 6.94 7.81 8.68 9.55 10.4 11.3 12.2 13.0 15.2 17.4 21.7 26.0 30.4 34.7

0.073 0.105 0.187 0.293 0.421 0.574 0.749 0.948 1.17 1.42 1.69 1.98 2.29 2.63 3.57 4.68 7.32 10.5 14.3 18.7

0.66 0.92 1.57 2.39 3.37 4.51 5.81 7.28 8.90 10.7 12.6 14.7 16.9 19.2 26.3 33.9 52.5 74.8 101 131

3 1⁄ 2 In. (3.548" I.D.) V

V2 2g

hf

1.95 2.60 3.25 3.89 4.54 5.19 5.84 6.49 7.14 7.79 8.44 9.09 9.74 11.3 13.0 16.2 19.5 22.7 26.0 32.5

0.059 0.105 0.164 0.236 0.321 0.419 0.530 0.655 0.792 0.943 1.11 1.28 1.47 2.00 2.62 4.09 5.89 8.02 10.5 16.44

0.45 0.77 1.17 1.64 2.18 2.80 3.50 4.27 5.12 6.04 7.04 8.11 9.26 12.4 16.2 25.0 35.6 48.0 62.3 96.4

4 In. (4.026" I.D.)

5 In. (5.047" I.D.)

6 In. (6.065" I.D.)

8 In. (7.981" I.D.)

V

V2

V2

V2

3.53 4.03 4.54 5.04 6.05 7.06 8.06 9.07 10.1 11.3 12.6 15.1 17.6 20.2 22.7 25.2 30.2 35.3

2g 0.193 0.253 0.320 0.395 0.569 0.774 1.01 1.28 1.58 2.00 2.47 3.55 4.84 6.32 8.00 9.87 14.2 19.3

TECH-C

hf

V

V2

1.16 1.49 1.86 2.27 3.21 4.30 5.51 6.92 8.47 10.5 13.0 18.6 25.0 32.4 40.8 50.2 72.0 97.6

2.25 2.57 2.89 3.21 3.85 4.49 5.13 5.77 6.41 7.23 8.02 9.62 11.2 12.8 14.4 16.0 19.2 22.5 25.7 28.8 32.1

2g 0.078 0.102 0.129 0.160 0.230 0.313 0.409 0.518 0.639 0.811 0.999 1.44 1.96 2.56 3.24 4.00 5.76 7.83 10.2 12.9 16.0

hf 0.38 0.49 0.61 0.74 1.03 1.38 1.78 2.22 2.72 3.42 4.16 5.88 7.93 10.2 12.9 15.8 22.5 30.4 39.5 49.7 61.0

V

2g

2.22 2.66 3.11 3.55 4.00 4.44 5.00 5.55 6.66 7.77 8.88 9.99 11.1 13.3 15.5 17.8 20.0 22.2 26.6 31.1 35.5

0.077 0.110 0.150 0.196 0.240 0.307 0.388 0.479 0.690 0.939 1.23 1.55 1.92 2.76 3.76 4.91 6.21 7.67 11.0 15.0 19.6

hf

0.30 0.42 0.56 0.72 0.90 1.09 1.37 1.66 2.34 3.13 4.03 5.05 6.17 8.76 11.8 15.4 19.4 23.8 34.2 46.1 59.9

V

2g

2.57 2.89 3.21 3.85 4.49 5.13 5.77 6.41 7.70 8.98 10.3 11.5 12.8 15.4 18.0 20.5 23.1 25.7

0.102 0.129 0.160 0.230 0.313 0.409 0.518 0.639 0.920 1.25 1.64 2.07 2.56 3.68 5.01 6.55 8.28 10.2

hf

0.28 0.35 0.42 0.60 0.80 1.02 1.27 1.56 2.20 2.95 3.82 4.79 5.86 8.31 11.2 14.5 18.4 22.6

U.S. Gallons per Minute 30 35 40 50 60 80 100 120 140 160 180 200 220 240 260 280 300 350 400 500 600 700 800 1000

U.S. Gallons per Minute 140 160 180 200 240 280 320 360 400 450 500 600 700 800 900 1000 1200 1400 1600 1800 2000 2400 2800 3200 3600 4000

U.S. Gallons per Minute 800 900 1000 1200 1400 1600 1800 2000 2500 3000 3500 4000 4500 5000 6000 7000 8000 9000 10,000 12,000 14,000 16,000 18,000 20,000

U.S. Gallons per Minute 2000 3000 4000 5000 6000 8000 10,000 12,000 14,000 16,000 18,000 20,000 22,000 24,000 26,000 28,000 30,000 34,000 38,000 42,000 46,000 50,000

10 In. (10.020" I.D.) V

V2 2g 0.165 0.208 0.257 0.370 0.504 0.659 0.834 1.03 1.62 2.32 3.13 4.12 5.21 6.43 9.26 12.6 16.5 20.8

3.25 3.66 4.07 4.88 5.70 6.51 7.32 8.14 10.2 12.2 14.2 16.3 18.3 20.3 24.4 28.5 32.5 36.6

12 In. (11.938" I.D.)

hf 0.328 0.410 0.500 0.703 0.940 1.21 1.52 1.86 2.86 4.06 5.46 7.07 8.88 10.9 15.6 21.1 27.5 34.6

14 In. (13.124" I.D.)

V

V2 2g

hf

2.58 2.87 3.44 4.01 4.59 5.16 5.73 7.17 8.60 10.0 11.5 12.9 14.3 17.2 20.1 22.9 25.8 28.7 34.4 40.1

0.103 0.128 0.184 0.250 0.327 0.414 0.511 0.799 1.15 1.55 2.04 2.59 3.19 4.60 6.26 8.17 10.3 12.8 18.3 25.0

0.173 0.210 0.296 0.395 0.609 0.636 0.776 1.19 1.68 2.25 2.92 3.65 4.47 6.39 8.63 11.2 14.1 17.4 24.8 33.5

16 In. (15.000" I.D.)

V

V2 2g

hf

2.37 2.85 3.32 3.79 4.27 4.74 5.93 7.11 8.30 9.48 10.7 11.9 14.2 16.6 19.0 21.3 23.7 28.5 33.2 37.9 42.7

0.087 0.126 0.171 0.224 0.283 0.349 0.546 0.786 1.07 1.40 1.77 2.18 3.14 4.28 5.59 7.08 8.74 12.6 17.1 22.4 28.3

0.131 0.185 0.247 0.317 0.395 0.483 0.738 1.04 1.40 1.81 2.27 2.78 3.95 5.32 6.90 8.7 10.7 15.2 20.7 26.8 33.9

V

V2 2g

hf

2.90 3.27 3.63 4.54 5.45 6.35 7.26 8.17 9.08 10.9 12.7 14.5 16.3 18.2 21.8 25.4 29.0 32.7 36.3

0.131 0.166 0.205 0.320 0.461 0.627 0.820 1.04 1.28 1.84 2.51 3.28 4.15 5.12 7.38 10.0 13.1 16.6 20.5

0.163 0.203 0.248 0.377 0.535 0.718 0.921 1.15 1.41 2.01 2.69 3.498 4.38 5.38 7.69 10.4 13.5 17.2 21.2

18 In. (16.876" I.D.)

20 In. (18.812" I.D.)

24 In. (22.624" I.D.)

V

V2

V2

V2

2.87 4.30 5.74 7.17 8.61 11.5 14.3 17.2 20.1 22.9 25.8 28.7 31.6 34.4 37.3 40.2 43.0

2g 0.128 0.288 0.512 0.799 1.15 2.05 3.20 4.60 6.27 8.19 10.4 12.8 15.5 18.4 21.6 25.1 28.8

hf 0.139 0.297 0.511 0.781 1.11 1.93 2.97 4.21 5.69 7.41 9.33 11.5 13.9 16.5 19.2 22.2 25.5

V

2g 3.46 4.62 5.77 6.92 9.23 11.5 13.8 16.2 18.5 20.8 23.1 25.4 27.7 30.0 32.3 34.6 39.2 43.9

0.186 0.331 0.517 0.745 1.32 2.07 2.98 4.06 5.30 6.71 8.28 10.0 11.9 14.0 16.2 18.6 23.9 29.9

hf

0.174 0.298 0.455 0.645 1.11 .70 2.44 3.29 4.26 5.35 6.56 7.91 9.39 11.0 12.7 14.6 18.7 23.2

V

2g

3.19 3.99 4.79 6.38 7.98 9.58 11.2 12.8 14.4 16.0 17.6 19.2 20.7 22.3 23.9 27.1 30.3 33.5 36.7 39.9

0.158 0.247 0.356 0.633 0.989 1.42 1.94 2.53 3.21 3.96 4.79 5.70 6.69 7.76 8.91 11.4 14.3 17.5 20.9 24.7

hf

0.120 0.181 0.257 0.441 0.671 0.959 1.29 1.67 2.10 2.58 3.10 3.67 4.29 4.96 5.68 7.22 9.00 11.0 13.2 15.5

U.S. Gallons per Minute 800 900 1000 1200 1400 1600 1800 2000 2500 3000 3500 4000 4500 5000 6000 7000 8000 9000 10,000 12,000 14,000 16,000 18,000 20,000

U.S. Gallons per Minute 2000 3000 4000 5000 6000 8000 10,000 12,000 14,000 16,000 18,000 20,000 22,000 24,000 26,000 28,000 30,000 34,000 38,000 42,000 46,000 50,000

Reprinted from PIPE FRICTION MANUAL, Third Edition. Copyright 1961 by Hydraulic Institute.

TECH-C

U.S. Gallons per Minute 5,000 6,000 7,000 8,000 9,000 10,000 12,000 14,000 16,000 18,000 20,000 25,000 30,000 35,000 40,000 45,000 50,000 60,000 70,000 80,000 90,000 100,000 120,000 140,000 160,000 180,000

U.S. Gallons per Minute 16,000 18,000 20,000 25,000 30,000 35,000 40,000 45,000 50,000 60,000 70,000 80,000 90,000 100,000 120,000 140,000 160,000 180,000 200,000 250,000 300,000 350,000

30 In. V 2.43 2.91 3.40 3.89 4.37 4.86 5.83 6.80 7.77 8.74 9.71 12.1 14.6 17.0 19.4 21.9 24.3 29.1 34.0 38.9

V2 2g 0.0917 0.132 0.180 0.235 0.297 0.367 0.528 0.719 0.939 1.19 1.47 2.29 3.30 4.49 5.87 7.42 9.17 13.2 18.0 23.5

36 In. hf 0.0535 0.0750 0.100 0.129 0.161 0.196 0.277 0.371 0.478 0.598 0.732 1.13 1.61 2.17 2.83 3.56 4.38 6.23 8.43 11.0

42 In.

V

V2 2g

hf

2.52 2.84 3.15 3.78 4.41 5.04 5.67 6.30 7.88 9.46 11.03 12.6 14.1 15.8 18.9 22.1 25.2 28.4 31.5 37.8

0.0988 0.125 0.154 0.222 0.303 0.395 0.500 0.618 0.965 1.39 1.89 2.47 3.13 3.86 5.56 7.56 9.88 12.5 15.4 22.2

0.0442 0.0551 0.0670 0.0942 0.126 0.162 0.203 0.248 0.378 0.540 0.724 0.941 1.18 1.45 2.07 2.81 3.66 4.59 5.64 8.05

V

V2 2g

hf

2.78 3.24 3.71 4.17 4.63 5.79 6.95 8.11 9.26 10.42 11.6 13.9 16.2 18.5 20.8 23.2 27.8 32.4 37.1 41.7

0.120 0.163 0.213 0.270 0.333 0.521 0.750 1.02 1.33 1.69 2.08 3.00 4.08 5.33 6.75 8.33 12.0 16.3 21.3 27.0

0.0441 0.0591 0.0758 0.0944 0.115 0.176 0.250 0.334 0.433 0.545 0.668 0.946 1.27 1.66 2.08 2.57 3.67 4.98 6.46 8.12

48 In.

54 In.

60 In.

V

V2

V2

V2

2.84 3.19 3.55 4.43 5.32 6.21 7.09 7.98 8.87 10.64 12.4 14.2 16.0 17.7 21.3 24.8 28.4 31.9 35.5

2g 0.125 0.158 0.195 0.305 0.440 0.598 0.782 0.989 1.221 1.76 2.39 3.13 3.96 4.89 7.03 9.57 12.5 15.8 19.5

TECH-C

hf 0.0391 0.0488 0.0598 0.0910 0.128 0.172 0.222 0.278 0.341 0.484 0.652 0.849 1.06 1.30 1.87 2.51 3.26 4.11 5.05

V

2g

2.80 3.50 4.20 4.90 5.60 6.30 7.00 8.40 9.81 11.21 12.6 14.0 16.8 19.6 22.4 25.2 28.0 35.0 42.0

0.122 0.191 0.274 0.374 0.488 0.618 0.762 1.098 1.49 1.95 2.47 3.05 4.39 5.98 7.81 9.88 12.2 19.1 27.4

hf

0.0333 0.0504 0.0713 0.0958 0.124 0.155 0.189 0.267 0.358 0.465 0.586 0.715 1.02 1.38 1.80 2.26 2.77 4.32 6.19

V

2g

2.84 3.40 3.97 4.54 5.11 5.67 6.81 7.94 9.08 10.21 11.3 13.6 15.9 18.2 20.4 22.7 28.4 34.0 39.7

0.125 0.180 0.245 0.320 0.405 0.500 0.720 0.980 1.28 1.62 2.00 2.88 3.92 5.12 6.48 8.00 12.5 18.0 24.5

hf

0.0301 0.0424 0.0567 0.0730 0.0916 0.112 0.158 0.213 0.275 0.344 0.420 0.600 0.806 1.04 1.32 1.62 2.52 3.60 4.88

U.S. Gallons per Minute 5,000 6,000 7,000 8,000 9,000 10,000 12,000 14,000 16,000 18,000 20,000 25,000 30,000 35,000 40,000 45,000 50,000 60,000 70,000 80,000 90,000 100,000 120,000 140,000 160,000 180,000

U.S. Gallons per Minute 16,000 18,000 20,000 25,000 30,000 35,000 40,000 45,000 50,000 60,000 70,000 80,000 90,000 100,000 120,000 140,000 160,000 180,000 200,000 250,000 300,000 350,000

TECH-C-2 Resistance Coefficients for Valves and Fittings REGULAR SCREWED 45° ELL.

BELL-MOUTH INLET OR REDUCER K = 0.05

LONG RADIUS FLANGED 45° ELL.

SQUARE EDGED INLET K = 0.5

INWARD PROJECTING PIPE K = 1.0

SCREWED RETURN BEND

NOTE: K DECREASES WITH INCREASING WALL THICKNESS OF PIPE AND ROUNDING OF EDGES FLANGED RETURN BEND

REGULAR SCREWED 90° ELL.

LONG RADIUS SCREWED 90° ELL.

LINE FLOW

SCREWED TEE

BRANCH FLOW

REGULAR FLANGED 90° ELL.

LINE FLOW FLANGED TEE

LONG RADIUS FLANGED 90° ELL. BRANCH FLOW

Chart 1 Where: h = Frictional Resistance in Feet of Liquid V = Average Velocity in Feet/Second in a Pipe of Corresponding Diameter

V2 h = K 2g g = 32.17 Feet/Second/Second K = Resistance Coefficient For Valve or Fitting

TECH-C

BASKET STRAINER

SCREWED

GLOBE VALVE FLANGED

FOOT VALVE

SCREWED

GATE VALVE

FLANGED

COUPLINGS AND UNIONS

SCREWED REDUCING BUSHING AND COUPLING V2 h=K 2 2g SWING CHECK VALVE

USED AS A REDUCER K = 0.05 – 2.0 SEE ALSO FIG. 3 USED AS INCREASER LOSS IS UP TO 40% MORE THAN THAT CAUSED BY A SUDDEN ENLARGEMENT

FLANGED

SUDDEN ENLARGEMENT SCREWED

h = (V1 – V2)2 FEET OF FLUID 2g SEE ALSO EQUATION(5) IF A2 –  SO THAT V2 = 0 h = V12 FEET OF FLUID

ANGLE VALVE

FLANGED

V2 h = K 2g

Chart 2 Reprinted from PIPE FRICTION MANUAL, Third Edition, Copyright 1961 by Hydraulic Institute.

TECH-C

2g

TECH-C-3 Resistance Coefficients for Increasers and Diffusers

Reprinted from PIPE FRICTION MANUAL, Third Edition. Copyright 1961 by Hydraulic Institute.

TECH-C-4 Resistance Coefficients for Reducers

Reprinted from PIPE FRICTION MANUAL, Third Edition. Copyright 1961 by Hydraulic Institute.

TECH-C

TECH-C-5 Properties of Water at Various Temperatures from 32° to 705.4°F Temp. F

Temp. C

32 40 45 50 55

0 4.4 7.2 10.0 12.8

SPECIFIC GRAVITY 60 F Reference 1.002 1.001 1.001 1.001 1.000

60 65 70 75 80

15.6 18.3 21.1 23.9 26.7

85 90 95 100 110

Wt. in Lb/Cu Ft

Vapor Pressure Psi Abs

62.42 62.42 62.40 62.38 62.36

0.0885 0.1217 0.1471 0.1781 0.2141

Vapor Pressure* Feet Abs. (At Temp.) 0.204 0.281 0.340 0.411 0.494

1.000 .999 .999 .998 .998

62.34 62.31 62.27 62.24 62.19

0.2653 0.3056 0.3631 0.4298 0.5069

0.591 0.706 0.839 0.994 1.172

29.4 32.2 35.0 37.8 43.3

.997 .996 .995 .994 .992

62.16 62.11 62.06 62.00 61.84

0.5959 0.9682 0.8153 0.9492 1.275

1.379 1.167 1.890 2.203 2.965

120 130 140 150 160

48.9 54.4 60.0 65.5 71.1

.990 .987 .985 .982 .979

61.73 61.54 61.39 61.20 61.01

1.692 2.223 2.889 3.718 4.741

3.943 5.196 6.766 8.735 11.172

170 180 190 200 212

76.7 82.2 87.7 93.3 100.0

.975 .972 .968 .966 .959

60.79 60.57 60.35 60.13 59.81

5.992 7.510 9.339 11.526 14.696

14.178 17.825 22.257 27.584 35.353

220 240 260 280 300

104.4 115.6 126.7 137.8 148.9

.956 .948 .939 .929 .919

59.63 59.10 58.51 58.00 57.31

17.186 24.97 35.43 49.20 67.01

41.343 60.77 87.05 122.18 168.22

320 340 360 380 400

160.0 171.1 182.2 193.3 204.4

.909 .898 .886 .874 .860

56.66 55.96 55.22 54.47 53.65

89.66 118.01 153.04 195.77 247.31

227.55 303.17 398.49 516.75 663.42

420 440 460 480

215.6 226.7 237.8 248.9

.847 .833 .818 .802

52.80 51.92 51.02 50.00

308.83 381.59 466.9 566.1

841.17 1056.8 1317.8 1628.4

500 520 540 560

260.0 271.1 282.2 293.3

.786 .766 .747 .727

49.02 47.85 46.51 45.3

680.8 812.4 962.5 1133.1

1998.2 2446.7 2972.5 3595.7

580 600 620 640

304.4 315.6 326.7 337.8

.704 .679 .650 .618

43.9 42.3 40.5 38.5

1325.8 1524.9 1786.6 2059.7

4345. 5242. 6341. 7689.

660 680 700 705.4

348.9 360.0 371.1 374.1

.577 .526 .435 .319

36.0 32.8 27.1 19.9

2365.4 2708.1 3039.7 3206.2

9458. 11878. 16407. 23187.

* Vapor pressure in feet of wate (Abs.) Converted from PSIA using sp. gr. at temperature.

TECH-C

TECH-C-6 Atmospheric Pressure, Barometer Reading and Boiling Point of Water at Various Altitudes Altitude Feet — 1000 — 500 0 — 500 I — I 1000 1500 2000 2500 3000 3500 4000 4500 5000 5500 6000 6500 7000 7500 8000 8500 9000 9500 10000 15000

— — — I

Barometric Reading

Atmos. Pressure

Meters

In. Hg.

Mm. Hg.

psia

Ft. Water

Boiling Pt. Of Water °F

304.8 152.4 0.0 152.4 304.8 457.2 609.6 762.0 914.4 1066.8 1219.2 1371.6 1524.0 1676.4 1828.8 1981.2 2133.6 2286.0 2438.4 2590.8 2743.2 2895.6 3048.0 4572.0

31.0 30.5 29.9 29.4 28.9 28.3 27.8 27.3 26.8 26.3 25.8 25.4 24.9 24.4 24.0 23.5 23.1 22.7 22.2 21.8 21.4 21.0 20.6 16.9

788 775 760 747 734 719 706 694 681 668 655 645 633 620 610 597 587 577 564 554 544 533 523 429

15.2 15.0 14.7 14.4 14.2 13.9 13.7 13.4 13.2 12.9 12.7 12.4 12.2 12.0 11.8 11.5 11.3 11.1 10.9 10.7 10.5 10.3 10.1 8.3

35.2 34.6 33.9 33.3 32.8 32.1 31.5 31.0 30.4 29.8 29.2 28.8 28.2 27.6 27.2 26.7 26.2 25.7 25.2 24.7 24.3 23.8 23.4 19.2

213.8 212.9 212.0 211.1 210.2 209.3 208.4 207.4 206.5 205.6 204.7 203.8 202.9 201.9 201.0 200.1 199.2 198.3 197.4 196.5 195.5 194.6 193.7 184.0

TECH-C

TECH-C-7 Saturation: Temperatures Steam Data Temp. F t 32 35 40 45 50 60 70 80 90 100 110 120 130 140 150 160 170 180 190 200 210 212 220 230 240 250 260 270 280 290 300 320 340 360 380 400 420 440 460 480 500 520 540 560 580 600 620 640 660 680 700 705.4

Abs. press. Specific Volume Lb Sat. Sat. Sq. In. Liquid Evap Vapor p vf vfg vg 0.08854 0.09995 0.12170 0.14752 0.17811 0.2563 0.3631 0.5069 0.6982 0.9492 1.2748 1.6924 2.2225 2.8886 3.718 4.741 5.992 7.510 9.339 11.526 14.123 14.696 17.186 20.780 24.969 29.825 35.429 41.858 49.203 57.556 67.013 89.66 118.01 153.04 195.77 247.31 308.83 381.59 466.9 566.1 680.8 812.4 962.5 1133.1 1325.8 1542.9 1786.6 2059.7 2365.4 2708.1 3093.7 3206.2

TECH-C

0.01602 0.01602 0.01602 0.01602 0.01603 0.01604 0.01606 0.01608 0.01610 0.01613 0.01617 0.01620 0.01625 0.01629 0.01634 0.01639 0.01645 0.01651 0.01657 0.01663 0.01670 0.01672 0.01677 0.01684 0.01692 0.01700 0.01709 0.01717 0.01726 0.01735 0.01745 0.01765 0.01787 0.01811 0.01836 0.01864 0.01894 0.01926 0.0196 0.0200 0.0204 0.0209 0.0215 0.0221 0.0228 0.0236 0.0247 0.0260 0.0278 0.0305 0.0369 0.0503

3306 2947 2444 2036.4 1703.2 1206.6 867.8 633.1 468.0 350.3 265.3 203.25 157.32 122.99 97.06 77.27 62.04 50.21 40.94 33.62 27.80 26.78 23.13 19.365 16.306 13.804 11.746 10.044 8.628 7.444 6.449 4.896 3.770 2.939 2.317 1.8447 1.4811 1.1979 0.9748 0.7972 0.6545 0.5385 0.4434 0.3647 0.2989 0.2432 0.1955 0.1538 0.1165 0.0810 0.0392 0

3306 2947 2444 2036.4 1703.2 1206.7 867.9 633.1 468.0 350.4 265.4 203.27 157.34 123.01 97.07 77.29 62.06 50.23 40.96 33.64 27.82 26.80 23.15 19.382 16.323 13.821 11.763 10.061 8.645 7.461 6.446 4.914 3.788 2.957 2.335 1.8633 1.5000 1.2171 0.9944 0.8172 0.6749 0.5594 0.4649 0.3868 0.3217 0.2668 0.2201 0.1798 0.1442 0.1115 0.0761 0.0503

Enthalpy Sat. Liquid hf 0.00 3.02 8.05 13.06 18.07 28.06 38.04 48.02 57.99 67.97 77.94 87.92 97.90 107.89 117.89 127.89 137.90 147.92 157.95 167.99 178.05 180.07 188.13 198.23 208.34 218.48 228.64 238.84 249.06 259.31 269.59 290.28 311.13 332.18 353.45 374.97 396.77 418.90 441.4 464.4 487.8 511.9 536.6 562.2 588.9 617.0 646.7 678.6 714.2 757.3 823.3 902.7

Entropy

Evap hfg

Sat. Vapor hg

Sat. Liquid sf

Sfg sfg

Sat Vapor sg

1075.8 1074.1 1071.3 1068.4 1065.6 1059.9 1054.3 1048.6 1042.9 1037.2 1031.6 1025.8 1020.0 1041.1 1008.2 1002.3 996.3 990.2 984.1 977.9 971.6 970.3 965.2 958.8 952.2 945.5 938.7 931.8 924.7 917.5 910.1 894.9 879.0 862.2 844.6 826.0 806.3 785.4 763.2 739.4 713.9 686.4 656.6 624.2 588.4 548.5 503.6 452.0 390.2 309.9 172.1 0

1075.8 1077.1 1079.3 1081.5 1083.7 1088.0 1092.3 1096.6 1100.9 1105.2 1109.5 1113.7 1117.9 1122.0 1126.1 1130.2 1134.2 1138.1 1142.0 1145.9 1149.7 1150.4 1153.4 1157.0 1160.5 1164.0 1167.3 1170.6 1173.8 1176.8 1179.7 1185.2 1190.1 1194.4 1198.1 1201.0 1203.1 1204.3 1204.6 1203.7 1201.7 1198.2 1193.2 1186.4 1177.3 1165.5 1150.3 1130.5 1104.4 1067.2 995.4 902.7

0.0000 0.0061 0.0162 0.0262 0.0361 0.0555 0.0745 0.0932 0.1115 0.1295 0.1471 0.1645 0.1816 0.1984 0.2149 0.2311 0.2472 0.2630 0.2785 0.2938 0.3090 0.3120 0.3239 0.3387 0.3531 0.3675 0.3817 0.3958 0.4096 0.4234 0.4369 0.4637 0.4900 0.5158 0.5413 0.5664 0.5912 0.6158 0.6402 0.6645 0.6887 0.7130 0.7374 0.7621 0.7872 0.8131 0.8398 0.8679 0.8987 0.9351 0.9905 1.0680

2.1877 2.1709 2.1435 2.1167 2.0903 2.0393 1.9902 1.9428 1.8972 1.8531 1.8106 1.7694 1.7296 1.6910 1.6537 1.6174 1.5822 1.5480 1.5147 1.4824 1.4508 1.4446 1.4201 1.3901 1.3609 1.3323 1.3043 1.2769 1.2501 1.2238 1.1980 1.1478 1.0992 1.0519 1.0059 0.9608 0.9166 0.8730 0.8298 0.7868 0.7438 0.7006 0.6568 0.6121 0.5659 0.5176 0.4664 0.4110 0.3485 0.2719 0.1484 0

2.1877 2.1770 2.1597 2.1429 2.1264 2.0948 2.0647 2.0360 2.0087 1.9826 1.9577 1.9339 1.9112 1.8894 1.8685 1.8485 1.8293 1.8109 1.7932 1.7762 1.7598 1.7566 1.7440 1.7288 1.7140 1.6998 1.6860 1.6727 1.6597 1.6472 1.6350 1.6115 1.5891 1.5677 1.5471 1.5272 1.5078 1.4887 1.4700 1.4513 1.4325 1.4136 1.3942 1.3742 1.3532 1.3307 1.3062 1.2789 1.2472 1.2071 1.1389 1.0580

Temp F t 32 35 40 45 50 60 70 80 90 100 110 120 130 140 150 160 170 180 190 200 210 212 220 230 240 250 260 270 280 290 300 320 340 360 380 400 420 440 460 480 500 520 540 560 580 600 620 640 660 680 700 705.4

TECH-C-7 Saturation: Pressures Steam Data Abs. press. Lb Temp. Sq. In. Liquid p t 1.0 2.0 3.0 4.0 5.0 6.0 7.0 8.0 9.0 10 14.696 15 20 30 40 50 60 70 80 90 100 120 140 160 180 200 250 300 350 400 450 500 550 600 700 800 900 1000 1100 1200 1300 1400 1500 2000 2500 3000 3206.2

101.74 126.08 141.48 152.97 162.24 170.06 176.85 182.86 188.28 193.21 212.00 213.03 227.96 250.33 267.25 281.01 292.71 302.92 312.03 320.27 327.81 341.25 353.02 363.53 373.06 381.79 400.95 417.33 431.72 444.59 456.28 467.01 476.93 486.21 503.10 518.23 531.98 544.61 556.31 567.22 577.46 587.10 596.23 635.82 668.13 695.36 705.40

Specific Volume Sat. Sat. Vapor Liquid vf vg

Sat. Liquid hr

0.01614 0.01623 0.01630 0.01636 0.01640 0.01645 0.01649 0.01653 0.01656 0.01659 0.01672 0.01672 0.01683 0.01701 0.01715 0.01727 0.01738 0.01748 0.01757 0.01766 0.01774 0.01789 0.01802 0.01815 0.01827 0.01839 0.01865 0.01890 0.01913 0.0193 0.0195 0.0197 0.0199 0.0201 0.0205 0.0209 0.0212 0.0216 0.0220 0.0223 0.0227 0.0231 0.0235 0.0257 0.0287 0.0346 0.0503

69.70 93.99 109.37 120.86 130.13 137.96 144.76 150.79 156.22 161.17 180.07 181.11 196.16 218.82 236.03 250.09 262.09 272.61 282.02 290.56 298.40 312.44 324.82 335.93 346.03 355.36 376.00 393.84 409.69 424.0 437.2 449.4 460.8 471.6 491.5 509.7 526.6 542.4 557.4 571.7 585.4 598.7 611.6 671.7 730.6 802.5 902.7

333.6 173.73 118.71 90.63 73.52 61.98 53.64 47.34 42.40 38.42 26.80 26.29 20.089 13.746 10.498 8.515 7.175 6.206 5.472 4.896 4.432 3.728 3.220 2.834 2.532 2.288 1.8438 1.5433 1.3260 1.1613 1.0320 0.9278 0.8422 0.7698 0.6554 0.5687 0.5006 0.4456 0.4001 0.3619 0.3293 0.3012 0.2765 0.1878 0.1307 0.0858 0.0503

Enthalpy

Entropy

Evap hfg

Sat. Vapor hg

Sat. Liquid sf

1036.3 1022.2 1031.2 1006.4 1001.0 996.2 992.1 988.5 985.2 982.1 970.3 969.7 960.1 945.3 933.7 924.0 915.5 907.9 901.1 894.7 888.8 877.9 868.2 859.2 850.8 843.0 825.1 809.1 794.2 780.5 767.4 755.0 743.1 731.6 709.7 688.9 668.8 649.4 630.4 611.7 593.2 574.7 556.3 463.4 360.5 217.8 0

1106.0 1116.2 1122.6 1127.3 1131.1 1134.2 1136.9 1139.3 1141.4 1143.3 1150.4 1150.8 1156.3 1164.1 1169.7 1174.1 1177.6 1180.6 1183.1 1185.3 1187.2 1190.4 1193.0 1195.1 1196.9 1198.4 1201.1 1202.8 1203.9 1204.5 1204.6 1204.4 1203.9 1203.2 1201.2 1198.6 1195.4 1191.8 1187.8 1183.4 1178.6 1173.4 1167.9 1135.1 1091.1 1020.3 902.7

0.1326 0.1749 0.2008 0.2198 0.2347 0.2472 0.2581 0.2674 0.2759 0.2835 0.3120 0.3135 0.3356 0.3680 0.3919 0.4110 0.4270 0.4409 0.4531 0.4641 0.4740 0.4916 0.5069 0.5204 0.5325 0.5435 0.5676 0.5879 0.6056 0.6214 0.6356 0.6487 0.6608 0.6720 0.6925 0.7108 0.7275 0.7430 0.7575 0.7711 0.7840 0.7963 0.8082 0.8619 0.9126 0.9731 1.0580

Evap s fg

Sat Vapor sg

Internal Energy Sat. Liquid Evap uf ufg

Abs. Sat press. Lb Vapor Sq. In. ug P

1.8456 1.7451 1.6855 1.6427 1.6094 1.5820 1.5586 1.5383 1.5203 1.5041 1.4446 1.4115 1.3962 1.3313 1.2844 1.2474 1.2168 1.1906 1.1676 1.1471 1.1286 1.0962 1.0682 1.0436 1.0217 1.0018 0.9588 0.9225 0.8910 0.8630 0.8378 0.8147 0.7934 0.7734 0.7371 0.7054 0.6744 0.6467 0.6205 0.5956 0.5719 0.5491 0.5269 0.4230 0.3197 0.1885 0

1.9782 1.9200 1.8863 1.8625 1.8441 1.8292 1.8167 1.8057 1.7962 1.7876 1.7566 1.7549 1.7319 1.6993 1.6763 1.6585 1.6438 1.6315 1.6207 1.6112 1.6026 1.5878 1.5751 1.5640 1.5542 1.5453 1.5263 1.5104 1.4966 1.4844 1.4734 1.4634 1.4542 1.4454 1.4296 1.4235 1.4020 1.3897 1.3780 1.3667 1.3559 1.3454 1.3351 1.2849 1.2322 1.1615 1.0580

69.70 93.98 109.36 120.85 130.12 137.94 144.74 150.77 156.19 161.14 180.02 181.06 196.10 218.73 235.90 249.93 261.90 272.38 281.76 290.27 298.08 312./05 324.35 335.39 345.42 354.68 375.14 392.79 408.55 422.6 435.5 447.6 458.8 469.4 488.8 506.6 523.1 538.4 552.9 566.7 580.0 592.7 605.1 662.2 717.3 783.4 872.9

1044.2 1051.9 1056.7 1060.2 1063.1 1065.4 1067.4 1069.2 1070.8 1072.2 1077.5 1077.8 1081.9 1087.8 1092.0 1095.3 1097.9 1100.2 1102.1 1103.7 1105.2 1107.6 1109.6 1111.2 1112.5 1113.7 1115.8 1117.1 1118.0 1118.5 1118.7 1118.6 1118.2 1117.7 1116.3 1114.4 1112.1 1109.4 1106.4 1103.0 1099.4 1095.4 1091.2 1065.6 1030.6 972.7 872.9

974.6 957.9 947.3 939.3 933.0 927.5 922.7 918.4 914.6 911.1 897.5 896.7 885.8 869.1 856.1 845.4 836.0 827.8 820.3 813.4 807.1 795.6 785.2 775.8 767.1 759.0 740.7 724.3 709.6 695.9 683.2 671.0 659.4 648.3 627.5 607.8 589.0 571.0 553.5 536.3 519.4 502.7 486.1 403.4 313.3 189.3 0

1.0 2.0 3.0 4.0 5.0 6.0 7.0 8.0 9.0 10 14.696 15 20 30 40 50 60 70 80 90 100 120 140 160 180 200 250 300 350 400 450 500 550 600 700 800 900 1000 1100 1200 1300 1400 1500 2000 2500 3000 3206.2

TECH-C

Section TECH-D Properties of Liquids TECH-D-1 Viscosity The viscosity of a fluid is that property which tends to resist a shearing force. It can be thought of as the internal friction resulting when one layer of fluid is made to move in relation to another layer. Consider the model shown in Fig. 1, which was used by Isaac Newton in first defining viscosity. It shows two parallel planes of fluid of area A separated by a distance dx and moving in the same direction at different velocities V1 and V2.

Fig. 2 Newtonian Liquid

Fig. 1 The velocity distribution will be linear over the distance dx, and dv experiments show that the velocity gradient, dx , is directly f proportional to the force per unit area, a . f = n x dv Where n is constant for a given liquid and A dx is called its viscosity. dv The velocity gradient, dx , describes the shearing experienced by the intermediate layers as they move with respect to each other. Therefore, it can be called the "rate of shear", S. Also, the F force per unit area, A , can be simplified and called the "shear force" or "shear stress," F. With these simplified terms, viscosity can be defined as follows:

Fig. 3 shows graphically the three most common types of NonNewtonian liquids. Group A shows a decreasing viscosity with an increasing rate of shear. This is known as a pseudo-plastic material. Examples of this type are grease, molasses, paint, soap, starch, and most emulsions. They present no serious pumping problems since they tend to thin out with the high rates of shear present in a pump. Group B shows a dilatant material or one in which the viscosity increases with an increasing rate of shear. Clay slurries and candy compounds are examples of dilatant liquids. Pumps must be selected with extreme care since these liquids can become almost solid if the shear rate is high enough. The normal procedure would be to oversize the pump somewhat and open up the internal clearances in an effort to reduce the shear rate. Group C shows a plastic material, The viscosity decreases with increasing rate of shear. However, a certain force must be applied before any movement is produced. This force is called the yield value of the material. Tomato catsup is a good example of this type of material. It behaves similar to a pseudo-plastic material from a pumping standpoint.

Group A

Group B

Group C

F= nXS Viscosity = n = F = shear stress S rate of shear Isaac Newton made the assumption that all materials have, at a given temperature, a viscosity that is independent of the rate of shear. In other words, a force twice as large would be required to move a liquid twice as fast. Fluids which behave this way are called Newtonian fluids. There are, of course, fluids which do not behave this way, in other words their viscosity is dependent on the rate of shear. These are known as Non-Newtonian fluids. Fig. 2 shows graphically the relationships between shear Stress (F,) rate of shear (S,) and viscosity (n) for a Newtonian liquid. The viscosity remains constant as shown in sketch 2, and in absolute units, the viscosity is the inverse slope of the line in sketch 1. Water and light oils are good examples of Newtonian liquids.

Pseudo-Plastic

Dilitant Fig. 3 Non-Newtonian Liquids

TECH-D

Plastic

The viscosity of some Non-Newtonian liquids is dependent upon time as well as shear rate. In other words, the viscosity at any particular time depends upon the amount of previous agitation or shearing of the liquid. A liquid whose viscosity decreases with time at a given shear rate is called a thixotropic liquid. Examples are asphalts, glues, molasses, paint, soap, starch, and grease. Liquids whose viscosity increases with time are called rheopectic liquids, but they are seldom encountered in pumping applications.

viscosity. The basic unit of kinematic viscosity is the stoke which is equal to a square centimeter per second in the Metric system. The corresponding English unit is square foot per second. The centistoke which is one-hundredth of a stoke is normally used in the charts. The following formula is used to obtain the kinematic viscosity when the dynamic or absolute viscosity is known: centistokes = centipoise sp. gr.

There are two basic viscosity parameters: dynamic (or absolute) viscosity and kinematic viscosity. Dynamic viscosities are given in terms of force required to move a unit area a unit distance. This is usually expressed in pound-seconds per square foot in the English system which is equal to slugs per foot-second. The Metric system is more commonly used, however, in which the unit is the dyne-second per square centimeter called the Poise. This is numerically equal to the gram per centimeter-second. For convenience, numerical values are normally expressed in centipoise, which are equal to onehundredth of a poise.

There are numerous types of viscometers available for determining liquid viscosities, most of which are designed for specific liquids or viscosity ranges. The Saybolt viscometers are probably the most widely used in the United States. The Saybolt Universal Viscometer measures low to medium viscosity, and the Saybolt Furol Viscometer measures high viscosities. The corresponding units are the SSU (Seconds Saybolt Universal) and the SSF (Seconds Saybolt Furol.) These units are found on most pipe friction and pump correction charts in addition to centistokes. A conversion chart for these and other units is shown in Fig. 4.

Most pipe friction charts and pump correction charts list kinematic

TECH-D-2A Viscosity Conversion Table The following table will give an approximate comparison of various viscosity ratings so that if the viscosity is given in terms other than Saybolt Universal, it can be translated quickly by following horizontally to the Saybolt Universal column. Seconds Kine- Seconds Seconds Seconds Saybolt matic Saybolt RedRedDegrees Degrees Seconds Seconds Seconds Seconds Seconds Seconds Universal Viscosity Furol wood 1 wood 2 Engler Barbey Parlin Parlin Parlin Parlin Ford Ford ssu Centissf (Stan(AdmirCup #7 Cup #10 Cup #15 Cup #20 Cup #3 Cup #4 stokes* dard) alty) 31 35 40 50

1.00 2.56 4.30 7.40

-

29 32.1 36.2 44.3

5.10 5.83

1.00 1.16 1.31 1.58

6200 2420 1440 838

-

-

-

-

-

-

60 70 80 90

10.3 13.1 15.7 18.2

12.95 13.70 14.44

52.3 60.9 69.2 77.6

6.77 7.60 8.44 9.30

1.88 2.17 2.45 2.73

618 483 404 348

-

-

-

-

-

-

100 150 200 250

20.6 32.1 43.2 54.0

15.24 19.30 23.5 28.0

85.6 128 170 212

10.12 14.48 18.90 23.45

3.02 4.48 5.92 7.35

307 195 144 114

40 46

-

-

-

-

-

300 400 500 600

65.0 87.60 110.0 132

32.5 41.9 51.6 61.4

254 338 423 508

28.0 37.1 46.2 55.4

8.79 11.70 14.60 17.50

95 70.8 56.4 47.0

52.5 66 79 92

15 21 25 30

6.0 7.2 7.8 8.5

3.0 3.2 3.4 3.6

30 42 50 58

20 28 34 40

700 800 900 1000

154 176 198 220

71.1 81.0 91.0 100.7

592 677 762 896

64.6 73.8 83.0 92.1

20.45 23.35 26.30 29.20

40.3 35.2 31.3 28.2

106 120 135 149

35 39 41 43

9.0 9.8 10.7 11.5

3.9 4.1 4.3 4.5

67 74 82 90

45 50 57 62

1500 2000 2500 3000

330 440 550 660

150 200 250 300

1270 1690 2120 2540

138.2 184.2 230 276

43.80 58.40 73.0 87.60

18.7 14.1 11.3 9.4

-

65 86 108 129

15.2 19.5 24 28.5

6.3 7.5 9 11

132 172 218 258

90 118 147 172

4000 5000 6000 7000

880 1100 1320 1540

400 500 600 700

3380 4230 5080 5920

368 461 553 645

117.0 146 175 204.5

7.05 5.64 4.70 4.03

-

172 215 258 300

37 47 57 67

14 18 22 25

337 425 520 600

230 290 350 410

8000 9000 10000

1760 1980 2200

800 900 1000

6770 7620 8460

737 829 921

233.5 263 292

3.52 3.13 2.82

-

344 387 430

76 86 96

29 32 35

680 780 850

465 520 575

15000 20000

3300 4400

1500 2000

13700 18400

-

438 584

2.50 1.40

-

650 860

147 203

53 70

1280 1715

860 1150

Reprinted from PIPE FRICTION MANUAL. Third Edition Copyright 1961 by Hydraulic institute

Fig. 4A

TECH-D

*Kinematic Viscosity (in centistokes) = Absolute Viscosity (in centipoises) Density

For values of 70 centistokes and above, use the following conversion:

When the Metric System terms centistokes and centipoises are are used, the density is numerically equal to the specific gravity. Therefore, the following expression can be used which will be sufficiently accurate for most calculations:

Above the range of this table and within the range of the viscosimeter, multiply the particular value by the following approximate factors to convert to SSU:

SSU = centistokes x 4.635

Viscosimeter

*Kinematic Viscosity (in centistokes) = Absolute Viscosity (in centipoises) Specific Gravity When the English System units are used, the density must be used rather than the specific gravity.

Factor

Saybolt Furol Redwood Standard Redwood Admiralty Engler – Degrees

10. 1.095 10.87 34.5

Viscosimeter Parlin cup #15 Parlin cup #20 Ford cup #4

Factor 98.2 187.0 17.4

TECH-D-2B Viscosity Conversion Table The following table will give an approximate comparison of various viscosity ratings so that if the viscosity is given in terms other than Saybolt Universal, it can be translated quickly by following horizontally to the Saybolt Universal column. Seconds Kine- Approx. Approx. Saybolt matic Seconds Gardner Universal Viscosity Mac Holt ssu Centi- Michael Bubble stokes* 31 35 40 50 60 70 80 90 100 150 200 250 300 400 500 600 700 800 900 1000 1500 2000 2500 3000 4000 5000 6000 7000 8000 9000 10000 15000 20000

1.00 2.56 4.30 7.40 10.3 13.1 15.7 18.2 20.6 32.1 43.2 54.0 65.0 87.0 110.0 132 154 176 198 220 330 440 550 660 880 1100 1320 1540 1760 1980 2200 3300 4400

125 145 165 198 225 270 320 370 420 470 515 570 805 1070 1325 1690 2110 2635 3145 3760 4170 4700 5220 7720 10500

A A B C D F G H I M Q T U V W X Y Z Z2 Z3

Seconds Seconds Seconds Seconds Seconds Seconds Seconds Approx. Seconds Zahn Zahn Zahn Zahn Zahn Demmier Demmier Seconds Pratt Cup #1 Cup #2 Cup #3 Cup #4 Cup #5 Cup #1 Cup #10 Stormer and 100 gpm Lambert Load "F" 38 47 54 62 73 90 -

18 20 23 26 29 37 46 55 63 72 80 88 -

22.5 24.5 27 29 40 51 63 75 -

Fig. 4B Above the range of this table and within the range of the viscosimeter, multiply the particular value by the following approximate factors to convert to SSU: Viscosimeter Mac Michael Demmier #1 Demmier #10 Stormer

TECH-D

Factor 1.92 (approx.) 14.6 146. 13. (approx.)

18 20 28 34 41 48 63 77 -

13 18 24 29 33 43 50 65 75 86 96 -

1.3 2.3 3.2 4.1 4.9 5.7 6.5 10.0 13.5 16.9 20.4 27.4 34.5 41 48 55 62 69 103 137 172 206 275 344 413 481 550 620 690 1030 1370

1.0 1.4 1.7 2.0 2.7 3.5 4.1 4.8 5.5 6.2 6.9 10.3 13.7 17.2 20.6 27.5 34.4 41.3 48 55 62 69 103 137

2.6 3.6 4.6 5.5 6.4 7.3 11.3 15.2 19 23 31 39 46 54 62 70 77 116 154 193 232 308 385 462 540 618 695 770 1160 1540

7 8 9 9.5 10.8 11.9 12.4 16.8 22 27.6 33.7 45 55.8 65.5 77 89 102 113 172 234

TECH-D-3 Determination of Pump Performance When Handling Viscous Liquids Reprinted from HYDRAULIC INSTITUTE STANDARDS. Twelfth Edition. Copyright 1969 by Hydraulic Institute.

The performance of centrifugal pumps is affected when handling viscous liquids. A marked increase in brake horsepower, a reduction in head, and some reduction in capacity occur with moderate and high viscosities.

Limitations on Use of Viscous Liquid Performance Correction Chart

Fig. 5 provides a means of determining the performance of a conventional centrifugal pump handling a viscous liquid when its performance on water is known. It can also be used as an aid in selecting a pump for a given application. The values shown in Fig. 5 are averaged from tests of conventional single stage pumps of 2-inch to 8inch size, handling petroleum oils. The correction curves are, therefore, not exact for any particular pump.

Use only for pumps of conventional hydraulic design, in the normal operating range, with open or closed impellers. Do not use for mixed flow or axial flow pumps or for pumps of special hydraulic design for either viscous or non-uniform liquids.

When accurate information is essential, performance tests should be conducted with the particular viscous liquid to be handled.

Use only on Newtonian (uniform) liquids. Gels, slurries, paper stock and other non-uniform liquids may produce widely varying results, depending on the particular characteristics of the liquids.

Reference is made to Fig. 5. This chart is to be used only within the scales shown. Do not extrapolate.

Use only where adequate NPSH is available in order to avoid the effect of cavitation.

Fig. 5 Performance Correction Chart

TECH-D

Symbols and Definitions Used in Determination of Pump Performance When Handling Viscous Liquids.

The viscous efficiency and the viscous brake horsepower may then be calculated.

These symbols and definitions are:

This procedure is approximate as the scales for capacity and head on the lower half of Fig. 5 are based on the water performance. However, the procedure has sufficient accuracy for most pump selection purposes. Where the corrections are appreciable, it is desirable to check the selection by the method described below.

Qvis

= Viscous Capacity, gpm The capacity when pumping a viscous liquid.

Hvis

= Viscous Head, feet The head when pumping a viscous liquid.

Evis

= Viscous Efficiency, per cent The efficiency when pumping a viscous liquid.

EXAMPLE. Select a pump to deliver 750 gpm at 100 feet total head of a liquid having a viscosity of 1000 SSU and a specific gravity of 0.90 at the pumping temperature.

bhpvis

= Viscous Brake Horsepower The horsepower required by the pump for the viscous conditions.

Enter the chart (Fig. 5) with 750 gpm, go up to 100 feet head, over to 1000 SSU, and then up to the correction factors:

QW

= Water Capacity, gpm The capacity when pumping water.

HW

= Water Head, feet The head when pumping water.

sp gr

= Specific Gravity

CQ

= Capacity correction factor

CH

= Head correction factor

CE

= Efficiency correction factor

1.0 Qw

= Water Capacity at which maximum efficiency is obtained.

The following equations are used for determining the viscous performance when the water performance of the pump is known:

Qvis

= CQ X Qw

Hvis

= CH x Hw

Evis

= CE x Ew

= = = =

Hw

=

0.95 0.92 (for 1.0 Qnw) 0.635 750 = 790 gpm 0.95 100 = 108.8  109 feet head 0.92

Select a pump for a water capacity of 790 gpm at 109 feet head. The selection should be at or close to the maximum efficiency point for water performance. If the pump selected has an efficiency on water of 81 per cent at 790 gpm, then the efficiency for the viscous liquid will be as follows: Evis = 0.635 x 81% = 51.5 per cent The brake horsepower for pumping the viscous liquid will be: bhpvis = 750 x 100 x 0.90 = 33.1 hp 3960 x 0.515 For performance curves of the pump selected, correct the water performance as shown below. Instructions for Determining Pump Performance on a Viscous Liquid When Performance on Water is Known

bhpvis = Qvis x Hvis x sp gr 3960 x Evis CQ, CH and CE are determined from Fig. 5 which is based on the water performance. The following equations are used for approximating the water performance when the desired viscous capacity and head are given and the values of CQ and CH must be estimated from Fig. 5 using Qvis and Hvis, as: QW(approx.) = Qvis CQ HW(approx.) = Hvis CH Instructions for Preliminary Selection of a Pump for a Given Head-Capacity-Viscosity Condition Given the desired capacity and head of the viscous liquid to be pumped and the viscosity and specific gravity at the pumping temperature, Fig. 5 can be used to find approximate equivalent capacity and head when pumping water. Enter the chart (Fig. 5) at the bottom with the desired viscous capacity, (Qvis) and proceed upward to the desired viscous head (Hvis) in feet of liquid. For multistage pumps, use head per stage. Proceed horizontally (either left or right) to the fluid viscosity, and then go upward to the correction curves. Divide the viscous capacity (Qvis) by the capacity correction factor (CQ) to get the approximate equivalent water capacity (Qw approximately). Divide the viscous head (Hvis) by the head correction factor (CH) from the curve marked "1.0 x Qw" to get the approximate equivalent water head (Hw approximately). Using this new equivalent water headcapacity point, select a pump in the usual manner.

TECH-D

CQ CH CE Qw

Given the complete performance characteristics of a pump handling water, determine the performance when pumping a liquid of a specified viscosity. From the efficiency curve, locate the water capacity (1.0 x Qw) at which maximum efficiency is obtained. From this capacity, determine the capacities (0.6 x Qw). (0.8 x Qw) and (1.2 x Qw). Enter the chart at the bottom with the capacity at best efficiency (1.0 x Qw), go upward to the head developed (in one stage) (Hw) at this capacity, then horizontally (either left or right) to the desired viscosity, and then proceed upward to the various correction curves. Read the values of (CE) and (CQ), and of (CH) for all four capacities. Multiply each head by its corresponding head correction factor to obtain the corrected heads. Multiply each efficiency value by (CE) to obtain the corrected efficiency values which apply at the corresponding corrected capacities. Plot corrected head and corrected efficiency against corrected capacity. Draw smooth curves through these points. The head at shut-off can be taken as approximately the same as that for water. Calculate the viscous brake horsepower (bhpvis) from the formula given above. Plot these points and draw a smooth curve through them which should be similar to and approximately parallel to the brake horsepower (bhp) curve for water.

EXAMPLE. Given the performance of a pump (Fig. 6) obtained by test on water, plot the performance of this pump when handling oil with a specific gravity of 0.90 and a viscosity of 1000 SSU at pumping temperature. On the performance curve (Fig. 6) locate the best efficiency point which determines (Qw). In this sample this is 750 gpm. Tabulate capacity, head and efficiency for (0.6 x 750), (0.8 x 750) and (1.2 x 750).

Using 750 gpm, 100 feet head and 1000 SSU, enter the chart and determine the correction factors. These are tabulated in Table of Sample Calculations. Multiply each value of head, capacity and efficiency by its correction factor to get the corrected values. Using the corrected values and the specific gravity, calculate brake horsepower. These calculations are shown on Table 6. Calculated points are plotted in Fig. 6 and corrected performance is represented by dashed curves.

TECH-D-4 Viscosity Corrections for Capacities of 100 GPM or Less

Fig. 5A

TECH-D

Fig. 6 Sample Performance Chart

TABLE 6

TECH-D

TECH-D-5A Viscosity of Common Liquids Reprinted from PIPE FRICTION MANUAL, Third Edition. Copyright 1961 by Hydraulic Institute.

VISCOSITY Liquid

*Sp Gr at 60 F

SSU

Centistokes

At F

2,950 813

.27-.32 648 176

70 68.6 100 70 70 70 70 68 70 100 65 100 100 68

Freon Glycerine (100%)

1.37 to 1.49 @ 70 F 1.26 @ 68F

Glycol: Propylene Triethylene Diethylene Ethylene Hydrochloric Acid(31.5) Mercury

1.038 @ 68F 1.125@ 68 F 1.12 1.125 1.05 @ 68 F 13.6

240.6 185.7 149.7 88.4

.95 to 1.08 40 Baume 42 Baume 1.83

65 365 637.6 75.7

52 40 32 17.8 1.9 .118 .11 11.7 79 138 14.6

220 65 150 95 287 160 190 to 220 112 to 128 140 90 230 130 110 78 163 to 184 97 to 112

47.5 11.6 32.1 19.4 62.1 34.3 41 to 47.5 23.4 to 27.1 29.8 18.2 49.7 27.5 23.0 15.2 35 to 39.6 19.9 to 23.4

130 212 100 130 100 130 100 130 100 130 100 130 100 130 100 130

165 to 240 90 to 120 240 to 400 120 to 185 400 to 580 185 to 255 580 to 950 255 to 80 950 to 1,600 80 to 105 1,600 to 2,300 105 to 125 2,300 to 3,100 125 to 150 5,000 to 10,000 10,000 to 40,000

35.4 to 51.9 18.2 to 25.3 51.9 to 86.6 25.3 to 39.9 86.6 to 125.5 39.9 to 55.1 125.5 to 205.6 55.1 to 15.6 205.6 to 352 15.6 to 21.6 352 to 507 21.6 to 26.2 507 to 682 26.2 to 31.8 1,100 t o2,200 2,200 TO 8,800

100 130 100 130 100 130 100 130 210 100 210 100 210 100 210 0 0

100,000 max 800 To 1,500 300 to 500 950 to 2,300 120 to 200 Over 2,300 Over 200

22,000 max 173.2 to 324.7 64.5 to 108.2 205.6 to 507 25.1 to 42.9 Over 507 Over 42.9

0 100 130 130 210 130 210

40 to 783 34.2 to 210 74 to 1,215 46 to 320 40 to 4,480 34 to 700 46 to 216 38 to 86

4.28 to 169.5 2.45 to 4.53 14.1 to 263 6.16 to 69.3 4.28 to 1,063 2.4 to 151.5 6.16 to 46.7 3.64 to 17.2

60 100 60 100 60 100 60 100

165 to 240 90 to 120

35.4 to 51.9 18.2 to 25.3

100 130

Phenol (Carbonic Acid) Silicate of soda Sulfric Acid (100%) FISH AND ANIMAL OILS: Bone Oil

.918

Cod Oil

.928

Lard

.96

Lard Oil

.912 to .925

Menhaddden Oil

.933

Neatsfoot Oil

.917

Sperm Oil

.883

Whale Oil

.925

Mineral Oils: Automobile Crankcase Oils (Average Midcontinent Parrafin Base) SAE 10

**.880 to .935

SAE 20

**.880 to .935

SAE 30

**.880 to .935

SAE 40

**.880 to .935

SAE 50

**.880 to .935

SAE 60

**.880 to .935

SAE 70

**.880 to .935

SAE 10W SAE 20W Automobile Transmission Lubricants: SAE 80 SAE 90

**.880 to .935 **.880 to .935 **.880 to .935 **.880 to .935

SAE 140

**.880 to .935

SAE 250

**.880 to .935

Crude Oils: Texas, Oklahoma

.81 to .916

Wyoming, Montana

.86 to .88

California

.78 to .92

Pennsylvania

.8 to .85

Diesel Engine Lubricating Oils (Based on Average Midcontinent Parafin Base): Federal Specefication No. 9110 * Unless otherwise noted.

**.880 to .935

** Depends on origin or percent and type of solvent.

TECH-D

VISCOSITY Liquid

SSU

Centistokes

At F

300 to 410 140 to 180 470 to 590 200 to 255 800 to 1,100 320 to 430 490 to 600 92 to 105

64.5 to 88.8 29.8 to 38.8 101.8 to 127.8 43.2 to 55.1 173.2 to 238.1 69.3 to 93.1 106.1 to 129.9 18.54 to 21.6

100 130 100 130 100 130 130 210

32.6 to 45.5 39 45.5 to 65 39 to 48 140 max 70 max 400 max 165 max

2 to 6 1 to 3.97 6 to 11.75 3.97 to 6.78 29.8 max 13.1 max. 86.6 max 35.2 max

100 130 100 130 100 130 122 160

34 to 40 32 to 35 36 to 50 33 to 40 35 to 45 32.8 to 39 50 to 125 42 to 72 125 to 400 72 to 310 450 to 3,000 175 to 780 110 to 225 63 to 115 1,500 max 480 max

73 50

2.39 to 4.28 2.69 3.0 to 7.4 2.11 to 4.28 2.69 to .584 2.06 to 3.97 7.4 to 26.4 4.91 to 13.73 26.4 to 86.6 13.63 to 67.1 97.4 to 660 37.5 to 172 23 to 48.6 11.08 to 23.9 324.7 max 104 max .46 to .88 .40 to .71 .41 13.9 7.4

70 100 70 100 100 130 100 130 100 122 130 122 160 122 160 122 160 60 100 68 70 100

65 max 35 32.6

11.75 max 2.69 2

100 68 100

112 to 160 70 to 90 160 to 235 90 to 120 235 to 385 120 to 185 385 to 550 185 to 255

23.4 to 34.3 13.1 to 18.2 34.3 to 50.8 18.2 to 25.3 50.8 to 83.4 25.3 to 39.9 83.4 to 119 39.9 to 55.1

100 130 100 130 100 130 100 130

140 to 190 86 to 110 190 to 220 110 to 125 100 77

29.8 to 41 17.22 to 23 41 to 47.5 23 to 26.4 20.6 14.8

100 130 100 130 130 160

.91 Average

400 to 440 185 to 205

86.6 to 95.2 39.9 to 44.3

100 130

.96 @ 68 F

1,200 to 1,500 450 to 600 1,425 580 140 to 148 76 to 80 135 54 176 100

259.8 to 324.7 97.4 to 129.9 308.5 125.5 29.8 to 31.6 14.69 to 15.7 28.7 8.59 37.9 20.6

100 130 69 100 100 130 130 212 100 130

*Sp Gr at 60 F

Diesel Engine Lubricating Oils (Based on Average Midcontinent Parafin Base): Federal Specification No.9170

**.880 to .935

Federal Specification No. 9250

**.880 to .935

Federal Specification No. 9370

**.880 to .935

Federal Specification No. 9500

**.880 to .935

Diesel Fuel Oils: No. 2 D

**.82 to .95

No.3 D

**.82 to .95

No.4 D

**.82 to .95

No.5 D

**.82 to .95

Fuel Oils: No. 1

**.82 to .95

No. 2

**.82 to .95

No.3

**.82 to .95

No.5A

**.82 to .95

No.5B

**.82 to .95

No.6

**.82 to .95

Fuel Oil – Navy Specification

**.989 max

Fuel Oil – Navy II

1.0 max

Gasoline

.68 to .74

Gasoline (Natural) Gas Oil

76.5 degrees API 28 degrees Api

Insulating Oil: Transformer, switches and Circuit breakers Kerosene

.78 to .82

Machine Lubricating Oil (Average Pennsylvania Parafin Base): Federal Specification No.8

**.880 to .935

Federal Specification No. 10

**.880 to .935

Federal Specification No. 20

**.880 to .935

Federal Specification No. 30

**.880 to .935

Mineral Lard Cutting Oil: Federal Specefication Grade 1 Federal Specification Grade 2 Petrolatum Turbine Lubricating Oil: Federal Specification (Penn Base) VEGETABLE OILS: Castor Oil

.825

China Wood Oil

.943

Cocoanut Oil

.925

Corn Oil

.924

Cotton Seed Oil

.88 to .925

* Unless otherwise noted.

TECH-D

** Depends on origin or percent and type of solvent.

VISCOSITY Liquid VEGETABLE OILS: Linseed Oil, Raw

*Sp Gr at 60 F .925 to .939

Olive oil

.912 to .918

Palm oil

.924

Peanut Oil

.920

Rape Seed Qil

.919

Rosin Oil

.980

Rosin (Wood)

1.09 Avg

Sesame Oil

.923

Soja Bean Oil

.927 to.98

Turpentine

.86 to .87

SUGARS, SYRUPS, MOLASSES, ETC. Corn Syrups Glucose Honey (Raw) Molasses “A” (First) Molasses”B” (Second) Molasses “C” (Blackstrap or final) Sucrose Solutions(Sugar Syrups) 60 Brix

1.4 TO 1.47 1.35 to 1.44 140.6 to 146 1.43 to 1.48 1.46 to 1.49 1.29

62 Brix

1.30

64 Brix

1.31

66 Brix

1.326

68 Brix

1.338

70 Brix

1.35

72 Brix

1.36

74 Brix

1.376

76 Brix

1.39

TARS: Tar Coke Oven Tar Gas House

1.12+ 1.16 to 1.30

Road Tar: Grade RT-2

1.07+

Grade RT-4

1.08+

Grade RT-6

109+

Grade RT-8

1.13+

Grade RT-10

1.14+

Grade RT-12

1.15+

Pine Tar

1.06

MISCELLANEOUS Corn Starch Solutions: 22 Baume 24 Baume

1.18 1.20

SSU

Centistokes

At F

143 93 200 115 221 125 195 112 250 145 1,500 600 500 to 20,000 1,000 to 50,000 184 110 165 96 33 32.6

30.5 18.94 43.2 24.1 47.8 26.4 42 23.4 54.1 31 324.7 129.9 108.2 to 4,400 216.4 to 11,000 39.6 23 35.4 19.64 2.11 2.0

100 130 100 130 100 130 100 130 100 130 100 130 200 190 100 130 100 130 60 100

5,000 to 500,000 1,500 to 60,000 35,000 to 100,000 4,000 to 11,000 340 1,300 to 23,00 700 to 8,000 6,500 to 60,000 3,000 to 15,000 17,00 to 250,000 6,000 to 75,00

1,1000 324.7 7,700 880

to 110.000 to 13,200 to 22,000 to 2420 73.6 281.1 to 5,070 151.5 to 1,760 1,410 to 13,200 660 to 3,300 2,630 to 5,500 1,320 to 16,500

100 130 100 150 100 100 130 100 130 100 130

230 92 310 111 440 148 650 195 1,000 275 1,650 400 2,700 640 5,500 1,100 10,000 2,000

49.7 18.7 67.1 23.2 95.2 31.6 140.7 42.0 216.4 59.5 364 86.6 595 138.6 1,210 238 2,200 440

70 100 70 100 70 100 70 100 70 100 70 100 70 100 70 100 70 100

3,000 to 8,000 650 to 1,400 15,000 to 300,000 2,000 to 20,000

600 to 1,760 140.7 to 308 3,300 to 66,000 440 to 4,400

71 100 70 100

200 to 300 55 to 60 400 to 700 65 to 75 1,000 to 2,000 85 to 125 3,000 to 8,000 150 to 225 20,000 to 60,000 250 to 400 114,000 to 456,000 500 to 800 2,500 500

43.2 to 64.9 8.77 to 10.22 86.6 to 154 11.63 to 14.28 216.4 to 440 16.83 to 26.2 660 to 1,760 31.8 to 48.3 4,400 to 13,200 53.7 to 86.6 25,000 to 75,000 108.2 to 173.2 559 108.2

122 212 122 212 122 212 122 212 122 212 122 212 100 132

150 130 600 440

32.1 27.5 129.8 95.2

70 100 70 100

* Unless otherwise noted.

TECH-D

VISCOSITY Liquid

*Sp Gr at 60 F

MISCELLANEOUS Corn Starch Solutions: 25 Baume

1.2

Ink- Printers

1.00 to 1.38

Tallow Milk Varnish – Spar

.918 Avg. 1.02 to 1.05 .9

Water- Fresh

1.0

SSU

Centistokes

At F

1400 800 2,500 to 10,000 1,100 to 3,000 56

303 17.2 550 to 2,200 238.1 to 660 9.07 1.13 313 143 1.13 .55

70 100 100 130 212 68 68 100 60 130

1425 650

* Unless otherwise noted.

TECH-D-5B Physical Properties of Common Liquids Liquid Acetic Acid Glacial 8.8% (1N) .88% (.1N) .09 (.01N) Acetone Alum, 0.6% (0.1N) Ammonia 100% 26% 1.7% (1N) .17% (0.1N) .02% (.01N) Asphalt Unblended RS1 RC2 RC5 Emulsion Benzene Benzoic Acid 0.1% (.01N) Black liquor, 50%

Sp. Gr. 60° F (16°C)

Melting Point °F (°C)

1.05

63 (17)

- 137 (-94)

133 (56)

.77

-108 (-78)

-27 (-33)

80°F

120°F

160°F

4°C

Centipoise 27°C 49°C

71°C

1.6

1.2

.8

.6

.4

.3

.3

.2

.14

.1

.08

.06

1.8

1.2

86

34

17

.6

.5

.3

3.2 .91 11.6 11.1 10.6 1.1-1.5 1.0 1.0 1.0 1.0 .84

155-1,000 500,000 1,000-7,000 42 (6)

160 2,400-5,000 45,000

(12,000 at 250°F) 85 8,000

176 (80)

.8 3.1

1.3

Borax

1.7 (75)

1% (0.1N) Boric Acid

5,000

(80-150 at 250°F) (6,300 at 250°F)

(15-37 at 121°C) 1,400 at 121°C)

167 9.2 338 (171) 5.2

.59

.18 9.4

1.23

-21 (-29)

4.5

2.1

.9

.5

14.5

7.3

3.9

2.1

12.4

1.07

109 (43)

360 (182)

Reprinted with permission of the Durametallic Corporation.

TECH-D

40°F

2.4 2.9 3.4

.79

1.5

Calcium Hydroxide Sat. (Slaked Lime) Carbolic Acid (Phenol)

VISCOSITY SSU

244 (118)

1.01

70%

0.2% (0.1N) Butane Calcium Carbonate Sat. Calcium Chloride 25%

Boiling pH Point At 77° F °F (°C) (25°C)

60

Liquid Carbonic Acid Sat. Carbon Tetrachloride Citric Acid .6% (1n) Corn Oil Corn Starch, 22° Baume 25° Baume Corn Syrup

Sp. Gr. 60° F (16°C)

1.58

Ethane

.37

Ethyl Alcohol

.79

Ethyl Alcohol 95% Ethylene Glycol

.81 1.1

3.6% (1N) .36% (0.1N) .04% (.01N) Jet Fuel Lactic Acid

-95 (-71)

80°F

120°F

160°F

170 (77)

4°C

1.3

Centipoise 27°C 49°C

71°C

.9

.7

.6

1.6

(.05 at 16° C.) 1.0

.7

.5

2.0 44

1.3 19

.8 9

.6 4

2.4

.49 1.5

-

.8

3.3 4.6 15

2.1 2.6 7 1,00

1.4 1.6 4 155

0.9 1.2 3 40

.7

.6

.4

.3

6,260 11

490 5.4

130 2.8

56 1.5

2.5

1.8

1.4

1.1

.8 1.0

.5 .7 1.1

.4 .5

.4

135

1.18 1.21 1.4

1.1

31.5%

40°F

2.2

Dowtherm C

Glycerine (Glycerol) 50% Hydrochloric Acid, 38%

VISCOSITY SSU

.92

.8 .9 .79 .86 .99

Diesel 2D 3D 5D Gasoline Glucose

Boiling pH Point At 77° F °F (°C) 25°C) 3.8

Cotton Seed Oil Crude Oil Pennsylvania Wyoming 48° API 32.6° API Dowtherm A

Ethyl Acetate Formic Acid, 1.22 100% .5% (.1N) Fuel Oil No. 1 (Kerosene) No. 2 No. 3 No. 6 (Bunker C)

Melting Point °F (°C)

150 1,400

130 800 5,000500,000 176

200 1,100

86 320

.9

.9 122 1.0

2.8 20.0 54 (12) 70 (21)

500 (260) 600 (316)

- 173 (- 144)

173 (78)

9 (- 13)

387 (198)

47 (8) -

213 (100) -

185

53

39

2.3

.81 .86 .89 .96

40 43 84

.82-.95 .82-.95 .82-.95 .6-.7 1.4

100 200 15,000 30

36 36 52 4,50020,000 53 80 2,000

31 33 41 680- 1,900 40 50 400

30 32 37 180500 35 40 160

35,000100,000

1.26 1.13

64 (18)

1.20

- 13 (-25) -115 (-46)

1.15

86

554 (290)

25,000

3,100

700

230

0.1 1.1 2.0 .7-.8

Methyl Alcohol 80% Milk, 3.5% Molasses A

.80 .82 1.03 1.40

Molasses C

1.49

35 63 (17) – - 144 (-98)

252 (122) – 149 (65)

2.4 6.3-6.6 10,000 300,000

2,60060,000 25,000250,000

TECH-D

Liquid Nitric Acid, 95% 60%

Sp. Gr. 60° F (16°C)

Melting Point °F (°C)

1.50

-44 (-44) -9 (-23)

1.37

Oil, 5W 10W 20W 30W 50W 70W Oleic Acid

.9 .9 .9 .9 .9 0.89

Olive Oil Palmetic Acid

.9 0.85

Parafin

.9

Peanut Oil Propane Propylene Gylcol Potassium Hydroxide 5.7% (1N) 0.57% (0.1N) 0.06% (0.01N) Rosin Sodium Bicarbonate 0.4% (0.1N) Sodium Chloride, 25% Sodium Hydroxide, 50% 30% 4% (1N) 0.4% (0.1N) .04% (.01N) Stearic Acid

.9 .51 1.0

Sucrose, 60% 40 % Sugar Syrup 60 Brix 70 Brix 76 Brix Sulfur Molten Sulfric Acid 110% (Fuming, Oleum) 100%

40°F

80°F

120°F

160°F

187 (86)

13 (-11)

547 (286)

146 (63) 100 (38)

520 (271) 660 (349)

550 1,500 2,900 5,000 23,000 120,000

160 265 500 870 3,600 10,000

74 120 170 260 720 1,800

51 64 80 110 225 500

1,500

320

150

80

1,200

300

150

80

4°C

Centipoise 27°C 49°C

71°C

1.4

1.0

.8

.6

3.4

2.2

1.5

1.0

110 170 580 1,200 -

30 50 98 200 400 4,000 26

12 22 33 60 100 -

7 11 14 25 45 -

.12 241

1.09

500-20,000 8.4

1.19 1.53 1.33

950

240 58

84

500

150

68

46

3.3

2.1

1.3

.9

250

77 10

26 4.5

10 2.5

156

41

14

7

120

5

2.5

1.6

(11 at 123°C)

(9 at 159°C)

82

41

22

12

46

23

12

6

8.9

5.8

3.9

2.7

2.5

1.4

0.8

0.6

.8

.6

.4

.4

.7

.6

.5

.4

1.9

1.4

.9

.7

1.6

.9

.6

.4

14.0 13.0 12.0 .85 1.29 1.18 1.29 1.35 1.39 2.06

1.83 1.84

60%

1.50

20%

1.14

4.9% (1N) .49% (.1N) .05 (.01N) Toluene

.86

Trichloroethylene

1.62

Turpentine

.86

Vinegar Water

1.0

TECH-D

VISCOSITY SSU

14.0 13.0 12.0

98%

Wines

Boiling pH Point At 77° F °F (°C) 25°C)

157 (69) 10 (- 12) 25 (-4)

721 (383) 218 (103) 214 (101)

230 1,650 10,000 239 (115)

832 (445)

92

342 (33)

50 (10) 37 (3) -83 (-64) 8 (-13)

280

100

92 400 2,000

55

(22 at (16,000 at 160°C) 184°C)

(172) 75

554 (290) 282 (139) 218 (103)

118

68

45

37

0.3 1.2 2.1 -139 (-95) -99 (-72) 140 (-10)

231 (111) 189 (87) 320 (160)

32 (0) 1.03

212 (100)

34 2.4-3.4 6.5-8.0

32 2.8-3.8

33

32

32

TECH-D-6 Friction Loss for Viscous Liquids. Loss in Feet of Liquid per 100 Feet of New Schedule 40 Steel Pipe GPM

3

Nom. Pipe Size 1⁄2 3⁄4

1 3⁄4

5 10 15 20 30 40 60 80 100 125

150

175 200 250 300 400 600 800 1000

1 11⁄4 1 11⁄4 11⁄2 1 11⁄4 11⁄2 1 11⁄2 2 11⁄2 2 21⁄2 11⁄2 2 21⁄2 2 21⁄2 3 21⁄2 3 4 21⁄2 3 4 3 4 6 3 4 6 3 4 6 3 4 6 4 6 8 6 8 10 6 8 10 6 8 10 6 8 10 8 10 12

Kinematic Viscosity – Seconds Saybolt Universal Water

100

200

10.0 2.50 0.77 6.32 1.93 0.51 6.86 1.77 0.83 14.6 3.72 1.73 25.1 2.94 0.87 6.26 1.82 0.75 10.8 3.10 1.28 6.59 2.72 0.92 4.66 1.57 0.41 7.11 2.39 0.62 3.62 0.94 0.12 5.14 1.32 0.18 6.9 1.76 0.23 8.90 2.27 0.30 3.46 0.45 0.12 1.09 0.28 0.09 1.09 0.28 0.09 2.34 0.60 0.19 4.03 1.02 0.33 1.56 0.50 0.21

25.7 8.5 3.2 14.1 5.3 1.8 11.2 3.6 1.9 26 6.4 2.8 46 5.3 1.5 11.6 3.2 1.4 19.6 5.8 2.5 11.6 5.1 1.8 8.3 3.0 0.83 12.2 4.4 1.2 6.5 1.8 0.25 9.2 2.4 0.34 11.7 3.2 0.44 15.0 4.2 0.58 6.0 0.83 0.21 8.5 1.2 0.30 1.9 0.53 0.18 4.2 1.1 0.37 6.5 1.8 0.60 2.5 0.88 0.39

54.4 17.5 6.6 29.3 11.0 3.7 22.4 7.5 4.2 34 11.3 6.2 46 8.1 3.0 12.2 4.4 2.2 20.8 5.8 3.0 13.4 5.5 1.8 9.7 3.2 0.83 14.1 5.1 1.3 7.8 2.1 0.28 10.4 2.9 0.39 13.8 4.0 0.52 17.8 5.1 0.69 7.4 0.99 0.28 9.9 1.4 0.39 2.3 0.62 0.21 5.1 1.3 0.42 8.1 2.2 0.69 3.2 1.0 0.46

300

400

500

600

83 108 135 162 26.7 35.5 44 53 10.2 13.4 16.6 20.0 44 59 74 88 16.8 22.4 28 33 5.5 7.6 9.5 11.1 33.5 45 56 66 11.2 14.9 19.1 22.4 6.0 8.1 10.2 12.3 50 67 85 104 16.9 22.4 29 34 9.2 12.4 15.3 18.4 67 90 111 133 12.2 16.2 20.3 25 4.4 6.0 7.4 9.0 18.2 24.3 30 37 6.7 9.0 11.1 13.2 3.2 4.4 5.5 6.5 24 32 40 50 9.0 11.8 14.8 17.7 4.4 5.8 7.4 8.8 13.4 17.8 22.2 27 6.5 8.8 10.9 13.1 2.8 3.7 4.6 5.6 9.7 11.8 14.6 17.6 3.7 4.8 6.2 7.3 1.2 1.7 2.1 2.5 14.8 14.8 18.5 22 5.1 6.2 7.6 9.1 1.5 2.1 2.5 3.1 8.1 8.1 9.7 11.5 2.1 2.6 3.2 3.9 0.39 0.52 0.63 0.78 11.5 11.5 11.5 13.7 2.9 3.1 3.9 4.6 0.46 0.62 0.77 0.9 15.8 15.8 15.8 15.9 4.0 4.0 4.6 5.4 0.54 0.7 0.9 1.1 20.3 20.3 20.3 20.3 5.1 5.1 5.1 6.2 0.69 0.81 1.0 1.2 8.0 8.0 8.0 8.0 1.0 1.0 1.2 1.5 0.28 0.35 0.42 0.51 11.6 11.6 11.6 11.6 1.5 1.5 1.5 1.8 0.39 0.42 0.51 0.61 2.5 2.8 2.8 2.8 0.67 0.67 0.67 0.81 0.23 0.23 0.28 0.32 5.3 5.5 6.0 6.2 1.4 1.5 1.5 1.5 0.46 0.51 0.51 0.51 8.5 9.2 9.7 11.1 2.3 2.5 2.8 2.8 0.78 0.88 0.92 0.92 3.5 3.7 4.2 4.4 1.2 1.3 1.4 1.4 0.51 0.55 0.58 0.58

800

1000

1500

2000

218 71 26.6 117 44 14.8 89 30 16.5 137 45 25 180 33 11.9 50 17.8 8.8 65 24 11.8 36 17.8 7.3 24 9.7 3.3 29 12.1 4.1 15.3 5.2 1.0 18.4 6.2 1.2 21.4 7.4 1.4 25 8.3 1.6 10.2 2.1 0.67 12.4 2.5 0.82 3.2 1.1 0.43 6.2 1.7 0.65 11.1 2.8 0.92 4.4 1.4 0.58

273 88 34 147 56 18.5 112 37 20.3 172 57 30 220 40 14.8 61 22.2 10.9 81 30 14.6 45 22.0 9.2 29 12.2 4.2 36 15.2 5.1 19.4 6.4 1.3 23 7.8 1.5 27 9.2 1.8 31 10.4 2.0 12.9 2.5 0.83 15.5 3.0 1.0 3.9 1.3 0.53 6.2 2.0 0.81 11.1 2.8 1.1 4.4 1.4 0.67

411 131 50 219 83 28 165 55 31 84 46 61 22.4 91 33 16.6 121 44 22.2 67 34 13.8 44 18.3 6.2 55 23 7.8 29 9.8 1.9 35 11.5 2.3 40 13.7 2.6 46 15.5 3.0 19.4 3.7 1.2 23 4.6 1.5 6.0 2.0 0.81 9.0 3.0 1.2 12.0 3.9 1.6 5.1 2.0 1.0

545 176 67 293 111 37 223 74 41 112 61 81 30 122 45 22.0 162 59 29 89 44 18.5 58 24 8.3 73 31 10.4 39 12.7 2.6 46 15.4 3.0 53 18.2 3.5 61 20.6 3.9 26 5.1 1.7 31 6.0 2.0 8.1 2.8 1.1 12.0 3.9 1.6 16.0 5.3 2.1 6.7 2.8 1.3

3000

5000 10,000

820 1350 265 440 880 100 167 440 740 1470 167 56 94 187 112 190 62 102 207 167 92 152 122 203 45 74 147 182 67 178 222 33 55 110 243 400 810 89 148 44 73 145 134 220 66 109 220 27 46 92 87 145 37 61 122 12.5 20.6 41 109 183 46 77 150 15.5 26 51 58 97 193 19.3 32 65 3.9 6.4 13.0 69 115 230 23 39 78 4.6 7.6 15.2 80 133 28 46 92 5.3 8.8 17.8 91 152 31 51 103 6.2 9.9 20.1 39 64 130 7.6 12.5 2.5 4.2 8.3 46 77 155 9.1 15.0 30 3.0 5.1 9.9 12.1 20.1 4.1 6.7 13.5 1.6 2.8 5.3 18.5 6.2 9.9 20 2.4 4.2 8.1 8.2 13.4 3.2 5.3 10.9 10.2 16.6 4.0 6.7 13.4 2.0 3.5 6.7

Extracted from PIPE FRICTION MANUAL. Third Edition. Copyright 1961 by Hydraulic Institute.

TECH-D

TECH-D-7 Pumping Liquids with Entrained Gas Pump applications in many industrial processes involve handling liquid and gas mixtures. The entrained gas may be an essential part of an industrial process, or it may be unwanted. The Pulp and Paper industry, for example, injects from between 4% and 10% air into a dilute pulp slurry as part of the ink removal process in a flote cell used in paper recycling. Many chemical and petrochemical processes also involve pumping a two phase flow. Unwanted entrained gas can result from excess agitation or vortexing due to inadequate submergence on the suction of a pump. The proper selection of a centrifugal pump for liquid and gas (two phase) mixtures is highly dependent on the amount of gas and the characteristics of the liquid. The presence of entrained gases will reduce the output of centrifugal pumps and can potentially cause loss of prime. Conventional pump designs can be used for low percentages by volume (up to 4%), while special modified impellers can be used effectively for up to 10% gas by volume. Performance corrections are required in all cases with gas content above approximately 2%. Gas concentrations above 10% can also be handled, but only with special design pumps (pumps with inducers, vortex pumps, or pumps with gas extraction). Virtually any type of centrifugal pump can handle some amount of entrained gas. The problem to be addressed is the tendency for the

gas to accumulate in the pump suction inhibiting flow and head generation. If gas continues to accumulate, the pump may lose prime. Fig. 1 shows how the performance of a standard end suction pump is affected by various amounts of air. With a minor performance correction, this type of pump is reasonably efficient in handling up to approximately 4% entrained gas. As the percentage of gas exceeds 4% by volume, the performance of a conventional pump begins to degrade drastically (Fig. 1) until the pump becomes unstable, eventually losing prime. It has been found beneficial to increase the impeller running clearance (0.090 to 0.180 in.) allowing for greater leakage. This is effective in preventing loss of prime with gas concentrations up to 10%. Fig. 2 shows a standard end suction open impeller pump with clearances opened for gas handling. Numerous tests have been conducted in an effort to quantify the performance corrections for various gas concentrations for both standard pumps and pumps with open clearances. The performance corrections are affected by many variables, including pump specific speed, operating speed, impeller design and number of vanes, operating point on the curve, and suction pressure. Performance correction charts are not presented here due to the numerous variables, but Goulds Applications Department can make recommendations and selections for most specific applications.

Standard Clearance (Typically .015") Increased Running Clearance (Typically .090" - .180")

Fig. 1 Head and Power vs Capacity Zero to Ten Percent Air by Volume for Normal Running Clearance

TECH-D

Fig. 2 Open Impeller End Suction Pump with Normal Running Clearance and Increased Running Clearance.

TECH-D-8A Solids and Slurries - Definition of Terms APPARENT VISCOSITY

PERCENT SOLIDS BY WEIGHT

The viscosity of a non-Newtonian slurry at a particular rate of shear, expressed in terms applicable to Newtonian fluids.

The weight of dry solids in a given volume of slurry, divided by the total weight of that volume of slurry, multiplied by 100.

CRITICAL CARRYING VELOCITY

SALTATON

The mean velocity of the specific slurry in a particular conduit, above which the solids phase remains in suspension, and below which solid-liquid separation occurs.

A condition which exists in a moving stream of slurry when solids settle in the bottom of the stream in random agglomerations which build up and wash away with irregular frequency.

EFFECTIVE PARTICLE DIAMETER

SETTLING SLURRY

The single or average particle size used to represent the behavior of a mixture of various sizes of particles in a slurry. This designation is used to calculate system requirements and pump performance.

A slurry in which the solids will move to the bottom of the containing vessel or conduit at a discernible rate, but which will remain in suspension if the slurry Is agitated constantly.

FRICTION CHARACTERISTIC

SETTLING VELOCITY

A term used to describe the resistance to flow which is exhibited by solid-liquid mixtures at various rates of flow.

The rate at which the solids in a slurry will move to the bottom of a container of liquid that is not in motion. (Not to be confused with the velocity of a slurry that is less than the critical carrying velocity as defined above.)

HETEROGENEOUS MIXTURE A mixture of solids and a liquid in which the solids are net uniformly distributed. HOMOGENEOUS FLOW (FULLY SUSPENDED SOLIDS) A type of slurry flow in which the solids are thoroughly mixed in the flowing stream and a negligible amount of the solids are sliding along the conduit wall.

SQUARE ROOT LAW A rule used to calculate the approximate increase in critical carrying velocity for a given slurry when pipe size is increased. It states: 1/ 2 VL = Vs = DL Ds

()

Where: VL DL Vs Ds

HOMOGENEOUS MIXTURE A mixture of solids and a liquid in which the solids are uniformly distributed. NON-HOMOGENEOUS FLOW (PARTIALLY SUSPENDED SOLIDS)

NOTE:

= Critical carrying velocity in larger pipe = Diameter of larger pipe = Critical carrying velocity in smaller pipe = Diameter of smaller pipe

This rule should not be used when pipe size is decreased.

A type of slurry flow in which the solids are stratified, with a portion of the solids sliding along the conduit wall. Sometimes called "heterogeneous flow” or “flow with partially suspended solids.”

VISCOSITY TYPES

NON-SETTLING SLURRY

YIELD VALUE (STRESS)

A slurry In which the solids will not settle to the bottom of the containing vessel or conduit, but will remain in suspension, without agitation, for long periods of time.

The stress at which many non-Newtonian slurries will start to deform and below which there will be no relative motion between adjacent particles in the slurry.

(For definitions of the various types of viscosities applicable to slurries, see Rheological Definitions.)

PERCENT SOLIDS BY VOLUME The actual volume of the solid material in a given volume of slurry, divided by the given volume of slurry, multiplied by 100.

TECH-D-8B Solids and Slurries - Slurry Pump Applications Determining the when to use a slurry style centrifugal pump can be a challenging decision. Often the cost of a slurry pump is many times that of a standard water pump and this can make the decision to use a slurry pump very difficult. One problem in selecting a pump type is determining whether or not the fluid to be pumped is actually a slurry. We can define a slurry as any fluid which contains more solids than that of potable water. Now, this does not mean that a slurry pump must be used for every application with a trace amount of solids, but at least a slurry pump should be considered. Slurry pumping in its simplest form can be divided into three categories: the light, medium and heavy slurry. In general, light slurries are slurries that are not intended to carry solids. The presence of the solids occurs more by accident than design. On the other hand, heavy slurries are slurries that are designed to transport material from one location to another. Very often the carrying fluid in a heavy slurry is just a necessary evil in helping to transport the desired

material. The medium slurry is one that falls somewhere in between. Generally, the Percent solids in a medium slurry will range from 5% to 20% by weight. After a determination has been made as to whether or not you are dealing with a heavy, medium, or light slurry, it is then time to match a pump to the application. Below is a general listing of the different characteristics of a light, medium, and heavy slurry. Light Slurry Characteristics: • • • • •

Presence of solids is primarily by accident Solids Size < 200 microns Non-settling slurry The slurry specific gravity < 1.05 Less than 5% solids by weight

TECH-D

Medium Slurry Characteristics: • Solids size 200 microns to 1/4 inch (6.4mm) • Settling or non-settling slurry • The slurry specific gravity < 1.15 • 5% to 20% solids by weight Heavy Slurry Characteristics: • Slurry’s main purpose is to transport material • Solids > 1/4 inch (6.4mm) • Settling or non-settling slurry • The slurry specific gravity > 1.15 • Greater than 20% solids by weight The previous listing is just a quick guideline to help classify various pump applications. Other considerations that need to be addressed when selecting a pump model are: • Abrasive hardness • Particle shape

It should be noted, however, that a hard metal pump can also be used for services that are outlined for the rubber-lined pump. After a decision has been made whether to use a hard metal pump or a rubber-lined pump, it is then time to select a particular pump model. A pump model should be selected by reviewing the application and determining which model pump will work best in the service. Light Slurries AF HS HSU HSUL VHS JC JCU VJC 5100 5800 Linapump

• Particle size • Particle velocity and direction • Particle density • Particle sharpness The designers of slurry pumps have taken all of the above factors into consideration and have designed pumps to give the end user maximum expected life. Unfortunately, there are some compromises that are made in order to provide an acceptable pump life. The following short table shows the design feature, benefit, and compromise of the slurry pump. SLURRY PUMP DESIGN Design Feature Thick Wear Sections Larger Impellers Specialty Materials Semi Volute or Concentric Casing Extra Rigid Power Ends

Benefit Longer component life Slower pump speeds longer component life Longer component life

Compromise Heavier, more expensive parts Heavier, more expensive parts Expensive parts

Improved pump life Loss in efficiency Improved bearing lives

More expensive shafts and bearings

Although selecting the proper slurry pump for a particular application can be quite complex, the selection task can be broken down into a simplified three-step process: 1. Determine which group of possible pump selections best matches your specific application. 2. Plot the system curve depicting the required pump head at various capacities. 3. Match the correct pump performance curve with the system curve. Slurry pumps can be broken down into two main categories. The rubber-lined pump and the hard metal pump. However, because of the elastomer lining, the rubber-lined pumps have a somewhat limited application range. Below is a general guideline which helps distinguish when to apply the rubber-lined pumps. Rubber Lined Solids < 1/2 inch (13mm) Temperature < 300° F (150°C) Low Head service < 150 feet (46m) Rounded particles Complete pH range

TECH-D

Hard Metal Pump Solids > 1/4 inch (6.4mm) Temperature < 250° F (120°C) Heads above 150 feet (46m) Sharp/Jagged particles pH range from 4 to 12 Hydrocarbon based slurry

Slurry Pump Break Down Medium Slurries Heavy Slurries AF HS HSU HSUL VHS JC JCU VJC 5100 5000 5150 RX SP SRL CW

5000 5150 RX CKX 5500 SRL-C SRL-XT

NOTES: The Model HS pump is a unique pump in that it is a recessed impeller or “vortex" pump. This style pump is well suited to handle light pulpy or fibrous slurries. The recessed impeller used in the HS family of pumps will pass large stringy fibers and should be considered when pump plugging is a concern. The Model AF is a specialized pump with an axial flow design. This design of pump is built specifically for high flow, low head applications. In general, slurry pumps have been designed to handle fluids with abrasive solids, and will give extended lives over standard water or process pumps. Although many features have been designed into the slurry pump, there are still two factors which directly relate to the pump's life that can be determined. The first choice to make is determining the metallurgy of the pump. In most cases, a hard metal slurry pump will be constructed of some hardened metal with a Brinell hardness of at least 500. Goulds standard slurry pump material is a 28% chrome iron with a minimum hardness of 600 Brinell. This material is used for most abrasive services and can also be used in some corrosive fluids as well. If a more corrosive resistant material is required, then the pump may be constructed out of a duplex Stainless steel Such as CD4MCu. Please check with your nearest Goulds sales office if you are unsure what material will be best suited for a particular application. PUMP RUNNING SPEED The other factor that can be controlled by the sales or end user engineer is the pump running speed. The running speed of a slurry pump is one of the most important factors which determines the life of the pump. Through testing, it has been proven that a slurry pump's wear rate is proportional to the speed of the pump raised to the 2 1⁄2 power. EXAMPLE: If Pump (A) is running at 1000 RPM and Pump (B) is running at 800 RPM, then the life factor for Pump (B) as compared to Pump (A) is (1000/800)2.5 or Pump (B) will last 1.75 times as long as Pump (A). With the above ratio in mind, it can be shown that by cutting a slurry pump speed in half, you get approximately 6 times the wear life. For this reason, most slurry pumps are V-belt driven with a full diameter impeller. This allows the pump to run at the slowest possible running speed and, therefore, providing the maximum pump life.

WHY USE A V-BELT DRIVE? In most ANSI pump applications it is a reasonable practice to control condition point by trimming the impeller and direct connecting the motor. However, this is not always sound practice in slurry applications. The abrasive solids present, wear life is enhanced by applying the pump at the slowest speed possible. Another situation where V-belts are beneficial is in the application of axial flow pumps. Axial flow pumps cannot be trimmed to reduce the condition point because they depend on close clearances between the vane tips and the casing for their function. The generally low RPM range for axial flow application also makes it beneficial to use a speed reduction from the point of view of motor cost.

The types of V-belt drives available for use in pump applications are termed fixed speed, or fixed pitch, and variable speed. The fixed pitch drive consists of two sheaves; each machined to a specific diameter, and a number of belts between them to transmit the torque. The speed ratio is roughly equal to the diameter ratio of the sheaves. The variable speed drive is similar to the fixed speed except that the motor sheave can be adjusted to a range of effective or pitch diameters to achieve a band of speed ratios. This pitch adjustment is made by changing the width of the Vgrooves on the sheave. Variable speed drives are useful in applications where an exact flow rate is required or when the true condition point is not well defined at the time that the pump is picked. V-belt drives can be applied up to about 2000 horsepower, but, pump applications are usually at or below 350 HP.

TECH-D-8C Solids and Slurries - Useful Formulas a. The formula for specific gravity of a solids-liquids mixture or slurry, Sm is: Ss x S1 Sm = Ss + Cw (S1 – Ss )

c. Slurry flow requirements can be determined from the expression: Qm = 4 x dry solids (tons per hour) Cw = Sm

where,

Qm = slurry flow (U.S. gallons per minute) 1 ton = 2000 lbs.

Sm = S1 = Ss = Cw = Cv =

specific gravity of mixture or slurry specific gravity of liquid phase specific gravity of solids phase concentration of solids by weight concentration of solids by volume

EXAMPLE: if the liquid has a specific gravity of 1.2 and the concentration of solids by weight is 35% with the solids having a specific gravity of 2.2, then: 2.2 x 1.2 Sm = = 1.43 2.2 + .35 (1.2 – 2.2) b. Basic relationships among concentration and specific gravities of solid liquid mixtures are shown below: In Terms of Cv

Ss, Sm, S1 Sm-S1

Cv

Cw Sm Ss

Ss-S1 Cw

(Sm – S1) Ss x (Ss – S1) Sm

Cw

Cv Ss Sm

Where pumps are to be applied to mixtures which are both corrosive and abrasive, the predominant factor causing wear should be identified and the materials of construction selected accordingly. This often results in a compromise and in many cases can only be decided as a result of test or operational experience. For any slurry pump application a complete description of the mixture components is required in order to select the correct type of pump and materials of construction. weight of dry solids CW = weight of dry solids + weight of liquid phase Cv =

volume of dry solids volume of dry solids + volume of liquid phase

See nomograph for the relationship of concentration to specific gravity of dry solids in water shown in Fig. B.

where,

EXAMPLE: 2,400 tons of dry solids is processed in 24 hours in water with a specific gravity of 1.0 and the concentration of solids by weight is 30% with the solids having a specific gravity of 2.7 then: 2.7 x 1.0 Sm = = .123 2.7 + .3 (1-2.7)

Qm = 4 x 100 = 1,084 U.S. GPM .3 x 1.23 d. Abrasive wear: Wear on metal pumps increases rapidly when the particle hardness exceeds that of the metal surfaces being abraded. If an elastomer lined pump cannot be selected, always select metals with a higher relative hardness to that of the particle hardness. There is little to be gained by increasing the hardness of the metal unless it can be made to exceed that of the particles. The effective abrasion resistance of any metal will depend on its position on the mohs or knoop hardness scale. The relationships of various common ore minerals and metals is shown in Fig. A. Wear increases rapidly when the particle size increases. The life of the pump parts can be extended by choosing the correct materials of construction. Sharp angular particles cause about twice the wear of rounded particles. Austenetic maganese steel is used when pumping large dense solids where the impact is high. Hard irons are used to resist erosion and, to a lesser extent, impact wear. Castable ceramic materials have excellent resistance to cutting erosion but impeller tip velocities are usually restricted to 100 ft./sec. Elastomer lined pumps offer the best wear life for slurries with solids under 1⁄4" for the SRL/SRL-C and under 1⁄2" for the SRL-XT. Several Elastomers are available for different applications. Hypalon is acceptable in the range of 1-14 pH. There is a single stage head limitation of about 150' due to tip speed limitations of elastomer impellers. See the Classification of Pumps according to Solids Size chart (Fig. C) and Elastomer Quick Selection Guide (Section TECH-B-2) for more information.

TECH-D

Solids and Slurries Approximate Comparison of Hardness Values of Common Ores and Minerals

Fig. A

TECH-D

Solids and Slurries Nomograph of the Relationship of Concentration to Specific Gravity in Aqueous Slurries

Ss Solids Specific Gravity

Cv % Solids by Volume

Cw % Solids by Weight

Sm Slurry Specific Gravity Fig. B

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Solids and Slurries Classification of Pumps According to Solid Size

Grade Mesh Very large boulders Large boulders Medium boulders Small boulders Large cobbles

Austenetic Manganese Steel Dredge Pump

Small cobbles Very coarse gravel Coarse Gravel Hard Iron Medium Gravel SRL-XT Fine Gravel

Very Fine Gravel

Severe Duty Slurry Pump

SRL-C

Very Coarse Sand

Sand Pump

Coarse Sand SRL/ SRL-C Medium Sand

Slurry Pump

Fine Sand

Silt Slimes

Note: This tabulation is for general guidance only since the selection of pump type and materials of construction also depends on the total head to be generated and the abrasivity of the slurry i.e. concentration, solids specific gravity, etc.

Mud Clay

* Theoretical values Micron = .001 mm

Fig. C

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Sand & Gravel Pump

Ceramic Lined

2.5 3 3.5 4 5 6 7 8 9 10 12 14 16 20 24 28 32 35 42 48 60 65 80 100 115 150 170 200 250 270 325 400 *500 *625 *1250 *2500 *12500

Pulverized

Tyler Standard Sieve Series Aperture Inch mm 160 4060 80 2030 40 1016 20 508 10 254 3 76.2 2 50.8 1.5 38.1 1.050 26.67 .883 22.43 .742 18.85 .624 15.85 .524 13.33 .441 11.20 .371 9.423 .321 7.925 .263 6.680 .221 5.613 .185 4.699 .156 3.962 .131 3.327 .110 2.794 .093 2.362 .078 1.981 .065 1.651 .055 1.397 0.46 1.168 0.39 .991 0.328 .833 0.276 .701 .0232 .589 .0195 .495 .0164 0.417 .0138 .351 .116 .295 .0097 .248 .0082 .204 .0069 .175 .0058 .147 .0049 .124 .0041 .104 .0035 .089 .0029 .074 .0024 .061 .0021 .053 .0017 .043 .0015 .038 .025 .020 .10 .005 .001 .0005 .0024

Solids and Slurries Standard Screen Sizes Comparison Chart U.S. Bureau of Standard Screens

Tyler Screens

Aperture

British Standard Screens

Aperture

Mesh

Inches

mm

Mesh 21⁄2 3

Inches .321 .263 .221 .185 .156 .131 .110 .093 .078 .065

mm 7.925 6.680 5.613 4.699 3.962 3.327 2.794 2.362 1.981 1.651

3 31⁄2 4 5 6 7 8 10 12

.265 .223 .187 .157 .132 .111 .0937 .0787 .0661

6.73 5.66 4.76 4.00 3.36 2.83 2.38 2.00 1.68

14

.0555

1.41

.055

1.397

16

.0469

1.19

14

.046

1.168

18 20

.0394 .0331

1.00 .84

20

.039 .0328

.991 .883

25

.0280

.71

.0276

.701

30 35 40 45

.0232 .0197 .0165 .0138

.59 .50 .42 .35

28

.0232 .0195 .0164 .0138

.589 .495 .417 .351

50 60 70 80 100

.0117 .0098 .0083 .0070 .0059

.297 .250 .210 .177 .149

48

.0116 .0097 .0082 .0069 .0058

.295 .246 .208 .175 .147

120 140 170 200 230 270 325

.0049 .0041 .0035 .0029 .0024 .0021 .0017

.125 .105 .088 .074 .062 .053 .044

.0049 .0041 .0035 .0029 .0024 .0021 .0017 .0015

.124 .104 .088 .074 .061 .053 .043 .037

4 6 8 10

35

65 100

150 200 270 400

I.M.M. Screens

Apeture Mesh Double Tyler Series

Aperture

Mesh

Inches

mm

Mesh

Inches

mm

5 6 7 8 10

.1320 .1107 .0949 .0810 0.660

3.34 2.81 2.41 2.05 1.67

5

.100

2.54

8

.062

1.574

12

.0553

1.40 10

.050

1.270

14

.0474

1.20 12

.0416

1.056

16

16 18

.0395 .0336

1.00 .85 16

.0312

.792

24

22

.0275

.70 20

.025

.635

25 30 36 44

.0236 .0197 .0166 .0139

.60 .50 .421 .353

25 30 35 40

.020 .0166 .0142 .0125

.508 .421 .361 .317

52 60 72 85 100

.0166 .0099 .0083 .0070 .0060

.295 .252 .211 .177 .152

120 150 170 200 240 300

.0049 .0041 .0035 .0030 .0026 .0021

.125 .105 .088 .076 .065 .053

50 60 70 80 90 100 120 150 170 200

.01 .0083 .0071 .0062 .0055 .0050 .0042 .0033 .0029 .0025

.254 .211 .180 .157 .139 .127 .107 .084 .074 .063

31⁄2 5 7 9

12

32 42

60 80

115 170 250 325

Fig. D

TECH-D

Solids and Slurries Specific Gravities of Rocks, Minerals and Ores Material

Specific Gravity Mohs Hardness

Aluminum Amber Ambylgonite Andesine Aragonite, CaCO3 Argentite Asbestos Asphaltum Asphalt Rock Barite Basalt Bauxite Bentonite Bertrandite Beryl Biotite Bone Borax Bornite Braggite Braunite Brick Calcite Carnotite Cassiterite

2.55- 2.75 1.06-1.11 3-3.1 2.66- 2.94 2.94-2.95 7.2-7.4 2.1-2.4 1.1-1.5 2.41 4.5 2.4-3.1 2.55-2.73 1.6 2.6 2.66- 2.83 2.7-3.1 1.7-2 1.71-1.73 5.06-5.08 10 4.72- 4.83 1.4-2.2 2.72-2.94 2.47 6.99-7.12

Carbon, Amorphous Graphitic

1.88-2.25

Celluloid Cerussite Chalcocite Chalcopyrite Chalk Charcoal, Pine Charcoal, Oak Chromite Chrysoberyl Cinnabar Clay Coal, Anthracite Coal, Bituminous Coal, Lignite Cobaltite Coke Colemanite Columbite Copper Cork Covellite Cuprite Diabase Diatomaceous Earth Diorite Dolomite Enargite Epidote Feldspar Fluorite Fly Ash Galena Glass Goethite Gold Granite Graphite Gravel, Dry Gravel, Wet Gypsum Halite Hausmannite Helvite Hematite

1.4 6.5- 6.57 5.5-5.8 4.1-4.3 1.9-2.8 0.28-0.44 0.47-0.57 4.5 3.65-3.85 8.09 1.8-2.6 1.4-1.8 1.2-1.5 1.1-1.4 6.2 1-1.7 1.73 5.15-5.25 8.95 0.22-0.26 4.6-4.76 6 2.94 0.4-0.72 2.86 2.8-2.86 4.4-4.5 3.25-3.5 2.55-2.75 3.18 2.07 7.3-7.6 2.4-2.8 3.3-4.3 19.3 2.6-2.9 2.2-2.72 1.55 2 2.3-2.37 2.2 4.83-4.85 3.2-3.44 4.9-5.3

Material

1-2

Hessite Ice Ilmenite Iron, Slag Lepidolite Lime, slaked Limestone Limonite Linnaeite Magnetite Manganite Marble Marl Millerite Monazite Molybdenite Muscovite Niccolite Orpiment Pentlandite Petalite Phosophite Phosphorus, white Polybasite

5.5-6 6-6.5 3.5-4 2-2.5 2 3-3.5 8-9 6 7.5-8 2.5-3 2-2.5 3 6-6.5 3 1-2 6-7

Potash Powellite Proustiie Psilomelane Pumice Pyragyrite Pyrites Pyrolusite Quartz Quartzite Realgar Rhodochrosite Rhodonite Rutile Sand (see Quartz) Sandstone Scheelite Schist Serpentine Shale Siderite Silica, fused trans. Slag, Furnace Slate Smaltite Soapstone, talc Sodium Nitrate Sperrylite Spodumene Sphalerite Stannite Starch Stibnite Sugar Sulfur Sylvanite Taconite Tallow, beef Tantalite Tetrahedrite Titanite Trap Rock Uraninite Witherite Wolframite Zinc Blende Zincite

3-3.5 2.5-3 3.5-4

5.5 8.5 2-2.5 2 2 5.5 4.5 6 2.5-3 1.5-2 3.5-4

3.5-4 3 6 4 2.5-2.75 7 5-5.5 2.5-3 1-2 4-5 2 2.5 5.5 6 5-6 Fig. E

TECH-D

Specific Gravity Mohs Hardness 8.24- 8.45 0.917 4.68-4.76 2.5-3 2.8-2.9 1.3- 1.4 2.4-2.7 3.6-4 4.89 4.9-5.2 4.3-4.4 2.5-2.78 2.23 5.3-5.7 5.1 4.62-4.73 2.77- 2.88 7.784 3.5 4.8 2.412-2.422 3.21 1.83 6-6.2 Porphyry 2.2 4.21-4.25 5.57 4.71 0.37-0.9 5.85 4.95-5.1 4.8 2.5-2.8 2.68 3.56 3.7 3.57-3.76 4.2-5.5 1.7-3.2 2-3.2 6.08-6.12 2.6-3 2.5 1.6-2.9 3.9-4 2.21 2-3.9 2.8-2.9 6.48 2.6-2.8 2.2 10.58 3.03-3.22 3.9-4.1 4.3-4.5 1.53 4.61-4.65 1.59 1.93-2.07 8.161 3.18 0.94 7.9-8 4.6-5.1 3.5 2.79 8-11 4.29-4.3 7.12-7.51 4.02 5.64-5.68

2-3 5-6 2.5-4 2-5 5.5-6.5 4 4 3-3.5 5 1-1.5 2.5-3 5-5.5 1.5-2 2.5-3 6.5 2.3 2.6-2.9 3.5-4 2-2.5 5-6 2.5 3.5-4.5 6-6.5 7-8 7 1.5-2 3.5-4 5.5-6.5 6-6.5 7 7 4.5-5 2.5-3.5 4-4.5

2 6-7 6.5-7 3.5-4 4 2 1.5-2.5 1.5-2 6.5 3-4.5 5-6 3.5 4-4.5 4 4

Solids and Slurries Hardness Conversion Table for Carbon and Alloy Steels Brinell Hardness Number (Carbide Ball) 722 688 654 615 577 543 512 481 455 443 432 421 409 400 309 381 371 362 353 344 336 327 319 311 301 294 286 279 271 264 258 253 247 243 240 234 222 210 200 195 185 176 169

Rockwell Hardness Numbers

C Scale

A Scale

15N Scale Superficial

66 64 62 60 58 56 54 52 50 48 47 46 45 44 43 42 41 40 39 38 37 36 35 34 33 32 31 30 29 28 27 26 25 24 23

84.5 83.4 82.3 81.2 80.1 79 78 76.8 75.9 74.7 74.1 73.6 73.1 72.5 72 71.5 70.9 70.4 69.9 69.4 68.9 68.4 67.9 67.4 66.8 66.3 65.8 65.3 64.6 64.3 63.8 63.3 62.8 62.4 62

92.5 91.8 91.1 90.2 89.3 88.3 87.4 86.4 85.5 84.5 83.9 83.5 83 82.5 82 81.5 80.9 80.4 79.9 79.4 78.8 78.3 77.7 77.2 76.6 76.1 75.6 75 74.5 73.9 73.3 72.8 72.2 71.6 71

Tensile Strength

B Scale

100 99 97 95 93 92 90 88 86

30T Scale Superficial

ksl

MPa

83.1 82.5 81.1 79.8 78.4 77.8 76.4 75.1 73.8

313 292 273 255 238 229 221 215 208 201 194 188 182 177 171 166 161 156 152 149 146 141 138 135 131 128 125 123 119 117 116 114 104 100 94 92 89 86 83

2160 2010 1880 1760 1640 1580 1520 1480 1430 1390 1340 1300 1250 1220 1180 1140 1110 1080 1050 1030 1010 970 950 930 900 880 860 850 820 810 800 785 715 690 650 635 615 590 570

Fig. F

TECH-D

Solids and Slurries Slurry Pump Materials MTL CODE

COMMON NAME

ASTM NUMBER

BRINELL HARDNESS

CHARACTERISTICS AND TYPICAL APPLICATIONS

pH RANGE

1002

Cast Iron

196-228

Offers moderate resistance to abrasion and corrosion. It is suitable for light slurry applications, particularly those for intermittent service.

6-9

1228

HC600

550-650

Hardened HC600 (High Chromium Iron)

5-12

1245

316SS

159-190

Used for high corrosive, mildly abrasive applications.

3-11

1247

CD4MCu

A48 CI. 35B A532 CI. III Type A A743 GR. CF-8M A734 Gr. CD4MCu

224-325

This is a high strength corrosion resistant alloy for mildly abrasive applications.

MTL CODE

Cr

Ni

1002 1228 1245 1247

23.0-28.0 18.0-21.0 25.0-27.0

15 Max 9.0-12.0 5.0-6.0

PRINCIPAL ALLOYING ELEMENTS (%, Bal Fe) C Mn Si 3.25-3.35 2.3-3.0 0.08 Max 0.4 Max

0.45-0.70 0.5-1.5 1.5 Max -

1.70-1.90 1.0 Max 2.0 Max -

Mo

Others

1.5 Max 2.0-3.0 2.0

Cu 3.0

Fig. G

Slurry Pump Application Guidelines Slurry

Solid Size Larger 1/4"

Solids Size 1/2" Smaller

Impeller Tip Speed > 5500 FPM (High Head)

Solids Size Larger than 1/2"

5500

Solids Round in Shape

Solids Sharp & Angular

SRL-XT 5500

5500 SP

Solids Sharp & Angular

SP, JC, SRL-XT (with metal Inpeller)

Slurry Contains Stringy Material

Solids Round in Shape

SRL, SRL-C (with froth factor sizing)

> 60 Mesh or > 25% Wt.

> 60 Mesh and > 25% Wt.

SRL-C

SRL SRL-C

SRL-C/SRL-XT With Metal or Urethane Impallers or Series Operation

TECH-D

Slurry Contains Entrained Air (Froth)

Solids Size 1/4" Smaller

SRL, SRL-X (Shearpeller)

TECH-D-9A Vapor Pressure – Various Liquids

TECH-D

VACUUM–INCHES OF MERCURY

ABSOLUTE PRESSURE–LBS. PER SQ. IN.

GAUGE PRESSURE–LBS. PER SQ. IN.

TECH-D-9A Vapor Pressure – Various Liquids

TECH-D

Section TECH-E Paper Stock TECH-E-1 Paper Stock, Discussion Centrifugal pumps are used with complete success in handling paper stock and other fibrous suspensions. However, the nature of a stock suspension requires certain special considerations. All of the factors affecting pump operation discussed below must be carefully considered for a good installation.

AIR IN STOCK

SUCTION PIPING

EXCESSIVE DISCHARGE THROTTLING

The stock must be delivered freely to the impeller for the pump to operate. The suction pipe should be as short and direct as possible. The suction pipe and entrance from the stock chest should never be smaller than the pump suction connection, and should be level with no air pockets. Always keep the direction of flow in a straight line.

While it is realized that excess capacity is normally required over the paper machine output in tons per day, "over-selection" of pumps on the basis of capacity and head usually results in the necessity of throttling the pump at the valve in the discharge line. Since the valve is normally located adjacent to the pump, the restriction of the valve and the high velocity within the valve will result in some dehydration and cause vibration due to slugs of stock. Vibration at the valve due to throttling is transmitted to the pump and may reduce the normal life of the pump-rotating element.

Inadequate suction design with undersize pipe and excessive fittings can prevent the pump from delivering rated capacity, or from operating at all on high consistency stocks. SUCTION HEAD Stock pumps will not operate when a vacuum is required to maintain flow into the pump. Thus, there must be a static suction head sufficient to overcome suction line friction losses. PERCENT CONSISTENCY The consistency of a pulp and water suspension is the percent by weight of pulp in the mixture. Oven Dry (O.D.) consistency is the amount of pulp left in a sample after drying in an oven at 212°F. Air Dry (A.D.) consistency is an arbitrary convention used by papermakers, and is the amount of pulp left in a sample after drying in atmosphere. Air Dry stock contains 10% more moisture than Bone Dry stock, i.e. 6% O.D. is 6.67% A.D. Traditional paper stock pumps will handle stock up to approximately 6% O.D. consistency. The absolute maximum limit is a function of many factors including stock fiber length, pulping process, degree of refining, available suction head, etc. In certain situations, consistencies as high as 8% O.D. can be successfully handled with a standard paper stock pump. Recent testing on various types of stock has indicated that pump performance is the same as on water for stock consistencies up to 6% O.D. In other words, water curves can be used to select stock pumps, as the capacity, head and efficiency are the same as for water. Medium consistency paper stock is a term generally used to describe stock between 7% and 15% O.D. consistency. Pumping of medium consistency paper stock with a centrifugal pump is possible, but requires a special design due to the fiber network strength and the inherently high air content.

Entrained air is detrimental to good operation of any centrifugal pump, and can result in reduced capacity, increased erosion and shaft breakage. Obviously every effort must be made to prevent the over-entrainment of air throughout the process.

Centrifugal pumps operating at greatly reduced capacity have more severe loading internally due to hydraulic radial thrust. Hence pumps selected too greatly oversize in both capacity and head have the combination of the vibration due to throttling plus the greater internal radial load acting to reduce the life of the rotating element. As a general rule, stock pumps should not be operated for extended periods at less than one quarter of their capacity at maximum efficiency. When excessive throttling is required, one of the two methods below should be employed. 1. Review capacity requirements and check the static and friction head required for the capacity desired. Reduce the impeller diameter to meet the maximum operating conditions. This will also result in considerable power saving. 2. Install a by-pass line upstream from the discharge valve back to the suction chest below the minimum chest level, if possible, and at a point opposite the chest opening to the pump suction. This by-pass line should include a valve for flow regulation. This method is suggested where mill production includes variation in weight of sheet. FILLERS AND ADDITIVES The presence of fillers and chemical additives such as clay, size and caustics can materially increase the ability of paper stock to remain in suspension. However, overdosing with additives such as alum may cause gas formation on the stock fibers resulting in interruption of pumping.

TECH-E

TECH-E-2 Conversion Chart of Mill Output in Tons per 24 Hours To U.S. Gallons per Minute of Paper Stock of Various Densities

EXAMPLE: Find the capacity in gallons per minute of a pump handling 4% stock for a mill producing 200 tons per 24 hours.

Enter chart at 200 tons per day, read horizontally to 4% stock, then downward to find pump capacity of 840 GPM.

TECH-E-2.1 Definitions / Conversion Factors A.D. = Air Dry stock (Contains 10% Water)

T/ D or TPD or S. T/ D = Short Tons Per Day

O.D. = Oven Dry stock (All Water Removed) Also Called Bone Dry (B.D.)

M. T/ D = Metric Tons per Day

A.D. = 1.11 x O.D.

One Short Ton = 2000 lbs.

One Metric Ton = 2205 lbs.

O.D. = 0.90 x A.D.

A.D.S. T/ D = Air Dry Short Tons/Day

A.D. = 1.11 O.D.T/ D

A.D.M. T/ D = Alr Dry Metric Tons/Day

O.D. = 0.90 x A.D. T/ D

S. T/ D = 1.1025 x M. T/ D

A.D. Consistency = 1.11 x O.D. Consistency O.D. Consistency = 0.90 x A.D. Consistency

Production in A. D. S. T/ D x 15 = Flow in GPM % O.D. Cons. Production in A. D. S. T/ D x 16.67 = Flow in GPM % A.D. Cons.

TECH-E

TECH-E-3 Friction Loss of Pulp Suspensions in Pipe I. INTRODUCTION In any stock piping system, the pump provides flow and develops hydraulic pressure (head) to overcome the differential in head between two points. This total head differential consists of pressure head, static head, velocity head and total friction head produced by friction between the pulp suspension and the pipe, bends, and fittings. The total friction head is the most difficult to determine because of the complex, nonlinear nature of the friction loss curve. This curve can be affected by many factors. The following analytical method for determining pipe friction loss is based on the recently published TAPPI Technical Information Sheet

Figure 1 – Friction loss curves for chemical pulp (C2 > C1).

(TIS) 408-4 (Reference 1), and is applicable to stock consistencies (oven-dried) from 2 to 6 percent. Normally, stock consistencies of less than 2% (oven-dried) are considered to have the same friction loss characteristic as water. The friction loss of pulp suspensions in pipe, as presented here, is intended to supersede the various methods previously issued. II. BACKGROUND Figure 1 and Figure 2 show typical friction loss curves for two different consistencies (C2>C1) of chemical pulp and mechanical pulp, respectively.

Figure 2 – Friction loss curves for mechanical pulp (C2 > C1).

The friction loss curve for chemical pulp can be conveniently divided into three regions, as illustrated by the shaded areas of Figure 3.

Figure 3 – Friction loss curves for chemical pulp, shaded to show individual regions.

Figure 4 – Friction loss curves for mechanical pulp, shaded to show individual regions.

TECH-E

These regions may be described as follows:

IV. PIPE FRICTION ESTIMATION PROCEDURE

Region 1 (Curve AB) is a linear region where friction loss for a given pulp is a function of consistency, velocity, and pipe diameter. The velocity at the upper limit of this linear region (Point B) is designated Vmax.

The bulk velocity (V) will depend on the daily mass flow rate and the pipe diameter (D) selected. The final value of V can be optimized to give the lowest capital investment and operating cost with due consideration of future demands or possible system expansion.

Region 2 (Curve BCD) shows an initial decrease in friction loss (to Point C) after which the friction loss again increases. The intersection of the pulp friction loss curve and the water friction loss curve (Point D) is termed the onset of drag reduction. The velocity at this point is designated Vw.

The bulk velocity will fall into one of the regions previously discussed. Once it has been determined in which region the design velocity will occur, the appropriate correlations for determining pipe friction loss value(s) may be selected. The following describes the procedure to be used for estimating pipe friction loss in each of the regions.

Region 3 (Curve DE) shows the friction loss curve for pulp fiber suspensions below the water curve. This is due to a phenomenon called drag reduction. Reference 2 describes the mechanisms which occur in this region.

Region 1 The upper limit of Region 1 in Figure 3 (Point B) is designated Vmax. The value of Vmax is determined using Equation 1 and data given in Table I or IA.

Regions 2 and 3 are separated by the friction loss curve for water, which is a straight line with a slope approximately equal to 2.

Vmax = K' C  (ft/s),

■

1 ■

where K' = numerical coefficient (constant for a given pulp is attained from Table I or IA.

The friction loss curve for mechanical pulp, as illustrated in Figure 4, is divided into only two regions:

C = consistency (oven-dried, expressed as a percentage, not decimally), and

Regions 1 and 3. For this pulp type, the friction loss curve crosses the water curve at VW and there is no true Vmax.

 = exponent (constant for a given pulp), obtained from Table I or IA.

III. DESIGN PARAMETERS To determine the pipe friction loss component for a specified design basis (usually daily mass flow rate), the following parameters must be defined: a)

Pulp Type - Chemical or mechanical pulp, long or short fibered, never dried or dried and reslurried, etc. This is required to choose the proper coefficients which define the pulp friction curve.

b)

Consistency, C (oven-dried) - Often a design constraint in an existing system. NOTE: If air-dried consistency is known, multiply by 0.9 to convert to oven-dried consistency.

c)

d)

Internal pipe diameter, D - Lowering D reduces initial capital investment, but increases pump operating costs. Once the pipe diameter is selected. it fixes the velocity for a prespecified mass flow rate. Bulk velocity, V - Usually based on a prespecified daily mass flow rate. Note that both V and D are interdependent for a constant mass flow rate.

e)

Stock temperature, T - Required to adjust for the effect of changes in viscosity of water (the suspending medium) on pipe friction loss.

f)

Freeness - Used to indicate the degree of refining or to define the pulp for comparison purposes.

g)

Pipe material - Important to specify design correlations and compare design values.

It the proposed design velocity (V) is less than Vmax, the value of flow resistance (▲ ▲H/ L) may be calculated using Equation 2 and data given in Table II or IIA, and the appendices.

■

2 ■

H/L = F K V C Dy (ft/100 ft), where F = factor to correct for temperature, pipe roughness, pulp type, freeness, or safety factor (refer to Appendix D), K = numerical coefficient (constant for a given pulp), obtained from Table II or IIA, V = bulk velocity (ft/s), C = consistency (oven-dried, expressed as a percentage, not decimally), D = pipe inside diameter (in), and

, , y =exponents (constant for a given pulp), obtained from Table II or IIA. For mechanical pumps, there is no true Vmax. The upper limit of the correlation equation (Equation 2 ) is also given by Equation 1 . In this case, the upper velocity is actually Vw.

■

■

Region 2 The lower limit of Region 2 in Figure 3 (Point B) is Vmax and the upper limit (Point D) is Vw. The velocity of the stock at the onset of drag reduction is determined using Equation 3

■

VW = 4.00 C1.40 (ft/s),

3 ■

where C = consistency (oven-dried, expressed as a percentage, not decimally). If V is between Vmax and Vw, Equation 2 may be used to determine ▲H/ L at the maximum point (Vmax). Because the system must cope with the worst flow condition, ▲H/ L at the maximum point (Vmax) can be used for all design velocities between Vmax and Vw.

TECH-E

Region 3 A conservative estimate of friction loss is obtained by using the water curve. (▲ ▲H/ L)w can be obtained from a Friction Factor vs. Reynolds Number plot (Reference 3, for example), or approximated from the following equation (based on the Blasius equation). (▲ ▲H/ L)w = 0.58. V1.75 D-1.25 (ft/100 ft),

4 ■

where V = bulk velocity (ft/s), and

Previously published methods for calculating pipe friction loss of pulp suspensions gave a very conservative estimate of head loss. The method just described gives a more accurate estimate of head loss due to friction, and has been used successfully in systems in North America and world-wide. Please refer to Appendix A for equivalent equations for use with metric (SI) units. Tables I and IA are located in Appendix B; Tables II and IIA are located in Appendix C. Pertinent equations, in addition to those herein presented, are located in Appendix D. Example problems are located in Appendix E.

1M ■

Vmax = K' C  (m/s) where K = numerical coefficient (constant for a given pulp), obtained from Table I or IA,

 = exponent (constant for a given pulp), obtained from Table I or IA. 2M ■

▲H/ L = F K V  C D y (m/100m),

where F = factor to correct for temperature, pipe roughness, pulp type, freeness, or safety factor (refer to Appendix D), K = numerical coefficient (constant for a given pulp), obtained from Table II or IIA, V = bulk velocity (m/s), C = consistency (oven-dried, expressed as a percentage, not decimally),

V. HEAD LOSSES IN BENDS AND FITTINGS The friction head loss of pulp suspensions in bends and fittings may be determined from the basic equation for head loss, Equation 5 .

■

When metric (SI) units are utilized, the following replace the corresponding equations in the main text.

C = consistency (oven-dried, expressed as a percentage, not decimally), and

D = pipe diameter (in).

H = K V12/ 2g (ft), where K = loss coefficient for a given fitting,

APPENDIX A

5 ■

V1 = inlet velocity (ft/s), and g = acceleration due to gravity (32.2 ft/s2). Values of K for the flow of water through various types of bends and fittings are tabulated in numerous reference sources (Reference 3, for example). The loss coefficient for valves may be obtained from the valve manufacturer. The loss coefficient for pulp suspensions in a given bend or fitting generally exceeds the loss coefficient for water in the same bend or fitting. As an approximate rule, the loss coefficient (K) increases 20 percent for each 1 percent increase in oven-dried stock consistency. Please note that this is an approximation; actual values of K may differ, depending on the type of bend or fitting under consideration (4).

D = pipe inside diameter (mm), and

, , y = exponents (constant for a given pulp), obtained from Table II or IIA. VW = 1.22 C1.40 (m/s),

3M ■

where C = consistency (oven-dried, expressed as a percentage, not decimally). (▲ ▲H/ L)w = 264 V1.75 D -1.25 (m/100m),

4M ■

where V = bulk velocity (m/s), and D = pipe inside diameter (mm). H = K V12/ 2g (m), where K = loss coefficient for a given fitting,

5M ■

V1 = inlet velocity (m/s), and g = acceleration due to gravity (9.81 m/s2).

TECH-E

APPENDIX B

TABLE I Data for use with Equation 1 or Equation 1M to determine velocity limit, Vmax (1).

■

■

Pulp Type

Pipe Material

K'



Unbeaten aspen sulfite never dried Long fibered kraft never dried CSF = 725 (6)

Stainless Steel PVC Stainless Steel PVC PVC PVC PVC Stainless Steel PVC PVC PVC PVC PVC PVC PVC PVC

0.85 (0.26) 0.98 (0.3) 0.89 (0.27) 0.85 (0.26) 0.75 (0.23) 0.75 (0.23) 0.79 (0.24) 0.59 (0.18) 0.49 (0.15) 0.69 (0.21) 4.0 (1.22) 4.0 (1.22) 4.0 (1.22) 4.0 (1.22) 4.0 (1.22) 0.59 (0.18)

1.6 1.85 1.5 1.9 1.65 1.8 1.5 1.45 1.8 1.3 1.40 1.40 1.40 1.40 1.40 1.8

Long fibered kraft never dried CSF = 650 (6) Long fibered kraft never dried CSF = 550 (6) Long fibered kraft never dried CSF = 260 (6) Bleached kraft never dried and reslurried (6) Long fibered kraft dried and reslurried (6) Kraft birch dried and reslurried (6) Stone groundwood CSF = 114 Refiner groundwood CSF = 150 Newsprint broke CSF = 75 Refiner groundwood (hardboard) Refiner groundwood (insulating board) Hardwood NSSC CSF = 620

NOTES: 1. When metric (SI) units are utilized. use the value of K' given in parentheses. When the metric values are used, diameter (D) must be in millimetres (mm) and velocity (V) in metres per second (m/s). 2. Original data obtained in stainless steel and PVC pipe. PVC is taken to be hydraulically smooth pipe. 3. Stainless steel may be hydraulically smooth although some manufacturing processes may destroy the surface and hydraulic smoothness is lost. 4. For cast iron and galvanized pipe, the K' values will be reduced. No systematic data are available for the effects of surface roughness. 5. It pulps are not identical to those shown, some engineering judgement is required. 6. Wood is New Zealand Kraft pulp.

TABLE IA Data (5, 6) for use with Equation 1 or Equation 1M to determine velocity limit, Vmax.

■

Pulp Type (5) Unbleached sulphite Bleached sulphite Kraft Bleached straw Unbleached straw

■

Pipe Material Copper Copper Copper Copper Copper

K' 0.98 0.98 0.98 0.98 0.98

(0.3) (0.3) (0.3) (0.3) (0.3)

 1.2 1.2 1.2 1.2 1.2

Estimates for other pulps based on published literature.

Pulp Type (5, 6) Cooked groundwood Soda NOTE:

Pipe Material

K'



Copper Steel

0.75 (0.23) 4.0 (1.22)

1.8 1.4

When metric (SI) units are utilized, use the value of K' given in parentheses, When the metric values are used, diameter (D) must be millimeters (mm) and velocity (V) in meters per second (m/s)

TECH-E

APPENDIX C

TABLE II Data for use with Equation 2 or Equation 2M to determine head loss, ▲H/ L (1).

■

■

Pulp Type

K





Unbeaten aspen sulfite never dried Long fibered kraft never dried CSF = 725 (5) Long fibered kraft never dried CSF = 650 (5) Long fibered kraft never dried CSF = 550 (5) Long fibered kraft never dried CSF = 260 (5) Bleached kraft bleached and reslurred (5) Long fibered kraft dried and reslurred (5) Kraft birch dried and reslurred (5) Stone groundwood CSF = 114 Refiner groundwood CSF = 150 Newspaper broke CSF = 75 Refiner groundwood CSF (hardboard) Refiner groundwood CSF (insulating board) Hardwood NSSF CSF = 620

5.30 (235) 11.80 (1301) 11.30 (1246) 12.10 (1334) 17.00 (1874) 8.80 (970) 9.40 (1036) 5.20 (236) 3.81 (82) 3.40 (143) 5.19 (113) 2.30 (196) 1.40 (87) 4.56 (369)

0.36 0.31 0.31 0.31 0.31 0.31 0.31 0.27 0.27 0.18 0.36 0.23 0.32 0.43

2.14 1.81 1.81 1.81 1.81 1.81 1.81 1.78 2.37 2.34 1.91 2.21 2.19 2.31

y -1.04 -1.34 -1.34 -1.34 -1.34 -1.34 -1.34 -1.08 -0.85 -1.09 -0.82 -1.29 -1.16 -1.20

NOTES: 1. When metric (SI) units are utilized, use the value of K given in parentheses. When the metric values are used, diameter (D) must be in millimetres (mm) and velocity must be in metres per second (m/s). 2. Original data obtained in stainless steel and PVC pipe (7,8, 9). 3. No safety factors are included in the above correlations. 4. The friction loss depends considerably on the condition of the inside of the pipe surface (10). 5. Wood is New Zealand Kraft pulp. TABLE IA Data (5, 6) for use with Equation 2 or Equation 2M to determine head loss, ▲H/ L.

■

K



12.69 (1438) 11.40 (1291) 1140 (1291) 11.40 (1291) 5.70 (646)

0.36 0.36 0.36 0.36 0.36

K



6.20 (501) 6.50 (288)

0.43 0.36

Pulp Type (5) Unbleached sulfite Bleached sulfite Kraft Bleached straw Unbleached straw

■

 1.89 1.89 1.89 1.89 1.89

y -1.33 -1.33 -1.33 -1.33 -1.33

Estimates for other pulps based on published literature.

Pulp Type (5, 6) Cooked groundwood Soda NOTE:

 2.13 1.85

y -1.20 -1.04

When metric (SI) units are utilized, use the value of K given in parentheses, When the metric values are used, diameter (D) must be millimeters (mm) and velocity (V) in meters per second (m/s)

APPENDIX D The following gives supplemental information to that where I.P.D. mill capacity (metric tons per day), provided in the main text. 1. Capacity (flow), Q — Q = 16.65 (T.P.D.) (U.S. GPM), C

(i)

Where T.P.D. = mill capacity (short tons per day), and C = consistency (oven-dried, expressed as a percentage, not decimally). If SI units are used, the following would apply: -3 Q = 1.157 (10 ) (T.P.D.) (m3/s), C

Where T.P.D. = mill capacity (metric tons per day), and C = consistency (oven-dried, expressed as a percentage, not decimally). 2. Bulk velocity, V — V = 0.321 Q (ft/s), or A

(ii)

V = 0.4085 Q D2

(ii)

(ft/s),

Where Q = capacity (U.S. GPM) A = inside area of pipe (in2), and D = inside diameter of pipe (in) (iM)

TECH-E

The following would apply if SI units are used: 6 V = 1 (10 ) Q (m/s), or A 6 V = 1.273 (10 ) Q (m/s), D2

APPENDIX E (iiM) (iiM)

Where Q = capacity (m3/s), A = inside area of pipe (mm2), and D = inside diameter of pipe (mm) 3.Multiplication Factor, F (.included in Equation 2 ) F = F1• F2 • F3 • F4 • F5, (iv) where F1 =correction factor for temperature. Friction loss calculations are normally based on a reference pulp temperature of 95° F (35°C). The flow resistance may be increased or decreased by 1 percent for each 1.8°F (1°C) below or above 95°F (35°C), respectively. This may be expressed as follows: F1 = 1.528 - 0.00556 T, (v) where T = pulp temperature (° F), or F1 = 1.35 - 0.01 T, (vM) where T = pulp temperature (°C). F2 = correction factor for pipe roughness. This factor may vary due to manufacturing processes of the piping, surface roughness, age, etc. Typical values for PVC and stainless steel piping are listed below: F2 = 1.0 for PVC piping, F2 = 1.25 for stainless steel piping. Please note that the above are typical values; experience and/or additional data may modify the above factors. F3 = correction factor for pulp type. Typical values are listed below: F3 = 1.0 for pulps that have never been dried and reslurried, F3 = 0.8 for pulps that have been dried and reslurried. Note: This factor has been incorporated in the numerical coefficient, K, for the pulps listed in Table II. When using Table II, F3 should not be used. F4 = correction factor for beating. Data have shown that progressive beating causes, initially, a small decrease in friction loss, followed by a substantial increase. For a kraft pine pulp initially at 725 CSF and F4 = 1.0, beating caused the freeness to decrease to 636 CSF and F4 to decrease to 0.96. Progressive beating decreased the freeness to 300 CSF and increased F4 to 1.37 (see K values in Table II). Some engineering judgement may be required. F5 = design safety factor. This is usually specified by company policy with consideration given to future requirements.

The following are three examples which illustrate the method for determination of pipe friction loss in each of the three regions shown in Figure 3. Example 1. Determine the friction loss (per 100 ft of pipe) for 1000 U.S. GPM of 4.5% oven-dried unbeaten aspen sulfite stock, never dried, in 8 inch schedule 40 stainless steel pipe (pipe inside diameter = 7.981 in). Assume the pulp temperature to be 95° F. Solution: a) The bulk velocity, V, is V = 0.4085 Q, D2 and Q = flow = 1000 U.S. GPM. D = pipe inside diameter = 7.981 in. 0.4085 (1000) = 6.41 ft/s. V= 7.9812 b) It must be determined in which region (1, 2, or 3) this velocity falls. Therefore, the next step is to determine the velocity at the upper limit of the linear region, Vmax. Vmax = K' C,

1 ■

and K' = numerical coefficient = 0.85 (from Appendix B, Table I), C = consistency = 4.5%,

 = exponent = 1.6 (from Appendix B, Table I). Vmax = 0.85 (4.51.6) = 9.43 ft/s. c) Since Vmax exceeds V, the friction loss, ▲H/ L, falls within the linear region, Region 1. The friction loss is given by the correlation: ▲H/ L =F K V C Dy

2 ■

and F = correction factor = F1• F2 • F3 • F4 • F5, F1 = correction factor for pulp temperature. Since the pulp temperature is 95° F, F1 = 1.0, F2 = correction factor for pipe roughness. For stainless steel pipe, F2 = 1.25 (from Appendix D), F3 = correction factor for pulp type. Numerical coefficients for this pulp are contained in Appendix C, Table II, and have already incorporated this factor. F4 = correction factor for beating. No additional beating has taken place, therefore F4 = 1.0 (from Appendix D), F5 = design safety factor. This has been assumed to be unity. F5 = 1.0. F = (1.0) (1.25) (1.0) (1.0) (1.0) = 1.25, K = numerical coefficient = 5.30 (from Appendix C, Table II), , , y = exponents = 0.36, 2.14, and -1.04, respectively (from Appendix C, Table II), V, C, D have been evaluated previously.

TECH-E

(ii)

▲H/ L

= (1.25) (5.30) (6.410.36) (4.52.14) (7.981-1.04)

Example 3.

=(1.25) (5.30) (1.952) (25.0) (0.1153) = 37.28 ft head loss/100 ft of pipe. This is a rather substantial head loss, but may be acceptable for short piping runs. In a large system, the economics of initial piping costs versus power costs should be weighed, however, before using piping which gives a friction loss of this magnitude. Example 2. Determine the friction loss (per 100 ft of pipe) of 2500 U.S. GPM of 3% oven-dried bleached kraft pine, dried and reslurried, in 12 inch schedule 10 stainless steel pipe (pipe inside diameter = 12.39 in). Stock temperature is 1250F. Solution:

Determine the friction loss (per 100 ft of pipe) for 2% oven-dried bleached kraft pine, dried and reslurried, through 6 inch schedule 40 stainless steel pipe (inside diameter = 6.065 in). The pulp temperature is 90° F; the flow rate 1100 U.S. GPM. Solution: a)The bulk velocity is V = 0.4085 Q, D2 = 0.4085 (1100) = 12.22 ft/s. 6.0652 b) It must be determined in which region (1, 2 or 3) this velocity falls. To obtain an initial indication, determine Vmax. 1 ■

Vmax = K' C ,

a) V, the bulk velocity, is V = 0.4085 Q, D2

(ii)

and K' = 0.59 (from Appendix B, Table I),



= 0.4085 (2500) = 6.65 ft/s. 12.392

= 1.45 (from Appendix B, Table I).

Vmax = 0.59 (201.40) = 1.61 ft/s.

b) The velocity at the upper limit of the linear region, Vmax, is Vmax = K' C ,

1 ■

and K' = 0.59 (from Appendix B, Table I),

c) Since V exceeds Vmax, Region 1 (the linear region) is eliminated. To determine whether V lies in Region 2 or 3, the velocity at the onset of drag reduction, Vw, must be calculated.

3 ■

VW = 4.00 C1.40 = 4.00 (2.01.40) = 10.56 ft/s.

= 1.45 (from Appendix B, Table I). Vmax = 0.59 (3.01.45) = 2.90 ft/s.

d) V exceeds Vw, indicating that it falls in Region 3. The friction loss is calculated as that of water flowing at the same velocity.

c) Region 1 (the linear region) has been eliminated, since the bulk velocity, V, exceeds Vmax.

4 ■

(▲ ▲H/ L) w = 0.579 V1.75 D-1.25,

The next step requires calculation of Vw. VW = 4.00 C1.40

3 ■

= 0.579 (12.221.75) (6.065-1.25) = 4.85 ft head loss/100 ft of pipe.

= 4.00 (3.01.40) = 18.62 ft/s.

This will be a conservative estimate, as the actual friction loss curve for pulp suspensions under these conditions will be below the water curve.

d) V exceeds Vmax, but is less than Vw, indicating that it falls in Region 2. The friction loss in this region is calculated by substituting Vmax into the equation for head loss, Equation 2 . ▲H/ L = F K (Vmax)  C Dy,

(ii)

■

and F1 • F2 • F3 • F4 • F5; F1 = 1.528 - 0.00556T, and T = stock temperature = 125° F F1 = 1.528 - 0.00556 (125) = 0.833,

(iv) (v)

REFERENCES (1)

TAPPI Technical Information Sheet (TIS) 408-4. Technical Association of the Pulp and Paper Industry, Atlanta, Georgia (1981). (2) K. Molter and G.G. Duffy, TAPPI 61,1, 63 (1978).

(3)

Hydraulic Institute Engineering Data Book. First Edition, Hydraulic Institute, Cleveland, Ohio (1979).

F2 = 1.25 (from Appendix D), F3 = F4 = F5 = 1.0,

(4)

K. Molter and G. Elmqvist, TAPPI 63. 3,101 (1980).

F = 0.833 (1.25) (1.0) = 1.041,

(5)

W. Brecht and H. Helte, TAPPI 33, 9, 14A (1950).

K = 8.80 (from Appendix C, Table II),

(6)

R.E. Durat and L.C. Jenness. TAPPI 39, 5, 277 (1956)

, , y =

(7)

K. Molter, G.G. Duffy and AL Titchener, APPITA 26, 4, 278 (1973)

Vmax, C, and D have been defined previously.

(8)

G.G. Duffy and A.L. Titchener, TAPPI 57, 5, 162 (1974)

▲H/ L

(9)

G.G. Duffy, K. Molter, P.F.W. Lee. and S.W.A. Milne, APPITA 27, 5, 327 (1974).

0.31,1.81, and -1.34, respectively (from Appendix C, Table II),

= 1.041 (8.80) (2.900.31) (3.01.81) (12.39-1.34) = 1.041 (8.80) (1.391) (7.304) (0.03430) = 3.19 ft head loss/100 ft of pipe.

(10) G.G. Duffy, TAPPI 59, 8, 124 (1976). (11) G.G. Duffy, Company Communications. Goulds Pumps. Inc.. (1980-1981)

TECH-E

TECH-E-4 Pump Types Used in the Pulp & Paper Industry Mill Area

Woodyard

Pump Mill

Bleach Plant

Stock Prep

Paper Machine (Wet End)

Paper Machine (Dry End) Coater

Kraft Recovery

Utility (Power House)

Miscellaneous

Recycle

TECH-E

Typical Services

Typical Pump Construction

Log Flume Log/Chip Pile spray Chip Washer

Al/316SS Trim AI/316SS trim Al/316SS Trim

Shower Supply Dilution Supply Screen Supply Cleaner Supply Decker Supply Hi/Med. Density Storage Transfer Medium Cons. Storage Chip Chute Circulation White Liquor Circulation Condensate Wash Liquor Circulation Brown Stock Storage Bleach Tower Storage Bleach Chemical Mixing High Density Storage Chemical Feed Washer Supply Washer Shower Water Dilution Water Medium Consistency O2 Reactor CI02 Generator Circulation Refiner Supply Deflaker Supply Machine Chest Supply Fan Pumps Couch Pit Saveall Sweetner Shower Dryer Drainage Condensate Trim Squirt Broke Chest Coating Slurries Kaolin Clay (Fillers) Weak Black Liquor Evaporator Circulation Concentrated Black Liquor Condensate Injection Black Liquor Transfer Pumps Smelt Spout Cooling Water Collection Weak Wash Scrubber Green Liquor (Storage Transfer) Lime Mud Dregs Feedwater Condensate Deaerator Booster

Al/31SS Al316SS Al316SS 316SS 316SS/317SS 316SS/317SS Various 316SS/317SS CD4MCu CD4MCu Al/316SS 316SS 316SS 316SS 317SS, 254 SMO, Titanium 316SS/317SS 316SS 316SS 316SS 316SS 316SS Titanium 316SS 316SS 316SS Al/316SS Trim, All 316SS Al/316SS Trim, All 316SS Al/316SS Trim, All 316SS Al/316SS Trim, All 316SS A/316SS Trim, All 316SS Al/316SS Trim, Al/316SS Trim Al/316SS Trim Al/316SS Trim 316SS/CD4MCu 316SS/CD4MCu 316SS 316SS 316SS 316SS 316SS CD4MCu Al/316SS Trim Al/316SS Trim 316SS/CD4MCu/28% Chrome 316SS/CD4MCu/28% Chrome 316SS/CD4MCu/28% Chrome 316SS/CD4MCu/28% Chrome CS/Chrome Trim/All Chrome 316SS 316SS

Mill Water Supply Sump Pumps

Al/316SS Trim Al/316SS Trim

Hole/Slot Screen Supply Rejects Floating Cell Medium Consistency Storage Hydro Pulper Dilution Water

316SS/CD4MCu 316SS/CD4MCu 316SS 316SS/317SS 316SS/CD4MCu Al/316SS Trim

Pump Type

Goulds Model

Mixed Flow Vertical Turbine Stock ANSI Double Suction Stock ANSI Double Suction Medium Consistency Hi Temp/Press Stock

MF VIT 3175, 3180/85 3196 3410, 3415, 3420 3175, 3180/85 3196 3410, 3415, 3420 3500 3181/86

Stock ANSI Medium Consistency Axial Flow Non-metallic

3175, 3180/85 3196 3500 AF NM 3196

Stock ANSI

3175, 3180/85 3196

Double Suction Stock Low Flow High Pressure Two-Stage ANSI Low Flow Stock

3415, 3420 3175, 3180/85 LF3196 3310H 3316 3196 LF 3196 3175, 3180/85

ANSI 3196 Medium Duty Slurry JC ANSI 3196 Stock 3175, 3180/85 Medium Duty Slurry JC High Temp/Pressure Stock 3181/86 High Pressure 3316 Multi Stage

Multi-Stage ANSI High Pressure Vertical Can Double Suction Vertical Turbine Self-Priming Vertical Sumps Vertical Sump; Recessed Submersible Stock Recessed ANSI Medium Consistency

3310H, 3600 3196 3700, VIC 3410, 3415, 3420 VIT 3796 3171 VHS HSU 3175, 3180/85 CV 3196,HS 3196 3500

Section TECH-F Mechanical Data TECH-F-1 Standard Weights and Dimensions of Mechanical Joint Cast Iron Pipe, Centrifugally Cast Extracted from USA Standard Cast Iron Pipe Flanges and Flanged Fittings (USAS B16. 1–1967), with the permission of the publisher, The American Society of Mechanical Engineers, United Engineering Center, 345 East 47th Street, New York, New York 10017.

Nom. Size & (Outside Diam), In.

Thickness, In.

Wall Weight Per Foot*

Average Thickness Class

3 (3.96)

0.32 0.35 0.38 0.35 0.38 0.41 0.44 0.38 0.41 0.44 0.48 0.52 0.41 0.44 0.48 0.52 0.56 0.60 0.44 0.48 0.52 0.56 0.60 0.65 0.48 0.52 0.56 0.60 0.65 0.70 0.76 0.48 0.51 0.55 0.59 0.64 0.69 0.75 0.81

11.9 12.9 13.8 16.1 17.3 18.4 19.6 25.4 27,2 29.0 31.3 33.6 36.2 38.6 41.8 45.0 48.1 51.2 48.0 52.0 55.9 59.9 63.8 68.6 62.3 67.1 59.9 76.6 82.5 88.3 95.2 73.6 77.8 83.4 89.0 95.9 102.7 110.9 118.9

22 23 24 22 23 24 25 22 23 24 25 26 22 23 24 25 26 27 22 23 24 25 26 27 22 23 25 25 26 27 28 21 22 23 24 25 26 27 28

4 (4.80)

6 (6.90)

8 (9.05)

10 (11.10)

12 (13.20)

14 (15.30)

Nom. Size & (Outside Diam), In.

16 (17.40)

18 (19.50)

20 (21.60)

24 (25.80)

Thickness, In.

Wall Weight Per Foot*

Average Thickness Class

0.50 0.54 0.58 0.63 0.68 0.73 0.79 0.85 0.54 0.58 0.63 0.68 0.73 0.79 0.85 0.92 0.57 0.62 0.67 0.72 0.78 0.84 0.91 0.98 0.63 0.68 0.73 0.79 0.85 0.92 0.99 1.07

87.6 94.0 100.3 108.3 116.2 124.0 133.3 142.7 106.0 113.2 122.2 131.0 140.0 150.6 161.0 173.2 124.2 134.2 144.2 154.1 165.9 177.6 191.2 214.8 164.2 176.2 188.2 202.6 216.8 233.2 249.7 268.2

21 22 23 24 25 26 27 28 21 22 23 24 25 26 27 28 21 22 23 24 25 26 27 28 21 22 23 24 25 26 27 28

*Based on 20 Ft. Laying Length of Mech. Joint Pipe including Bell.

TECH-F

TECH-F-2 125 Lb. & 250 Lb. Cast Iron Pipe Flanges and Flanged Fittings Thickness of Flange (Min.)

Nomi- Diam. nal of Pipe Flange Size

41⁄4 4 5⁄8 5 6 7 71⁄2 81⁄2 9 10 11 131⁄2 16 19 21 231⁄2 25 271⁄2 32 383⁄4 46 53 591⁄2

1 11⁄4 11⁄2 2 21⁄2 3 31⁄2 4 5 6 8 10 12 14 16 18 20 24 30 36 42 48

Diam. of Bolt Circle

31⁄8 31⁄2 37⁄8 43⁄4 51⁄2 6 7 71⁄2 81⁄2 91⁄2 113⁄4 141⁄4 17 183⁄4 211⁄2 223⁄4 25 291⁄2 36 423⁄4 491⁄2 56

7⁄16 1⁄2 9 ⁄16 5 ⁄8 11 ⁄16 3⁄4 13⁄16 15 ⁄16 15 ⁄16

1 11⁄8 13⁄16 11⁄4 13⁄8 17⁄16 19⁄16 111⁄16 17⁄8 21⁄8 23⁄8 25⁄8 23⁄4

Number of Bolts

4 4 4 4 4 4 8 8 8 8 8 12 12 12 16 16 20 20 28 32 36 44

Diam. of Bolts

Diam. of Length Drilled of Bolt Bolts Holes

1 ⁄2 1 ⁄2 1 ⁄2 5 ⁄8 5 ⁄8 5 ⁄8 5 ⁄8 5 ⁄8 3 ⁄4 3 ⁄4 3 ⁄4 7 ⁄8 7 ⁄8

13⁄4 2 2 21⁄4 21⁄2 21⁄2 23⁄4 3 3 31⁄4 31⁄2 33⁄4 33⁄4 41⁄4 41⁄2 43⁄4 5 51⁄2 61⁄4 7 71⁄2 73⁄4

5⁄8 5⁄8 5⁄8 3⁄4 3⁄4 3⁄4 3⁄4 3⁄4 7⁄8 7⁄8 7⁄8

1 1 11⁄8 11⁄8 11⁄4 11⁄4 13⁄8 13⁄8 13⁄8 13⁄8 13⁄8

1 1 11⁄8 11⁄8 11⁄4 11⁄4 11⁄2 11⁄2 11⁄2

ThickNomi- Diam. ness nal of of Pipe Flange Flange3 Size (Min.)

Diam. of Bolt Circle

Diam. of Bolt Holes1

47⁄8 51⁄4 61⁄8 61⁄2 71⁄2 81⁄4 9 10 11 121⁄2 15 171⁄2 201⁄2 23 251⁄2 28 301⁄2 36 43 50 57 65

31⁄2 37⁄8 41⁄2 5 57⁄8 65⁄8 71⁄4 77⁄8 91⁄4 105⁄8 13 151⁄4 173⁄4 201⁄4 221⁄2 243⁄4 27 32 391⁄4 46 523⁄2 603⁄4

3⁄4 3⁄4 7⁄8 3 ⁄4 7⁄8 7⁄8 7⁄8 7⁄8 7⁄8 7⁄8

1 11⁄4 11⁄2 2 21⁄2 3 31⁄2 4 5 6 8 10 12 14 16 18 20 24 *30 *36 *42 *48

1 11⁄4 11⁄2 2 21⁄2 3 31⁄2 4 5 6 8 10 12 14 16 18 20 24 30 36 42 48

Center to Face

A

B

C

D

E

Face to Face F

31⁄2 33⁄4 4 41⁄2 5 51⁄2 6 61⁄2 71⁄2 8 9 11 12 14 15 161⁄2 18 22 25 28* 31* 34*

5 51⁄2 6 61⁄2 7 73⁄4 81⁄2 9 101⁄4 111⁄2 14 161⁄2 19 211⁄2 24 261⁄2 29 34 411⁄2 49 561⁄2 64

13⁄4 2 21⁄4 21⁄2 3 3 31⁄2 4 41⁄2 5 51⁄2 61⁄2 71⁄2 7 1⁄2 8 81⁄2 91⁄2 11 15 18 21 24

53⁄4 61⁄4 7 8 1 9 ⁄2 10 111⁄2 12 131⁄2 141⁄2 171⁄2 201⁄2 241⁄2 27 30 32 35 401⁄2 49 …. …. ….

13⁄4 13⁄4 2 2 1⁄2 21⁄2 3 3 3 31⁄2 31⁄2 41⁄2 5 51⁄2 6 61⁄2 7 8 9 10 …. …. ….

…. …. …. 5 51⁄2 6 61⁄2 7 8 9 11 14 14 16 18 19 20 24 30 36 42 48

1 11⁄8 13⁄16 11⁄4 13⁄8 17⁄16 15⁄8 17⁄8 2 21⁄8 21⁄4 23⁄8 21⁄2 23⁄4 3 33⁄8 311⁄16 4

Body Wall Thick nesst 5 ⁄16 5 ⁄16 5 ⁄16 5 ⁄16 5⁄16 3 ⁄8 7⁄16 1 ⁄2 1 ⁄2 9 ⁄16 5 ⁄8 3 ⁄4 13/16 7 ⁄8

Nomi- Inside Wall Diam. nal Diam. Thickof Pipe of ness Raised Size Fitting of Face (Min.) Body*

2 2 21⁄2 21⁄2 3 3 31⁄2 31⁄2 4 4 5 5 6 6 8 8 10 10 12 12 14 131⁄4 16 151⁄4 18 17 20 19 24 23

1 11⁄16 11⁄8 11⁄4 17⁄16 15⁄8 113⁄16 2

A

A

1 11⁄8 11⁄8 11⁄4 11⁄4 11⁄4 11⁄2 11⁄2 2 2 2

7 ⁄16 1⁄2 9 ⁄16 9⁄16 5 ⁄8 11 ⁄16 3 ⁄4 13 ⁄16 15 ⁄16

1 11⁄8 11⁄4 13⁄8 11⁄2 15⁄8

4 3⁄16 4 15⁄16 5 11⁄16 6 5⁄16 615⁄16 8 5⁄16 911⁄16 1115⁄16 141⁄6 167⁄16 1815⁄16 211⁄16 235⁄16 259⁄16 301⁄4

A

5 51⁄2 6 61⁄2 7 8 81⁄2 10 111⁄2 13 15 161⁄2 18 191⁄2 221⁄2

B

C

6 1⁄2 3 7 31⁄2 3 7 ⁄4 31⁄2 81⁄2 4 9 41⁄2 1 10 ⁄4 5 111⁄2 51⁄2 14 6 16 1⁄2 7 19 8 1 21 ⁄2 81⁄2 24 91⁄2 261⁄2 10 29 101⁄2 34 12

D

Face to Face F

E

9 21⁄2 101⁄2 21⁄2 11 3 121⁄2 3 1 13 ⁄2 3 15 31⁄2 1 17 ⁄2 4 201⁄2 5 24 51⁄2 1 27 ⁄2 6 31 61⁄2 1 34 ⁄2 71⁄2 371⁄2 8 401⁄2 8 1⁄2 1 47 ⁄2 10

5 51⁄2 6 6 1⁄2 7 8 9 11 12 14 16 18 19 20 24

B A

A C B

A

C 90° ELBOW

A

21⁄2 21⁄2 23⁄4 23⁄4 31⁄4 31⁄2 31⁄2 3 3⁄4 4 4 41⁄2 51⁄4 51⁄2 6 61⁄4 61⁄2 63⁄4 73⁄4 81⁄2 91⁄2 101⁄4 103⁄4

Center to Face

A

A

5 ⁄8 5 ⁄8 3 ⁄4 5 ⁄8 3 ⁄4 3 ⁄4 3 ⁄4 3 ⁄4 3 ⁄4 3 ⁄4 7 ⁄8

4 4 4 8 8 8 8 8 8 12 12 16 16 20 20 24 24 24 28 32 36 40

1 11⁄8 11⁄4 11⁄4 13⁄8 13⁄8 13⁄8 15⁄8 2 21⁄4 21⁄4 21⁄4

Length of Bolts2

Chart 5 American Standard Class 250 Cast Iron Flanged Fittings (ASA B16b)

Chart 4 American Standard Class 125 Cast Iron Flanged Fittings (ASA B16.1)

A

Size of Bolt

Chart 3 American Standard Class 250 Cast Iron Pipe Flanges (ASA B16b)

Chart 2 American Standard Class 125 Cast Iron Pipe Flanges (ASA B16.1) Nominal Pipe Size

11⁄16 3⁄4 13⁄16 7⁄8

Number of Bolts1

90° LONG RADIUS ELBOW

45° ELBOW

SIDE OUTLET ELBOW

A A

A

A

A

A

A

A

A D

90°

45° D

A

DOUBLE BRANCH ELBOW

TECH-F

TEE

CROSS

SIDE OUTLET TEE OR CROSS

E

F

F

REDUCER

ECCENTRIC REDUCER

TRUE Y

E 45° LATERAL

TECH-F-3 Steel Pipe, Dimensions and Weights Size: Nom. & (Outside Diam.), In.* 1 ⁄8 (0.405) 1 ⁄4 (0.540) 3 ⁄8 (0.675) 1 ⁄2 (0.840)

3

⁄4 (1.050) 1 (1.315) 11⁄4 (1.660) 11⁄2 (1.900) 2 (2.375) 21⁄2 (2.875) 3 (3.500) 31⁄2 (4.000) 4 (4.500) 5 (5.563)

6 (6.625)

8 (8.625)

10 (10.750)

Wall Thickness, In.

Weight per Foot, Plain Ends, Lb.

0.068 0.095 0.088 0.119 0.091 0.126 0.109 0.147 0.188 0.294 0.113 0.154 0.219 0.308 0.133 0.179 0.250 0.308 0.140 0.191 0.250 0.382 0.145 0.200 0.281 0.400 0.154 0.218 0.344 0.436 0.203 0.276 0.375 0.552 0.216 0.300 0.438 0.600 0.226 0.318 0.237 0.337 0.438 0.531 0.674 0.258 0.375 0.500 0.625 0.750 0.280 0.432 0.562 0.719 0.864 0.250 0.277 0.322 0.406 0.500 0.594 0.719 0.812 0.875 0.906 0.250 0.307 0.365 0.500 0.594 0.719 0.844 1.000 1.125

0.24 0.31 0.42 0.54 0.57 0.74 0.85 1.09 1.31 1.71 1.13 1.47 1.94 2.44 1.68 2.17 2.84 2.44 2.27 3.00 3.76 5.21 2.72 3.63 4.86 6.41 3.65 5.02 7.46 9.03 5.79 7.66 10.01 13.70 7.58 10.25 14.31 18.58 9.11 12.51 10.79 14.98 18.98 22.52 27.54 14.62 20.78 27.04 32.96 38.55 18.97 28.57 36.42 45.34 53.16 22.36 24.70 28.55 35.66 43.39 50.93 45.34 67.79 72.42 74.71 28.04 34.24 40.48 54.74 64.40 77.00 89.27 104.13 115.65

Schedule No. 40 80 40 80 40 80 40 80 160

S XS S XS S XS S XS

XXS 40 S 80 XS 160 XXS 40 S 80 XS 160 XXS 40 S 80 XS 160 XXS 40 S 80 XS 160 XXS 40 S 80 XS 160 XXS 40 S 80 XS 160 XXS 40 S 80 XS 160 XXS 40 S 80 XS 40 S 80 XS 120 160 XXS 40 S 80 XS 120 160 XXS 40 S 80 XS 120 160 XXS 20 30 40 S 60 80 XS 100 160 140 XXS 160 20 30 40 S 60 XS 80 100 120 140 XXS 160

Size: Nom. & (Outside Diam.), In.*

12 (12.750)

14 (14.000)

16 (16.000)

18 (18.000)

20 (20.000)

22 (22.000)

24 (24.000)

Wall Thickness, In.

Weight per Foot, Plain Ends, Lb.

0.250 0.330 0.375 0.406 0.500 0.562 0.688 0.844 1.000 1.125 1.312 0.250 0.312 0.375 0.438 0.500 0.594 0.750 0.938 1.094 1.250 1.406 0.250 0.312 0.375 0.500 0.656 0.844 1.031 1.219 1.438 1.594 0.250 0.312 0.375 0.438 0.500 0.562 0.750 0.938 1.156 1.375 1.562 1.781 0.250 0.375 0.500 0.594 0.812 1.031 1.281 1.500 1.750 1.969 0.250 0.375 0.500 0.875 1.125 1.375 1.625 1.875 2.125 0.250 0.375 0.250 0.375 0.500 0.562 0.688 0.969 1.219 1.531 1.812 2.062 2.344

33.38 43.77 49.56 53.56 65.42 73.22 88.57 107.29 125.49 139.68 160.33 36.71 45.68 54.57 63.37 72.09 85.01 106.13 130.79 150.76 170.22 189.15 42.05 52.36 62.58 82.77 107.54 136.58 164.86 192.40 223.57 245.22 47.39 59.03 70.59 82.06 93.45 104.76 138.17 170.84 208.00 244.14 274.30 308.55 47.39 78.60 93.45 123.06 166.50 208.92 256.15 296.37 341.10 379.14 58.07 86.61 114.81 197.42 250.82 302.88 353.61 403.01 451.07 63.41 94.62 63.41 94.62 125.49 140.80 171.17 238.29 296.53 367.45 429.50 483.24 542.09

Schedule No. 20 30 S 40 XS 60 80 100 120 XXS 140 160 10 20 30 S 40 XS 60 80 100 120 140 160 10 20 30 S 40 XS 60 80 100 120 140 160 10 20 S 30 XS 40 60 80 100 120 140 160 10 20 S XS 40 60 80 100 120 140 160 10 20 S 30 XS 60 80 100 120 140 160 10 20 S 10 20 S XS 30 40 60 80 100 120 140 160

TECH-F

TECH-F-4 150 Lb. & 300 Lb. Steel Pipe Flanges and Fittings Extracted from USA Standard Cast Iron Pipe Flanges and Flanged Fittings (USAS, B16. 5-1968), with the permission of the publisher, The American Society of Mechanical Engineers, United Engineering Center, 345 East 47th Street, New York NY 10017.

Nomi- Diam. nal of Pipe Flange Size O 1 ⁄2 3⁄4

31⁄2 37⁄8 41⁄4 45⁄8 5 6 7 71⁄2 81⁄2 9 10 11 131⁄2 16 19 21 231⁄2 25 271⁄2 32

1 11⁄4 11⁄2 2 21⁄2 3 31⁄2 4 5 6 8 10 12 14 16 18 20 24

Thickness of Flange (Min.)*

Diam. of Bolt Circle

Diam. of Bolt Holes

23⁄8 21⁄4 31⁄8 31⁄2 37⁄8 43⁄4 51⁄2 6 7 71⁄2 81⁄2 91⁄2 113⁄4 141⁄4 17 183⁄4 211⁄4 223⁄4 25 291⁄2

5⁄8 5⁄8 5⁄8 5⁄8 5⁄8 3⁄4 3⁄4 3 ⁄4 3 ⁄4 3⁄4 7⁄8 7⁄8 7⁄8

7⁄16 1⁄2 9⁄16 5⁄8 11⁄16 3 ⁄4 7 ⁄8 15⁄16 15⁄16 15 ⁄16 15 ⁄16

1 11⁄8 13⁄16 11⁄4 13⁄8 17⁄16 19⁄16 111⁄16 17⁄8

Number of Bolts

Diam. of Bolts

4 4 4 4 4 4 4 4 8 8 8 8 8 12 12 12 16 16 20 20

1 ⁄2 1 ⁄2 1 ⁄2 1 ⁄2 1 ⁄2 5 ⁄8 5 ⁄8 5 ⁄8 5 ⁄8 5 ⁄8 3 ⁄4 3 ⁄4 3 ⁄4

1 1

11⁄8 11⁄8 11⁄4 11⁄4 3⁄8

7/8 7/8 1 1 11⁄8 11⁄8 11⁄4

Length of (with 1 ⁄16" Raised Face

13⁄4 2 2 2 1⁄4 21⁄4 23⁄4 3 3 3 3 31⁄4 31⁄4 31⁄2 33⁄4 4 41⁄4 41⁄2 43⁄4 51⁄4 53⁄4

Nominal Pipe Size

AA

BB

CC

EE

FF

GG

1 11⁄4 11⁄2 2 21⁄2 3 31⁄2 4 5 6 8 10 12 14 16 18 20 24

31⁄2 33⁄4 4 41⁄2 5 51⁄2 6 61⁄2 71⁄2 8 9 11 12 14 15 161⁄2 18 22

5 51⁄2 6 61⁄2 7 73⁄4 81⁄2 9 101⁄4 111⁄2 14 161⁄2 19 211⁄2 24 261⁄2 29 34

13⁄4 2 21⁄4 21⁄2 3 3 31⁄2 4 41⁄2 5 51⁄2 61⁄2 71⁄2 71⁄2 8 81⁄2 91⁄2 11

53⁄4 61⁄4 7 8 91⁄2 10 111⁄2 12 131⁄2 141⁄2 171⁄2 201⁄2 241⁄2 27 30 32 35 401⁄2

13⁄4 13⁄4 2 21⁄2 21⁄2 3 3 3 31⁄2 31⁄2 41⁄2 5 51⁄2 6 6 1⁄2 7 8 9

41⁄2 41⁄2 41⁄2 5 51⁄2 6 61⁄2 7 8 9 11 12 14 16 18 19 20 24

Chart 8 150 Lb. Steel Flanged Fittings

BB

AA

AA CC

AA

BB

AA

AA

CC

Chart 6 150 Lb. Steel Pipe Flanges ELBOW

AA

Nominal Pipe Size 1 11⁄4 11⁄2 2 21⁄2 3 31⁄2 4 5 6 8 10 12 14 16 18 20 24

Flange Diam.

Flange Thickness (Min.)*

Bolt Circle Diam.

Diam. of Bolt Holes

7

11⁄16

1

3⁄4

4 ⁄8 51⁄4 61⁄8 61⁄2 71⁄2 81⁄4 9 10 11 121⁄2 15 171⁄2 201⁄2 23 251⁄2 28 301⁄2 36

3⁄4 13⁄16 7⁄8

1 11⁄8 3 1 ⁄16 11⁄4 13⁄8 17⁄16 15⁄8 17⁄8 2 21⁄8 21⁄4 2 3/8 21⁄2 23⁄4

No. of Bolts

LONG RADIUS ELBOW

45° ELBOW

TEE

AA

Size of Bolts

AA

45° EE EE

3 ⁄2 37⁄8 41⁄2 5 57⁄8 63⁄8 71⁄4 71⁄8 91⁄4 105⁄8 13 151⁄4 173⁄4 201⁄4 221⁄2 243⁄4 27 32

3⁄4 7⁄8

_ 7⁄8 7⁄8 7⁄8 7⁄8 7⁄8 7⁄8

1 11⁄8 11⁄4 11⁄4 13⁄8 13⁄8 13⁄8 15⁄8

4 4 4 8 8 8 8 8 8 12 12 16 16 20 20 24 24 24

AA

5

⁄8 ⁄8 3 ⁄4 5 ⁄8 3 ⁄4 3 ⁄4 3 ⁄4 3 ⁄4 3 ⁄4 3 ⁄4 7 ⁄8 1 1 1 ⁄8 11⁄8 11⁄4 11⁄4 11⁄4 11⁄2

FF

5

Chart 7 300 Lb. Steel Pipe Flanges

* A raised face of 1/16 inch is included in (a) minimum thickness of flanges, and (b) "center to contact surface" dimension of fitting. Where facings other then 1/16 inch raised face are used, the "center to contact surface" dimensions shall remain unchanged.

CROSS

GG

GG

REDUCER

ECCENTRIC REDUCER

45° LATERAL

Nominal Pipe Size

AA

BB

CC

EE

FF

GG

1 11⁄4 11⁄2 2 21⁄2 3 31⁄2 4 5 6 8 10 12 14 16 18 20 24

4 41⁄4 41⁄2 5 51⁄2 6 61⁄2 7 8 81⁄2 10 111⁄2 13 15 161⁄2 18 191⁄2 221⁄2

5 51⁄2 6 61⁄2 7 73⁄4 81⁄2 9 101⁄4 111⁄2 14 161⁄2 19 211⁄2 24 261⁄2 29 34

21⁄4 21⁄2 23⁄4 3 31⁄2 31⁄2 4 41⁄2 5 51⁄2 6 7 8 81⁄2 91⁄2 10 101⁄2 12

61⁄2 71⁄4 81⁄2 9 101⁄2 11 121⁄2 131⁄2 15 171⁄2 201⁄2 24 271⁄2 31 343⁄4 371⁄2 401⁄2 471⁄2

2 21⁄4 21⁄2 21⁄2 21⁄2 3 3 3 31⁄2 4 5 51⁄2 6 61⁄2 71⁄2 8 81⁄2 10

41⁄2 41⁄2 41⁄2 5 51⁄2 6 61⁄2 7 8 9 11 12 14 16 18 19 20 24

Chart 9 300 Lb. Steel Flanged Fittings

TECH-F

TECH-F-5 150 Lb. ANSI / Metric Flange Comparison Flange Nom. I.D.

Outside Diameter ANSI ISO JIS 150 10 lb. Bar 10 K

1.00 4.25 4.53 25

108

115

1.50 5.00 5.91 40

127

150

2.00 6.00 6.50 50

52

165

2.50 7.00 7.28 65

178

185

3.00 7.50 7.87 80

191

200

3.50 8.50 0.00 90

216

0

Bolt Circle ANSI 150 lb.

4.92 3.12 125

79

ISO 10 Bar

Thickness (Min.) JIS ANSI ISO JIS 150 10 10 K lb. Bar 10 K

3.35 3.54 0.56 0.63 85

90

14

16

5.51 3.88

4.33 4.13 0.69 0.71

140

110

98

105

17

18

6.10 4.75

4.92 4.72 0.75 0.79

155

125

121

120

19

20

6.89 5.50

5.71 5.51 0.88 0.79

175

145

140

140

22

20

7.28 6.00

6.30 5.91 0.94 0.79

185

160

152

7.68 7.00 195

178

150

24

20

0.00 6.30 0.94 0.00 0

160

24

0

4.00 9.00 8.66

8.27 7.50

7.09 6.89 0.94 0.87

100

210

180

229

220

191

175

24

22

6.00 11.00 11.22 11.02 9.50

9.45 9.45 1.00 0.94

150

240

279

285

280

241

240

25

24

8.00 13.50 13.39 12.99 11.75 11.61 11.42 1.12 0.94 200

343

340

330

298

295

290

28

24

10.00 16.00 15.55 15.75 14.25 13.78 13.98 1.19 1.02 250

406

395

400

362

350

355

30

26

12.00 19.00 17.52 17.52 17.00 15.75 15.75 1.25 1.10 300

483

445

445

432

400

400

32

28

14.00 21.00 19.88 19.29 18.75 18.11 17.52 1.38 1.18 350

533

505

490

476

460

445

35

30

16.00 23.50 22.24 22.05 21.25 20.28 20.08 1.44 1.26 400

597

565

560

540

515

510

37

32

18.00 25.00 24.21 24.41 22.75 22.24 22.24 1.56 1.38 450

635

615

620

578

565

565

40

35

20.00 27.50 26.38 26.57 25.00 24.41 24.41 1.69 1.50 500

699

670

675

635

620

620

43

38

24.00 32.00 30.71 31.30 29.50 28.54 28.74 1.88 1.65 600

813

780

795

749

725

730

48

42

30.00 38.75 0.00 38.19 36.00 0.00 35.43 2.12 0.00 750

984

0

970

914

0

900

54

0

36.00 46.00 43.90 44.09 42.75 41.34 41.34 2.38 1.34 900 1168 1115 1120 1086 1050 1050 60

34

42.00 53.00 48.43 48.62 49.50 45.67 45.67 2.62 1.34 1000 1230 1230 1235 1257 1160 1160 67

34

48.00 59.50 57.28 57.68 56.00 54.33 54.33 2.75 1.50 1200 1230 1455 1465 1422 1380 1380 70

38

0.55 14 0.63 16 0.63 16 0.71 18 0.71 18 0.71 18 0.71 18 0.87 22 0.87 22 0.94 24 0.94 24 1.02 26 1.10 28 1.18 30 1.18 30 1.26 32 1.42 36 1.50 38 1.57 40 1.73 44

10 K

ANSI 150 lb.

ISO 10 Bar

10 K

Raised Face Diameter ANSI ISO JIS 150 10 lb. Bar 10 K

-

-

0.5

-

-

2.00 2.68 2.64

-

4

4

-

4

-

-

0.5

-

4

4

-

Bolt Hole ANSI 150 lb.

ISO 10 Bar

Bolts Quantity JIS 10 K

0.62 0.55 0.75 16

14

19

0.62 0.71 0.75 16

18

19

0.75 0.71 0.75 19

18

19

0.75 0.71 0.75 19

18

19

0.75 0.71 0.75 19

18

19

0.75 0.00 0.75 19

0

19

0.75 0.71 0.75 19

18

19

0.88 0.87 0.91 22

22

23

0.88 0.87 0.91 22

22

23

1.00 0.87 0.98 25

22

25

1.00 0.87 0.98 25

22

25

1.12 0.87 0.98 28

22

25

1.12 1.02 1.06 28

26

27

1.25 1.02 1.06 32

26

27

1.25 1.02 1.06 32

26

27

1.38 1.16 1.30 35

29.5

33

1.38 0.00 1.30 35

0

33

1.62 1.28 1.30 41

32.5

33

1.62 1.40 1.54 41

35.5

39

1.62 1.54 1.54 41

39

39

ANSI 150 lb.

ISO 10 Bar

4

JIS

Bolt Size JIS

M12 M16 -

-

M16 M16

4

-

-

0.62

-

4

4

-

4

-

-

0.62

-

8

4

-

4

-

-

0.62

-

8

8

-

8

-

-

0.62

-

-

-

-

8

-

-

M16

8

-

-

0.62

-

-

-

8

8

-

8

-

-

0.75

-

8

8

-

8

-

-

0.75

-

8

12

-

12

-

-

0.88

-

12

12

-

12

-

-

0.88

-

12

16

-

12

-

-

1.00

-

16

16

-

16

-

-

1.00

-

16

16

-

16

-

-

1.12

-

20

20

-

20

-

-

1.12

-

20

20

-

20

-

-

1.25

-

20

24

-

28

-

-

1.25

-

-

M16 M16 -

-

M16 M16 -

-

M16 M16

M16 M16 -

-

M20 M20 -

-

M20 M20 -

-

M20 M22 -

-

M20 M22 -

-

M20 M22 -

-

M24 M24 -

-

M24 M24 -

-

M24 M24 -

-

M27 M30 -

-

-

0

24

-

-

M30

32

-

-

1.50

-

-

-

28

28

-

36

-

-

1.50

-

28

28

-

44

-

-

1.50

-

32

32

-

51

68

67

2.88 3.46 3.19 73

88

81

3.62 4.02 3.78 92

102

96

4.12 4.80 4.57 105

122

116

5.00 5.24 4.96 127

133

126

5.50 0.00 5.35 140

0

136

6.19 6.22 5.94 157

158

151

8.50 8.35 8.35 216

212

212

10.62 10.55 10.31 270

268

262

12.75 12.60 12.76 324

320

324

15.00 14.57 14.49 381

370

368

16.25 16.93 16.26 413

430

413

18.50 18.98 18.70 470

482

475

21.00 20.94 20.87 533

532

530

23.00 23.03 23.03 584

585

585

27.25 26.97 27.17 692 685.0 690 33.75 0.00 33.66 857

0

855

40.25 39.57 39.57

M30 M30 1022 1005.0 1005 -

-

47.00 43.70 43.70

M33 M36 1194 1110.0 1110 -

-

53.50 52.36 52.17

M36 M36 1359 1330 1325

TECH-F

TECH-F-6 300 Lb. ANSI / Metric Flange Comparison Flange Nom. I.D.

Outside Diameter ANSI ISO JIS 300 16 lb. Bar 16 K

Bolt Circle ANSI 300 lb.

ISO 16 Bar

Thickness (Min.) JIS ANSI ISO JIS 300 16 16 K lb. Bar 16 K

1.00 4.88 4.53 4.92 3.50 3.35 3.54 0.69 0.63 25

124

115

125

90

85

90

17

16

1.50 6.12 5.91 5.51 4.50 4.33 4.13 0.81 0.71 40

156

150

140

114

110

105

21

18

2.00 6.50 6.50 6.10 5.00 4.92 4.72 0.88 0.79 50

165

165

155 127.0 125

120

22

20

2.50 7.50 7.28 6.89 5.88 5.71 5.51 1.00 0.79 65

191

185

175

149

145

140

25

20

3.00 8.25 7.87 7.87 6.62 6.30 6.30 1.12 0.79 80

210

200

200

169

160

160

29

20

3.50 9.00 0.00 8.27 7.25 0.00 6.69 1.19 0.00 90

229

-

210

184

-

170

30

-

4.00 10.00 8.66 8.86 7.88 7.09 7.28 1.25 0.87 100

254

220

225

200

180

185

32

22

6.00 12.50 11.22 12.01 10.62 9.54 10.24 1.44 0.94 150

381

285

305

270

240

260

37

24

8.00 15.00 13.39 13.78 13.00 11.61 12.01 1.62 1.02 200

381

340

350

330

295

305

41

26

10.00 17.50 15.94 16.93 15.25 13.98 14.96 1.88 1.10 250

445

405

430

387

355

380

48

28

12.00 20.50 18.11 18.90 17.75 16.14 16.93 2.00 1.26 300

521

460

480

451

410

430

51

32

14.00 23.00 20.47 21.26 20.25 18.50 18.90 2.12 1.38 350

584

520

540

514

470

480

54

35

16.00 25.50 22.83 23.82 22.50 20.67 21.26 2.25 1.50 400

648

580

605

572

525

540

57

38

18.00 28.00 25.20 26.57 24.75 23.03 23.82 2.83 1.65 450

711

640

675

629

585

605

60

42

20.00 30.50 28.15 28.74 27.00 25.59 25.98 2.50 1.81 500

775

715

730

686

650

660

64

46

24.00 36.00 33.07 33.27 32.00 30.31 30.31 2.75 2.05 600

914

840

845

813

770

770

70

52

30.00 43.00 0.00 40.16 39.25 0.00 36.81 3.00 0.00 750 1092

0

1020 997

0

935

76

0

36.00 50.00 44.29 46.65 46.00 41.34 42.91 3.38 2.99 900 1270 1125 1185 1168 1050 1090 86

76

42.00 57.00 49.41 51.97 52.75 46.06 47.64 3.69 3.31 1000 1448 1255 1320 1340 1170 1210 94

84

48.00 65.00 58.46 60.24 60.75 54.72 55.91 4.00 3.86 1200 1651 1485 1530 1543 1390 1420 102

TECH-F

98

0.55 14 0.63 16 0.63 16 0.71 18 0.79 20 0.79 20 0.87 22 0.94 24 1.02 26 1.10 28 1.18 30 1.34 34 1.50 38 1.57 40 1.65 42 1.81 46 2.05 52 2.28 58 2.44 62 2.76 70

Bolt Hole ANSI 300 lb.

ISO 16 Bar

Bolts Quantity JIS 16 K

0.75 0.55 0.75 19

14

19

0.88 0.71 0.75 22

18

19

0.75 0.71 0.75 19

18

19

0.88 0.71 0.75 22

18

19

0.88 0.71 0.91 22

18

23

0.88 0.00 0.91 22

-

23

0.88 0.71 0.91 22

18

23

0.88 0.87 0.98 22

22

25

1.00 0.87 0.98 25

22

25

1.12 1.02 1.06 28

26

27

1.25 1.02 1.06 32

26

27

1.25 1.02 1.30 32

26

33

1.38 1.16 1.30 35

29.5

33

1.38 1.16 1.30 35

29.5

33

1.38 1.28 1.30 35

32.5

33

1.62 1.40 1.54 41

35.5

39

1.88 0.00 1.65 48

0

42

2.12 1.54 1.89 54

39

48

2.12 1.65 2.20 54

42

56

2.12 1.89 2.20 54

48

56

ANSI 300 lb.

ISO 16 Bar

4

JIS

Bolt Size

16 K

ANSI 300 lb.

ISO 16 Bar

16 K

-

-

0.62

-

-

-

4

4

-

4

-

-

0.75

-

4

4

-

8

-

-

0.62

-

4

8

-

8

-

-

0.75

-

8

8

-

8

-

-

0.75

M12 M16 -

-

-

-

8

-

-

-

0.75

-

-

-

8

-

-

8

-

-

0.75

-

8

-

-

-

0.75

-

8

12

-

12

-

-

0.88

-

12

12

-

16

-

-

1.00

-

12

12

-

16

-

-

1.12

-

12

16

-

20

-

-

1.12

-

16

16

-

20

-

-

1.25

-

16

16

-

24

-

-

1.25

-

20

20

-

24

-

-

1.25

-

-

-

-

-

-

-

-

20

24

-

28

-

-

1.75

-

-

0

24

-

0

-

-

2.00

-

-

28

28

-

36

-

-

2.00

-

132 5.71

-

-

-

-

-

0

6.19 6.22

145 6.30

158

160

8.50 8.35

9.06

212

230

10.62 10.55 10.83 268

275

12.75 12.60 13.58 320

345

15.00 14.57 15.55 370

395

16.25 16.93 17.32 430

440

18.50 18.98 19.49 482

495

21.00 20.94 22.05 532

560

23.00 23.03 24.21 585

615

27.25 26.97 28.35

M33 M36 692 685.0 720

-

32

133

M30 M30 584

32

32

-

96 4.57

5.50 0.00

M27 M30 533

-

-

-

3.78

5.20

M27 M30 470

-

-

-

81

5.00 5.24

M24 M30 413

1.50

2.00

-

67 3.19

116

M20 140 -

102

2.64

122

M24 M24 381

-

-

92

M24 M24 324

20

28

-

88

4.12 4.80

M20 M22 270

-

-

-

68

3.62 4.02

M20 M22 216

20

28

73

M16 M20 157

-

-

51

2.88 3.46

M16 M20 127

24

40

-

2.00 2.68

M16 M16 105

8

8

-

M16 M16

-

-

-

M16 M16

8

12

JIS

Raised Face Diameter ANSI ISO JIS 300 16 lb. Bar 16 K

-

33.75 0.00 34.65

M39 857 -

0

880

40.25 39.57 40.55

M36 M45 1022 1005.0 1030 -

-

47.00 43.70 44.88

M39 M52 1194 1110.0 1140 -

-

58.44 52.36 53.15

M45 M52 1484 1330 1350

TECH-F-7 Weights and Dimensions of Steel & Wrought Iron Pipe Recommended for Use as Permanent Well Casings Reprinted from American Water Works Association Standard A100-66 by permission of the Association. Copyrighted 1966 by the American Water Works Association, Inc., 2 Park Avenue, New Yok, NY 10016. Steel Pipe, Black or Galvanized Diameter - In.

Size In.

External

Internal

Thickness In.

6 8 8 8 10 10 10 12 12 14 14 16 16 18 18 20 20 22 22 22 24 24 24 26 26 28 28 30 30 32 32 34 34 36 36

6.625 8.625 8.625 8.625 10.750 10.750 10.750 12.750 12.750 14.000 14.000 16.000 16.000 18.000 18.000 20.000 20.000 22.000 22.000 22.000 24.000 24.000 24.000 26.000 26.000 28.000 28.000 30.000 30.000 32.000 32.000 34.000 34.000 36.000 36.000

6.065 8.249 8.071 7.981 10.192 10.136 10.020 12.090 12.000 13.500 13.250 15.376 15.250 17.376 17.250 19.376 19.250 21.376 21.250 21.000 23.376 23.250 23.000 25.376 25.000 27.376 27.000 29.376 29.000 31.376 31.000 33.376 33.000 35.376 35.000

0.280 0.188 0.277 0.322 0.279 0.307 0.365# 0.330 0.375# 0.250 0.375# 0.312 0.375# 0.312 0.375# 0.312 0.375# 0.312 0.375 0.500 0.312 0.375 0.500# 0.312 0.500# 0.312 0.500# 0.312 0.500# 0.312 0.500# 0.312 0.500# 0.312 0.500#

Weight Per Foot - Lb 1 Plain Ends With Threads (Calculated) and Couplings (Nominal)2

18.97 16.90 24.70 28.55 31.20 34.24 40.48 43.77 49.56 36.71 54.57 52.36 62.58 59.03 70.59 65.71 78.60 72.38 86.61 114.81 79.06 94.62 125.49 85.73 136.17 92.41 146.85 99.08 157.53 105.76 168.21 112.43 178.89 119.11 189.57

19.18 17.80 25.55 29.35 32.75 35.75 41.85 45.45 51.15 57.00 65.30 73.00 81.00

#Thickness indicated is believed to be best practice. If soil and water conditions are unusually favorable, lighter pipe may be used if permitted in the purchaser's specifications. 1Manufacturing

weight tolerance is 10 per cent over and 3,5 per cent under nominal weight for pipe 6-20 in. in size and +/- per cent of nominal weight for

larger sizes. 2 Nominal

weights of pipe with threads and couplings (based on lengths of 20 ft. including coupling) are shown for purposes of specification. Thread data are contained in the various standards covering sizes which can be purchased with threads. Wrought-Iron Pipe, Black or Galvanized Diameter - In.

Size In.

External

Internal

Thickness In.

6 8 10 12 14 16 18 20 20 22 22 24 24 26 26 28 28 30 30

6.625 8.625 10.750 12.750 14.000 16.000 18.000 20.000 20.000 22.000 22.000 24.000 24.000 26.000 26.000 28.000 28.000 30.000 30.000

6.053 7.967 10.005 11.985 13.234 15.324 17.165 19.125 19.000 21.125 21.000 23.125 23.000 25.125 25.000 27.125 27.000 29.125 29.000

0.286 0.329 0.372 0.383 0.383 0.383 0.417 0.438 0.500* 0.438 0.500* 0.438 0.500* 0.438 0.500* 0.438 0.500* 0.438 0.500*

Weight Per Foot - Lb 1 Plain Ends With Threads (Calculated) and Couplings (Nominal)2

18.97 28.55 40.48 49.56 54.56 62.58 76.84 89.63 102.10 98.77 112.57 107.96 123.04 117.12 133.51 126.27 143.99 135.42 154.46

19.45 29.35 41.85 51.15 57.00 65.30 81.20 94.38 106.62

1Manufacturing

weight tolerance is 10 per cent over and 3.5 per cent under nominal weight for pipe ~20 in. in size and +10 per cent of nominal weight for larger sizes.

2Based

on length of 20 ft. including coupling. Threaded pipe has 8 threads per inch.

*Thickness indicated is believed to be best practice. If soil and water conditions are unusually favorable tighter pipe may be used if permitted in the purchaser's specifications. Note: Welded joints advocated for pipe larger than 20 in. in diameter; also for smaller diameter pipe, where applicable, to obtain clearance and maintain uniform grout thickness.

TECH-F

TECH-F-8 Capacities of Tanks of Various Dimensions Diam.

Gals.

Area Sq. Ft.

Diam.

Gals.

Area Sq. Ft.

Diam.

Gals.

Area Sq. Ft.

Diam.

Gals.

Area Sq. Ft.

1' 1' 1” 1' 2" 1' 3" 1' 4" 1' 5" 1' 6" 1' 7" 1' 8" 1' 9" 1' 10" 1' 11" 2' 2' 1" 2' 2" 2' 3" 2' 4" 2' 5" 2' 6" 2' 7" 2' 8" 2' 9" 2' 10" 2' 11" 3' 3' 1" 3' 2" 3' 3" 3' 4" 3' 5" 3' 6" 3' 7" 3' 8" 3' 9" 3' 10" 3' 11" 4' 4' 1"

5.87 6.89 8.00 9.18 10.44 11.79 13.22 14.73 16.32 17.99 19.75 21.58 23.50 25.50 27.58 29.74 31.99 34.31 36.72 39.21 41.78 44.43 47.16 49.98 52.88 55.86 58.92 62.06 65.28 68.58 71.97 75.44 78.99 82.62 86.33 90.13 94.00 97.96

.785 .922 1.069 1.277 1.396 1.576 1.767 1.969 2.182 2.405 2.640 2.885 3.142 3.409 3.687 3.976 4.276 4.587 4.909 5.241 5.585 5.940 6.305 6.681 7.069 7.467 7.876 8.296 8.727 9.168 9.621 10.085 10.559 11.045 11.541 12.048 12.566 13.095

4' 2” 4' 3" 4' 4" 4' 5" 4' 6" 4' 7" 4' 8" 4' 9" 4' 10" 4' 11" 5' 5' 1" 5' 2" 5' 3" 5' 4" 5' 5" 5' 6" 5' 7" 5' 8" 5' 9" 5' 10" 5' 11" 6" 6' 3" 6' 6" 6' 9" 7' 7' 3" 7' 6" 7' 9" 8' 8' 3" 8' 6" 8' 9" 9" 9' 3" 9' 6" 9' 9"

102.00 106.12 110.32 114.61 118.97 123.42 127.95 132.56 137.25 142.02 146.91 151.81 156.83 161.94 167.11 172.38 177.71 183.14 188.66 194.25 199.92 205.67 211.51 229.50 248.23 267.69 287.88 308.81 330.48 352.88 376.01 399.80 424.48 449.82 475.89 502.70 530.24 558.51

13.635 14.186 14.748 15.321 15.90 16.50 17.10 17.72 18.35 18.99 19.64 20.30 20.97 21.65 22.34 23.04 23.76 24.48 25.22 25.97 26.73 27.49 28.27 30.68 35.18 35.78 38.48 41.28 44.18 47.17 50.27 53.46 56.75 60.13 63.62 67.20 70.88 74.66

10' 10' 3" 10' 6" 10' 9" 11' 11' 3" 11' 6" 11' 9" 12' 12' 3" 12' 6" 12' 9" 13' 13' 3" 13' 6" 13' 9" 14' 14' 3" 14 ‘6" 14' 9" 15' 15' 3" 15' 6" 15' 9" 16' 16' 3" 16' 6" 16' 9" 19' 19' 3" 19' 6" 19' 9" 20' 20' 3" 20' 6" 20' 9" 21' 21' 3"

587.52 617.26 640.74 678.95 710.90 743.58 776.99 811.14 846.03 881.65 918.00 955.09 992.91 1031.50 1070.80 1110.80 1151.50 1193.00 1235.30 1278.20 1321.90 1366.40 1411.50 1457.40 1504.10 1551.40 1599.50 1648.40 2120.90 2177.10 2234.00 2291.70 2350.10 2409.20 2469.10 2529.60 2591.00 2653.00

78.54 82.52 86.59 90.76 95.03 99.40 103.87 108.43 113.10 117.86 122.72 127.68 132.73 137.89 142.14 148.49 153.94 159.48 165.13 170.87 176.71 182.65 188.69 194.83 201.06 207.39 213.82 220.35 283.53 291.04 298.65 306.35 314.16 322.06 330.06 338.16 346.36 346.36

21' 6” 21' 9" 22' 22' 3' 22' 6' 22' 9" 23' 23' 3" 23' 6" 23' 9" 24' 24' 3" 24' 6" 24' 9" 25' 25' 3" 25' 6" 25' 9" 26' 26' 3" 26' 6" 26' 9" 27' 27' 3" 27' 6" 27' 9" 28' 28' 3" 28' 6" 28' 9" 29' 29' 3" 29' 6" 29' 9" 30' 30' 3" 30' 6" 30' 9"

2715.80 2779.30 2843.60 2908.60 2974.30 3040.80 3108.00 3175.90 3244.60 3314.00 3384.10 3455.00 3526.60 3598.90 3672.00 3745.80 3820.30 3895.60 3971.60 4048.40 4125. 90 4204.10 4283.00 4362.70 4443.10 4524.30 4606.20 4688.80 4772.10 4856.20 4941.00 5026.60 5112.90 5199.90 5287.70 5376.20 5465.40 5555.40

363.05 371.54 380.13 388.82 397.61 406.49 415.48 424.56 433.74 443.01 452.39 461.86 471.44 481.11 490.87 500.74 510.71 527.77 530.93 541.19 551.55 562.00 572.66 583.21 593.96 604.81 615.75 626.80 637.94 649.18 660.52 671.96 683.49 695.13 706.86 718.69 730.62 742.64

To find the capacity of tanks greater than shown above, find a tank of one-half the size desired, and multiply its capacity by four, or find one one-third the size desired and multiply its capacity by 9. Chart 10 Capacity of Round Tanks (per foot of depth)

Dimensions in Feet 4 5 6 7 8 9 10 11 12

X X X X X X X X X

4 5 6 7 8 9 10 11 12

1'

4'

119.68 187.00 269.28 366.52 478.72 605.88 748.08 905.08 1077.12

479. 748. 1077. 1466. 1915. 2424. 2992. 3620. 4308.

Contents in Gallons for Depth in Feet of: 5' 6' 8' 10' 598. 935. 1346. 1833. 2394. 3029. 3740. 4525. 5386

718. 1202. 1616. 2199. 2872. 3635. 4488. 5430. 6463.

957. 1516. 2154. 2922. 3830. 4847. 5984. 7241. 8617.

1197. 1870 2693. 3665. 4787. 6059. 7480. 9051. 10771

To find the capacity of a depth not given, multiply the capacity for one foot by the required depth in feet. Chart 11 Capacity of Square Tanks

TECH-F

11'

12'

1316. 2057. 2968 4032. 5266. 6665. 8228. 9956. 11848.

1436. 2244 3231. 4398 5745. 7272. 8976. 10861. 12925.

Capacities of Tanks of Various Dimensions Gallons Per Foot of Length When Tank is Filled 3/10 2/5 1/2 3/5 7/10

Diameter

1/10

1/5

1 ft. 2 ft 3 ft. 4 ft. 5 ft. 6 ft. 7 ft 8 ft. 9 ft. 10 ft. 11 ft. 12 ft. 13 ft. 14 ft. 15 ft.

.3 1.2 2.7 4.9 7.6 11.0 15.0 19.0 25.0 30.0 37.0 44.0 51.0 60.0 68.0

.8 3.3 7.5 13.4 20.0 30.0 41.0 52.0 67.0 83.0 101.0 120.0 141.0 164.0 188.0

1.4 5.9 13.6 23.8 37.0 53.0 73.0 96.0 112.0 149.0 179.0 214.0 250.0 291.0 334.0

2.1 8.8 19.8 35.0 55.0 78.0 107.0 140.0 178.0 219.0 265.0 315.0 370.0 430.0 494.0

2.9 11.7 26.4 47.0 73.0 106.0 144.0 188.0 238.0 294.0 356.0 423.0 496.0 576.0 661.0

3.6 14.7 33.0 59.0 92.0 133.0 181.0 235.0 298.0 368.0 445.0 530.0 621.0 722.0 829.0

4.3 17.5 39.4 70.2 110.0 158.0 215.0 281.0 352.0 440.0 531.0 632.0 740.0 862.0 988.0

4/5

9/10

4.9 20.6 45.2 80.5 126.0 182.0 247.0 322.0 408.0 504.0 610.0 741.0 850.0 989.0 1134.0

5.5 22.2 50.1 89.0 139.0 201.0 272.0 356.0 450.0 556.0 672.0 800.0 940.0 1084.0 1253.0

Chart 12 Cylindrical Tanks Set Horizontally and Partially Filled Diam. In.

1"

1'

5'

6'

7'

8'

9'

10'

Length of Cylinder 11' 12' 13' 14'

1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17 18 19 20 21 22 23 24 25 26 27 28 29 30 32 34 36

0.01 0.03 0.05 0.08 0.12 0.17 0.22 0.28 0.34 0.41 0.49 0.57 0.67 0.77 0.87 0.98 1.10 1.23 1.36 1.50 1.65 1.80 1.96 2.12 2.30 2.48 2.67 2.86 3.06 3.48 3.93 4.41

0.04 0.16 0.37 0.65 1.02 1.47 2.00 2.61 3.31 4.08 4.94 5.88 6.90 8.00 9.18 10.4 11.8 13.2 14.7 16.3 18.0 19.8 21.6 23.5 25.5 27.6 29.7 32.0 34.3 36.7 41.8 47.2 52.9

0.20 0.80 1.84 3.26 5.10 7.34 10.0 13.0 16.5 20.4 24.6 29.4 34.6 40.0 46.0 52.0 59.0 66.0 73.6 81.6 90.0 99.0 108. 118. 128. 138. 148. 160 171. 183 209 236. 264.

0.24 0.96 2.20 3.92 6.12 8.80 12.0 15.6 19.8 24.4 29.6 35.2 41.6 48.0 55.2 62.4 70.8 79.2 88.4 98.0 108 119. 130. 141. 153. 166. 178. 192. 206. 220. 251. 283. 317.

0.28 1.12 2.56 4.58 7.14 10.3 14.0 18.2 23.1 28.4 34.6 41.0 48.6 56.0 64.4 72.8 81.6 92.4 103. 114. 126 139. 151. 165. 179. 193. 208. 224. 240. 257. 293. 330. 370.

0.32 1.28 2.92 5.24 8.16 11.8 16.0 20.8 26.4 32.6 39.4 46.8 55.2 64.0 73.6 83.2 94.4 106. 118. 130 144. 158. 173. 188. 204 221. 238. 256. 274. 294. 334. 378. 422.

0.36 1.44 3.30 5.88 9.18 13.2 18.0 23.4 29.8 36.8 44.4 52.8 62.2 72.0 82.8 93.6 106. 119. 132. 147. 162. 178. 194. 212. 230. 248. 267. 288. 309. 330. 376. 424. 476.

0.40 1.60 3.68 6.52 10.2 14.7 20.0 26.0 33.0 40.8 49.2 58.8 69.2 80.0 92.0 104. 118. 132. 147. 163. 180. 198. 216. 235. 255. 276. 297. 320. 343. 367. 418. 472. 528.

0.44 1.76 4.04 7.18 11.2 16.1 22.0 28.6 36.4 44.8 54.2 64.6 76.2 88.0 101. 114 130. 145. 162. 180. 198. 218. 238. 259. 281. 304. 326. 352. 377. 404. 460. 520. 582.

0.48 1.92 4.40 7.84 12.2 17.6 24.0 31.2 39.6 48.8 59.2 70.4 83.2 96.0 110. 125. 142. 158. 177. 196. 216. 238. 259. 282. 306. 331. 356. 384. 412. 440. 502. 566. 634.

0.52 2.08 4.76 8.50 13.3 19.1 26.0 33.8 43.0 52.8 64.2 76.2 90.2 104. 120. 135. 153. 172. 192. 212. 238. 257. 281. 306. 332. 359. 386. 416. 446. 476. 544. 614. 688.

0.56 2.24 5.12 9.16 14.3 20.6 28.0 36.4 46.2 56.8 69.2 82.0 97.2 112. 129. 146. 163. 185. 206. 229 252. 277. 302. 330. 358. 386. 416. 448. 480. 514. 586. 660. 740.

15

16'

17'

18'

20'

0.60 2.40 5.48 9.82 15.3 22.0 30.0 39.0 49.6 61.0 74.0 87.8 104. 120. 138. 156. 177. 198. 221. 245. 270. 297. 324. 353. 383. 414. 426. 480. 514. 550. 628. 708. 792.

0.64 2.56 5.84 10.5 16.3 23.6 32.0 41.6 52.8 65.2 78.8 93.6 110. 128. 147. 166. 189. 211. 235. 261. 288. 317. 346. 376. 408. 442. 476. 512. 548. 588. 668. 756. 844.

0.68 2.72 6.22 11.1 17.3 25.0 34.0 44.2 56.2 69.4 83.8 99.6 117. 136. 156. 177. 201. 224. 250. 277. 306. 337. 367. 400. 434. 470. 504. 544. 584. 624. 710. 802. 898.

0.72 0.80 2.88 3.20 6.60 7.36 11.8 13.0 18.4 20.4 26.4 29.4 36.0 40.0 46.8 52.0 60.0 66.0 73.6 81.6 88.8 98.4 106 118. 124. 138. 144. 160. 166. 184. 187. 208. 212. 236. 240. 264. 265. 294. 294. 326. 324. 360. 356. 396. 389. 432. 424 470. 460. 510. 496. 552. 534. 594. 576. 640. 618. 686. 660. 734. 752. 836. 848. 944. 952. 1056.

22'

24'

Diam. In.

0.88 3.52 8.08 14.4 22.4 32.2 44.0 57.2 72.4 89.6 104. 129. 152. 176. 202. 229. 260. 290. 324. 359. 396. 436. 476. 518. 562. 608. 652. 704. 754. 808. 920. 1040. 1164.

0.96 3.84 8.80 15.7 24.4 35.2 48.0 62.4 79.2 97.6 118. 1411 166. 192. 220. 250. 283. 317. 354. 392. 432. 476. 518. 564. 612. 662. 712. 768. 824. 880. 1004. 1132. 1268.

1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17 18 19 20 21 22 23 24 25 26 27 28 29 30 32 34 36

Chart 13 Capacities, in U.S. Gallons of Cylinders of Various Diameters and Lengths

TECH-F

Section TECH-G Motor Data TECH-G-1 Motor Enclosures The selection of a motor enclosure depends upon the ambient and surrounding conditions. The two general classifications of motor enclosures are open and totally enclosed. An open motor has ventilating openings which permit passage of external air over and around the motor windings. A totally enclosed motor is constructed to prevent the free exchange of air between the inside and outside of the frame, but not sufficiently enclosed to be termed air-tight. These two categories are further broken down by enclosure design, type of insulation, and/or cooling method. The most common of these types are listed below. Open Dripproof - An open motor in which all ventilating openings are so constructed that drops of liquid or solid particles falling on the motor at any angle from 0 to 15 degrees from vertical cannot enter the machine. This is the most common type and is designed for use in nonhazardous, relatively clean, industrial areas. Encapsulated - A dripproof motor with the stator windings completely surrounded by a protective coating. An encapsulated motor offers more resistance to moisture and/or corrosive environments than an ODP motor.

Totally Enclosed, Fan-Cooled - A enclosed motor equipped for external cooling by means of a fan integral with the motor, but external to the enclosed parts. TEFC motors are designed for use in extremely wet, dirty, or dusty areas. Explosion-Proof, Dust-Ignition-Proof - An enclosed motor whose enclosure is designed to withstand an explosion of a specified dust, gas, or vapor which may occur within the motor and to prevent the ignition of this dust, gas, or vapor surrounding the motor. A motor manufacturer should be consulted regarding the various classes and groups of explosion-proof motors available and the application of each. Motor insulation is classified according to the total allowable temperature. This is made up of a maximum ambient temperature plus a maximum temperature rise plus allowances for hot spots and service factors. Class B insulation is the standard and allows for a total temperature of 130°C. The maximum ambient is 40°C, and the temperature rise is 70°C, for ODP motors and 75°C for TEFC motors.

Design L, 60 cycles, class B insulation system, open type, 1.15 service factor.

TECH-G-2 NEMA Frame Assignments

hp 3

SINGLE-PHASE MOTORS Horizontal and Vertical open _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _

POLYPHASE SQUIRREL-CAGE MOTORS Horizontal and Vertical open type_ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _

1

⁄2 3 ⁄4 1 1 1 ⁄2 2 3 5 71⁄2 10 15 20 25 30 40 50 60 75 100 125 150 200 250

TECH-G

3600

143T 145T 145T 182T 184T 213T 215T 254T 256T 284TS 286TS 324TS 326TS 364TS 365TS 404TS 405TS 444TS 445TS*

speed, rpm 1800 1200

143T 145T 145T 182T 184T 213T 215T 254T 256T 284T 286T 324T 326T 364TS 365TS 404TS 405TS 444TS 454TS -

143T 145T 182T 184T 213T 215T 254T 256T 284T 286T 324T 326T 364T 365T 404T 405T 444T 445T -

1200

143T 145T 182T 184T 213T

143T 145T 182T 184T 213T 215T

145T 182T 184T -

Designs A and B - class B insulation system totally-enclosed fan-cooled type, 1.00 service factor, 60-cycles. hp

900

143T 145T 182T 184T 213T 215T 254T 256T 284T 286T 324T 326T 364T 365T 404T 405T 444T 445T -

speed, rpm 1800

fan cooled _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _

Designs A and B - class B insulation system, open type 1.15 service factor, 60 cycles. hp

⁄4 1 11⁄2 2 3 5 71⁄2

3600

1

*The 250 hp rating at the 3600 rpm speed has a 1.0 service factor

⁄2 3 ⁄4 1 1 1 ⁄2 2 3 5 71⁄2 10 15 20 25 30 40 50 60 75 100 125 150

3600

143T 145T 182T 184T 213T 215T 254T 256T 284TS 286TS 324TS 326TS 364TS 365TS 405TS 444TS 445TS

speed, rpm 1800 1200

143T 145T 145T 182T 184T 213T 215T 254T 256T 284T 286T 324T 326T 364TS 365TS 405TS 444TS 445TS

143T 145T 182T 184T 213T 215T 254T 256T 284T 286T 324T 326T 364T 365T 404T 405T 444T 445T -

900

143T 145T 182T 184T 213T 215T 254T 256T 284T 286T 324T 326T 364T 365T 404T 405T 444T 445T -

TECH-G-3 NEMA Frame Dimensions

Motor H.P. (Open) H.P. (Enclosed) A B C (Approx.) Frame 900 1200 1800 3600 900 1200 1800 3600 Max. Max. Open Encl. 143T 1⁄2 145T 3⁄4 182T 1 184T 11⁄2 213T 2 215T 3 254T 5 256T 71⁄2 284T 10 284TS 286T 15 286TS 324T 20 324TS 326T 25 326TS 364T 30 364TS 365T 40 365TS 404T 50 404TS 405T 60 405TS 444T 75 444TS 445T 100 445TS 447T 447TS 56 1⁄2 182 3⁄4 184 213 1-11⁄2 215 2 254U 3 256U 5 284U 71⁄2 286U 10 324U 324S 326U 15 326S 364U 20 364US 365U 25 365US 404U 30 404US 405U 40 405US 444U 50 444US 445U 60 445US

3⁄4

1 1 11⁄2 - 2 1 1 ⁄2 3 2 5 3 71⁄2 5 10 71⁄2 15 10 20 15 25

11⁄2 2-3 5 71⁄2 10 15 20 25

1⁄2 3

⁄4 1 1 1 ⁄2 2 3 5 71⁄2 10

30 20

30

25

40

30

50

15 40 20 50 25 60

40

30 60

75

50

40 75

100

60

50 100

125

75

60 125

75

150

100

75 150

200

200

250

125

3⁄4

1 11⁄2 1 11⁄2- 2 2 11⁄2 3 3 2 5 5 3 71⁄2 71⁄2 5 10 10 7 1⁄2 15 15 10 20 20 15 25 25 20 30 30 25 40 40 30 50 50 40 60 60 50 75 75 60 3⁄4

100

1 11⁄2 1⁄2 1-11⁄2 11⁄2-2 2-3 3⁄4 2 3 5 1-11⁄2 3 5 71⁄2 2 3 5 71⁄2 10 71⁄2 10 15 5 10 15 20 71⁄2 20 25 10 15 25 30 20 30 15 40 25 40 20 50 30 25 50 60 40 30 60 75 50 40 75 100 60 50 100 125 75 60 125 150

100

100

125

125

150

150

100 125

3⁄4

1 11⁄2 1-11⁄2 11⁄2-2 2-3 2 3 5 3 5 71⁄2 5 7 1⁄2 10 71⁄2 10 15 10 15 20 20 15 25 25 20

30

25

40

30 40 30 50

50

60

60

75

75

100

100

40 50 60 75

7 7 9 9 101⁄2 101⁄2 121⁄2 121⁄2 14 14 14 14 16 16 16 16 18 18 18 18 20 20 20 20 22 22 22 22 22 22 61⁄2 9 9 101⁄2 101⁄2 121⁄2 121⁄2 14 14 16 16 16 16 18 18 18 18 20 20 20 20 22 22 22 22

6 6 61⁄2 71⁄2 71⁄2 9 103⁄4 121⁄2 121⁄2 121⁄2 14 14 14 14 151⁄2 151⁄2 151⁄4 151⁄4 161⁄4 161⁄4 161⁄4 161⁄4 173⁄4 173⁄4 181⁄2 181⁄2 201⁄2 201⁄2 231⁄4 231⁄4 3 7⁄8 61⁄2 71⁄2 71⁄2 9 103⁄4 121⁄2 121⁄2 14 14 14 151⁄2 151⁄2 151⁄4 151⁄4 161⁄4 161⁄4 161⁄4 161⁄4 173⁄4 173⁄4 181⁄2 181⁄2 201⁄2 201⁄2

12 121⁄2 13 14 16 171⁄2 201⁄2 221⁄2 231⁄2 22 25 231⁄2 26 241⁄2 271⁄2 26 29 27 30 28 321⁄2 291⁄2 34 31 38 34 40 36 431⁄2 401⁄2 101⁄2 121⁄2 131⁄2 151⁄2 17 201⁄2 221⁄2 24 251⁄2 261⁄2 241⁄2 28 26 291⁄2 27 301⁄2 28 321⁄2 30 34 311⁄2 38 34 40 36

121⁄2 131⁄2 141⁄2 151⁄2 18 191⁄2 221⁄2 24 251⁄2 241⁄2 27 26 281⁄2 27 30 281⁄2 33 31 34 32 37 34 381⁄2 351⁄2 421⁄2 381⁄2 441⁄2 41 48 461⁄2 141⁄2 151⁄2 171⁄2 19 22 24 25 261⁄2 28 251⁄2 291⁄2 27 34 31 35 32 371⁄2 341⁄2 39 36 43 381⁄2 45 401⁄2

D

E

31⁄2 23⁄4 33⁄4 23⁄4 41⁄2 33⁄4 41⁄2 33⁄4 51⁄4 41⁄4 51⁄4 41⁄4 61⁄4 5 61⁄4 5 7 51⁄2 7 51⁄2 7 51⁄2 7 51⁄2 8 61⁄4 8 61⁄4 8 61⁄4 8 61⁄4 9 7 9 7 9 7 9 7 10 8 10 8 10 8 10 8 11 9 11 9 11 9 11 9 11 9 11 9 1 7 3 ⁄2 2 ⁄16 1 4 ⁄2 33⁄4 41⁄2 33⁄4 51⁄4 41⁄4 51⁄4 41⁄4 61⁄4 5 61⁄4 5 7 51⁄2 7 51⁄2 8 6 1/4 8 61⁄4 8 61⁄4 8 61⁄4 9 7 9 7 9 7 9 7 10 8 10 8 10 8 10 8 11 9 11 9 11 9 11 9

F

H

2 23⁄4 21⁄4 23⁄4 23⁄4 33⁄4 41⁄8 5 43⁄4 43⁄4 51⁄2 51⁄2 51⁄4 51⁄4 6 6 55⁄8 55⁄8 61⁄8 61⁄8 61⁄8 61⁄8 67⁄8 67⁄8 71⁄4 71⁄4 81⁄4 81⁄4 10 10 11⁄2 21⁄4 23⁄4 23⁄4 31⁄2 41⁄8 5 43⁄4 51⁄2 51⁄4 51⁄4 6 6 55⁄8 55⁄8 61⁄8 61⁄8 61⁄8 61⁄8 67⁄8 67⁄8 71⁄4 71⁄4 81⁄4 81⁄4

11 ⁄32 11⁄32 13⁄32 13⁄32 13⁄32 13⁄32 17⁄32 17

⁄32 17⁄32 17⁄32 17⁄32 17⁄32 21⁄32 21⁄32 21⁄32 21⁄32 21⁄32 21⁄32 21⁄32 21⁄32 13⁄16 13⁄16 13⁄16 13⁄16 13⁄16 13⁄16 13⁄16 13⁄16 13⁄16 13 ⁄16 11⁄32 13⁄32 13⁄32 13⁄32 13⁄32 17⁄32 17 ⁄32 17⁄32 17⁄32 21⁄32 21⁄32 21 ⁄32 21⁄32 21 32 ⁄ 21⁄32 21⁄32 21⁄32 13⁄16 13 16 ⁄ 13⁄16 13⁄16 13⁄16 13⁄16 13 16 ⁄ 13⁄16

O (Approx.) Open Encl. 67⁄8 67⁄8 91⁄8 91⁄8 103⁄4 103⁄4 125⁄8 125⁄8 14 14 14 14 16 16 16 16 18 18 18 18 20 20 20 20 223⁄8 223⁄8 223⁄8 223⁄8 223⁄8 223⁄8 67⁄8 9 9 101⁄2 101⁄2 125⁄8 125⁄8 14 14 16 16 16 16 181⁄4 181⁄4 181⁄4 181⁄4 201⁄4 201⁄4 201⁄4 201⁄4 221⁄4 221⁄4 221⁄4 221⁄4

7 7 91⁄4 91⁄4 107⁄8 107⁄8 123⁄4 123⁄4 143⁄8 143⁄8 143⁄8 143⁄8 165⁄8 165⁄8 165⁄8 165⁄8 181⁄2 181⁄2 181⁄2 181⁄2 205⁄8 205⁄8 205⁄8 205⁄8 231⁄8 231⁄8 231⁄8 231⁄8 231⁄8 231⁄8 9 9 105⁄8 105⁄8 131⁄8 131⁄8 145⁄8 145⁄8 163⁄4 163⁄4 163⁄4 163⁄4 183⁄4 183⁄4 183⁄4 183⁄4 207⁄8 207⁄8 207⁄8 207⁄8 231⁄8 231⁄8 231⁄8 231⁄8

U 7

V Keyway Min. AC

⁄8 3⁄16 x 3⁄32 2 ⁄8 3⁄16 x 3⁄32 2 11⁄8 1⁄4 x 1⁄8 21⁄2 11⁄8 1⁄4 x 1⁄8 21⁄2 13⁄8 5⁄16 x 5⁄32 31⁄8 13⁄8 5⁄16 x 5⁄32 31⁄8 15⁄8 3/8 x 3⁄16 33⁄4 15⁄8 3/8 X 3⁄16 33⁄4 17⁄8 1⁄2 x 1⁄4 43⁄8 15⁄8 3⁄8 x 3⁄16 3 17⁄8 1⁄2 x 1⁄4 43⁄8 5 3 1 ⁄8 3/8 x ⁄16 3 21⁄8 1⁄2 x 1⁄4 5 17⁄8 1⁄2 x 1⁄4 31⁄2 21⁄8 1⁄2 x 1⁄4 5 17⁄8 1⁄2 x 1⁄4 31⁄2 23⁄8 5⁄8 x 5⁄16 55⁄8 17⁄8 1⁄2 x 1⁄4 31⁄2 23⁄8 5⁄8 x 5⁄16 55⁄8 17⁄8 1⁄2 x 1⁄4 31⁄2 27⁄8 3⁄4 x 3⁄8 7 21⁄8 1⁄2 x 1⁄4 4 7 3 3 2 ⁄8 ⁄4 x ⁄8 7 1 1 1 2 ⁄8 ⁄2 x ⁄4 4 33⁄8 7⁄8 x 7⁄16 81⁄4 23⁄8 5⁄8 x 5⁄16 41⁄2 33⁄8 7⁄8 x 7⁄16 81⁄4 23⁄8 5⁄8 x 5⁄16 41⁄2 33⁄8 7⁄8 x 7⁄16 81⁄4 23⁄8 5⁄8 x 5⁄16 41⁄2 5⁄8 3⁄16 x 3⁄32 17⁄8 7⁄8 3⁄16 x 3⁄32 2 7⁄8 3⁄16 X 3⁄32 2 1 1 1 1 ⁄8 ⁄2 x ⁄8 23⁄4 11⁄8 1⁄2 x 1⁄8 23⁄4 13⁄8 5⁄16 x 5⁄32 31⁄2 13⁄8 5⁄16 x 5⁄32 31⁄2 15⁄8 3⁄8 x 3⁄16 45⁄8 15⁄8 3⁄8 X 3⁄16 45⁄8 17⁄8 1⁄2 x 1⁄4 55⁄8 15⁄8 3⁄8 X 3⁄16 3 17⁄8 1⁄2 x 1⁄4 53⁄8 15⁄8 3⁄8 X 3⁄16 3 21⁄8 1⁄2 x 1⁄4 61⁄8 17⁄8 1⁄2 x 1⁄4 3⁄2 21⁄8 1⁄2 x 1⁄4 61⁄8 17⁄8 1⁄2 x 1⁄4 3⁄2 23⁄8 5⁄8 X 5⁄16 67⁄8 21⁄8 1⁄2 x 1⁄4 4 23⁄8 5⁄8 X 5⁄16 67⁄8 1 1 1 2 ⁄8 ⁄2 x ⁄4 4 27⁄8 3⁄4 X 3⁄8 83⁄8 1 1 1 2 ⁄8 ⁄2 x ⁄4 4 27⁄8 3⁄4 X 3⁄8 83⁄8 21⁄8 1⁄2 x 1⁄4 4 7

41⁄2 41⁄2 51⁄2 51⁄2 67⁄8 67⁄8 81⁄4 81⁄4 93⁄8 8 93⁄8 8 101⁄2 9 101⁄2 9 113⁄4 95⁄8 113⁄4 95⁄8 137⁄8 107⁄8 137⁄8 107⁄8 16 121⁄4 16 121⁄4 16 121⁄4 45⁄8 5 5 61⁄2 61⁄2 8 8 95⁄8 95⁄8 107⁄8 81⁄2 107⁄8 81⁄2 121⁄4 95⁄8 121⁄4 95⁄8 133⁄4 107⁄8 133⁄4 107⁄8 161⁄8 113⁄4 161⁄8 113⁄4

Bolts Wt. (Approx.) Dia. Lg. Open Encl. 1⁄4 1⁄4 5⁄16 5⁄16 5⁄16 5⁄16 3⁄8 3⁄8 3⁄8 3 ⁄8 3⁄8 3⁄8 1⁄2 1⁄2 1⁄2 1⁄2 1⁄2 1⁄2 1⁄2 1⁄2 5⁄8 5⁄8 5⁄8 5⁄8 5 ⁄8 5⁄8 5⁄8 5⁄8 5⁄8 5⁄8 1⁄4 5⁄16 5⁄16 5⁄16 5⁄16 3

⁄8 ⁄8 3⁄8 3⁄8 1⁄2 1⁄2 1⁄2 1⁄2 12 ⁄ 1⁄2 1⁄2 1⁄2 5⁄8 58 ⁄ 5⁄8 5⁄8 5⁄8 5⁄8 58 ⁄ 5⁄8 3

40 45 65 80 120 140 200 235 295 255 340 295 440 445 435 480 605 670 665 730 830 870 930 950 1165 1050 1370 1250 1800 1800

45 50 79 95 140 160 235 270 370 340 405 395 520 500 580 560 755 740 835 820 1050 1050 1160 1150 1440 1440 1650 1615 2260 2260

1 1 1 1 11⁄4 11⁄4 11⁄2 11⁄2 11⁄2 11⁄2 13⁄4 13⁄4 13⁄4 13⁄4 13⁄4 13⁄4 13⁄4 13⁄4 21⁄4 21⁄4 21⁄4 21⁄4 21⁄4 21⁄4 21⁄4 21⁄4 21⁄4 21⁄4 1 1 60 70 1 70 80 1 105 125 1 115 140 11⁄4 180 210 11⁄4 210 245 11⁄2 280 330 11⁄2 325 365 13⁄4 380 480 13⁄4 380 480 13⁄4 430 560 13⁄4 430 560 13⁄4 525 720 13⁄4 670 710 13⁄4 580 785 13⁄4 730 780 21⁄4 725 965 21⁄4 860 1075 21⁄4 810 1110 2v 970 1165 21⁄4 985 1315 21⁄4 1175 1355 21⁄4 1135 1550 21⁄4 1340 1620

TECH-G

TECH-G-4 Synchronous and Approximate Full Load Speed of Standard A.C. Induction Motors NUMBER of POLES

60 CYCLE RPM

50 CYCLE RPM

SYNC.

F.L.

SYNC.

F.L.

2 4 6 8 10 12 14 16 18 20 22 24 26 28 30

3600 1800 1200 900 720 600 515 450 400 360 327 300 277 257 240

3500 1770 1170 870 690 575 490 430 380 340 310 285 265 245 230

3000 1500 1000 750 600 500 429 375 333 300 273 240 231 214 200

2900 1450 960 720 575 480 410 360 319 285 260 230 222 205 192

TECH-G-5 Full Load Amperes at Motor Terminals* Average Values for All Speeds and Frequencies MOTOR HP

1

⁄2 ⁄4 1 11⁄2 2 3 5 71⁄2 10 15 20 25 30 40 50 60 75 100 125 150 200 250 3

SINGLE-PHASE A-C 115 VOLTS

230 VOLTS**

9.8 13.8 16 20 24 34 56 80 100

4.9 6.9 8 10 12 17 28 40 50

THREE PHASE A-C INDUCTION TYPE SQUIRREL CAGE & WOUND ROTOR 230 460 575 VOLTS** VOLTS VOLTS

2.0 2.8 3.6 5.2 6.8 9.6 15.2 22 28 42 54 68 80 104 130 154 192 240 296 350 456 558

1.0 1.4 1.8 2.6 3.4 4.8 7.6 11 14 21 27 34 40 52 65 77 96 120 148 175 228 279

.8 1.1 1.4 2.1 2.7 3.9 6.1 9 11 17 22 27 32 41 52 62 77 96 118 140 182 223

DIRECT CURRENT 120 VOLTS

240 VOLTS

5.2 7.4 9.4 13.2 17 25 40 58 76 112 148 184 220 292 360 430 536

2.6 3.7 4.7 6.6 8.5 12.2 20 29 29 55 72 89 106 140 173 206 255 350 440 530 710

* These values for full load current are for running at speeds usual for belted motors and motors with normal torque characteristics. Motors built for especially low speeds or high torques may require more running current, in which case the nameplate current rating should be used. ** For full-load currents of 208 and 200 volt motors, increase the corresponding 230 volt motor full-load current by 10 and 15 per cent respectively.

TECH-G

TECH-G-6 Motor Terms AMPERE: a unit of intensity of electric current being produced in a conductor by the applied voltage.

SERVICE FACTOR: a safety factor in some motors which allows the motor, when necessary, to deliver greater than rated horsepower.

FREQUENCY: the number of complete cycles per second of alternating current, e.g., 60 Hertz.

SYNCHRONOUS SPEED & SLIP: the speed of an a-c motor at which the motor would operate if the rotor turned at the exact speed of the rotating magnetic field. However, in a-c induction motors, the rotor actually turns slightly slower. This difference is defined as slip and is expressed in percent of synchronous speed. Most induction motors have a slip of 1-3%.

HORSEPOWER: the rate at which work is done. It is the result of the work done (stated in foot-pounds) divided by the time involved. INERTIA: the property of physical matter to remain at rest unless acted on by some external force. Inertia usually concerns the driven load. MOTOR EFFICIENCY: a measure of how effectively the motor turns electrical energy into mechanical energy. Motor efficiency is never 100% and is normally in the neighborhood of 85%.

TORQUE: that force which tends to produce torsion or rotation. In motors, it is considered to be the amount of force produced to turn the load, it is measured in lb.-ft. VOLTAGE: a unit of electro-motive force. It is a force which, when applied to a conductor, will produce a current in the conductor.

POWER FACTOR: the ratio of the true power to the volt-amperes in an alternating current circuit or apparatus. APPROXIMATE RULES OF THUMB

At 1800 rpm, a motor develops 3 lb.- ft per hp. At 1200 rpm, a motor develops 4.5 lb-ft per hp. At 575 volts, a 3-phase motor draws 1 amp per hp.

MECHANICAL FORMULAS

Torque in lb-ft = HP x 5250 RPM Hp= Torque x RPM 5250 RPM = 120 x Frequency No. of poles

At 230 volts, a single- phase motor draws 2.5 amp per hp. At 230 volts, a single- phase motor draws 5 amp per hp. At 115 volts, a single- phase motor draws 10 amp per hp.

At 460 volts, a 3-phase motor draws 1.25 amp per hp. Average Efficiencies and Power Factors of Electric Motors Efficiency % Power Factor kW

Full Load

0.75 1.5 3 5.5 7.5 11 18.5 30 45 75

74 79 82.5 84.5 85.5 87 88.5 90 91 92

3

⁄4 Load 73 78.5 82 84.5 85.5 87 88.5 89.5 90.5 91.5

1

⁄2 Load

Full Load

69 76 80.5 83.5 84.5 85.5 87 88 89 90

0.72 0.83 0.85 0.87 0.87 0.88 0.89 0.89 0.89 0.90

3

⁄2 Load

Full Load Amps on 3ph 415V

0.53 0.69 0.73 0.75 0.76 0.77 0.79 0.80 0.80 0.81

2.0 3.2 6.0 10.5 14 20 33 52 77 126

1

⁄4 Load 0.65 0.78 0.80 0.82 0.83 0.84 0.85 0.86 0.86 0.87

Required Value

Direct Current

Single Phases

Two-Phase 4-Wire

Three Phase

HP Output

I x E x Eff 746

I x E x Eff x PF 746

I x E x 2 x Eff x Pf 746

I x E x 1.73 x Eff x PF 746

TECH-G-7 Electrical Conversion Formulae TO FIND

DIRECT CURRENT

Amperes when horsepower (input) is known

HP x 746 E x Efff kW x 1000 E

Amperes when kilowatts is known Amperes when kva is known

IxE 1000

Kilowatts Kva P.F.

I x E x Eff 746

Horespower (output) I = Amperes E = Volts HP= Horsepower

Eff= Effiency (decimal) P.F = Power Factor

ALTERNATING CURRENT Single Phase Three Phase HP x 746 E x Eff x P.F. kW x 1000 E x P.F. kva x 1000 E I x E x P.F. 1000 IxE 1000 KW Kva I x E x Eff x P.F. 746

Kva = Kilovolt- amperes kW = Kilowatts

HP x 746 1.73 x E x Eff x P.F. kW x 1000 1.73 x E x P.F. kvax 1000 1.73 x E 1.73 x I x E x P.F. 1000 1.73 x I x E 1000 KW Kva 1.73 x I x E x Eff x P.F. 746

TECH-G

TECH-G-8 Vertical Motors

VHS VERTICAL HOLLOWSHAFT Pump shaft thru motor and coupled below motor with impeller adjustment made at top of motor.

VHS VERTICAL SOLID SHAFT Pump shaft coupled to shaft extension below motor. Impeller adjustment at coupling

NOTE: The following dimensions may vary upon vendor selection and design: XC, CD, AG, AF, BV, C.

DIMENSIONS Top Shaft Dia. 3

⁄4

1 3

1 ⁄16 1

1 ⁄2 15

1 ⁄16 3

2 ⁄16

VERTICAL HOLLOWSHAFT NEMA dimensions for common top drive coupling sizes.

TECH-G

BX Bore

BZ Dia. BC

0.751

13⁄8

3

1.001

3

1

10-32

1.188

3

1 ⁄4

1

1

1.501

1

2 ⁄8

3

1

1.938

1

2 ⁄2

1⁄2

1⁄4

- 20

2.188

1

1⁄2

3⁄8

- 16

⁄8

3 ⁄4

SQ Key Size ⁄16 ⁄4 ⁄4 ⁄8

BY Tap Size 10-32

⁄4 - 20 ⁄4 - 20

NEMA SOLID SHAFT NEMA DIMENSIONS FOR COMMON SOLID SHAFT EXTENSION SIZES.

DIMENSIONS Motor Shaft Dia. AH U 7

V

H

B

C

D

Nominal Pump Shaft Keyway Diameters

⁄8

23⁄4

23⁄4

5⁄8

3⁄8

3⁄4

11⁄16

3⁄16

x 3⁄22

11⁄8

23⁄4

23⁄4

1

3

⁄8

3⁄4

15⁄16

1⁄4

x 1⁄8

15⁄8

41⁄2

41⁄4

25⁄8

3⁄8

3⁄4

11⁄4

3⁄8

x 3⁄16

7⁄8, 1, 13⁄16, 1 1⁄2

21⁄8

41⁄2

41⁄4

25⁄8

3⁄8

3⁄4

13⁄4

1⁄2

x 1⁄4

1, 13⁄16, 11⁄2, 115⁄16

25⁄8

5

5

31⁄2

3⁄8

3⁄4

21⁄4

5⁄8

x 5⁄16

23⁄16

27⁄8

7

61⁄2

5

1

⁄2

1

23⁄8

3⁄4

x 3⁄8

23⁄16, 211⁄16

31⁄8

7

7

43⁄4

3⁄4

11⁄2

25⁄8

3⁄4

x 3⁄8

23⁄16, 211⁄16, 215⁄16

7⁄8 7⁄8,

1

HEADSHAFT COUPLINGS WITH VERTICAL HOLLOWSHAFT MOTOR: Impeller adjustment made on adjusting nut above motor (under motor canopy and bolted to top drive coupling). 1. Sleeve type (lineshaft) coupling. 2. Rigid flanged coupling (Type AR). 3. No coupling-straight shaft (not recommended due to difficult Installation/disassembly of head and motor).

WITH VERTICAL SOLID SHAFT MOTOR: Impeller adjustment made on adjusting plate of coupling without removal of motor canopy. (VSS motors also provide a lesser tolerance of shaft run-out which coincides with mechanical seal recommendations). 1. Adjustable coupling (Type A). 2. Adjustable spacer coupling (Type AS-recommended for applications with mechanical seals. The mechanical seal can be removed without disengaging motor).

TECH-G

TECH-G-9 I.E.C. Motor Frames IPP44 TOTALLY ENCLOSED & FLAMEPROOF (Similar to NEMA TEFC & Explosion Proof) C M=N

O

E

U

N-W

D

F

E

H-SIZE HOLE

A

F

AC

B

DIMENSIONS I.E.C. Frames

Poles

Units

A B C Max. Max. Approx.

D80-19 E80-19 D90S24 E900S24 D90L24 E90L24 D100L28 E100L28 D112M28 E112M28 D132S38 E132S38 D132M38 E132M38 D160M42 E160M42 D160L42 E160L42 D180M48 E180M48 D180L48 E180L48 D200L55 E200L55 D225S55 E225S55 D225M60 E225M60 D250M60 E250M60 D250M65 E250M65 D280S65 E280S65 D280S75 E280S75 D280M65 E280M65 D280M75 E280M75 D315S65 E315S65 D315S80 E315S80 D315S80 E315M65 D315M80 E315M80

All

mm Inches mm Inches mm Inches mm Inches mm Inches mm Inches mm inches mm Inches mm Inches mm Inches mm Inches mm Inches mm Inches mm Inches mm Inches mm Inches mm Inches mm Inches mm Inches mm Inches mm Inches mm Inches mm Inches mm Inches

157 61⁄8 180 7 180 7 205 8 240 91⁄2 266 101⁄2 266 101⁄2 318 121⁄2 318 121⁄2 356 14 356 14 400 153⁄4 457 18 457 18 508 20 508 20 570 221⁄2 570 221⁄2 570 221⁄2 570 221⁄2 635 25 635 25 635 25 635 25

“ “ “ “ “ “ “ “ “ “ “ 2 4 to 8 2 4 to 8 2 4 to 8 2 4 to 8 2 4 to 8 2 4 to 8

TECH-G

130 51⁄8 130 51⁄8 155 61⁄8 180 7 185 71⁄4 185 71⁄4 225 83⁄4 267 101⁄2 311 121⁄4 300 113⁄4 340 133⁄8 368 141⁄2 370 141⁄2 395 151⁄2 426 163⁄4 426 163⁄4 470 181⁄2 470 181⁄2 520 201⁄2 520 201⁄2 520 201⁄2 520 201⁄2 570 221⁄2 570 221⁄2

245 10 300 10 320 121⁄2 380 15 380 15 440 171⁄2 480 19 580 23 620 241⁄2 650 251⁄2 685 27 760 30 810 32 835 33 925 361⁄2 925 361⁄2 1000 391⁄2 1000 391⁄2 1060 42 1060 42 1140 45 1140 45 1190 47 1190 47

D

E

F

H

M&N

O Approx.

80 3.15 90 3.54 90 3.54 100 3.94 112 4.41 132 5.20 132 5.20 160 6.30 160 6.30 180 7.09 180 7.09 200 7.87 225 8.86 225 8.86 250 9.84 250 9.84 280 11.02 280 11.02 280 11.02 280 11.02 315 12.41 315 12.41 315 12.41 315 12.41

63 21⁄2 70 23⁄4 70 23⁄4 80 31⁄8 95 33⁄4 108 41⁄4 108 41⁄4 127 5 127 5 140 51⁄2 140 51⁄2 159 61⁄4 178 7 178 7 203 8 203 8 229 9 229 9 229 9 229 9 254 10 254 10 254 10 254 10

50 2 50 2 63 211⁄2 70 23⁄4 70 23⁄4 70 23⁄4 89 31⁄2 105 41⁄8 127 5 121 43⁄4 140 51⁄2 153 6 143 55⁄8 156 61⁄8 175 67⁄8 175 67⁄8 184 71⁄4 184 71⁄4 210 81⁄4 210 81⁄4 203 8 203 8 229 9 229 9

10 3 ⁄8 10 3 ⁄8 10 3⁄8 12 15⁄32 12 15⁄32 12 15⁄32 12 15⁄32 15 19⁄32 15 19 ⁄32 15 19⁄32 15 19⁄32 19 3 ⁄4 19 3⁄4 19 3⁄4 24 15⁄16 24 15⁄16 24 15⁄16 24 15⁄16 24 15⁄16 24 15⁄16 28 13⁄32 28 13⁄32 28 13⁄32 28 13⁄32

140 51⁄2 156 6 3⁄16 169 611⁄16 193 75⁄8 200 77⁄8 239 93⁄8 258 101⁄8 323 123⁄4 345 135⁄8 352 137⁄8 371 145⁄8 396 151⁄2 402 157⁄8 445 171⁄2 483 19 483 19 514 201⁄4 514 201⁄4 540 211⁄4 540 211⁄4 559 22 589 231⁄4 585 23 615 241⁄4

185 71⁄4 210 81⁄4 210 81⁄4 230 9 250 10 290 111⁄2 290 111⁄2 360 14 360 14 400 153⁄4 400 153⁄4 440 171⁄2 490 191⁄4 490 191⁄4 550 215⁄8 550 215⁄8 630 243⁄4 630 243⁄4 630 243⁄4 630 243⁄4 725 281⁄2 725 281⁄2 725 281⁄2 725 281⁄2

U Nominal Tolerance 19 7890 24 9459 24 .9499 28 1.1024 28 1.1024 38 1.4961 38 1.4961 42 1.6539 42 1.6539 48 1.8898 48 1.8898 55 2.1654 55 2.1654 60 2.3622 60 2.3622 65 2.5591 65 2.5591 75 2.9528 65 2.5591 75 2.9528 65 2.5591 80 3.1945 65 2.5591 80 3.1495

j6 j6 j6 j6 j6 k6 k6 k6 k6 k6 k6 m6 m6 m6 m6 m6 m6 m6 m6 m6 m6 m6 m6 m6

N&W

AC

Weight Approx.

40 11⁄2 50 2 50 2 60 23⁄8 60 23⁄8 80 31⁄8 80 31⁄8 110 43⁄8 110 43⁄8 110 43⁄8 110 43⁄8 110 43⁄8 110 43⁄8 140 51⁄2 140 51⁄2 140 51⁄2 140 51⁄2 140 51⁄2 140 51⁄2 140 51⁄2 140 51⁄2 170 611⁄16 140 51⁄2 170 611⁄16

90 31⁄2 106 43⁄16 106 43⁄16 123 47⁄8 130 51⁄8 169 65⁄8 169 65⁄8 218 85⁄8 218 85⁄8 231 91⁄8 231 91⁄8 243 91⁄2 259 101⁄4 289 113⁄8 308 121⁄8 308 121⁄8 330 13 330 13 330 13 330 113 356 14 386 151⁄4 356 14 386 151⁄4

10 kg 20 Lbs 20 kg. 45 kg. 22 kg. 50 Lbs. 30 kg. 65 Lbs. 44 kg. 100 Lbs 65 kg. 145 Lbs 90 kg. 100 Lbs. 120 kg. 265 Lbs. 150 kg. 330 Lbs 175 kg. 385 Lbs. 190 kg. 420 Lbs. 255 kg. 560 Lbs. 290 kg. 640 Lbs 350 kg 770 Lbs. 440 kg. 970 Lbs. 440 kg. 970 Lbs. 615 kg 1355 Lbs. 615 kg. 1355 Lbs. 675 kg. 1500 Lbs. 675 kg. 1500 Lbs. 800 kg. 1760 Lbs. 800 kg. 1760 Lbs 900 kg. 1985 Lbs. 900 kg. 1985 Lbs.

I.E.C. Motor Frames (cont'd) IP23 ENCLOSED VENTILATED (Similar to NEMA Open Drip Proof)

C M=N

O U D

E

N-W

F

E H-SIZE HOLE

AC

F B

A

DIMENSIONS I.E.C. Frames

Poles

C160M48

All

C160L48

All

C180M55

All

C180L55

All

C200M60

All

C200L60

All

C225M60

2

C225M65

4 to 8

C250S65

2

C250S75

4 to 8

C250M65

2

C250M75

4 to 8

C280S65

2

C280S80

4 to 8

C280M65

2

C280M80

4 to 8

C315S70

2

C315S90

4 to 8

C315M7C

2

C315M90

4 to 8

Units

A B C Max. Max. Approx.

mm 318 inches 121⁄2 mm 318 inches 121⁄2 mm 356 inches 14 mm 356 inches 14 mm 400 inches 153⁄4 mm 400 inches 153⁄4 mm 457 inches 18 mm 457 inches 18 mm 508 inches 20 mm 508 inches 20 mm 508 inches 20 mm 508 inches 20 mm 570 inches 221⁄2 mm 570 inches 221⁄2 mm 570 inches 22 1⁄2 mm 570 inches 221⁄2 mm 635 inches 25 mm 635 inches 25 mm 635 inches 25 mm 635 inches 25

267 101⁄2 311 121⁄4 300 113⁄4 340 133⁄8 326 127⁄8 368 141⁄2 395 151⁄2 395 151⁄2 388 151⁄4 388 151⁄4 426 163⁄4 426 163⁄4 470 181⁄2 470 181⁄2 520 201⁄2 520 201⁄2 520 201⁄2 520 201⁄2 570 221⁄2 570 221⁄2

700 271⁄2 750 291⁄2 770 301⁄4 810 317⁄8 870 341⁄4 900 351⁄2 970 38 970 38 1100 431⁄4 1100 431⁄4 1140 447⁄8 1140 447⁄8 1265 493⁄4 1265 493⁄4 1315 513⁄4 1315 513⁄4 1475 58 1475 58 1525 60 1525 60

D

E

F

160 6.30 160 6.30 180 7.09 180 7.09 200 7.87 200 7.87 225 8.86 225 8.86 250 9.84 250 9.84 250 9.84 250 9.84 280 11.02 280 11.02 280 11.02 280 11.02 315 12.40 315 12.40 315 12.40 315 12.40

127 5 127 5 140 51⁄2 140 51⁄2 159 61⁄4 159 61⁄4 178 7 178 7 203 8 203 8 203 8 203 8 229 9 229 9 229 9 229 9 254 10 254 10 254 10 254 10

105 41⁄8 127 5 121 43⁄4 140 51⁄2 133 51⁄4 152 6 156 61⁄8 156 61⁄8 154 61⁄8 154 61⁄8 175 67⁄8 175 67⁄8 184 71⁄4 184 71⁄4 210 81⁄4 210 81⁄4 203 8 203 8 229 9 229 9

H

M&N

O Approx.

15

323 123⁄4 345 135⁄8 352 137⁄8 371 145⁄8 406 16 425 163⁄4 445 171⁄2 445 171⁄2 464 181⁄4 464 181⁄4 483 19 483 19 514 201⁄4 544 217⁄16 540 211⁄4 570 227⁄16 559 22 589 231⁄4 585 23 615 241⁄4

330 13 330 13 370 141⁄2 370 141⁄2 410 16 410 16 490 191⁄4 490 191⁄4 550 215⁄8 550 215⁄6 550 215⁄8 550 215⁄8 630 243⁄4 630 243⁄4 630 243⁄4 630 243⁄4 725 281⁄2 725 281⁄2 725 281⁄2 725 281⁄2

19⁄32

15 ⁄32 15 19⁄32 15 19⁄32 19 3⁄4 19 3 ⁄4 19 3⁄4 19 3⁄4 24 15⁄16 24 15⁄16 24 15⁄16 24 15⁄16 24 15⁄16 24 15⁄16 24 15⁄16 24 15⁄16 28 13⁄32 28 13⁄32 28 13⁄32 28 13⁄32 19

U Nominal Tolerance 48 1.8898 48 1.8898 55 2.1654 55 2.1654 60 2.3622 60 2.3622 60 2.3622 65 2.5591 65 2.5591 75 2.9528 65 2.5591 75 2.9528 65 2.5591 80 3.1496 65 2.5591 80 3.1496 70 2.7559 90 3.5433 70 2.7559 90 3.5433

k6 k6 m6 m6 m6 m6 m6 m6 m6 m6 m6 m6 m6 m6 m6 m6 m6 m6 m6 m6

N&W

AC

Weight Approx.

110 43⁄8 110 43⁄8 110 43⁄8 110 43⁄8 140 51⁄2 140 51⁄2 140 51⁄2 140 51⁄2 140 51⁄2 140 51⁄2 140 51⁄2 140 51⁄2 140 51⁄2 170 611⁄16 140 51⁄2 170 611⁄16 140 51⁄2 170 611⁄16 140 51⁄2 170 611⁄16

218 85⁄8 218 85⁄8 231 91⁄8 231 91⁄8 273 103⁄4 273 103⁄4 289 113⁄8 289 113⁄8 308 121⁄8 308 121⁄8 308 121⁄8 308 121⁄8 330 13 360 143⁄16 330 13 360 143⁄16 356 14 386 151⁄4 356 14 386 151⁄4

120 kg 265 Lbs. 150 kg 330 Lbs. 200 kg 440 Lbs. 210 kg 465 Lbs. 270 kg 595 Lbs. 285 kg 630 Lbs. 350 kg 770 Lbs. 350 kg 770 Lbs. 450 kg 990 Lbs. 450 kg 990 Lbs. 500 kg 1100 Lbs. 500 kg 1100 Lbs. 650 kg 1435 Lbs. 650 kg 1435 Lbs. 700 kg 1545 Lbs. 700 kg 1545 Lbs. 850 kg 1875 Lbs. 850 kg 1875 Lbs. 950 kg 2100 Lbs. 950 kg 2100 Lbs.

TECH-G

TECH-G-10 TEFC IP55 Metric IEC Motors (Conversion NEMA to Metric) HP

kW

RPM

FRAME

NEMA Equivalent Frame

1 1 1 1.5 1.5 1.5 2 2 2 3 3 3 4 4 4 5.5 5.5 5.5 7.5 7.5 7.5 10 10 10 15 15 15 20 20 20 25 25 25 30 30 30 40 40 40 50 50 50 60 60 60 75 75 75 100 100 100 125 125 125 150 150 150

.75 .75 .75 1.1 1.1 1.1 1.5 1.5 1.5 2.2 2.2 2.2 3.0 3.0 3.0 4.0 4.0 4.0 5.5 5.5 5.5 7.5 7.5 7.5 11 11 11 15 15 15 18.5 18.5 18.5 22 22 22 30 30 30 37 37 37 45 45 45 55 55 55 75 75 75 90 90 90 110 110 110

3000 1500 1000 3000 1500 1000 3000 1500 1000 3000 1500 1000 3000 1500 1000 3000 1500 1000 3000 1500 1000 3000 1500 1000 3000 1500 1000 3000 1500 1000 3000 1500 1000 3000 1500 1000 3000 1500 1000 3000 1500 1000 3000 1500 1000 3000 1500 1000 3000 1500 1000 3000 1500 1000 3000 1500 1000

80 80 90S 80 90S 90L 90S 90L 100L 90L 100L 112M 100L 100L 132S 112M 112M 132M 132S 132S 132M 132S 132M 160M 160M 160M 160L 160M 160L 180L 160L 180M 200L 180M 180L 200L 200L 200L 225M 200L 225S 250S 225M 225M 250M 250S 250S 280S 250M 250M 280M 280S 280S 315S 280M 280M 315M

56 56 143T 56 143T 145T 143T 145T 182T 145T 182T 184T 182T 182T 213T 184T 184T 215T 213T 213T 215T 213T 215T 254T 254T 254T 256T 254T 256T 284T 256T 284T 326T 284T 286T 326T 326T 326T 365T 326T 364T 404T 354T 365T 405T 404T 404T 444T 405T 405T 445T 444T 444T 504Z 445T 445T 505Z

TECH-G

Section TECH-H Conversion Factors TECH-H-1 Temperature Conversion Chart {Centigrade (Celsius)-Fahrenheit} C

F

-40 -38 -36 -34 -32

-40.0 -36.4 -32.8 -29.2 -25.6

-30 -28 -26 -24 -22

C

F

C

C

F

C

+5 6 7 8 9

+41.0 42.8 44.6 46.4 48.2

+40 41 42 43 44

+104.0 105.8 107.6 109.4 111.2

+175 180 185 190 195

+347 356 365 374 383

+350 355 360 365 370

+662 671 680 689 698

+750 800 850 900 950

+1382 1472 1562 1652 1742

-22.0 -18.4 -14.8 11.2 -7.6

10 11 12 13 14

50.0 51.8 53.6 55.4 57.2

45 46 47 48 49

113.0 114.8 116.6 118.4 120.2

200 205 210 215 220

392 401 410 419 428

375 380 385 390 395

707 716 725 734 743

1000 1050 1100 1150 1200

1832 1922 2012 2102 2192

-20 -19 -18 -17 -16

-4.0 -2.2 -0.4 +1.4 3.2

15 16 17 18 19

59.0 60.8 62.6 64.4 66.2

50 55 60 65 70

122.0 131.0 140.0 149.0 158.0

225 230 235 240 245

437 446 455 464 473

400 405 410 415 420

752 761 770 779 788

1250 1300 1350 1400 1450

2282 2372 2462 2552 2642

-15 -14 -13 -12 -11

5.0 6.8 8.6 10.4 12.2

20 21 22 23 24

68.0 69.8 71.6 73.4 75.2

75 80 85 90 95

167.0 176.0 185.0 194.0 203.0

250 255 260 265 270

482 491 500 509 518

425 430 435 440 445

797 806 815 824 833

1500 1550 1600 1650 1700

2732 2822 2912 3002 3092

-10 -9 -8 -7 -6

14.0 15.8 17.6 19.4 21.2

25 26 27 28 29

77.0 78.8 80.6 82.4 84.2

100 105 110 115 120

212.0 221.0 230.0 239.0 248.0

275 280 285 290 295

527 536 545 554 563

450 455 460 465 470

842 851 860 869 878

1750 1800 1850 1900 1950

3182 3272 3362 3452 3542

-5 -4 -3 -2 -1

23.0 24.8 26.6 28.4 30.2

30 31 32 33 34

86.0 87.8 89.6 91.4 93.2

125 130 135 140 145

257.0 266.0 275.0 284.0 293.0

300 305 310 315 320

572 581 590 599 608

475 480 485 490 495

887 896 905 914 923

2000 2050 2100 2150 2200

3632 3722 3812 3902 3992

0 +1 2 3 4

32.0 33.8 35.6 47.4 39.2

35 36 37 38 39

95.0 96.8 98.6 100.4 102.2

150 155 160 165 170

302.0 311.0 320.0 329.0 338.0

325 330 335 340 345

617 626 635 644 653

500 550 600 650 700

932 1022 1112 1202 1292

2250 2300 2350 2400 2450

4082 4172 4262 4352 4442

Degrees Celsius = (Degrees Fahrenheit - 32) x 5 9

F

C

F

F

Degrees Kelvin (K) = Degrees Celsius + 273.15 Degrees Rankine (R) = Degrees Fahrenheit + 459.69

Degrees Fahrenheit = (Degrees Celsius x 9) + 32 5

(0 degrees K or R = absolute zero)

TECH-H

TECH-H-2 A.P.I. and Baumé Gravity Tables and Weight Factors A.P.I Gravity

Baume Gravity

Specific Gravity

Lbs. Per U.S. Gal.

U.S. Gals. per Lb.

0 1 2 3 4 5

10.247 9.223 8.198 7.173 6.148 5.124

1.0760 1.0679 1.0599 1.0520 1.0443 1.0366

8.962 8.895 8.828 8.762 8.698 8.634

0.1116 0.1124 0.1133 0.1141 0.1150 0.1158

6 7 8 9 10

4.099 3.074 2.049 1.025 10.00

1.0291 1.0217 1.0143 1.0071 1.0000

8.571 8.509 8.448 8.388 8.328

0.1167 0.1175 0.1184 0.1192 0.1201

11 12 13 14 15

10.99 11.98 12.97 13.96 14.95

0.9930 0.9861 0.9792 0.9725 9.9659

8.270 8.212 8.155 8.099 8.044

0.1209 0.1218 0.1226 0.1235 0.1243

16 17 18 19 20

15.94 16.93 17.92 18.90 19.89

0.9593 0.9529 0.9465 0.9402 0.9340

7.989 7.935 7.882 7.830 7.778

0.1252 0.1260 0.1269 0.1277 0.1286

21 22 23 24 25

20.88 21.87 22.86 23.85 24.84

0.9279 0.9218 0.9159 0.9100 0.9024

7.727 7.676 7.627 7.578 7.529

0.1294 0.1303 0.1311 0.1320 0.1328

26 27 28 29 30

25.83 26.82 27.81 28.80 29.79

0.8984 0.8927 0.8871 0.8816 0.8762

7.481 7.434 7.387 7.341 7.296

0.1337 0.1345 0.1354 0.1362 0.1371

31 32 33 34 35

30.78 31.77 32.76 33.75 34.73

0.8708 0.8654 0.8602 0.8850 0.8498

7.251 7.206 7.163 7.119 7.076

0.1379 0.1388 0.1396 0.1405 0.1413

36 37 38 39 40

35.72 36.71 37.70 38.69 39.68

0.8448 0.8398 0.8348 0.8299 0.8251

7.034 6.993 6.951 6.910 6.870

0.1422 0.1430 0.1439 0.1447 0.1456

41 42 43 44 45

40.67 41.66 42.65 43.64 44.63

0.8203 0.8155 0.8109 0.8063 0.8017

6.830 6.790 6.752 6.713 6.675

0.1464 0.1473 0.1481 0.1490 0.1498

46 47 48 49 50

45.62 50.61 50.60 50.59 50.58

0.7972 0.7927 0.7883 0.7839 0.7796

6.637 6.600 6.563 6.526 6.490

0.1507 0.1515 0.1524 0.1532 0.1541

The relation of Degrees Baume or A.P.I. to Specific Gravity is expressed by the following formulas: For liquids lighter than water: Degrees Baume = 140 - 130, G Degrees A.P.I. = 141.5 - 131.5, G

G=

140 130 + Degrees Baume

141.5 G= 131.5 + Degrees A.P.I.

For liquids heavier than water: Degrees Baume = 145 - 145 , G

G=

145 145 - Degrees Baume

G = Specific Gravity = ratio of the weight of a given volume of oil at 60° Fahrenheit to the weight of the same volume of water at 60° Fahrenheit.

TECH-H

A.P.I Gravity

Baume Gravity

Specific Gravity

Lbs. Per U.S. Gal.

U.S. Gals. per Lb.

51 52 53 54 55

50.57 51.55 52.54 53.53 54.52

0.7753 0.7711 0.7669 0.7628 0.7587

6.455 6.420 6.385 6.350 6.316

0.1549 0.1558 0.1566 0.1575 0.1583

56 57 58 59 60

55.51 56.50 57.49 58.48 59.47

0.7547 0.7507 0.7467 0.7428 0.7389

6.283 6.249 6.216 6.184 6.151

0.1592 0.1600 0.1609 0.1617 0.1626

61 62 63 64 65

60.46 61.45 62.44 63.43 64.42

0.7351 0.7313 0.7275 0.7238 0.7201

6.119 6.087 6.056 6.025 5.994

0.1634 0.1643 0.1651 0.1660 0.1668

66 67 68 69 70

65.41 66.40 67.39 68.37 69.36

0.7165 0.7128 0.7093 0.7057 0.7022

5.964 5.934 5.904 5.874 5.845

0.1677 0.1685 0.1694 0.1702 0.1711

71 72 73 74 75

70.35 71.34 72.33 73.32 74.31

0.6988 0.6953 0.6919 0.6886 0.6852

5.817 5.788 5.759 5.731 5.703

0.1719 0.1728 0.1736 0.1745 0.1753

76 77 78 79 80

75.30 76.29 77.28 78.27 79.26

0.6819 0.6787 0.6754 0.6722 0.6690

5.676 5.649 5.622 5.595 5.568

0.1762 0.1770 0.1779 0.1787 0.1796

81 82 83 84 85

80.25 81.24 82.23 83.22 84.20

0.6659 0.6628 0.6597 0.6566 0.6536

5.542 5.516 5.491 5.465 5.440

0.1804 0.1813 0.1821 0.1830 0.1838

86 87 88 89 90

85.19 86.18 87.17 88.16 89.15

0.6506 0.6476 0.6446 0.6417 0.6388

5.415 5.390 5.365 5.341 5.316

0.1847 0.1855 0.1864 0.1872 0.1881

91 92 93 94 95

90.14 91.13 92.12 93.11 94.10

0.6360 0.6331 0.6303 0.6275 0.6247

5.293 5.269 5.246 5.222 5.199

0.1889 0.1898 0.1906 0.1915 0.1924

96 97 98 99 100

95.09 96.08 97.07 98.06 99.05

0.6220 0.6193 0.6166 0.6139 0.6112

5.176 5.154 5.131 5.109 5.086

0.1932 0.1940 0.1949 0.1957 0.1966

The above tables are based on the weight of 1 gallon (U.S.) of oil with a volume of 231 cubic inches at 60 degrees Fahrenheit in air at 760 m.m. pressure and 50% humidity. Assumed weight of 1 gallon of water at 60° Fahrenheit in air is 8.32828 pounds. To determine the resulting gravity by missing oils of different gravities: D = md1 - nd2 m+n D = Density or Specific Gravity of mixture m = Proportion of oil of d1 density n = Proportion of oil of d2 density d1 = Specific Gravity of m oil d2 = Specific Gravity of n oil

TECH-H-3 Approximate Conversion Table for Hardness Numbers Obtained by Different Methods* Brinell Number 10 mm. Ball 3000 Kg. Load 682 653 633 614 596 578 560 543 527 500 475 451 432 409 390 371 353 336 319 301 286 271 258 247 237 226 212 194 179 158 141 125 110 99 89

Rockwell Number C-Scale

B-Scale

61.7 60 59 58 57 56 55 54 53 52 50 48 46 44 42 40 38 36 34 32 30 28 26 24 22 20 16 12 8 2

Shore Scieroscope Number

Vickers Pyramid Number

84 81 79 78 77 75 73 72 71 69 67 64 62 58 56 54 51 49 47 44 42 41 38 37 35 34 32 29 27 24 21 18

99 98 95 92 89 83 77 70 62 55 47

737 697 674 654 636 615 596 578 561 544 513 484 458 434 412 392 372 354 336 318 302 286 272 260 248 238 222 204 188 166 141 125 110 99 89

*Compiled from various manufacturers' Tables.

TECH-H-4 Conversion Factors English measures - unless otherwise designated, are those used in the United States, and the units of weight and mass are avoirdupois units. Gallon - designates the U.S. gallon. To convert into the Imperial gallon, multiply the U.S. gallon by 0.83267. Likewise, the word ton designates a short ton, 2,000 pounds.

Multiply Acres Acres Acres Acres Acre-feet Acre-feet Acre-feet Atmospheres Atmospheres Atmospheres Atmospheres

By 43,560 4047 1.562 x 10-3 4840 43,560 325,851 1233,48 1.0332 1.01325 76.0 29.92

To Obtain Square feet Square meters Square miles Square yards Cubic feet Gallons Cubic Meters Atmospheres (metric) Bars Cms. Of mercury Inches of mercury

Properties of water- it freezes at 32°F., and is at its maximum density at 39.2° F. In the multipliers using the properties of water, calculations are based on water at 39.2° F. in a vacuum, weighing 62.427 pounds per cubic foot, or 8.345 pounds per U.S. gallon.

Multiply Atmospheres Atmospheres Atmospheres Atmospheres Atmospheres (metric) Atmospheres (metric) Bars Bars Bars Bars Bars

By 33.90 10,332 14.70 1.058 0.9678 980,665. .98692 33.456 29.530 1.0197 2088.6

To Obtain Feet of water kgs/sq. ft Lbs./ sq. inch Tons/sq. ft. Atmospheres Bars Atmospheres Feet H2O @39°F. In. Hg @ 32° F. kg/cm2 Pounds/ ft.2

TECH-H

Multiply Bars Barrels- oil Barrels- beer Barrels- whiskey Barrels/day- oil Bags or sacks-cement Board feet British Thermal Units British Thermal Units British Thermal Units British Thermal Units British Thermal Units B.T.U./min. B.T.U./min. B.T.U./min. B.T.U./min. Centares (Centiares) Centigrams Centiliters Centimeters Centimeters Centimeters Centimeters of mercury Centimeters of mercury Centimeters of mercury Centimeters of mercury Centimeters of mercury Centimeters of mercury Centimeters of mercury Centimeters/sec. Centimeters/sec. Centimeters/sec. Centimeters/sec. Centimeters/sec. Centimeters/sec. Cms./sec./sec. Centipoises Centipoises Centistokes Centistokes Cubic centimeters Cubic centimeters Cubic centimeters Cubic centimeters Cubic centimeters Cubic centimeters Cubic centimeters Cubic centimeters Cubic cm/sec. Cubic cm/sec. Cubic feet Cubic feet Cubic feet Cubic feet Cubic feet Cubic feet Cubic feet Cubic feet Cubic feet Cubic feet/min.

TECH-H

By 14.504 42 31 45 0.02917 94 144 sq. in. x 1 in. 0.2520 777.6 3.927 x 104 107.5 2.928 x 104 12.96 0.02356 0.01757 17.57 1 0.01 0.01 0.3937 0.01 10 0.01316 0.013332 0.013595 0.4461 136.0 27.85 0.1934 1.969 0.03281 0.036 0.6 0.02237 3.728 x 10-4 0.03281 0.001 0.01 0.01 0.01 3.531 x 10-5 6.102 x 10-2 10-6 1.308 x 10-6 2.642 x 10-4 9.999 x 10-4 2.113 x 10-3 1.057 x 10-3 0.0158502 0.001 0.1781 2.832 x 10-4 1728 0.02832 0.03704 7.48052 28.32 59.84 29.92 472.0

To Obtain Pounds/in.2 Gallons- oil Gallons- beer Gallons- whiskey Gallons/min.- oil Pounds/cement Cubic inches Kilogram- calories Foot- lbs. Horsepower- hrs. Kilogram- meters Kilowatt- hrs. Foot-lbs./sec. Horsepower Kilowatts Watts Square meters Grams Liters Inches Meters Millimeters Atmosphere Bars kg/cm2 Feet of water kgs/sq. meter Lbs./sq. ft. Lbs./sq. inch Feet/min. Feet/sec. Kilometers/hr. Meters/min. Miles/hr. Miles/min. Feet/sec./sec. Pascal-second Poises Sq. cm/sec. Stokes Cubic feet Cubic inches Cubic meters Cubic yards Gallons Liters Pints (liq.) Quarts (liq.) Gallons/minute Liters/sec. Barrels (42 US Gal.) Cubic cms. Cubic inches Cubic meters Cubic yards Gallons Liters Pints (liq.) Quarts (liq.) Cubic cms./sec.

Multiply Cubic feet/min. Cubic feet/min. Cubic feet/min. Cubic feet/sec. Cubic feet/sec. Cubic inches Cubic inches Cubic inches Cubic inches Cubic inches Cubic inches Cubic inches Cubic inches Cubic meters Cubic meters Cubic meters Cubic meters Cubic meters Cubic meters Cubic meters Cubic meters Cubic meters/hr. Cubic yards Cubic yards Cubic yards Cubic yards Cubic yards Cubic yards Cubic yards Cubic yards Cubic yards Cubic yards/min. Cubic yards/min. Cubic yards/min. Cubic yards/min. Decigrams Deciliters Decimeters Degrees (angle) Degrees (angle) Degrees (angle) Degrees/sec. Degrees/sec. Degrees/sec. Dekagrams Dekaliters Dekameters Drams Drams Drams Fathoms Feet Feet Feet Feet Feet Feet Feet Feet of water Feet of water

By

To Obtain

0.1247 0.4719 62.43 0.646317 448.831 16.39 5.787 x 10-4 1.639 x 10-5 2.143 x 10-5 4.329 x 10-3 1.639 x 10-2 0.03463 0.01732 106 35.31 61023 1.308 264.2 999.97 2113 1057 4.40 4.8089 764,554.86 27 46, 656 0.7646 202.0 764.5 1616 807.9 0.45 202.0 3.366 12.74 0.1 0.1 0.1 60 0.01745 3600 0.01745 0.1667 0.002778 10 10 10 27.34375 0.0625 1.771845 6 30.48 0.166667 3.0480 x 10-4 304.80 12 0.3048 1/3 0.0295 0.8826

Gallons/sec. Liters/sec. Pounds of water/min. Millions gals./day Gallons/min. Cubic centimeters Cubic feet Cubic meters Cubic yards Gallons Liters Pints (liq.) Quarts (liq.) Cubic centimeters Cubic feet Cubic inches Cubic yards Gallons Liters Pints (liq.) Quarts (liq.) Gallons/min. Barrels (42 U.S. Gal.) Cubic centimeters Cubic feet Cubic inches Cubic meters Gallons Liters Pints (liq.) Quarts (liq.) Cubic feet/sec. Gallons/min. Gallons/sec. Liters/sec. Grams Liters Meters Minutes Radians Seconds Radians/sec. Revolutions/min. Revolutions/sec. Grams Liters Meters Grains Ounces Grams Feet Centimeters Fathoms Kilometers Millimeters Inches Meters Yards Atmospheres Inches of mercury

Multiply Feet of water Feet of water Feet of water Feet/min. Feet/min. Feet/min. Feet/min. Feet/min. Feet/sec. Feet/sec. Feet/sec. Feet/sec. Feet/sec. Feet/sec. Feet/sec./sec. Feet/sec./sec. Feet/sec./sec. Foot- pounds Foot- pounds Foot- pounds Foot- pounds Foot- pounds Foot- pounds/min. Foot- pounds/min. Foot- pounds/min. Foot- pounds/min. Foot- pounds/min. Foot- pounds/sec. Foot- pounds/sec. Foot- pounds/sec. Foot- pounds/sec. G's (Accel. due to grav.) G's (Accel. due to grav.) G's (Accel. due to grav.) G's (Accel. due to grav.) Gallons Gallons Gallons Gallons Gallons Gallons Gallons Gallons Gallons-Imperial Gallons- US Gallons water Gallons per day Gallons per day Gallons per day Gallons per day Gallons per hour Gallons per hour Gallons per hour Gallons per hour Gallons per hour Gallons per hour Gallons per hour Gallons/min.

By 304.8 62.43 0.4335 0.5080 0.01667 0.01829 0.3048 0.01136 30.48 1.09726 0.5924 18.29 0.6818 0.01136 30.48 0.3048 0.0310810 1.286 x 10-3 5.050 x 10-7 3.240 x 10-4 0.1383 3.766 x 10-7 2.140 x 10-5 0.01667 3.030 x 10-5 5.393 x 10-3 2.280 x 10-5 7.704 x 10-2 1.818 x 10-3 1.941 x 10-2 1.356 x 10-3 32.174 35.3034 9.80665 21.9371 3785 0.1337 231 3.785 x 10-3 4.951 x 10-3 3.785 8 4 1.20095 0.83267 8.345 9.284 x 10-5 1.5472 x 10-6 2.6289 x 10-6 0.09284 0.1337 0.002228 3.71 x 10-5 6.309 x 10-5 .016667 2.7778 x 10-4 0.06309 34.286

To Obtain kgs./sq. meter Lbs./sq. ft. Lbs./sq. inch Centimeters/sec. Feet/sec. Kilometers/hr. Meters/min. Miles/hr. Centimeters/sec. Kilometers/hr. Knots Meters/min. Miles/hr. Miles/min. Cms./sec./sec. Meters/sec./sec. g's (gravity) British Thermal Units Horsepower-hrs. Kilogram- calories Kilogram- meters Kilowatt- hours B.T.U/sec. Foot-pounds/sec. Horsepower Gm.-calories/sec. Kilowatts B.T.U/min. Horsepower kg.-calories/min. Kilowatts Feet/sec.2 Km/hr.-sec. Meters/sec.2 Miles/hr.-sec. Cubic centimeters Cubic feet Cubic inches Cubic meters Cubic yards Liters Pints (liq.) Quarts (liq.) US Gallons Imperial Gallons Pounds of water Cubic ft./min. Cubic ft./sec. Cubic meters/min. Liters/min. Cubic ft./hr. Cubic ft./min. Cubic ft./sec. Cubic meters/min. Gallons/min. Gallons/sec. Liters/min. Barrels (42 US Gal.)/day

Multiply Gallons/min. Gallons/min. Gallons/min. Gallons/min. Gallons/min. Gallons/min. Gallons/sec. Gallons/sec. Grains (troy) Grains (troy) Grains (troy) Grains/US gal. Grains/US gal. Grains/Imp. gal. Grams Grams Grams Grams Grams Grams Grams Grams/cm. Grams/cu. cm. Grams/cu. cm. Grams/liter Grams/liter Grams/liter Grams/liter Hectares Hectares Hectograms Hectoliters Hectometers Hectowatts Horsepower Horsepower Horsepower Horsepower Horsepower Horsepower Horsepower Horsepower (boiler) Horsepower (boiler) Horsepower (boiler) Horsepower (boiler) Horsepower-hours Horsepower-hours Horsepower-hours Horsepower-hours Horsepower-hours Inches Inches Inches Inches Inches Inches of mercury Inches of mercury Inches of mercury

By

To Obtain

1.4286 0.02381 1440 2.228 x 10-3 0.06308 8.0208 60 227.12 0.06480 0.04167 2.0833 x 10-3 17.118 142.86 14.254 980.7 15.43 .001 1000 0.03527 0.03215 2.205 x 10-3 5.600 x 10-3 62.43 0.03613 58.416 8.345 0.06242 1000 2.471 1.076 x 105 100 100 100 100 42.44 33,000 550 1.014 10.547 0.7457 745.7 33, 493 9.809 9.2994 9809.5 2546 1.98 x 106 641.6 2.737 x 105 0.7457 2.540 0.083333 0.0254 25.4 0.0277778 0.03342 0.03386 13.6

Barrels (42 US Gal.)/hr. Barrels (42 USGal.)/min. Gallons/day Cubic feet/sec. Liters/sec. Cu. ft./hr. Gallons/min. Liters/min. Grams Pennyweights (troy) Ounces Parts/million Lbs./million gal. Parts/million Dynes Grains Kilograms Milligrams Ounces Ounces (troy) Pounds Pounds/ inch Pounds/cubic foot Pounds/cubic inch Grains/gal. Pounds/1000 gals. Pounds/cubic foot Parts/million Acres Square feet Grams Liters Meters Watts B.T.U./min. Foot-lbs./min. Foot-lbs./sec. Horsepower (metric) kg.-calories/min. Kilowatts Watts B.T.U./hr. Kilowatts B.T.U./sec. Watts B.T.U Foot-lbs. Kilogram-calories Kilogram-meters Kilowatt-hours Centimeters Feet Meters Millimeters Yards Atmospheres Bars Inches H2O

TECH-H

Multiply Inches of mercury Inches of mercury Inches of mercury Inches of mercury Inches of mercury Inches of mercury Inches of mercury Inches of mercury (32° F) Inches of water Inches of water Inches of water nches of water Inches of water Inches of water Joules Joules Joules Joules Joules Joules Joules Kilograms Kilograms Kilograms Kilograms Kilograms Kilograms Kilograms Kilograms Kilograms

Kilograms-cal./sec. Kilograms-cal./sec Kilograms-cal./sec Kilograms-cal./sec Kilograms/cm Kilograms/cm Kilograms/cm Kilograms/cm Kilograms-cal./min. Kilograms-cal./min Kilograms-cal./min kgs/meter kgs/sq. meter kgs/sq. meter kgs/sq. meter kgs/sq. meter kgs/sq. meter kgs/sq. millimeter Kiloliters Kilometers Kilometers Kilometers Kilometers Kilometers Kilopascal Kilometers/hr. Kilometers/hr. Kilometers/hr. Kilometers/hr. Kilometers/hr.

TECH-H

By 0.034531 3374.1 70.727 0.49116 1.133 345.3 70.73 0.491 0.002458 0.07355 25.40 0.578 5.202 0.03613 9.479 x 10-4 0.239006 0.73756 3.725 x 10-7 2.7778 x 10-7 1 2.7778 x 10-4 35.274 32.151 980,665 2.205 1.102 x 10-3 34.286 9.8421 x 10-4 0.001 103

3.968 3086 5.6145 4186.7 0.96783 0.980665 28.959 14.223 3085.9 0.09351 69.733 0.6720 9.678 x 10-5 3.281 x 10-3 2.896 x 10-3 0.2048 1.422 x 10-3 106 103 105 3281 103 0.6214 1094 .145 27.78 54.68 0.9113 .5399 16.67

To Obtain kg/cm2 Pascals Pounds/ft.2 Pounds/in.2 Feet of water kgs./sq. meter Lbs./sq. ft. Lbs./sq. inch Atmospheres Inches of mercury kgs./sq. meter Ounces/sq. inch Lbs./sq. foot Lbs./sq. inch B.T.U Calories (Thermo) Foot-lb.f. HP-hr. (US) Kilowatt-hr. Newton-m Watt-hr. Ounces (avoir) Ounces (troy) Dynes Lbs. Tons (short) Tons (assay) Tons (long) Tons (metric) Grams

B.T.U./sec. Foot-lbs./sec. Horsepower Watts Atmospheres Bars Inches Hg@ 32° F Pounds/in.2 Foot-lbs./min. Horsepower Watts Lbs./foot Atmospheres Feet of water Inches of mercury Lbs./sq. foot Lbs./sq. inch kgs./sq. meter Liters Centimeters Feet Meters Miles Yards Pounds/in.2 Centimeters/sec. Feet/min. Feet/sec. Knots Meters/min.

Multiply Kilometers/hr. Kms./hr./sec. Kms./hr./sec. Kms./hr./sec. Kilowatts Kilowatts Kilowatts Kilowatts Kilowatts Kilowatts Kilowatts Kilowatts Kilowatts Kilowatts Kilowatt-hours Kilowatt-hours Kilowatt-hours Kilowatt-hours Kilowatt-hours Liters Liters Liters Liters Liters Liters Liters Liters Liters/min. Liters/min. Lumber width (in) x Thickness (in) 12 Meters Meters Meters Meters Meters Meters Meters/min. Meters/min. Meters/min. Meters/min. Meters/min. Meters/sec. Meters/sec. Meters/sec. Meters/sec. Meters/sec. Meters/sec. Meters/sec.2 Meters/sec.2 Meters/sec.2 Meters/sec.2 Meter-kg. (force) Microns Miles Miles Miles Miles Miles/hr.

By

To Obtain

0.6214 27.78 0.9113 0.2778 56.907 4.425 x 104 737.6 1.341 1.3597 1000 3412.9 0.94827 14.34 103 3414.4 2.655 x 106 1.341 860.4 3.671 x 105 103 0.03531 61.02 10-3 1.308 x 10-3 0.2642 2.113 1.057 5.886 x 10-4 4.403 x 10-3

Miles/hr. Cms./sec./sec. Ft./sec./sec. Meters/sec./sec. B.T.U./min. Foot-lbs./min. Foot-lbs./sec. Horsepower (US) Horsepower (metric) Joules/sec. B.T.U/hr. B.T.U./sec. kg.-calories/min. Watts B.T.U Foot-lbs. Horsepower-hrs. Kilogram-calories Kilogram-meters Cubic centimeters Cubic feet Cubic inches Cubic meters Cubic yards Gallons Pints (liq.) Quarts (liq.) Cubic ft./sec. Gals./sec.

Length (ft.)

Board feet

100 3.281 39.37 10-3 103 1.094 1.667 3.281 0.05468 0.06 0.03728 196.8 3.281 3.6 0.06 2.287 0.03728 3.2808 0.101972 39.37 134.214 9.8067 10-6 1.609 x 105 5280 1.609 1760 44.70

Centimeters Feet Inches Kilometers Millimeters Yards Centimeters/sec. Feet/min. Feet/sec. Kilometers/hr. Miles/hr. Feet/min. Feet/sec. Kilometers/hr. Kilometers/min. Miles/hr. Miles/min. Feet/sec.2 G (gravity) Inches/sec.2 Miles/hr.-min. Joules Meters Centimeters Feet Kilometers Yards Centimeters/sec.

Multiply Miles/hr. Miles/hr. Miles/hr. Miles/hr. Miles/hr. Miles/min. Miles/min. Miles/min. Miles/min. Milliers Milligrams Milliliters Millimeters Millimeters Milligrams/liter Million Gals./day Miner's inches Minutes (angle) Newtons (N) Ounces Ounces Ounces Ounces Ounces Ounces Ounces Ounces (troy) Ounces (troy) Ounces (troy) Ounces (troy) Ounces (troy) Ounces (fluid) Ounces (fluid) Ounces/sq. inch Ounces/gal (US) Ounces/gal (US) Ounces/gal (US) Ounces/gal (US) Parts/million Parts/million Parts/million Pennyweights (troy) Pennyweights (troy) Pennyweights (troy) Pennyweights (troy) Pounds Pounds Pounds Pounds Pounds Pounds Pounds Pounds (troy) Pounds (troy) Pounds (troy) Pounds (troy) Pounds (troy) Pounds (troy) Pounds (troy) Pounds (troy)

By 88 1.467 1.609 0.8689 26.82 2682 88 1.609 60 103 10-3 10-3 0.1 0.03937 1 1.54723 1.5 2.909 x 10-4 .225 16 437.5 0.0625 28.3495 0.9115 2.790 x 10-5 2.835 x 10-5 480 20 0.08333 31.10348 1.09714 1.805 0.02957 0.0625 7.4892 0.25 0.46753 2.7056 x 10-4 0.0584 0.07015 8.345 24 1.55517 0.05 4.1667 x 10-3 16 256 7000 0.0005 453.5924 1.21528 14.5833 5760 240 12 373.2417 0.822857 13.1657 3.6735 x 10-4 4.1143 x 10-4

To Obtain Feet/min. Feet/sec. Kilometers/hr. Knots Meter/min. Meters/min. Feet/sec. Kilometers/min. Miles/hr. Kilograms Grams Liters Centimeters Inches Parts/million Cubic ft./sec. Cubic ft./min. Radians Pounds-force Drams Grains Pounds Grams Ounces (troy) Tons (long) Tons (metric) Grains Pennyweights (troy) Pounds (troy) Grams Ounces (avoir) Cubic inches Liters Lbs./sq. inch kg/m3 Ounces/quart Pounds/ft.3 Pounds/in.3 Grains/US gal. Grains/Imp. gal. Lbs./million gal. Grains Grams Ounces (troy) Pounds (troy) Ounces Drams Grains Tons (short) Grams Pounds (troy) Ounces (troy) Grains Pennyweights (troy) Ounces (troy) Grams Pounds (avoir.) Ounces (avoir.) Tons (long) Tons (short)

Multiply Pounds (troy) Pounds of water Pounds of water Pounds of water Pounds of water/min. Pounds/cubic foot Pounds/cubic foot Pounds/cubic foot Pounds/cubic inch Pounds/cubic inch Pounds/cubic inch Pounds/foot Pounds/inch Pounds/sq. in. Pounds/sq. in. Pounds/sq. in. Pounds/sq. in. Pounds/sq. in. Pounds/sq. foot Pounds/sq. foot Pounds/sq. foot Pounds/sq. inch Pounds/sq. inch Pounds/sq. inch Pounds/sq. inch Pounds/sq. foot Pounds/sq. foot Pounds/sq. foot Pounds/sq. foot Pounds/sq. foot Pounds/sq. foot Quadrants (angle) Quadrants (angle) Quadrants (angle) Quarts (dry) Quarts (liq.) Quintal, Argentine Quintal, Brazil Quintal, Castile, Peru Quintal, Chile Quintal, Mexico Quintal, Metric Quires Radians Radians Radians Radians/sec. Radians/sec. Radians/sec. Radians/sec./sec. Radians/sec./sec. Reams Revolutions Revolutions Revolutions Revolutions/min. Revolutions/min. Revolutions/min. Revolutions/min./min. Revolutions/min./min.

By

To Obtain

3.7324 x 10-4 0.01602 27.68 0.1198 2.670 x 10-4 0.01602 16.02 5.787 x 10-4 27.68 2.768 x 10-4 1728 1.488 1152 0.06895 5.1715 0.070307 6895 6895 0.01602 4.882 6.944 x 10-3 0.06804 2.307 2.036 703.1 4.788 x 10-4 0.035913 0.014139 4.8824 x 10-4 47.880 47.880 90 5400 1.571 67.20 57.75 101.28 129.54 101.43 101.41 101.47 220.46 25 57.30 3438 0.637 57.30 0.1592 9.549 573.0 0.1592 500 360 4 6.283 6 0.1047 0.01667 1.745 x 10-3 2.778 x 10-4

Tons (metric) Cubic feet Cubic inches Gallons Cubic ft./sec Grams/cubic cm. kgs./cubic centimeters Lbs./cubic inch Grams/cubic inch kgs./cubic meter Lbs./cubic foot kgs/meter Grams/cm. Bars Cm Hg @ 0° C kg./cm2 Newtons/m2 Pascals Feet of water kgs./sq. meter Pounds/sq. inch Atmospheres Feet of water Inches of mercury kgs./sq. meter Bars Cm Hg @ 0°C In Hg @ 32°C kg/cm2 Newtons/m2 Pascals Degrees Minutes Radians Cubic inches Cubic inches Pounds Pounds Pounds Pounds Pounds Pounds Sheets Degrees Minutes Quadrants Degrees/sec. Revolutions/sec. Revolutions/min. Revs./min./min. Revs./sec./sec. Sheets Degrees Quadrants Radians Degrees/sec. Radians/sec. Revolutions/sec. Rads./sec./sec. Rev./sec./sec.

TECH-H

Multiply

By

To Obtain

Multiply

By

To Obtain

Square yards Square yards Temp. (°C.) + 273 Temp. (° C.) +17.78 Temp. (° F.) + 460 Temp (° F.) -32 Tons (long) Tons (long) Tons (long) Tons (metric) Tons (metric) Tons (short) Tons (short) Tons (short) Tons (short) Tons (short) Tons (short) Tons (short) Tons of water/ 24 hrs. Tons of water/24 hrs Tons of water/ 24 hrs Watts Watts Watts Watts Watts Watts Watts Watts Watt- hours Watt- hours Watt- hours Watt- hours Watt- hours Watt- hours Yards Yards Yards Yards

0.8361 3.228 X 10-7 1 1.8 1 5/9 1016 2240 1.12000 103 2205 2000 32,000 907. 1843 2430.56 2430.56 29166.66 0.90718 83.333 0.16643 1.3349 0.05686 44.25 0.7376 1.341 X 10-3 0.001360 1 0.01434 10-3 3.414 2655 1.341 X 10-3 0.8604 367.1 10-3 91.44 3 36 0.9144

Square Meters Square miles Abs. Temp. (° C.) Temp. (° F.) Abs. Temp (° F.) Temp. (° C.) Kilogams Pounds Tons (short) Kilogams Pounds Pounds Ounces Kilograms Pounds (troy) Tons (long) Ounces (troy) Tons (metric) Pounds water/ hr. Gallons/ min. Cu. Ft. / hr. B.T..U/ min Foot- Lbs. / min. Foot- Lb/sec. Horsepower (U .S) Horsepower( metric) Joules/ sec Kg- calories/ min. Kilowatts B.T.U Foot- Lbs Horsepower- hrs Kilogram-calories kilogram- meters Kilowatt- hours Centimeters Feet Inches Meters

Revolutions/ sec Revolutions/ sec Revolutions/ sec Revolutions/sec/sec Revolutions/ sec/sec. Seconds (angle) Square centimeters Square centimetera Square centimeters Square centimeters Square feet Square feet Square feet Square feet Square feet Square feet 1 Sq. ft./ gal. Min

360 6.283 60 6,283 3600 4.848 X 10-6 1.076 X10-3 0.1550 104 100 2.296 X 10-5 929.0 144 0.09290 3.587 X10-4 1/9

Degrees/ sec. Radians/ sec. Revolutions/ min. Radians/sec./sec Revs. / min/ min Radians Square feet Square inches Square meters Square milimeters Acres Square centimeters Square inches Square meters Square miles Square yards

8.0208

Square inches Square inches Square inches Square kilometers Square kilometers Square kilometers Square kilometers Square kilometers Square meters Square meters Square meters Square meters Square miles Square miles Square miles Square miles Square millimeters Square milimeters Square yards Square yards

6.542 6.944 X 10-3 645.2 247.1 10.76 X 106 106 0.3861 1.196 X 106 2.471 X10-4 10.76 3.861 X 10-7 1.196 640 27.88 x 106 2.590 3.098 x 106 0.01 1.550 x 10-3 2.066x 10-4 9

Overflow rate (ft. / hr.) Square centimeters Square feet Square millimeters Acres Square feet Square meters Square miles Square yards Acres Square feet Square miles Square yards Acres Square feet Square kilometers Square yards Square centimeters Square inchea Acres Square feet

TECH-H

TECH-H-5 Qwik Convert Tables AREA inch2 x 645.16- mm2 inch2 x 6.4516 = cm2

mm2 x .00155= inch2 cm2 x 0.1550 = inch2

cm2 = square centimetre mm2 = square millimetre

N • m x 8.85 = in-lbs

N • m= Newton- metre

m3/h x 4.403 = gpm liters/ second x 15.85 = gpm

m3/h= cubic metre per hour

BENDING MOMENT (Torque) in- lbf x 0.113 = N • m ft- lbf x 1.356 = N • m CAPACITY (Volume per Unit Time) gpm x 0.2271 = m3/h gpm x 0.638 = liters per second FORCE lbf x 0.00448 = kN

kN = kilonewton

HEAD ( & NPSH) foot x 0.3048 = m

m x 3.28084 = foot

m = metre

mm x 0.003281 = feet mm 0.03937= inch m x 3.281 = foot

mm= millimetre m = metre

kg x 2.205 = pound g x 0.03527 = ounce

kg = kilogram g =gram

kW x 1.340483 = hp

kW = kilowatt

kg/cm2 x 14.233578 = psi kPa x .145= psi kPa x 0.010197=kg/cm2 Bar x 14.50377 = psi

kg/cm2 = kilogram/ square centimetre

°F = (1.8 x °C ) + 32

°C = degrees Celsius

LENGTH foot x 304.8 = mm inch x 25.4 = mm foot x 0.3048 = m MASS (Weight) ounce x 0.02853 = kg pound x 0.4536 = kg ounce x 28.35 = g POWER hp x 0.7457= kW PRESSURE psi x 0.0703= kg/cm2 psi x 6.895 = kPa kg/cm2 x 98.07 = kPa psi x 0.06895 = Bar

kPa = kiloascal

TEMPERATURE °C= 0.556 (°F –32) VOLUME ft3 x 0.02832 = m3 Gallon x 0.003785= m3 Quart x 0.9464 = L Ounce x 29.57= mL Gallon x 3.7854 = L

m3 x 35.31 = ft3 m3 x 264 .17= gallon L x 1.057 = quart

m3 = cubic metre L = litre mL = millilitre

L X 0.26418 = gallon

TECH-H

TECH-H-6 Conversion Chart–Gallons Per Minute to Barrels Per Day

GALLONS PER MINUTE

1 GPM = 34.286 BPD

BARRELS PER DAY X 1000

TECH-H-7 Decimal and Millimeter Equivalents of Fractions Inches Fractions 1

⁄64 ⁄32 3 ⁄64 1 ⁄16 5 ⁄64 3 ⁄32 7 ⁄64 1 ⁄8 9 ⁄64 5 ⁄32 11 ⁄64 3 ⁄16 13 ⁄64 7 ⁄32 15 ⁄64 1 ⁄4 17 ⁄64 9 ⁄32 19 ⁄64 5 ⁄16 21 ⁄64 11 ⁄32 23 ⁄64 3 ⁄8 25 ⁄64 13 ⁄32 27 ⁄64 7 ⁄16 29 ⁄64 15 ⁄32 31 ⁄64 1 ⁄2 1

TECH-H

Inches

Millimeters Decimals .015625 .03125 .046875 .0625 .078125 .09375 .109375 .125 .140625 .15625 .171845 .1875 .203125 .21875 .234375 .250 .265625 .28125 .296875 .3125 .328125 .34375 .359375 .375 .390625 .40625 .421875 .4375 .453125 .46875 .484375 .500

Fractions .397 .794 1.191 1.588 1.984 2.381 2.778 3.175 3.572 3.969 4.366 4.763 5.159 5.556 5.953 6.350 6.747 7.144 7.541 7.938 8.334 8.731 9.128 9.525 9.922 10.319 10.716 11.113 11.509 11.906 12.303 12.700

33

⁄64 ⁄32 35 ⁄64 9 ⁄16 37 ⁄64 19 ⁄32 39 ⁄64 5 ⁄8 41 ⁄64 21 ⁄32 43 ⁄64 11 ⁄16 45 ⁄64 22 ⁄32 47 ⁄64 3 ⁄4 49 ⁄64 25 ⁄32 51 ⁄64 13 ⁄16 53 ⁄64 27 ⁄32 55 ⁄64 7 ⁄8 57 ⁄64 29 ⁄32 59 ⁄64 15 ⁄16 61 ⁄64 31 ⁄32 63 ⁄64 1 17

Millimeters Decimals .515625 .53125 .546875 .5625 .578125 .59375 .609375 .625 .640625 .65625 .671875 .6875 .703125 .71875 .734375 .750 .765625 .78125 .796875 .8125 .828125 .84375 .859375 .875 .890625 .90625 .921875 .9375 .953125 .96875 .984375 1.000

13.097 13.494 13.891 14.288 14.684 15.081 15.487 15.875 16.272 16.669 17.066 17.463 17.859 18.256 18.653 19.050 19.447 19.844 20.241 20.638 21.034 21.431 21.828 22.225 22.622 23.019 23.416 23.813 24.209 24.606 25.003 25.400

TECH-H-8 Atmospheric Pressures and Barometer Readings at Different Altitudes* Altitude Below or Above Sea Level (Feet) -1000 -500 0 +500 +1000 1500 2000 2500 3000 3500 4000 4500 5000 5500 6000 6500 7000 7500 8000 8500 9000 9500 10,000 15,000 20,000 30,000 40,000 50,000

Barometer Reading Inches Mercury at 32° F 31.02 30.47 29.921 29.38 28.86 28.33 27.82 27.31 26.81 26.32 25.84 25.36 24.89 24.43 23.98 23.53 23.09 22.65 22.22 21.80 21.38 20.98 20.58 16.88 13.75 8.88 5.54 3.44

Atmospheric Pressure (PSI)

Equivalent Head of Water (75°) (Feet)

Boiling Point of Water °F

15.2 15.0 14.7 14.4 14.2 13.9 13.7 13.4 13.2 12.9 12.7 12.4 12.2 12.0 11.8 11.5 11.3 11.1 10.9 10.7 10.5 10.3 10.1 8.3 6.7 4.4 2.7 1.7

35.2 34.7 34.0 33.4 32.8 32.2 31.6 31.0 30.5 29.9 29.4 28.8 28.3 27.8 27.3 26.7 26.2 25.7 25.2 24.8 24.3 23.8 23.4 19.1 15.2 10.2 6.3 3.9

213.8 212.9 212.0 211.1 210.2 209.3 208.4 207.4 206.5 205.6 204.7 203.8 202.9 201.9 201.0 200.1 199.2 198.3 197.4 196.5 195.5 194.6 193.7 184 -

°C 101.0 100.5 100.0 99.5 99.0 98.5 98.0 97.4 96.9 96.4 95.9 95.4 94.9 94.4 94.4 93.9 92.9 92.4 91.9 91.4 90.8 90.3 89.8 84.4 -

*Approximate Values

TECH-H

Section TECH-I Pump Operation and Maintenance TECH-I-1 Pump Safety Tips Maintenance personnel should be aware of potential hazards to reduce the risk of accidents... Safety Apparel:

• Ensure there are no missing fasteners. • Beware of corroded or loose fasteners. Operation:

• Insulated work gloves when handling hot bearings or using bearing heater.

• Do not operate below minimum rated flow, or with suction/discharge valves closed.

• Heavy work gloves when handling parts with sharp edges, especially impellers.

• Do not open vent or drain valves, or remove plugs while system is pressurized.

• Safety glasses (with side shields) for eye protection, especially in machine shop areas. • Steel-toed shoes for foot protection when handling parts, heavy tools, etc. • Other personal protective equipment to protect against hazardous/toxic fluids.

Maintenance Safety: • Always lockout power. • Ensure pump is isolated from system and pressure is relieved before disassembling pump, removing plugs, or disconnecting piping. • Use proper lifting and supporting equipment to prevent serious injury.

Couplings Guards: • Never operate pump without a coupling guard properly installed.

• Observe proper decontamination procedures. • Know and follow company safety regulations.

Flanged Connections: • Never force piping to make a connection with a pump. • Use only fasteners of the proper size and material.

• Never apply heat to remove impeller. • Observe all cautions and warnings highlighted in pump instruction manual.

TECH-I-2 PRO Service Centers: An Economical Alternative Goulds offers an economical alternative to high maintenance costs. Goulds PRO Service Centers are experienced with reconditioning all types of pumps and rotating equipment, restoring equipment to original specifications. Users continually utilize PRO Service Centers for economical repair versus replacement, decreased downtime, reduced inventory of replacement pants and the advantage of updated engineering technology.

Benefits/Services: • Factory trained service personnel • 24-hour emergency service • Machine shop facilities • Inventory of replacement parts • Repairs to all makes and manufacture of pumps • Pickup and delivery service • Pump installation supervision • Technical advisory services • Turnkey field service capability • Vertical turbine rebowling Contact your nearest Goulds sales office for location of your nearest PRO Service Center.

TECH-I

TECH-I-3 Symptoms and Causes of Hydraulic and Mechanical Pump Failure 5

6

7

8

Pump does not deliver sufficient pressure

Pump delivers flow intermittently

Bearings run hot and/or fail on a regular basis

High rate of mechanical seal failure

Packing has short life

Pump vibrates at higher-than-normal levels

9

10 Wear of internal wetted parts is accelerated

4

Pump is drawing too much power

3

Pump does not deliver sufficient capacity

Cause Pump not primed or prime lost Suction and/or discharge valves closed or clogged Suction piping incorrect Insufficient NPSH available Excessive air entrapped in liquid Speed (RPM) too low Incorrect rotation Broken impeller or bent vanes Incorrect impeller or impeller diameter System head too high Instruments give erroneous readings Air leaks in suction line Excessive shaft misalignment Inadequate lubrication Lubricant contamination Inadequate lubricant cooling Axial thrust or radial loads higher than bearing rating Improper coupling lubrication Suction pressure too high Bearing incorrectly installed Impeller out of balance Overheating of seal faces Excessive shaft deflection Lack of seal flush at seal faces Incorrect seal installation Pump is run dry Pump run off design point Shaft/shaft sleeve worn Packing gland not properly adjusted Packing not properly installed Impeller clogged Coupling out of balance Baseplate not installed properly Pump operating speed too close to system's natural frequency Bearing failing Piping not properly anchored Pump and/or driver not secured to baseplate Specific gravity higher than specified Viscosity higher than specified Internal clearances too tight Chemicals in liquid other than specified Pump assembled incorrectly Higher solids concentration than specified

Mechanical Failure

2

Pump does not deliver liquid

Hydraulic Failure 1

TECH-I

TECH-I-4 Troubleshooting Centrifugal Pumps Problem

Probable Cause Pump not primed.

No liquid delivered.

Pump not producing rated flow or head.

Pump starts then stops pumping.

Bearings run hot.

Pump is noisy or vibrates.

Excessive leakage from stuffing box/seal chamber.

Motor requires excessive power.

TECH-I

Suction line clogged. Impeller clogged with foreign material. Wrong direction of rotation. Foot valve or suction pipe opening not submerged enough. Suction lift to high. Air leak through gasket. Air leak through stuffing box. Impeller partly clogged. Worn suction sideplate or wear rings. Insufficient suction head. Worn or broken impeller. Improperly primed pump. Air or vapor pockets in suction line. Air leak in suction line. Improper alignment. Improper lubrication. Lube cooling. Improper pump/driver alignment. Partly clogged impeller causing imbalance. Broken or bent impeller or shaft. Foundation not rigid. Worn bearings. Suction or discharge piping not anchored or properly supported. Pump is cavitating. Packing gland improperly adjusted. Stuffing box improperly packed. Worn mechanical seal parts. Overheating mechanical seal. Shaft sleeve scored. Head lower than rating. Pumps too much liquid. Liquid heavier than expected. Stuffing packing too tight. Rotating parts bind.

Remedy Reprime pump, check that pump and suction line are full of liquid. Remove obstructions. Back flush pump to clean impeller. Change rotation to concur with direction indicated by arrow on bearing housing or pump casing. Consult factory for proper depth. Use baffler to eliminate vortices. Shorten suction pipe. Replace gasket. Replace or readjust packing/mechanical seal. Back flush pump to clean impeller. Replace defective part as required. Ensure that suction line shutoff valve is fully open and line is unobstructed. Inspect and replace if necessary. Reprime pump. Rearrange piping to eliminate air pockets. Repair (plug) leak. Re-align pump and drive. Check lubricate for suitability and level. Check cooling system. Align shafts. Back-flush pump to clean impeller. Replace as required. Tighten hold down bolts of pump and motor or adjust stilts. Replace. Anchor per Hydraulic Institute Standards Manual recommendation. System problem. Tighten gland nuts. Check packing and repack box. Replace worn parts. Check lubrication and cooling lines. Remachine or replace as required. Consult factory. Install throttle valve, trim impeller diameter. Check specific gravity and viscosity. Readjust packing. Replace if worn. Check internal wearing parts for proper clearances.

TECH-I-5 Abrasive Slurries and Pump Wear THE EFFECTS OF OPERATING AT DIFFERENT ZONES ON THE PUMP CHARACTERISTIC CURVE The rate of wear is directly influenced by the system point on the characteristic curve. These condition points can be divided into four significant zones of operation (Fig. 1).

PRINCIPAL WEAR AREAS As the abrasive mixture passes through the pump, all the wetted surfaces which come in contact will be subject to varying degrees of wear. It is very important to note that the performance of a conventional centrifugal pump, which has been misapplied to a slurry service, will be significantly effected by a relatively small degree of abrasive wear. The areas most prone to wear, in order of severity, are: 1. Suction sideplate, particularly at the nozzle region. 2. Impeller, particularly at the eye vane inlets, suction side impeller shroud, and the vane tips. 3. Casing cutwater and side walls adjacent to the impeller tip. 4. Stuffing box packing and sleeve. NOTE: In the case of a conventional pump with radial wear rings on the impeller, this is where the worst wear occurs. On severely abrasive services where there are high concentrations of hard, larger, sharp particles, the suction side liner life can be increased if it is rotated periodically to equalize the effects of wear.

Fig. 1 Slurry Pump Characteristic Curve Overcapacity Zone:

The velocities within the pump are usually very high and recirculation occurs causing excessive wear. The radial hydraulic loads on the impeller increase.

Recommended The velocities within the pump are reduced (but not Operation enough to cause settlement). Recirculation is Zone: minimal and the flow in the suction nozzle should be axial (no induced vortex). The radial hydraulic loads are minimized. Reduced Capacity Zone

The velocities within the pump are low, separation and recirculation occurs, causing excessive wear. Reducing the capacity should be limited because a certain minimum velocity must be maintained to avoid settling out; with the consequence of increased wear and clogging. The hydraulic radial loads will increase and the pump efficiency will decrease.

Shut Valve Zone:

This is the point of zero flow, and pump should not be operated at this point for any length of time. Wear and tear will be rapid due to separation and recirculation, the hydraulic forces will be at their highest, and settlement and plugging will occur. The pump will rapidly heat up, which is particularly serious in rubber constructed pumps.

In hard iron pumps applied to severely abrasive service, the relative wear rates of the suction side liner, casing, and impeller are in the order of 3 to 1.5 to 1, e.g. the life of the casing is three times that of a suction side wear plate. Recognizing that due to the nature of the mixtures being pumped, the complete elimination of wear is impossible, the life of the parts can be appreciably prolonged and the cost of maintenance reduced by a good pump design and selection, e.g.: •

Construct the pump with good abrasion resistant materials.

•

Provide generous wear allowances on all parts subject to excessive wear.

•

Adopt a hydraulic design which will minimize the effects causing wear.

•

Adopt a mechanical design which is suitable for the materials of construction and has ready access to the parts for renewal.

•

Limit the head to be generated and select a low speed pump.

TECH-I

TECH-I-6 Start-Up and Shut-Off Procedure for Heated and Unheated Mag Drive Pumps (This procedure does not replace the operation instruction handbook.) A. CHECKLIST BEFORE START-UP 1.

The nominal motor power must not exceed the pump's allowed maximum capacity (compare rating plates of motor and pump).

2.

Check direction of rotation with disconnected coupling.

3.

Check alignment of coupling.

4. 5. 6.

Check ease of pump operation by hand. Attach coupling protection. Connect thermocouples, dry run protection, pressure gauges, etc.

C. SHUT-OFF 1.

Close pressure valve.

2.

Shut off motor. Allow pump to slow down smoothly.

3.

In case of external cooling, shut off coolant flow.

4.

Close suction valve.

NOTE: •

Throttling must not be done with the suction valve.

•

Never shut off the pump with the suction valve.

•

Pump must never run dry. Never run the pump against a closed pressure valve.

7.

Connect heater for heated pumps.

•

8.

Connect cooling system (if required).

•

The pump motor unit must run vibration free.

Attention: Insulation must not cover roller bearings.

•

Temperature of roller bearings must not exceed tolerated limit.

9.

B. START-UP 1.

Preheat heated pumps for a minimum of 2 hours.

2.

Open pressure valve.

3.

Open suction valve completely and fill pump.

4.

After 2-3 minutes close pressure valve.

5.

In case of external cooling, switch on coolant flow.

6.

Start motor.

7.

Subsequently open pressure valve slowly until pump reaches specified performance level.

TECH-I

TECH-I-7 Raised Face and Flat Face Flanges (Mating Combinations) Pumps of cast iron construction are furnished with 125 or 250 lb. flat face (F.F.) flanges. Since industry normally uses fabricated steel piping, the pumps are often connected to 150 or 300 lb. 1⁄16" raised face (R.F.) steel flanges. Difficulty can occur with this flange mating combination. The pump flange tends to pivot around the edge of the raised face as the flange bolts are tightened. This can cause the pump flange to break allowing leakage at the joint. (Fig. 1). A similar problem can be encountered when a bronze pump with F.F. flanges is connected to R.F. steel flanges (Fig. 2). Since the materials are not of equal strength, the bronze flange may distort, resulting in leakage. To avoid problems when attaching bronze or cast iron F.F. pump flanges to R.F. steel pipe flanges, the following steps should be taken (refer to Fig. 3). 1. Machine off the raised face on the steel pipe flange. 2. Use a full face gasket. If the pump is steel or stainless steel with F.F. flanges, no problem arises since materials of equal strength are being connected. Many customers, however, specify R.F. flanges on steel pumps for mating to R.F. companion flanges. This arrangement is technically and practically not required.

The purpose of a R.F. flange is to concentrate more pressure on a smaller gasket area and thereby increase the pressure containment capability of the joint. To create this higher gasket load, it is only necessary to have one-half of the flanged joint supplied with a raised face - not both. The following illustrations show 4" steel R.F. and F.F. mating flange combinations and the gasket loading incurred in each instance. Assuming the force (F) from the flange bolts to be 10,000 lbs. and constant in each combination, the gasket stress is: P (Stress) = Bolt Force (F) Gasket Area P1 (Fig. 4) = 10,000 lbs. = 203 psi 49.4 sq. in. P2 (Fig. 5) = 10,000 lbs = 630 psi P3 (Fig. 6) = 15.9 sq. in. It can be readily seen that the smaller gasket, used with a raised face flange, increases the pressure containment capability of a flanged joint. However, it can also be noted that there is no difference in pressure capability between R.F.-to-R.F. and R.F.-to-F.F. flange combinations. In addition to being technically unnecessary to have a R.F.-to-R.F. mating combination, the advantages are: 1. The elimination of the extra cost for R.F. flanges. 2. The elimination of the extra delivery time required for a non-standard casing.

Figure 1

Figure 2

Figure 3 Steel Flange With Raised Face Machined Off

Steel R.F. Mating Flange

Full Face Gasket

Steel R.F. Mating Flange Cast Iron F.F. Pump Flange

Figure 4 Gasket Area 49.4 sq. in.

P1

P1

Bronze F.F. Pump Flange Figure 5 Gasket Area 15.9 sq. in.

P2

F.F. to F.F.

Cast Iron or Bronze F.F. Pump Flange Figure 6 Gasket Area 15.9 sq. in.

P3

P2

R.F. to R.F.

P3

F.F. to R.F.

TECH-I

TECH-I-8 Keep Air Out of Your Pump Most centrifugal pumps are not designed to operate on a mixture of liquid and gases. To do so is an invitation to serious mechanical trouble, shortened life and unsatisfactory operation. The presence of relatively small quantities of air can result in considerable reduction in capacity, since only 2% free air will cause a 10% reduction in capacity, and 4% free air will reduce the capacity by 43.5%. In addition to a serious loss in efficiency and wasted power, the pump may be noisy with destructive vibration. Entrained air is one of the most frequent causes of shaft breakage. It also may cause the pump to lose its prime and greatly accelerate corrosion.

When the source of suction supply is above the centerline of the pump, a check for air leaks can be made by collecting a sample in a "bubble bottle" as illustrated. Since the pressure at the suction chamber of the pump is above atmospheric pressure, a valve can be installed in one of the tapped openings at the high point in the chamber and liquid can be fed into the "bubble bottle." The presence of air or vapor will show itself in the "bubble bottle." Connect To Valve Installed At The High Point In Suction Chamber Or Discharge

Air may be present in the liquid being pumped due to leaky suction lines, stuffing boxes improperly packed, or inadequately sealed on suction lift or from other sources. Refer also to Section TECH-D-7, Pumping Liquids with Entrained Gas.

To Drain

On the other hand, very small amounts of entrained air (less than 1%) can actually quiet noisy pumps by cushioning the collapse of cavitation bubbles. TESTING FOR AIR IN CENTRIFUGAL PUMPS The amount of air which can be handled with reasonable pump life varies from pump to pump. The elimination of air has greatly improved the operation and life of many troublesome pumps. When trouble occurs, it is common to suspect everything but air, and to consider air last, if at all. In many cases a great deal of time, inconvenience, and expense can be saved by making a simple test for the presence of air. We will assume that calculations have already been made to determine that there is sufficient NPSH Margin (2 - 5 time the NPSHR) to insure that the noise is not due to cavitation. The next step should be to check for the presence of entrained air in the pumpage.

This test can also be made from a high point in the discharge side. Obviously, the next step is to eliminate the source of air since quantities present insufficient amount to be audible are almost certain to cause premature mechanical failure. NOTE: The absence of bubbles is not proof that the pumpage doesn't contain air.

TECH-I-9 Ball Bearings – Handling, Replacement and Maintenance Suggestions Ball bearings are carefully designed and made to watch-like tolerances. They give long, trouble-free service when property used. They will not stand abuse.

PULL BEARINGS CAREFULLY

KEEP CLEAN

1. Use sleeve or puller which contacts just inner race of bearing. (The only exception to this is some double suction pumps which use the housing to pull the bearing.)

Dirt causes 90% of early bearing failures. Cleanliness is a must when working on bearings. Some things which help:

2. Never press against the balls or ball cages, only against the races.

1. Do not open housings unless absolutely necessary.

3. Do not cock bearing. Use sleeve which is cut square, or puller which is adjusted square.

2. Spread clean newspapers on work benches and at pump. Set tools and bearings on papers only. 3. Wash hands. Wipe dirt, chips and grease off tools.

4. When using a bearing housing to pull a bearing, pull evenly, do not hammer on housing or shaft. With both races locked, shock will be carried to balls and ruin bearing.

4. Keep bearings, housings, and shaft covered with clean cloths whenever they are not being worked on.

INSPECT BEARINGS AND SHAFT

5. Do not unwrap new bearings until ready to install. 6. Flush shaft and housing with clean solvent before reassembly.

1. Look bearing over carefully. Scrap it if there are any flat spots, nicks or pits on the balls or races. Bearings should be in perfect shape. 2. Turn bearing over slowly by hand. It should turn smoothly and quietly. Scrap if "catchy" or noisy.

TECH-I

3. Whenever in doubt about the condition of the bearing, scrap it. Five or ten dollars worth of new bearings may prevent serious loss from downtime and pump damage. In severe or critical services, replace bearings at each overhaul. 4. Check condition of shaft. Bearing seats should be smooth and free from burrs. Smooth burrs with crocus cloth. Shaft shoulders should be square and not run over. CHECK NEW BEARINGS Be sure bearing is of correct size and type. For instance, an angular contact bearing which is dimensionally the same as a deep groove bearing may fit perfectly in the pump. However, the angular contact bearing is not suitable for end thrust in both directions, and may quickly fail. Also check to see that shields (if any) are the same as in the original unit. Refer to the pump instruction manual for the proper bearing to use.

INSTALL CAREFULLY 1. Oil bearing seat on shaft lightly. 2. Shielding, if any, must face in proper direction. Angular contact bearings, on pumps where they are used, must also face in the proper direction. Duplex bearings must be mounted with the proper faces together. Mounting arrangements vary from model to model. Consult instruction manual for specific pump. 3. Press bearing on squarely. Do not cock it on shaft. Be sure that the sleeve used to press the bearing on is clean, cut square, and contacts the inner race only. 4. Press bearing firmly against shaft shoulder. The shoulder helps support and square the bearing. 5. Be sure snap rings are properly installed, flat side against bearing, and that lock nuts are tight. 6. Lubricate properly, as directed in instruction manual.

TECH-I-10 Impeller Clearance IMPELLER CLEARANCE Open impeller centrifugal pumps offer several advantages. They're particularly suited but not restricted to liquids which contain abrasive solids. Abrasive wear on an open impeller is distributed over the diametrical area swept by the vanes. The resulting total wear has less effect on performance than the same total wear concentrated on the radial ring clearance of a closed impeller. The open impeller permits restoration of "new pump" running clearance after wear has occurred without parts replacement. Many of Goulds open impeller pumps feature a simple positive means for axial adjustment without necessity of disassembling the unit to add shims or gaskets.

7. Evenly tighten locking bolts, the jack bolts keeping indicator at proper setting. 8. Check shaft for free turning. *Established clearance may vary due to service temperature.

228

134A

423B

SETTING IMPELLER CLEARANCE (DIAL INDICATOR METHOD) 1. After locking out power, remove coupling guard and coupling. 2. Set dial indicator so that button contacts shaft end.

371A

3. Loosen jam nuts (423B) on jack bolts (371A) and back bolts out about two turns. 4. Tighten each locking bolt (370C) evenly, drawing the bearing housing toward the bearing frame until impeller contacts casing. 5. Set indicator to zero and back locking bolt about one turn.

370C DIAL INDICATOR METHOD

6. Thread jack bolts in until they evenly contact the bearing frame. Tighten evenly backing the bearing housing away from the frame until indicator shows the proper clearance established in instruction manual.*

TECH-I-11 Predictive and Preventative Maintenance Program This overview of Predictive and Preventative Maintenance (PPM) is intended to assist the pump users who are starting a PPM program or have an interest in the continuous improvement of their current programs. There are four areas that should be incorporated in a PPM program. Individually each one will provide information that gives an indication of the condition of the pump; collectively they will provide a complete picture as to the actual condition of the pump.

PUMP PERFORMANCE MONITORING There are six parameters that should be monitored to understand how a pump is performing. They are Suction pressure (Ps ), discharge pressure (Pd ), flow (Q), pump speed (Nr ), pumpage properties, and power. Power is easiest measured with a clip on amp meter but some facilities have continuous monitoring systems that can be utilized. In any event, the intent is to determine the BHP of the pump. When using a clip on amp meter the degree of accuracy is limited. It

TECH-I

should not be used to determine the efficiency of the pump. Clip on amp meters are best used for trouble shooting where the engineer is trying to determine the operating point of the pump. The most basic method of determining the TDH of the pump is by utilizing suction and discharge gauges to determine PS and Pd. The installation of the taps for the gauges is very important. Ideally, they should be located normal to the pipe wall and on the horizontal centerline of the pipe. They should also be in a straight section of pipe. Avoid locating the taps in elbows or reducers because the readings will not indicate the true static pressure due to the velocity head component. Avoid locating taps in the top or bottom of the pipe because the gauges can become air bound or clogged with solids.

Typically, readings are taken on the motor outboard and inboard bearing housings in the vertical and horizontal directions and on the pump outboard and inboard bearing housings in the vertical and horizontal directions. Additionally, an axial vibration measurement is taken on the pump. The inboard location is defined as the coupling end of the machine. It is critical that when the baseline vibration measurement is taken that the operating point of the pump is also recorded. The vibration level of a pump is directly related to where it is operating and in relation to its Best Efficiency Point (BEP). The further away from the BEP, the higher the vibrations will be. See the following chart for a graphical representation of vibration amplitudevs- flow.

Flow measurements can be difficult to obtain but every effort should be made to do so, especially when trouble shooting. In some new installations permanent flow meters are installed which make the job easier. When this is the case, make sure the flow meters are working properly and have been calibrated on a regular schedule. When flow meters are not installed, pitot tubes can be used. Pitot tubes provide a very accurate measure of flow, but this in an obtrusive device and provisions must be made to insert the tube into the piping. The other method of determining flow is with either a doppler or transitime device. Again, provisions must be made on the piping for these instruments, but these are non-obtrusive devices and are easier to use than the pitot tube. Caution must be exercised because each device must be calibrated, and independent testing has shown these devices are sensitive to the pumpage and are not 100% accurate. An accurate power measurement reading can also be difficult to obtain. Clip on map meters are the most common tool available to the Field Engineer who is trouble shooting a pump problem. In most cases this has proven to be accurate. However, as previously mentioned, this tool must be used and applied properly. Clip on map meters are not accurate enough to determine the actual efficiency of a pump. If accurate horsepower readings are necessary, a torque shaft must be installed but is not very practical in an actual field installation and lends itself to use in a laboratory environment much better. In some critical installations where the user has provided a permanent power monitor, these have varying degrees of accuracy and they must be understood up front. Finally, the properties of the pumpage must be known to accurately determine the actual pump performance. Pumpage temperature (Tp), viscosity, and specific gravity (S.G.), must be known. When all of the above parameters are known, it becomes a simple matter of calculating the pump performance. There are instances when it proves to be a very difficult if not an impossible task to determine all of the above parameters in the field, therefore, the Field Engineer must rely on his or her ability to understand where a compromise must be made to get the job done. The basic document the Field Engineer must have is the pump performance curve. With this it can be determined where the pump is performing in some cases without all of the information. PUMP VIBRATION AND BEARING ANALYSIS Vibration analysis is the cornerstone of all PPM programs. Perhaps the question asked most often is "What is the vibration level that indicates the pump is in distress?". The answer is that there is no absolute vibration amplitude level that is indicative of a pump in distress. However, there are several guidelines that have been developed as target values that enable the analyst to set alarm levels. Also many users have developed their own site criteria that is used as a guideline. Institutions such as the Hydraulic Institute and API have developed independent vibration criteria. Caution should be exercised when applying the published values...each installation is unique and should be handled accordingly. When a machine is initially started, a baseline vibration reading should be taken and trended over time.

TECH-I

The engineer must also look at the frequency where the amplitude is occurring. Frequency identifies what the defect is that is causing the problem, and the amplitude is an indication of the severity of the problem. These are general guidelines and do not cover every situation. The spectrum in the chart is a typical spectrum for a pump that has an unbalance condition. Bearing defect analysis is another useful tool that can be used in many condition monitoring programs. Each component of a roller bearing has its own unique defect frequency. Vibration equipment available today enables the engineer to isolate the unique bearing defects and determine if the bearing is in distress. This allows the user to shut the machine down prior to a catastrophic failure. There are several methods utilized but the most practical from a Field Engineering perspective is called bearing enveloping. In this method, special filters built into the analyzer are used to amplify the repetitive high frequency signals in the high frequency range and amplify them in the low frequency part of the vibration spectrum. Bearing manufacturers publish the bearing defect frequency as a function of running speed which allows the engineer to identify and monitor the defect frequency. Similar to conventional vibration analysis, a baseline must be established and then trended. There are other methods available such as High Frequency Detection (HFD), and Spike Energy but the enveloping technology is the latest development. It is a common practice to monitor bearing temperature. The most accurate method to monitor the actual bearing temperature is to use a device that will contact the outer race of the bearing. This requires holes to be drilled into the bearing housings which is not always practical. The other method is the use of an infrared 'gun' where the analyst aims the gun at a point on the bearing housing where the temperature reading is going to be taken. Obviously, this method is the most convenient but there is a downside. The temperature being measured is the outside surface of the bearing housing, not the actual bearing temperature. This must be considered when using this method.

To complete the condition monitoring portion of a PPM program, many users have begun an oil analysis program. There are several tests that can be performed on the lubricant to determine the condition of the bearing or determine why a bearing failed so appropriate corrective action can be taken. These tests include Spectrographic Analysis, Viscosity Analysis, Infrared Analysis, Total Acid Number, Wear Particle Analysis and Wear Particle Count. Most of these tests have to be performed under laboratory conditions. Portable instruments are now available that enable the user to perform the test on site. PUMP SYSTEM ANALYSIS Pump system analysis is often overlooked because it is assumed the system was constructed and operation of the pumps are in accordance with the design specifications. This is often not the case. A proper system analysis begins with a system head curve. System head curves are very difficult to obtain from the end user and, more often than not, are not available. On simple systems, they can be generated in the field but on more complicated systems this can't be done. As has been stated previously, it is imperative to know where the pumps are being operated to perform a correct analysis and this is dependent on the system.

A typical system analysis will include the following information; NPSHA, NPSHR, static head, friction loss through the system, and a complete review of the piping configuration and valving. The process must also be understood because it ultimately dictates how the pumps are being operated. All indicators may show the pump is in distress when the real problem is it is being run at low or high flows which will generate high hydraulic forces inside the pump. CONCLUSION A PPM program that incorporates all of the topics discussed will greatly enhance the effectiveness of the program. The more complete understanding the engineer has of the pumping system, the more effective the PPM program becomes.

TECH-I-12 Field Alignment Proper field alignment of pumps and drivers is critical to the life of the equipment. There are three methods used in industry: rim and face, reverse dial indicator, and laser alignment. RIM AND FACE This method should not be used when there is no fixed thrust bearing or on pumps/drivers that have axial shaft movement.

P

A

Y (Motor End)

X (Pump End)

Fig. 1 Rim and Face Dial Indicator Alignment (Criteria: 0.002 in. T.I.R. rim and face reading) REVERSE DIAL INDICATOR

Fig. 2 Reverse Dial Indicator Alignment (Criteria: 0.0005 in. per inch of dial indicator separation)

LASER ALIGNMENT Although a popular method, it's not any more accurate than either dial indicator method. Instruments are expensive and require frequent calibration.

This method is the most widely used and is recommended for most situations.

TECH-I

MAXIMUM DEVIATION AT EITHER DIAL INDICATOR (MILS/INCH OF INDICATOR SEPARATION)

2. Level the pump off of the shaft extension. Do not level off of the pump casing flanges. Remember, the piping must come to the pump. You are aligning the pump shaft and the driver shaft. Shafts are the datum, not flanges. a. Use a STARRET No.135 level to level the shaft. Unacceptable

b. Leveling the pump should be accomplished by shimming under the bearing frame toot. B. Motor 1. Set the motor on the baseplate. 2. Using a straight edge, approximate the shaft alignment.

Acceptable Excellent

Fig. 3 Guideline for Alignment Tolerances MECHANICAL ALIGNMENT PROCEDURE This procedure assumes the presenter knows how to align a pump and has a basic understanding of pump baseplates and piping installation. There are many alignment systems available. We will be using the plotting board with dial indicators developed by M.G. Murray. The plotting board is as accurate as any method available today and gives the best representation of the actual position of the machines that are being aligned. The actual procedure that will be discussed is the reverse dial indicator procedure because it is the most versatile and widely used alignment procedure used today. PREPARING FOR ALIGNMENT A. Baseplate Inspection 1. Inspect all mounting surfaces to make sure they are clean and free of any paint, rust, grime, burrs, etc.

a. This will require setting shims of the same thickness under the motor feet; you are just trying to get close so you can use the dial indicators. Get the rough alignment within 0.0625". b. If the motor is higher, there is something wrong or it is a special case. This situation must be inspected. Do not shim the pump. The pump is connected to the piping and it will present difficulties with future work on the installation. c. Make sure you have the proper shaft separation. 3. Remove soft toot. C. Alignment. (Reverse indicator Method) 1. Install reverse dial indicator tooling on shafts. 2. Measure and record the following dimensions on a worksheet, SA, Al, IO. These parameters are defined as follows: a. SA = Distance between the dial indicators which are located at the respective planes of correction. b. Al = Distance between the adjustable plane of correction and the inboard foot of the adjustable machine. c. IO = Distance between the inboard foot and outboard foot of the adjustable machine. 2. Correct for dial indicator sag.

a. Thoroughly clean mounting surfaces. Debar using a honing stone if necessary.

a. Remove dial indicator tooling from the unit.

b. At this point, it is assumed that the baseplate has been installed correctly and is level. B. Pump and Driver Inspection

b. Install reverse dial indicator tooling on a pipe or piece of round bar stock in the exact configuration that you removed it from the unit that is being aligned. The dial indicators must be set to the SA distance.

1. Inspect all mounting surfaces to make sure they are clean and free of any paint, rust, grime, burrs, etc.

c. Zero the dial indicator while they are in the vertical up position.

C. Shim InspectIan

d. Rotate the entire set-up 180° and record dial indicator readings. This is the sag, the correction will be made when you take the alignment readings.

1. Inspect all shims to make sure they are clean and free of any paint, rust, grime, burrs. etc. 2. Dimensionally inspect ALL shims to be used and record the reading on the individual shims. DO NOT ASSUME THAT THE SHIMS ARE TO THE EXACT DIMENSIONS THATARE RECORDED ON THEM. SETTING EQUIPMENT A. Pump 1. Set pump on pump mounting pads. Insert pump hold-down bolts but do not tighten. a. If there is existing piping, line up pump flanges with pipe flanges. DO NOT CONNECT THE PIPING AT THIS POINT.

TECH-I

3. Reinstall the reverse dial indicator tooling back to the configuration it was in Step 1. a. The SA dimension must be held. 4. Establishing the datums. a. You must take readings from the same position relative to the fixed machine or the moveable machine. Choose the position that is the most comfortable. DO NOT CHANGE THE ORIENTATION ONCE YOU BEGIN TO TAKE READINGS. b. All dial indicator readings must be taken 90° apart from each other and at the same relative position each time. Either mark the couplings in 80° increments or use a two dimension bubble level with a magnetic pad. The level is the most accurate method.

c. The shafts must be rotated together and readings taken from the same exact locations every time; therefore, if the coupling spacer is removed, the stationary and adjustable machines coupling hubs must be marked in 90°. increments. 5. Take the initial set of readings. a. Zero the dial indicators at the 0° position. b. Rotate the shafts simultaneously taking readings every 90°, (0°, 90°, 180°, 270°). Record readings on the reverse dial indicator worksheet. 6. Determine if the initial readings are good. a. Add top (T) and bottom (B) together for both planes and the two side readings (S) together for both planes. b. Take the difference of the two readings. If the difference exceeds 0.002", there is something wrong with the readings. Inspect the set up and make any necessary adjustments. 7. Algebraically zero the side readings. Be consistent on which side you zero; it is usually easier to zero the 90° side. 8. Make dial indicator sag correction on worksheet. a. Dial indicator sag only effects vertical readings. Since the dial indicator is going to read negative on the bottom, add the sag to the dial indicator reading on the bottom. 9. Divide all corrected readings by two because they are TIR readings taken on the outside of a circle. a. Remember, when the dial indicator reads positive, the probe is being pushed in. When it reads negative, the probe is extended. 10. Determine shim change. a. Lay out the machine dimensions on the plotting board transparency. 1. Once the scale is determined you must be consistent and use only that particular scale. b. Referring to our example, you must use the "C" scale on the bottom horizontal axis. The bottom horizontal axis represents the physical dimensions of the machine.

c. The left vertical axis represents the misalignment/shim correction scale. d. Locate , S, A, IB, OB, 1. S is located where the vertical and horizontal axis of the overlay intersect. S represents the location of the stationary reference plane. 2. A is marked on the horizontal axis and represents the location of the adjustable reference plane. In our example, it is marked at 7" on the C scale. 3. B is marked on the horizontal axis and represents the location of the inboard foot of the adjustable machine. In our example it is marked at 15" on the C scale. 4. OB is marked on the horizontal axis and represents the location of the outboard foot of the adjustable machine. In our example it is marked at 36" on the C scale. 5. Mark reference on the plotting board transparent vertical scale. 11. Plot shim change for vertical correction first. a. Transform worksheet data to the plotting board. 1. Set S at 0.009" low mark based on the E vertical scale 2. Set at 0.0035" high mark based on the E vertical scale b. Draw vertical lines from the lB and OB locations on the red line to the horizontal zero line on the plotting board. c. Count the vertical distances from the lB and OB marks to the horizontal zero line using the correct scale, in our case the E scale, these values are the shim changes at the inboard (lB) and outboard (OB) feet of the adjustable machine. 12. Make shim change. 13. Repeat Step 11 for horizontal correction. 14. Check alignment. a. The machines should be aligned at this point; if not, repeat Steps 11 and 12. 15. Inspect final alignment and record all results.

TECH-I

NOTES

Your ITT Industries Pump Manual ITT/Goulds Pumps is pleased to provide you with this copy of GPM-7. Since the first edition was published in 1973, GPM has earned a reputation as the most complete and useful source of pump information available. We’re proud of GPM and confident that you will find it to be a valuable tool for application and selection of pumps. But because we're continually improving our products or adding new pump lines to meet the ever changing needs of industry, your GPM can never be considered current. For this reason we've provided a GPM registration card so that we can keep you informed of the latest product information.

IPG Salesperson: Please fill out the following registration form upon presentation of this manual and mail to Goulds Pumps/ITT Industries, IPG Advertising Dept., 240 Fall Street, Seneca Falls, NY 13148 USA.

GPM Registration Name ________________________________________

Title _______________________________

Company ___________________________________________________________________________ Division or Dept. ________________________________

Phone _____________________________ FAX _______________________________

Address

___________________________________________________________________________ Street

Address

___________________________________________________________________________ City

State

Country

Zip

Goulds Sales Office or Representative

Sales Engineer

__________________________________________

___________________________________

Markets – Check one(s) that apply: ■ Chemical ■ Power ■ Hydrocarbon Processing Industry ■ Other

■ Pulp & Paper ■ Municipal ■ Mining & Materials

Customer Type ■ A&E ■ Distributor/ Rep ■ OEM ■ End User ■ Other

Place Postage Here

Goulds Pumps An ITT Industries Fluid Company IPG Advertising Dept. 240 Fall Street Seneca Falls, NY 13148 USA

Shell Chemical - Safety Literature

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SHELL CHEMICAL SAFETY GUIDE EMERGENCY RESPONSE NUMBER (CHECK WITH YOUR LOCAL REP)

The information contained in this Guide is for the protection of persons working with or using chemicals discussed in this Guide and who might be affected by emergency situations involving these chemicals. This information does not address and should not be relied upon to protect against all types of risks. Also, it does not necessarily provide all information essential to personal protection with respect to any particular situation. TABLE OF CONTENTS SECTION I INTRODUCTION HOW TO HANDLE CHEMICALS PROPERLY PERSONAL PROTECTIVE EQUIPMENT FIRE FIGHTING SPILL AND LEAK FIRST AID SECTION II INSTRUCTIONS CHEMICAL REFERENCE CHARTS A Acetone Alcohols, Combustible Alcohols, Flammable

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Allyl Chloride Azamide TM Resins B Benzene Bisphenol A Butadiene, 1,3Butene-1 Butyl Dioxitol TM Glycol Ether Butyl Oxitol TM Glycol Ether C CleartufTM Resins Cyclo-Sol TM 28 Solvent Cyclo-Sol 42, 47, 53, 63 Solvents D Diacetone Alcohol Dienes Diethylene Glycol E Eco-Cryl TM Acrylic Resin Dispersion 9790 Epichlorohydrin Epi-Cure TM Curing Agents (3000, 3100, 3200, 3300 and 3500 series) Epi-Cure Curing Agents (3400, 9000 series, W, Y and Z) Epi-Cure Curing Agent Solvent Solutions Epi-Cure Series P Curing Agents Epi-RezTM Epoxy Resins

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EponTM Resins Epon Resin Solutions/Eponol TM Resins Ethyl Hexanol, 2Ethylene Ethylene Glycol Ethylene Oxide G Glycols H HeloxyTM Modifiers Hexylene Glycol Hydrocarbon Solvents, Aromatic Hydrocarbon Solvent Blends, Combustible Hydrocarbon Solvent Blends, Extremely Flammable Hydrocarbon Solvent Blends, Flammable Hydrochloric Acid I Isobutyl Alcohol Isoprene Isopropyl Alcohol Isopropyl Ether K Ketones Kraton TM Theromoplastic Rubber M Methyl Ethyl Ketone Methyl Isobutyl Carbinol

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Methyl Isobutyl Ketone N Neodene TM Alpha and Internal Olefins 13, 1314, 14, 14/16, 14/16/18, 1518, 16, 16/18, 18, 20, 2024, 2024+ Neodene Alpha and Internal Olefins, Combustible, 10, 10-11/12-13,10-11/12-13/14, 10/1112, 10/1112/1314/12, 10/12/1314, 1012, 1014, 1112, 1112/12, 12, 12/14, 12/14/16, 1213, 12/1314, 1112/1314 Neodene Alpha and Internal Olefins, Flammable 6 HP, 6/8, 6/8/10, 6/12, 8, 810 IO, 810 AO Neodol TM Detergent Alcohols (C11 and less) 1, 91 Neodol TM Detergent Alcohols (C12 and greater) 23, 25, 45 Neodol TM Alcohol Ethoxylates Neoflex TM Plasticizer Alcohol NeosolvTM Solvents Normal Butyl Alcohol O Olefins P Phenol Plastics Polybutylene Propylene R Rubber Solvent, Rubber Solvent 332 S Secondary Butyl Alcohol Shell Mineral Spirits 135, 145-EC, 150 EC,

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146 HT, 200 HT, 160, 1168, 1169, 147 EC Shell Mineral Spirits 6855 Shell Sol BT-9, BT-67 EC, BT-4, BT 67W EC, B-8, B HT, BJ-8, BT-20, BT-40, BT-60, BT -75, BT -83, WRG Shell Sol LF, M-95 EC, 320, M-75 EC, MS-14, PS, RB, TV, ZP Shell Sol 71, 340 EC, 340 HT, 323 EC, 142 HT, 137, 130 B, WT, D60, FC Shell Solvent Blend SC -0612 Shell TS-28, TS -28B, TS-28R Shell VM&P Naphtha 3M, EC, High Flash, HB, HT ShellvisTM Viscosity Index Improvers Sulfolane Surfam TM Resin 58 T Thermoplastic Polyamide Resins Tolu-Sol TM Solvents Toluene Traytuf TM Resins Triethylene Glycol W Water Dispersible Resins X Xylene

Note: Bolded entries indicate major headings of chemicals or chemical groups.

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SECTION I INTRODUCTION This Guide is a supplement to the product labels and Material Safety Data Sheets (MSDS) for selected Shell Chemical products and product groups. It is a quick reference for basic safety information which: 1) describes the products and the hazards associated with them; 2) provides the general information for handling the products properly to reduce the potential risks associated with those hazards; and 3) provides the basic information needed for responding to emergencies involving fire, release, or personnel exposure. This Guide is current as of the publication date, however, it does not contain all the information needed to properly receive, ship, handle, or use the chemicals listed. You should refer to the most current MSDS for more detailed information on the product(s) you are about to handle. To ensure that you have the most current MSDS call your Shell sales representative or 1-800-240-MSDS (6737). HOW TO HANDLE CHEMICALS PROPERLY Treat all chemicals with respect! Don't take shortcuts! Follow your supervisor's directions and the manufacturer's instructions for the proper use of the chemical. Be serious when working around chemicals. "Horseplay" around chemicals can have serious, possibly fatal, consequences. Read and follow the warning signs in your work area! DO NOT EAT, SMOKE, or DRINK in areas where chemicals are being stored, handled or used.

STORE AND HANDLE CHEMICALS PROPERLY Read product label and MSDS for proper storage and handling requirements. Store and handle chemicals in dry, well-ventilated areas away from potential sources of ignition (heat, sparks and open flame) such as welding, smoking, grinding, cutting, spark-producing machinery, or open burning. Storage, handling and transfer equipment should be electrically grounded and bonded to prevent the buildup of an electrostatic charge. Eye wash stations and safety showers should be readily available. DO NOT remove or alter the container label. NEVER put chemicals into unmarked or mislabeled containers. Keep chemical containers tightly closed when not in use. "Empty" containers can be dangerous because they can contain chemical vapors or liquid residue that can be explosive. They should be disposed of properly or cleaned by a professional service before reuse. KEEP YOURSELF AND YOUR WORK AREA CLEAN Objects lying on the ground or small spills of chemicals can create falling hazards.

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Avoid creating dust clouds which can be potentially explosive. Wash your hands thoroughly with soap and water after handling any chemical or chemical container and before eating, smoking or using toilet facilities. DO NOT reuse contaminated clothing before it has been cleaned. WORK IN WELL VENTILATED AREAS DO NOT depend on your sense of smell to warn you about the presence of gases or vapors. Some chemicals can be hazardous at concentrations too low to be detected by smell. Other chemicals can actually deaden your sense of smell. Engineering controls, such as proper ventilation, are the best way to prevent overexposure to gases or vapors. BE PREPARED You should know your facility's emergency procedures. Know the emergency phone numbers in your area (911 or police, fire, poison control, ambulance) and post them at all phone locations. Know the location of the nearest: fire exit, fire alarm, fire fighting equipment, eye wash station, safety shower, first aid kit, and other safety equipment. Know how to use this equipment and periodically check to make sure it is in good working order. PERSONAL PROTECTIVE EQUIPMENT (PPE) Exposure Assessment/Training (OSHA requirement) An exposure assessment should be performed before beginning any job that may require PPE. This is necessary to determine the proper PPE for the job. You should be trained in the proper use of all PPE before using. RESPIRATORY PROTECTION Wear the proper respirator if ventilation does not provide adequate exposure control. You should be medically fit in order to wear a respirator. Consult your supervisor for more information on this requirement. Be sure you have been fitted with the type and size of respirator you need to ensure you are protected. You should be trained in how and when to wear the respirator and how to clean, maintain and store it properly. You should use either an air-purifying respirator with the proper cartridge(s) or an airsupplied respirator, depending on the chemical and its concentration in the air. Consult the product MSDS and your respirator supplier for further information on respirator use. In case of respirator failure, an escape route should be planned. SKIN PROTECTION

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Gloves that are resistant to the chemical being handled should be worn. More information on glove use, if available, can be found for each product in the Guide under "Recommended Personal Protective Equipment" or on the product MSDS. If protective clothing is required, it should also be chemical-resistant. The glove and protective clothing recommendations in this guide are provided from test data in published literature and do not take into account such factors as job task, type of exposure, or durability requirements. If more information is needed on glove or clothing use, contact your PPE supplier. EYE/FACE PROTECTION Protect your eyes and face. Wear chemical goggles if there is a potential for eye contact. Safety glasses may be worn in some cases where the exposure potential is less. Face shields may be needed in some cases but they are not a substitute for eye protection and should only be used with goggles or safety glasses. OTHER PROTECTION Other protection, such as hard hats, safety shoes and chemical-resistant boots may be required in some cases. DECONTAMINATION If you are overexposed from breathing chemical vapors or gases, get first aid immediately. If you get chemicals on your skin or in your eyes, flush with large amounts of water immediately. Use an eye wash station or safety shower as appropriate. Contaminated clothing should be removed immediately and washed thoroughly before reuse. Contaminated leather articles, such as belts and shoes, may not be decontaminated and may have to be discarded (refer to the product MSDS for specific guidance). Refer to the product MSDS for more information on first aid treatment. FIRE FIGHTING If a fire starts, act quickly! Activate the alarm, rescue injured persons if it is safe to do so, and call 911 or your local fire department. Small fires If the fire is small and contained, try to put it out if you can do so safely. Your quick action may save lives and prevent the fire from spreading. Be properly trained in the use of fire extinguishers. Large fires DO NOT attempt to put out large fires without help. Trained fire fighters should be called. If you are not trained in fire fighting, the best thing to do is to evacuate the area, notify your supervisor and the fire department. Fire fighters should not enter a fire space where chemicals are involved without full bunker gear (helmet with face shield, bunker coat, gloves and rubber boots) including a positive pressure NIOSH-approved self-contained breathing apparatus. Fire-exposed chemical containers should be cooled with water spray from a

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safe distance to reduce the risk of their rupture or explosion due to pressure buildup. FIRE EXTINGUISHERS Knowing the right type of fire extinguisher to use and how to use it is important. Using the wrong type of extinguisher, as in water spray on hydrocarbon solvents, can actually cause the fire to spread. Even if you have the right extinguisher, you must know how to use it correctly to be effective. Get training. General guidelines for fire extinguisher use include: Water Spray or Fog: Dry Chemical: Direct the Direct the water at the chemical spray at the base of base of the flames. the flames. CO 2 (Carbon Foam & Alcohol Foam: Allow the foam to fall lightly Dioxide): Direct the on the fire. Do not spray spray as close to the fire directly into the fire.* as possible

*NOTE: In general, foam is used for water insoluble chemicals, such as hydrocarbon solvents, and "alcohol" foam is used for water soluble chemicals, such as alcohols and ketones. The Emergency Action Guides presented with each chemical entry in the Guide provide product specific fire fighting information. SPILL AND LEAK Accidental releases to the environment can be hazardous and may require government notification! Notify your supervisor immediately in the event of a spill or leak so that immediate steps can be taken to contain the spill and a determination can be made if notifications are required. DO NOT flush chemical spills or residues into sewers or allow them to contaminate rivers, streams or public water supplies! Run-off from spill cleanup or fire fighting activities should be contained and disposed of properly. The Emergency Action Guides presented with each chemical entry in the Guide provide product specific spill and leak information. FIRST AID In the event of an emergency, act quickly! For serious injuries and exposures, activate your Emergency Medical System (EMS) immediately by calling 911 or your local emergency phone number so that qualified help can arrive quickly. Be prepared to provide them with the following information: location, sex and approximate age of the victim, the estimated time of the injury or exposure, and the nature of the injury or exposure.

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KEEP THE VICTIM CALM Know Basic Life Support - first aid, rescue breathing and CPR (Cardio Pulmonary Resuscitation)! Training is available from a variety of sources including the American Red Cross (ARC), the National Safety Council (NSC) and the American Heart Association (AHA). Do not attempt any procedure that you have not been trained to perform. MAINTAIN A PROPERLY EQUIPPED FIRST AID KIT Check the contents periodically to make sure the kit is properly stocked and in good condition. Ensure your own safety! DO NOT attempt to rescue or assist a victim until you are sure that the scene is safe or until you are equipped with the proper personal protective equipment. Dangerous concentrations of a chemical may still be present in the area. Decontaminate the victim as completely as possible before rendering aid to minimize your own exposure to chemicals that may be on the victim. Use a barrier device when performing rescue breathing or CPR to prevent exposure to chemicals and/or infectious disease. DO NOT attempt to move the injured victim unless they are in danger of further exposure or injury! Wait until qualified help arrives! Moving the injured victim unnecessarily can cause additional and potentially serious injury such as permanent paralysis if they have sustained a neck or spine injury. The Emergency Action Guides presented with each chemical entry in the Guide provide product specific first aid information. NOTE: Any child exposed to the chemicals listed in this Guide should receive immediate medical attention.

SECTION II INSTRUCTIONS This section of the Guide provides chemical specific information on most Shell Chemical products. The chemicals are listed alphabetically either as individual chemicals or chemical groups. Chemical groups were established based on similar chemical characteristics and hazards wherever possible in order to minimize the size of the Guide. The Table of Contents lists all chemicals in this Guide as well as the chemical group to which they belong. Thus, in the Table of Contents isopropyl alcohol can be located by either looking under "isopropyl alcohol" or its chemical group, "alcohols, flammable". PRODUCTS The first part of each page, identifies the product or product category, provides you with a brief physical description of the chemicals covered, followed by three columns headed

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"Product Name", "Product Synonyms", and "Flash Point". The product names listed are the names under which Shell Chemical markets the products. Below the product name, the Chemical Abstract Services (CAS) Registry Number is provided unless the product is a mixture and/or is regarded as proprietary information. Common product synonyms are provided in the next column. The flash points in degrees Fahrenheit and the method used (TCC = Tag Closed Cup; SF = Seta Flash) are listed in the last column. NOTE: An asterisk(s) following a product name indicates that different hazard ratings are noted in the next section, "Hazard Ratings". HAZARD RATINGS The hazard ratings for the two most common numerical hazard rating systems are given for each chemical or chemical group. The first is the Hazardous Material Identification System or HMIS which was developed in the mid 1980's by the National Paint and Coatings Association. It was principally designed to numerically rank the hazard severity of chemicals in the normal work environment. The second rating system is the National Fire Protection Association or NFPA rating system that was designed to rank the hazard severity of chemicals in emergency response situations such as fires or spills. Both systems use a number scale from 0 to 4 to rate Health, Fire and Reactivity hazards with 0 representing a minimal hazard and 4 representing the most severe hazard. The health hazard number represents the hazard from short-term exposure with resulting acute effects only; chronic health effects are not addressed. The criteria for assigning numbers to each hazard class are too lengthy, particularly in the case of health hazards, to explain in this Guide, however, it should be noted that HMIS and NFPA criteria are similar. You may have observed numerical hazard ratings on Shell Material Safety Data Sheets. This system was adapted from the NFPA system by Shell in the late 1970's. It is somewhat more conservative than both the HMIS and NFPA system. For instance, a corrosive material would be rated a 4 in the Shell system, a 3 in HMIS, and could be either a 3 or 4 in the NFPA system. Because of the wider recognition of the HMIS and NFPA hazard rating systems, the Shell ratings are not presented in this Guide and will be phased out from Shell Material Safety Data Sheets. CAUTION: Numerical rating systems should only be used as a quick reference for gauging the relative acute hazards of a chemical product. You must read the Material Safety Data Sheet to gain a full understanding of the hazards associated with a chemical. SPECIAL STORAGE AND HANDLING GUIDELINES Supplemental chemical specific storage and handling guidelines, if available, are provided under this heading. You should consider this information in addition to the general guidelines found in Section I. RECOMMENDED PERSONAL PROTECTIVE EQUIPMENT Chemical specific PPE, if available, is provided under this heading. This information includes recommended glove material and type of respirator cartridge. You should consider this information in addition to the general guidelines found in Section I. HEALTH HAZARDS The principal health hazards of the products are summarized. In some cases a product

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group may be so large and diverse that the health hazards cannot be summarized easily. You must review the Material Safety Data Sheet for each product for more detailed health hazard information. FIRE/EXPLOSION/REACTIVITY HAZARDS You will find important information regarding physical hazards, such as fire and explosion, under this heading. The flammability hazard statements in this guide are based on the American National Standard Institute (ANSI) Labeling Standard, OSHA definitions, and are designed for the workplace. For transportation classification under Department of Transportation (DOT) hazardous materials regulations, you must refer to the MSDS. EMERGENCY ACTION GUIDE The last item provides you with the basic information necessary to handle an emergency situation in the event of a fire, spill or leak, or personnel exposure. EMERGENCY RESPONSE NUMBER CHECK WITH YOUR LOCAL REP.

ALCOHOLS, COMBUSTIBLE (Clear, mobile liquids) Product Name Diacetone Alcohol* 123-42-2 Methyl Isobutyl Carbinol 108-11-2 2-Ethyl Hexanol 104-76-7 Neoflex TM Plasticizer Alcohol 68527-05-9

Product Synonyms DAA; 4-hydroxy- 4-methyl2-pentanone MIBC; 4-methyl2-pentanol; methyl amyl alcohol 2-EH;2-ethyl-1-hexanol; 2-ethyl hexyl alcohol Isononyl alcohol; INA; octene, hydroformylation products

Flash Point (°F, TCC) 133 103 166 195

HAZARD RATINGS Rating HMIS

Health 2

Fire 2

Reactivity 0

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2

0

RECOMMENDED PERSONAL PROTECTIVE EQUIPMENT (Refer to the Introduction for general PPE guidelines.) Gloves: diacetone alcohol 2-ethyl hexanol Respirator cartridge:

neoprene, butyl rubber, nitrile rubber neoprene, Viton, butyl rubber organic vapor

HEALTH HAZARDS Alcohols may be moderately to severely irritating to the eyes. Prolonged or repeated skin contact with alcohols can result in defatting or drying of the skin with resulting irritation and dermatitis. Prolonged breathing of high alcohol vapor concentrations may cause dizziness, lightheadedness, headache, nausea and loss of coordination (central nervous system depression). Alcohols may be toxic if swallowed, producing central nervous system depression. FIRE/EXPLOSION/REACTIVITY HAZARDS CAUTION! COMBUSTIBLE. These products may be ignited by heat, sparks, open flames or contact with strong oxidizing agents. Vapors may accumulate and travel to ignition sources distant from the handling site resulting in a flash fire. Contact with metallic aluminum at temperatures over 120 °F may cause a vigorous reaction. EMERGENCY ACTION GUIDE IF...FIRE THEN...Use water fog, "alcohol" foam, dry chemical or CO2 to extinguish flames. Cool fire exposed containers with water spray. IF...SPILL OR LEAK THEN...Eliminate ignition sources. Wear appropriate personal protective equipment. Shut off leak if safe to do so. If vapor cloud forms, use water fog to suppress or blanket spill with "alcohol" foam. Dike and contain spill. Monitor area with a combustible gas indicator. Remove with vacuum truck or pump to storage/recovery vessel. Soak up residue with suitable adsorbent material such as clay or sand. Flush area with water to remove trace residue. Contain run-off from water fog and

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residue flush and dispose of properly. IF...EYE CONTACT THEN...Flush eyes with large amounts of water for at least 15 minutes while holding eyelids open. Rest eyes for 30 minutes. If redness, burning, blurred vision or swelling persists get medical attention. IF...SKIN CONTACT THEN...Remove contaminated clothing and shoes. Flush affected area with large amounts of water. Wash with soap if available. If blisters appear on the skin get medical attention. IF...INHALATION THEN...Move victim to fresh air. If the victim has difficulty breathing or tightness of the chest, is dizzy, vomiting or unresponsive, give 100% oxygen with rescue breathing or CPR as required and transport to the nearest medical facility. IF...INGESTION THEN...DO NOT GIVE LIQUIDS TO A DROWSY, CONVULSING OR UNCONSCIOUS PERSON. Otherwise, induce vomiting by administering 2 glasses of water and 2 tablespoons of Syrup of Ipecac. Keep head below hips while vomiting. Get medical attention. ALCOHOLS, FLAMMABLE (Clear, mobile liquids) Product Name Normal Butyl Alcohol 71-36-3 Isobutyl Alcohol 78-83-1 Secondary Butyl Alcohol 78-92-2 Isopropyl Alcohol 67-63-0

Product Synonyms NBA; 1-butanol; propyl carbinol IBA; isopropyl carbinol; 2-methyl-1-propanol SBA; sec-butyl alcohol; 2-butanol; 2-hydroxybutane; methyl ethyl carbinol IPA; isopropanol; 2-propanol; secondary propyl alcohol; dimethyl carbinol

Flash Point (°F, TCC) 98 86 72

53

HAZARD RATINGS Rating

Health

Fire

Reactivity

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3 3

0 0

RECOMMENDED PERSONAL PROTECTIVE EQUIPMENT (Refer to the Introduction for general PPE guidelines.) Gloves: normal butyl alcohol isobutyl alcohol isopropyl alcohol Respirator cartridge:

neoprene, nitrile rubber, PVC neoprene, nitrile rubber, PVC neoprene, nitrile rubber, PVC organic vapor

HEALTH HAZARDS Alcohols may be from moderately to severely irritating to the eyes. Prolonged or repeated skin contact with alcohols can cause defatting or drying of the skin resulting in irritation and dermatitis. Prolonged breathing of high alcohol vapor concentrations may cause dizziness, lightheadedness, headache, nausea and loss of coordination (central nervous system depression). Alcohols may be toxic if swallowed, producing central nervous system depression. FIRE/EXPLOSION/REACTIVITY HAZARDS WARNING! FLAMMABLE. These products may be ignited by heat, sparks, open flames or contact with strong oxidizing agents. Vapors may accumulate and travel to ignition sources distant from the handling site resulting in a flash fire. Contact with metallic aluminum at temperatures over 120 °F may cause a vigorous reaction. EMERGENCY ACTION GUIDE

IF...FIRE THEN...Use water fog, "alcohol" foam, dry chemical or CO2 to extinguish flames. Cool fire exposed containers with water spray. IF...SPILL OR LEAK THEN...Eliminate ignition sources. Wear appropriate personal protective equipment. Shut off leak if safe to do so. If vapor cloud forms, use water fog to suppress or blanket spill with "alcohol" foam. Dike and contain spill. Monitor area with a combustible gas indicator. Remove with vacuum truck or pump to storage/recovery vessel. Soak up

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residue with suitable adsorbent material such as clay or sand. Flush area with water to remove trace residue. Contain run-off from water fog and residue flush and dispose of properly. IF...EYE CONTACT THEN...Flush eyes with large amounts of water for at least 15 minutes while holding eyelids open. Rest eyes for 30 minutes. If redness, burning, blurred vision or swelling persists, get medical attention. IF...SKIN CONTACT THEN...Remove contaminated clothing and shoes. Flush affected area with large amounts of water. Wash with soap if available. If blisters appear on the skin, get medical attention. IF...INHALATION THEN...Move victim to fresh air. If the victim has difficulty breathing or tightness of the chest, is dizzy, vomiting or unresponsive, give 100% oxygen with rescue breathing or CPR as required and transport to the nearest medical facility. IF...INGESTION THEN...DO NOT GIVE LIQUIDS TO A DROWSY, CONVULSING OR UNCONSCIOUS PERSON. Otherwise, induce vomiting by administering 2 glasses of water and 2 tablespoons of Syrup of Ipecac. Keep head below hips while vomiting. Get medical attention. ALLYL CHLORIDE (Clear to yellow, mobile liquid, irritating odor) Product Name

Flash Point (°F, TCC)

Product Synonyms Chlorallyene; 3-chloro-1-propene; 1-chloropropene-2

Allyl Chloride 107-05-1

-25

HAZARD RATINGS Rating HMIS NFPA

Health 3 3

Fire 3 3

Reactivity 1 1

RECOMMENDED PERSONAL PROTECTIVE EQUIPMENT

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(Refer to the Introduction for general PPE guidelines.) Full protective clothing should be worn when handling allyl chloride. Gloves: Respirator cartridge:

polyvinyl alcohol organic vapor

HEALTH HAZARDS Allyl chloride is severely irritating to the eyes. Allyl chloride is severely irritating to the skin and may cause an allergic skin reaction. If it is absorbed through the skin, it may cause a deep-seated pain, called "deep bone" ache, at the point of contact. Breathing allyl chloride vapors is severely irritating to the nose, throat and respiratory tract and is moderately toxic. It may cause liver, kidney, and lung damage, peripheral nerve damage, and central nervous system depression (dizziness, light-headedness, headache, nausea and loss of coordination). Allyl chloride may be harmful or fatal if swallowed; may produce liver and kidney damage. FIRE/EXPLOSION/REACTIVITY HAZARDS DANGER! EXTREMELY FLAMMABLE. Allyl chloride may be ignited by heat, sparks, open flames or contact with strong oxidizing agents. Vapors may accumulate and travel to ignition sources distant from the handling site resulting in a flash fire. Allyl chloride may undergo spontaneous hazardous polymerization reaction at elevated temperatures. Avoid allyl chloride contact with caustics, strong acids, amines, ammonia, ferric chloride, magnesium, aluminum, and galvanized steel. EMERGENCY ACTION GUIDE IF...FIRE THEN...Use water fog, foam, dry chemical or CO2 to extinguish flames. DO NOT use a direct stream of water; product will float and can be reignited on surface of water. Cool fire exposed containers with water spray. IF...SPILL OR LEAK THEN...Eliminate ignition sources. Wear appropriate personal protective equipment. Shut off leak if safe to do so. If vapor cloud forms, use water fog to suppress or blanket spill with foam. Dike and contain spill. Monitor area with a combustible gas indicator. Remove with vacuum truck or pump to storage/recovery vessel. Soak up residue with suitable adsorbent material such as clay or sand. Flush area with water to remove trace residue. Contain run-off from water fog and residue flush and dispose of properly.

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IF...EYE CONTACT THEN...Flush eyes with large amounts of water for at least 15 minutes while holding eyelids open. Rest eyes for 30 minutes. If redness, burning, blurred vision or swelling persists, get medical attention. IF...SKIN CONTACT THEN...Remove contaminated clothing and shoes. Flush affected area with large amounts of water for at least 15 minutes. Wash with soap if available. If blisters appear on the skin or deep pain develops, get medical attention. IF...INHALATION THEN...Move victim to fresh air. If the victim has difficulty breathing or tightness of the chest, is dizzy, vomiting or unresponsive, give 100% oxygen with rescue breathing or CPR as required and transport to the nearest medical facility. IF...INGESTION THEN...DO NOT GIVE LIQUIDS TO A DROWSY, CONVULSING OR UNCONSCIOUS PERSON. Otherwise, induce vomiting by administering 2 glasses of water and 2 tablespoons of Syrup of Ipecac. Keep head below hips while vomiting. Get medical attention. BENZENE (Clear, mobile liquid) Product Name Benzene 71-43-2

Flash Point (°F, TCC)

Product Synonyms Benzol; cyclohexatriene

15

HAZARD RATINGS Rating HMIS NFPA

Health 2 2

Fire 3 3

RECOMMENDED PERSONAL PROTECTIVE EQUIPMENT (Refer to the Introduction for general PPE guidelines.) Full protective clothing should be worn when handling benzene.

Reactivity 0 0

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Viton, polyvinyl alcohol, North Silver Shield (or equivalent) organic vapor

HEALTH HAZARDS Benzene is irritating to the eyes and irritating to the skin. Prolonged or repeated skin contact can result in defatting or drying of the skin with resulting irritation and dermatitis. Breathing of benzene vapors may cause central nervous system depression (dizziness, light-headedness, headache, nausea and loss of coordination). WARNING! Prolonged or repeated exposure to benzene may cause serious damage to the blood forming organs resulting in anemia and certain forms of leukemia. Liquid can directly enter the lungs (aspiration) when swallowed or vomited. Due to the low viscosity of this product, a serious and possibly fatal chemical pneumonia (aspiration pneumonitis) can develop if aspiration occurs. FIRE/EXPLOSION/REACTIVITY HAZARDS DANGER! EXTREMELY FLAMMABLE. This product may be ignited by heat, sparks, open flames or contact with strong oxidizing agents. Vapors may accumulate and travel to ignition sources distant from the handling site resulting in a flash fire. EMERGENCY ACTION GUIDE IF...FIRE THEN...Use water fog, foam, dry chemical or CO2 to extinguish flames. DO NOT use a direct stream of water; product will float and can be reignited on surface of water. Cool fire exposed containers with water spray. IF...SPILL OR LEAK THEN...Eliminate ignition sources. Wear appropriate personal protective equipment. Shut off leak if safe to do so. If vapor cloud forms, use water fog to suppress or blanket spill area with foam. Dike and contain spill. Monitor area with a combustible gas indicator. Remove with vacuum truck or pump to storage/recovery vessel. Soak up residue with suitable adsorbent material such as clay or sand. Flush area with water to remove trace residue. Contain run-off from water fog and residue flush and dispose of properly. IF...EYE CONTACT THEN...Flush eyes with large amounts of water for at least 15

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minutes while holding eyelids open. Rest eyes for 30 minutes. If redness, burning, blurred vision or swelling persists, get medical attention. IF...SKIN CONTACT THEN...Remove contaminated clothing and shoes. Flush affected area with large amounts of water. Wash with soap if available. If blisters appear on the skin, get medical attention. IF...INHALATION THEN...Move victim to fresh air. If the victim has difficulty breathing or tightness of the chest, is dizzy, vomiting or unresponsive, give 100% oxygen with rescue breathing or CPR as required and transport to the nearest medical facility. IF...INGESTION THEN...Have victim rinse mouth out with water, then drink sips of water to help remove the taste from the mouth. If the victim is coughing, choking, has shortness of breath, or difficulty breathing, get medical attention. If any of the following delayed signs and symptoms appear within the next 6 hours, get medical attention: fever greater than 101 °F, shortness of breath, chest congestion, coughing or wheezing. BISPHENOL A (White flakes) Product Name Bisphenol A 80-05-7

Flash Point (°F, TCC)

Product Synonyms BPA; bisphenol of acetone; 4,4'-(1- methylethylidene) bisphenol

405

HAZARD RATINGS Rating HMIS NFPA

Health 1 1

Fire 1 1

SPECIAL STORAGE AND HANDLING GUIDELINES (Refer to the Introduction for general guidelines.) Avoid creating dust clouds when handling. RECOMMENDED PERSONAL PROTECTIVE EQUIPMENT (Refer to the Introduction for general PPE guidelines.)

Reactivity 0 0

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HEALTH HAZARDS BPA dust may be moderately irritating to the eyes. BPA dust may be mildly irritating to the skin and may cause an allergic skin reaction after repeated exposure. BPA dust may be mildly irritating to the nose, throat and lungs. BPA may be moderately toxic if swallowed. FIRE/EXPLOSION/REACTIVITY HAZARDS BPA will not burn unless preheated. BPA dust may form explosive dust clouds. EMERGENCY ACTION GUIDE IF...FIRE THEN...Use water fog, foam, dry chemical or CO2 to extinguish flames. Cool fire exposed containers with water spray. IF...SPILL OR LEAK THEN...Wear appropriate personal protective equipment. Shovel and sweep up or use industrial vacuum cleaner; avoid creating dust clouds. Place in containers for recycling or disposal. IF...EYE CONTACT THEN...Flush eyes with water with large quantities of water for at least 15 minutes while holding eyelids open. Rest eyes for 30 minutes. If redness, burning, blurred vision or swelling persists, get medical attention. IF...SKIN CONTACT THEN...Remove contaminated clothing and shoes. Flush affected area with large amounts of water. Wash with soap if available. If blisters appear on the skin, get medical attention. IF...INHALATION THEN...Move victim to fresh air.

IF...INGESTION THEN...DO NOT GIVE LIQUIDS TO A DROWSY, CONVULSING OR UNCONSCIOUS PERSON. Otherwise, induce vomiting by administering 2 glasses of water followed by 2 tablespoons of Syrup of Ipecac. Keep head below hips while vomiting. Get

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medical attention. BUTYL DIOXITOLTM GLYCOL ETHER (Clear colorless liquid; slight odor) Product Name Butyl Dioxitol TM Glycol Ether 112-34-5

Flash Point (°F, TCC)

Product Synonyms Diethylene glycol monobutyl ether; DGBE

220

HAZARD RATINGS Rating HMIS NFPA

Health 1 1

Fire 1 2

Reactivity 0 0

SPECIAL STORAGE AND HANDLING GUIDELINES (Refer to the Introduction for general guidelines.) Avoid contact with aluminum surfaces. RECOMMENDED PERSONAL PROTECTIVE EQUIPMENT (Refer to the Introduction for general PPE guidelines.) Gloves: Respirator cartridge:

neoprene organic vapor

HEALTH HAZARDS Butyl Dioxitol is a moderate eye irritant. It may be slightly toxic if absorbed through skin and is a mild skin irritant. Butyl Dioxitol may be toxic if swallowed. FIRE/EXPLOSION/REACTIVITY HAZARDS This product will not burn unless preheated. Contact with aluminum may cause the release of flammable hydrogen gas. Butyl Dioxitol can form explosive peroxides on exposure to air. Runaway reaction can occur with Butyl Dioxitol at elevated temperatures in the presence of strong bases and salts of strong bases. EMERGENCY ACTION GUIDE IF...FIRE

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THEN...Use water fog, "alcohol" foam, dry chemical or CO2 to extinguish flames. Cool fire exposed containers with water spray. IF...SPILL OR LEAK THEN...Wear appropriate personal protective equipment. Shut off leak if safe to do so. Dike and contain spill. Remove with vacuum truck or pump to storage/recovery vessel. Soak up residue with suitable adsorbent material such as clay or sand. Flush area with water to remove trace residue. Contain run-off from residue flush and dispose of properly. IF...EYE CONTACT THEN...Flush eyes with large amounts of water for at least 15 minutes while holding eyelids open. Rest eyes for 30 minutes. If redness, burning, blurred vision or swelling persists, get medical attention. IF...SKIN CONTACT THEN...Remove contaminated clothing and shoes. Flush affected area with large amounts of water. Wash with soap if available. If blisters appear on the skin, get medical attention. IF...INHALATION THEN...Move victim to fresh air. If the victim has difficulty breathing or tightness of the chest, is dizzy, vomiting or unresponsive, give 100% oxygen with rescue breathing or CPR as appropriate and transport to the nearest medical facility. IF...INGESTION THEN...DO NOT GIVE LIQUIDS TO A DROWSY, CONVULSING OR UNCONSCIOUS PERSON. Otherwise, induce vomiting by administering 2 glasses of water and 2 tablespoons of Syrup of Ipecac. Keep head below hips while vomiting. Get medical attention. BUTYL OXITOL TM GLYCOL ETHER (Clear colorless liquid; slight odor) Product Name Butyl OxitolTM Glycol Ether 111-76-2

Product Synonyms 2-butoxyethanol; 2-BE; ethylene glycol monobutyl ether; EGBE

Flash Point (°F, TCC) 143

HAZARD RATINGS

file://C:\Shell%20Chemical%20-%20Online%20Literature.htm

21-04-01

Shell Chemical - Online Literature Rating HMIS NFPA

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Fire 2 2

Reactivity 0 0

SPECIAL STORAGE AND HANDLING GUIDELINES (Refer to the Introduction for general guidelines.) Avoid contact with aluminum surfaces. RECOMMENDED PERSONAL PROTECTIVE EQUIPMENT (Refer to the Introduction for general PPE guidelines.) Gloves: Respirator cartridge:

neoprene organic vapor

HEALTH HAZARDS Butyl Oxitol is a severe eye irritant. It is toxic if absorbed through skin and a mild skin irritant. Breathing Butyl Oxitol vapor may be toxic and may cause irritation of the nose, throat and lungs. Butyl Oxitol may be toxic if swallowed. FIRE/EXPLOSION/REACTIVITY HAZARDS CAUTION! COMBUSTIBLE. This product may be ignited by heat, sparks, open flames or contact with strong oxidizers. Vapors may accumulate and travel to ignition sources distant from the handling site resulting in a flash fire. Contact with aluminum may cause the release of flammable hydrogen gas. Butyl Oxitol can form explosive peroxides on exposure to air. Runaway reaction can occur at elevated temperatures in the presence of strong bases and salts of strong bases. EMERGENCY ACTION GUIDE IF...FIRE THEN...Use water fog, "alcohol" foam, dry chemical or CO2 to extinguish flames. Cool fire exposed containers with water spray. IF...SPILL OR LEAK THEN...Eliminate ignition sources. Wear appropriate personal protective equipment. Shut off leak if safe to do so. If vapor cloud forms, use water fog to suppress or blanket spill with "alcohol" foam.

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Dike and contain spill. Monitor area with a combustible gas indicator. Remove with vacuum truck or pump to storage/recovery vessel. Soak up residue with suitable adsorbent material such as clay or sand. Flush area with water to remove trace residue. Contain run-off from water fog and residue flush and dispose of properly. IF...EYE CONTACT THEN...Flush eyes with large amounts of water for at least 15 minutes while holding eyelids open. Rest eyes for 30 minutes. If redness, burning, blurred vision or swelling persists, get medical attention. IF...SKIN CONTACT THEN...Remove contaminated clothing and shoes. Flush affected area with large amounts of water. Wash with soap if available. If blisters appear on the skin, get medical attention. IF...INHALATION THEN...Move victim to fresh air. If the victim has difficulty breathing or tightness of the chest, is dizzy, vomiting or unresponsive, give 100% oxygen with rescue breathing or CPR as required and transport to the nearest medical facility. IF...INGESTION THEN...DO NOT GIVE LIQUIDS TO A DROWSY, CONVULSING OR UNCONSCIOUS PERSON. Otherwise, induce vomiting by administering 2 glasses of water and 2 tablespoons of Syrup of Ipecac. Keep head below hips while vomiting. Get medical attention. DIENES (Colorless gases; liquids under pressure) Product Name

Flash Point (°F, TCC)

Product Synonyms 1,3-butadiene;bivinyl; divinyl; vinylethylene; erythrene; biethylene; pyrrolylene 2-methyl-1, 3-butadiene

Butadiene 106-99-0 Isoprene 78-79-5

-105 -65

HAZARD RATINGS

Rating HMIS

Health 1

Fire 4

Reactivity 2

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4

2

SPECIAL STORAGE AND HANDLING GUIDELINES (Refer to the Introduction for general guidelines.) Storage and handling of these products should be in approved pressure vessels. Products must be chemically inhibited during storage and transport to prevent explosive peroxide formation. RECOMMENDED PERSONAL PROTECTIVE EQUIPMENT (Refer to the Introduction for general PPE guidelines.) Respirator cartridge:

organic vapors

HEALTH HAZARDS Direct contact with liquid butadiene can result in burns (frostbite) to the eyes and skin due to rapid evaporation of the liquid. Vapor may be irritating to the eyes. Isoprene may be irritating to the eyes. Prolonged or repeated liquid contact can result in defatting and drying of the skin which may result in skin irritation and dermatitis. Breathing high vapor concentrations are irritating to the nose, throat and respiratory tract, and may produce central nervous system depression (dizziness, light-headedness, headache, nausea and loss of coordination) and possibly loss of consciousness and death. 1,3-butadiene is listed by the National Toxicology Program as a chemical "reasonably anticipated to be a carcinogen" and the International Agency for Research on Cancer as a probable human carcinogen. FIRE/EXPLOSION/REACTIVITY HAZARDS DANGER! EXTREMELY FLAMMABLE, REACTIVE. These products may be ignited by heat, sparks, open flames or contact with strong oxidizing agents. Vapors may form explosive peroxides on exposure to air. Vapors may accumulate and travel to ignition sources distant from the handling site resulting in a flash fire and explosion. These products may undergo spontaneous hazardous polymerization. Exposure of containers to elevated temperatures will result in a build-up of pressure which may cause the container to rupture. EMERGENCY ACTION GUIDE IF...FIRE THEN...Allow gas to burn if flow cannot be shut off safely. Cool fire exposed containers with water spray from a safe distance. If a sufficient water supply cannot be sustained, withdraw from the area. IF...SPILL OR LEAK

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THEN...Evacuate surrounding areas. Eliminate ignition sources. Wear appropriate personal protective equipment. Disperse vapor cloud with sustained water spray until leak is stopped. If a sufficient water supply cannot be sustained, withdraw from the area. IF...EYE CONTACT THEN...Flush eyes with large amounts of water for at least 15 minutes while holding eyelids open. Rest eyes for 30 minutes. If redness, burning, blurred vision or swelling persists, get medical attention. IF...SKIN CONTACT THEN...Remove contaminated clothing and shoes. Flush affected area with large amounts of water for at least 15 minutes. Wash with soap if available. If blisters appear on the skin, get medical attention. IF...INHALATION THEN...Move victim to fresh air. If the victim has difficulty breathing or tightness of the chest, is dizzy, vomiting or unresponsive, give 100% oxygen with rescue breathing or CPR as required and transport to the nearest medical facility. IF...INGESTION THEN...No action required. Not applicable for gases or extremely volatile liquids. EPICHLOROHYDRIN (Clear, mobile liquid, irritating odor) Product Name Epichlorohydrin 106-89-8

Flash Point (°F, TCC)

Product Synonyms ECH; epi; 1-chloro-2, 3-epoxy propane; chloromethyloxirane

87

HAZARD RATINGS Rating HMIS NFPA

Health 3 3

Fire 3 3

Reactivity 1 2

RECOMMENDED PERSONAL PROTECTION EQUIPMENT

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(Refer to the Introduction for general PPE guidelines.) Full protective clothing should be worn when handling epichlorohydrin. Gloves: Respirator cartridge:

butyl rubber organic vapor

HEALTH HAZARDS DANGER! Epichlorohydrin is corrosive to the eyes. Epichlorohydrin is corrosive to the skin and may be toxic if absorbed through the skin; may cause kidney and liver damage. It may cause an allergic skin

reaction. Breathing epichlorohydrin vapor is severely irritating to the nose, throat and lungs. May cause respiratory allergic reaction after repeated exposure. Breathing epichlorohydrin vapor is toxic; may cause liver, kidney, and lung damage, and central nervous system depression (dizziness, light-headedness, headache, nausea and loss of coordination). Epichlorohydrin may be harmful or fatal if swallowed; may produce liver and kidney damage, and central nervous system depression. Epichlorohydrin is classified as a probable human carcinogen. FIRE/EXPLOSION/REACTIVITY HAZARDS WARNING! FLAMMABLE, REACTIVE. This product may be ignited by heat, sparks, open flames or contact with strong oxidizing agents. Vapors may accumulate and travel to ignition sources distant from the handling site resulting in a flash fire. Epichlorohydrin may undergo spontaneous hazardous polymerization reaction. Avoid epichlorohydrin contact with acids, bases, ammonia, amines, Lewis and Bronsted acids and bases, and metals such as aluminum, copper, magnesium, zinc, lead, and their alloys. EMERGENCY ACTION GUIDE IF...FIRE THEN...Use water fog, foam, dry chemical or CO2 to extinguish flames. Cool fire exposed containers with water spray. IF...SPILL OR LEAK THEN...Eliminate ignition sources. Wear appropriate personal protective equipment. Shut off leak if safe to do so. If vapor cloud forms, use water fog to suppress or blanket spill area with foam. Dike and contain spill. Monitor area with combustible gas indicator. Remove with vacuum truck or pump to storage/recovery vessel. Soak up residue with suitable adsorbent material such as sand; DO NOT use clay. Flush area with water to remove trace residue. Contain run-off from water fog and residue flush and dispose of properly. IF...EYE CONTACT

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THEN...Flush eyes with large amounts of water for at least 15 minutes while holding eyelids open. Continue to rinse with water while on the way to get medical attention. IF...SKIN CONTACT THEN...Remove contaminated clothing and shoes. Flush affected area with large amounts of water for at least 15 minutes. Get medical attention. IF...INHALATION THEN...Move victim to fresh air. If the victim has difficulty breathing or tightness of the chest, is dizzy, vomiting or unresponsive, give 100% oxygen with rescue breathing or CPR as required and transport to the nearest medical facility. IF...INGESTION THEN...DO NOT GIVE LIQUIDS TO A DROWSY, CONVULSING OR UNCONSCIOUS PERSON. Otherwise, induce vomiting by administering 2 glasses of water and 2 tablespoons of Syrup of Ipecac. Keep head below hips while vomiting. Get medical attention. EPI-CURE TM CURING AGENTS (3000), 3100, 3200, 3300 AND 3500 SERIES) (Clear to yellow liquids with amine or ammonia odor) Product Name Epi-CureTM Curing Agents 3000, 3100, 3200, 3300, 3500 Series

Flash Point (°F, TCC)

Product Synonyms None

>200

HAZARD RATINGS Rating HMIS NFPA

Health 2 or 3 2 or 3

Fire 1 1

SPECIAL STORAGE AND HANDLING GUIDELINES (Refer to the Introduction for general guidelines.) Do not reuse contaminated leather articles such as shoes as they cannot be decontaminated. RECOMMENDED PERSONAL PROTECTIVE EQUIPMENT (Refer to the Introduction for general PPE guidelines.)

Reactivity 0 0

Shell Chemical - Online Literature Respirator cartridge:

Page 30 of 84 organic vapor

HEALTH HAZARDS (Refer to the MSDS for product specific health hazard information.) WARNING! Depending on specific composition, these products may range from irritating to corrosive to the eyes, skin, respiratory tract and digestive

tract. These products may be harmful if absorbed through the skin, breathed or swallowed resulting in possible organ damage. There is a potential for skin and/or respiratory allergic reactions after repeated exposure to these curing agents. FIRE/EXPLOSION/REACTIVITY HAZARDS These products will not burn unless preheated. Heating in the presence of air may cause decomposition. Extreme heat producing and runaway reactions may occur in combination with some epoxy resins. EMERGENCY ACTION GUIDE IF...FIRE THEN...Use water fog, foam, dry chemical or CO2 to extinguish flames. Cool fire exposed containers with water spray. IF...SPILL OR LEAK THEN...Wear appropriate personal protective equipment. Shut off leak if safe to do so. Dike and contain spill. Remove with vacuum truck or pump to storage/recovery vessel. Soak up residue with suitable adsorbent material such as clay or sand. Flush area with water to remove trace residue. Contain run-off from water fog and residue flush and dispose of properly. IF...EYE CONTACT THEN...Flush eyes with large amounts of water for at least 15 minutes while holding eyelids open. Continue to rinse with water while on the way to get medical attention. IF...SKIN CONTACT THEN...Remove contaminated clothing and shoes. Flush affected area with large amounts of water for at least 15 minutes. Wash with soap if available. Get medical attention. IF...INHALATION THEN...Move victim to fresh air. If the victim has difficulty breathing or tightness of the chest, is dizzy, vomiting or unresponsive, give 100% oxygen with rescue breathing or CPR as required

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and transport to the nearest medical facility. IF...INGESTION THEN...DO NOT GIVE LIQUIDS TO A DROWSY, CONVULSING OR UNCONSCIOUS PERSON. Otherwise, induce vomiting by administering 2 glasses of water and 2 tablespoons of Syrup of Ipecac. Keep head below hips while vomiting. Get medical attention. EPI-CURE TM CURING AGENTS (3400 AND 9000 SERIES AND CURING AGENTS W, Y, Z) (Dark liquids with amine odor) Product Name

Flash Point (°F, TCC)

Product Synonyms

Epi-CureTM Curing Agents 3400, 9000 Series W, None Y, Z

>200

HAZARD RATINGS Rating HMIS NFPA

Health 2 or 3 2 or 3

Fire 1 1

Reactivity 0 0

SPECIAL STORAGE AND HANDLING GUIDELINES (Refer to the Introduction for general guidelines.) Do not reuse contaminated leather articles such as shoes as they cannot be decontaminated. RECOMMENDED PERSONAL PROTECTIVE EQUIPMENT (Refer to the Introduction for general PPE guidelines.) Respirator cartridge:

organic vapor

HEALTH HAZARDS (Refer to the MSDS for product specific health hazard information.) WARNING! Depending on specific composition, these products may range from irritating to corrosive to the eyes, skin, respiratory tract and digestive

tract. These products may be harmful if absorbed through the skin, breathed or swallowed resulting in possible organ damage. There is a potential for skin and/or respiratory allergic reactions after repeated exposure

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to these curing agents. These products may contain possible cancer causing agents. FIRE/EXPLOSION/REACTIVITY HAZARDS These products will not burn unless preheated. Avoid oxidizing agents and strong acids and bases. Extreme heat producing reactions may occur in combination with some epoxy resins. EMERGENCY ACTION GUIDE IF...FIRE THEN...Use water fog, foam, dry chemical or CO2 to extinguish flames. Cool fire exposed containers with water spray. IF...SPILL OR LEAK THEN...Wear appropriate personal protective equipment. Shut off leak if safe to do so. Dike and contain spill. Remove with vacuum truck or pump to storage/recovery vessel. Soak up residue with suitable adsorbent material such as clay or sand. Flush area with water to remove trace residue. Contain run-off from water fog and residue flush and dispose of properly. IF...EYE CONTACT THEN...Flush eyes with large amounts of water for at least 15 minutes while holding eyelids open. Continue to rinse with water while on the way to get medical attention. IF...SKIN CONTACT THEN...Remove contaminated clothing and shoes. Flush affected area with large amounts of water for at least 15 minutes. Wash with soap if available. Get medical attention. IF...INHALATION THEN...Move victim to fresh air. If the victim has difficulty breathing or tightness of the chest, is dizzy, vomiting or unresponsive, give 100% oxygen with rescue breathing or CPR as required and transport to the nearest medical facility. IF...INGESTION THEN...DO NOT GIVE LIQUIDS TO A DROWSY, CONVULSING OR UNCONSCIOUS PERSON. Otherwise, induce vomiting by administering 2 glasses of water and 2 tablespoons of Syrup of Ipecac. Keep head below hips while vomiting. Get medical attention. EPI-CURE TM CURING AGENT SOLVENT SOLUTIONS

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(Mobile liquids) Product Name Epi-CureTM Curing Agent Solvent Solutions

Flash Point (°F, TCC)

Product Synonyms None

<200

HAZARD RATINGS Rating HMIS NFPA

Health 2 or 3 2 or 3

Fire 2 or 3 2 or 3

Reactivity 0 0

SPECIAL STORAGE AND HANDLING GUIDELINES (Refer to the Introduction for general guidelines.) Do not reuse contaminated leather articles such as shoes as they cannot be decontaminated. RECOMMENDED PERSONAL PROTECTIVE EQUIPMENT (Refer to the Introduction for general PPE guidelines.) Respirator cartridge:

organic vapor

HEALTH HAZARDS (Refer to the MSDS for product specific health hazard information.)

tract.

WARNING! Depending on specific composition, these products may range from irritating to corrosive to the eyes, skin, respiratory tract and digestive

These products may be harmful if absorbed through the skin, breathed or swallowed resulting in possible organ damage. There is a potential for skin and/or respiratory allergic reactions after repeated exposure to these curing agents. FIRE/EXPLOSION/REACTIVITY HAZARDS (Refer to the product MSDS for specific information.) These products range from Combustible to Flammable depending on the solvent used. They may be ignited by heat, sparks or open flames. Avoid contact with oxidizing agents and strong acids and bases. Extreme heat producing reactions may occur in combination with some epoxy resins. EMERGENCY ACTION GUIDE

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IF...FIRE THEN...Use water fog, foam, dry chemical or CO2 to extinguish flames. Cool fire exposed containers with water spray. IF...SPILL OR LEAK THEN...Eliminate ignition sources. Wear appropriate personal protective equipment. Shut off leak if safe to do so. If vapor cloud forms, use water fog to suppress or blanket spill area with foam. Dike and contain spill. Monitor area with a combustible gas indicator. Remove with vacuum truck or pump to storage/recovery vessel. Soak up residue with suitable adsorbent material such as clay or sand. Flush area with water to remove trace residue. Contain run-off from water fog and residue flush and dispose of properly. IF...EYE CONTACT THEN...Flush eyes with large amounts of water for at least 15 minutes while holding eyelids open. Continue to rinse with water while on way to get medical attention. IF...SKIN CONTACT THEN...Remove contaminated clothing and shoes. Flush affected area with large amounts of water for at least 15 minutes. Wash with soap if available. Get medical attention. IF...INHALATION THEN...Move victim to fresh air. If the victim has difficulty breathing or tightness of the chest, is dizzy, vomiting or unresponsive, give 100% oxygen with rescue breathing or CPR as required and transport to the nearest medical facility. IF...INGESTION THEN...DO NOT INDUCE VOMITING. DO NOT GIVE LIQUIDS TO A DROWSY, CONVULSING OR UNCONSCIOUS PERSON. Otherwise give one glass of water. If vomiting occurs spontaneously, keep head below hips while vomiting. Get medical attention. EPI-CURE TM SERIES P CURING AGENTS (Powder) Product Name Epi-CureTM Series P Curing Agents

Product Synonyms None

Flash Point (°F, TCC) >200

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HAZARD RATINGS Rating HMIS NFPA

Health 1 1

Fire 1 1

SPECIAL STORAGE AND HANDLING GUIDELINES (Refer to the Introduction for general guidelines.) Avoid creating dust clouds when handling. RECOMMENDED PERSONAL PROTECTIVE EQUIPMENT (Refer to the Introduction for general PPE guidelines.) Respirator cartridge:

organic vapor

HEALTH HAZARDS These products may be irritating to the eyes. Series P curing agents may be irritating to the skin. Dust may be mildly irritating to the nose, throat and lungs. FIRE/EXPLOSION/REACTIVITY HAZARDS These products will not burn unless preheated. Products may be heat sensitive. EMERGENCY ACTION GUIDE IF...FIRE THEN...Use water fog, "alcohol" foam, dry chemical or CO2 to extinguish flames. Cool fire exposed containers with water spray. IF...SPILL OR LEAK THEN...Shovel and sweep up or use industrial vacuum cleaner; avoid creating dust clouds. Place in containers for recycling or disposal. IF...EYE CONTACT THEN...Flush eyes with water while holding eyelids open. Rest eyes for 30 minutes. If redness, burning, blurred vision or swelling persists, get medical attention. IF...SKIN CONTACT

Reactivity 0 0

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THEN...For contact with molten product, cool with water and treat as a thermal burn. DO NOT attempt to remove solidified material from the burn area. Get medical attention. IF...INHALATION THEN...Move victim to fresh air.

IF...INGESTION THEN...No action is ordinarily required. EPON TM RESINS (Viscous liquids) Product Name Epon TM Resins

Flash Point (°F, SF)

Product Synonyms Bisphenol A/epichlorohydrin based epoxy resins

>200

HAZARD RATINGS Rating HMIS NFPA

Health 2 2

Fire 1 1

Reactivity 0 0

SPECIAL STORAGE AND HANDLING GUIDELINES (Refer to the Introduction for general guidelines.) Do not reuse contaminated leather articles such as shoes as they cannot be decontaminated. HEALTH HAZARDS (Refer to the MSDS for product specific health hazard information.) Epon resins may be moderately irritating to eyes and skin. Contact with these products at elevated temperature may result in thermal burns. There is a potential for a skin allergy to develop after repeated exposure to some of these products. Epon resins are no more than slightly toxic if swallowed. FIRE/EXPLOSION/REACTIVITY HAZARDS Products will not burn unless preheated.

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These products can react vigorously with strong oxidizing agents, strong Lewis or mineral acids and strong mineral and organic bases (especially primary and secondary amines). EMERGENCY ACTION GUIDE IF...FIRE THEN...Use water fog, foam, dry chemical or CO2 to extinguish flames. DO NOT use a direct stream of water; product will float and can be reignited on surface of water. Cool fire exposed containers with water spray. IF...SPILL OR LEAK THEN...Wear appropriate personal protective equipment. Shut off leak if safe to do so. Dike and contain spill. Remove with vacuum truck or pump to storage/recovery vessel. Soak up residue with suitable adsorbent material such as clay or sand. Flush area with water to remove trace residue. Contain run-off from residue flush and dispose of properly. IF...EYE CONTACT THEN...Flush eyes with water while holding eyelids open. Rest eyes for 30 minutes. If redness, burning, blurred vision or swelling persists, get medical attention. IF...SKIN CONTACT THEN...Remove contaminated clothing and shoes. Flush affected area with large amounts of water. Wash with soap if available. If blisters appear on the skin, get medical attention. IF...INHALATION THEN...Remove victim to fresh air.

IF...INGESTION THEN...Have victim drink sips of water ot help remove the taste from the mouth. In general no treatment is necessary unless large quantities are ingested. However, get medical advice. EPON TM RESIN SOLUTIONS/EPONOL TM RESINS (Light colored, mobile liquids) Product Name Epon TM Resin Solutions/

Product Synonyms None

Flash Point (°F, SF) -15 to 126

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Eponol TM Resins

HAZARD RATINGS Rating HMIS NFPA

Health 2 2

Fire 2 or 3 2 or 3

Reactivity 0 0

SPECIAL STORAGE AND HANDLING GUIDELINES (Refer to the Introduction for general guidelines.) Do not reuse contaminated leather articles such as shoes as they cannot be decontaminated. RECOMMENDED PERSONAL PROTECTIVE EQUIPMENT (Refer to the Introduction for general PPE guidelines.) Respirator cartridge:

organic vapor

HEALTH HAZARDS (Refer to the MSDS for product specific health hazard information.) These products may be irritating to the eyes and skin. Prolonged or repeated skin contact with the solvents used in these products can result in defatting or drying of the skin with resulting irritation and dermatitis. Some of the solvents used in these products may be absorbed by the skin. Inhalation of solvent vapors may cause central nervous system depression (dizziness, light-headedness, headache, nausea and loss of coordination). These products may be toxic if swallowed producing central nervous system depression. Skin absorption, inhalation or ingestion of the solvent used in the Eponol Resin J series may produce damage to the liver, kidney, blood forming organs and testes; may produce reproductive toxicity. FIRE/EXPLOSION/REACTIVITY HAZARDS (Refer to the product MSDS for specific physical hazard information.) WARNING! COMBUSTIBLE TO FLAMMABLE. These products may be ignited by heat, sparks, open flames or contact with strong oxidizing agents. Vapors may accumulate and travel to ignition sources distant from the handling site resulting in a flash fire. Products may react vigorously with strong Lewis or mineral acids. EMERGENCY ACTION GUIDE

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IF...FIRE THEN...Use water fog, "alcohol" foam, dry chemical or CO2 to extinguish flames. Cool fire exposed containers with water spray. IF...SPILL OR LEAK THEN...Eliminate ignition sources. Wear appropriate personal protective equipment. Shut off leak if safe to do so. If vapor cloud forms, use water fog to suppress or blanket spill with "alcohol" foam. Dike and contain spill. Monitor area with a combustible gas indicator. Remove with vacuum truck or pump to storage/recovery vessel. Soak up residue with suitable adsorbent material such as clay or sand. Flush area with water to remove trace residue. Contain run-off from water fog and residue flush and dispose of properly. IF...EYE CONTACT THEN...Flush eyes with large amounts of water for at least 15 minutes while holding eyelids open. Rest eyes for 30 minutes. If redness, burning, blurred vision or swelling persists, get medical attention. IF...SKIN CONTACT THEN...Remove contaminated clothing and shoes. Flush affected area with large amounts of water. Wash with soap if available. If blisters appear on the skin, get medical attention. IF...INHALATION THEN...Move victim to fresh air. If the victim has difficulty breathing or tightness of the chest, is dizzy, vomiting or unresponsive, give 100% oxygen with rescue breathing or CPR as required and transport to the nearest medical facility. IF...INGESTION THEN...DO NOT GIVE LIQUIDS TO A DROWSY, CONVULSING OR UNCONSCIOUS PERSON. Otherwise, induce vomiting by administering 2 glasses of water and 2 tablespoons of Syrup of Ipecac. Keep head below hips while vomiting. Get medical attention. ETHYLENE OXIDE (Clear liquid below 51 °F, gas above; ether-like odor) Product Name Ethylene oxide 75-21-8

Product Synonyms EO; oxirane; 1, 2-epoxy ethane

Flash Point (°F, TCC) <0

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HAZARD RATINGS Rating HMIS NFPA

Health 3 2

Fire 4 4

Reactivity 3 3

SPECIAL STORAGE AND HANDLING GUIDELINES (Refer to the Introduction for general guidelines.) Storage and handling of this product should be in approved pressure vessels and pressure-rated systems. RECOMMENDED PERSONAL PROTECTIVE EQUIPMENT (Refer to the Introduction for general PPE guidelines.) Full protective clothing should be worn when handling ethylene oxide. FEPpolycarbonate composites provide the best protection for visors on EOresistant suits. Gloves: Respirator:

Safety 4-H, butyl rubber NIOSH has approved a full facepiece canister with end-of-service life indicator for concentrations <50 ppm.

CAUTION! The odor threshold for EO is above the occupational exposure limits. If you can smell EO, leave the area immediately. For additional information on protective clothing and respirator use consult the MSDS. HEALTH HAZARDS Ethylene oxide is severely irritating to the eyes. Ethylene oxide is severely irritating to the skin and may cause delayed skin burns even if dilute solutions of EO remain on the skin for short periods of time. Breathing ethylene oxide vapor is severely irritating to the nose, throat and respiratory tract and may be toxic; may cause central nervous system depression (dizziness, lightheadedness, headache, nausea and loss of coordination) and peripheral nerve damage. Classified as a probable human carcinogen and a reproductive hazard. FIRE/EXPLOSION/REACTIVITY HAZARDS DANGER! EXTREMELY FLAMMABLE, REACTIVE. This product may be ignited by heat, sparks, open flames or contact with strong oxidizing agents. EO vapors are explosive. Vapors may accumulate and travel to ignition sources distant from the handling site resulting in a flash fire and explosion. Ethylene oxide may undergo spontaneous hazardous polymerization. Avoid EO contact with acids, organic bases, ammonia, alkali metal hydroxides, metallic

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potassium, iron dust, iron and aluminum oxide and anhydrous chlorides of iron, tin and aluminum. EMERGENCY ACTION GUIDE IF...FIRE THEN...Allow gas to burn if flow cannot be shut off safely. Cool fire exposed containers with water spray from a safe distance. If a sufficient water supply cannot be sustained, withdraw from the area. IF...SPILL OR LEAK THEN...Evacuate surrounding areas in all directions to at least 600 feet and downwind to at least 2 miles. Eliminate ignition sources. Wear appropriate personal protective equipment. Disperse vapor cloud with sustained water spray (at least 30:1 water dilution) until leak is stopped. If a sufficient water supply cannot be sustained, withdraw from the area. IF...EYE CONTACT THEN...Flush eyes with large amounts of water for at least 15 minutes while holding eyelids open. Rest eyes for 30 minutes. If redness, burning, blurred vision or swelling persists, get medical attention. IF...SKIN CONTACT THEN...Remove contaminated clothing and shoes. Flush affected area with large amounts of water for at least 15 minutes. Wash with soap if available. If blisters appear on the skin, get medical attention. IF...INHALATION THEN...Move victim to fresh air. If the victim has difficulty breathing or tightness of the chest, is dizzy, vomiting or unresponsive, give 100% oxygen with rescue breathing or CPR as required and transport to the nearest medical facility. IF...INGESTION THEN...No action required. Not applicable for gases or extremely volatile liquids. GLYCOLS (Slightly viscous liquids; mild odor; sweet taste) Product Name Ethylene Glycol* 107-21-1 Diethylene Glycol

Product Synonyms EG; antifreeze; ShellzoneTM; 1,2-ethanediol DEG; 2,2'-oxybisethanol;

Flash Point (°F, TCC) 240

Shell Chemical - Online Literature Diethylene Glycol 111-46-6 Triethylene Glycol 112-27-6 Hexylene Glycol* 107-41-5 Shell Solvent Blend SC- 0612*

Page 42 of 84 DEG; 2,2'-oxybisethanol; 2,2'-oxydiethanol TEG; 2,2'ethylenedioxybis (ethanol) HG; 2-methyl-2, 4-pentanediol; pinakon Blend of diethylene glycol and hexylene glycol

290 330 211 211 to 244

HAZARD RATINGS Rating HMIS NFPA

Health 2 *, 1 1

Fire 1 1

Reactivity 0 0

RECOMMENDED PERSONAL PROTECTIVE EQUIPMENT (Refer to the Introduction for general PPE guidelines.) Gloves: Respirator cartridge:

ethylene glycol , neoprene, nitrile rubber, natural rubber, PVC organic vapor

HEALTH HAZARDS (Refer to the MSDS for product specific health hazard information.) Glycols are from moderately to severely irritating to the eyes. They are mild to moderate skin irritants and slightly toxic upon prolonged or repeated contact. Breathing glycol vapors may cause irritation of the nose, throat and lungs and may be toxic upon prolonged or repeated exposure. Swallowing glycols may be HARMFUL or FATAL producing central nervous system depression (dizziness, light-headedness, headache, nausea and loss of coordination), kidney damage and liver damage. Swallowing just a few ounces of ethylene glycol can be fatal. Glycols have a sweet taste making them a particularly dangerous ingestion hazard for children and pets. FIRE/EXPLOSION/REACTIVITY HAZARDS Products will not burn unless preheated. EMERGENCY ACTION GUIDE IF...FIRE

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THEN...Use water fog, "alcohol" foam, dry chemical or CO2 to extinguish flames. Cool fire exposed containers with water spray. IF...SPILL OR LEAK THEN...Wear appropriate personal protective equipment. Shut off leak if safe to do so. Dike and contain spill. Remove with vacuum truck or pump to storage/recovery vessel. Soak up residue with suitable adsorbent material such as clay or sand. Flush area with water to remove trace residue. Contain run-off from residue flush and dispose of properly. IF...EYE CONTACT THEN...Flush eyes with large amounts of water for at least 15 minutes while holding eyelids open. Rest eyes for 30 minutes. If redness, burning, blurred vision or swelling persists, get medical attention. IF...SKIN CONTACT THEN...Remove contaminated clothing and shoes. Flush affected area with large amounts of water. Wash with soap if available. If blisters appear on the skin, get medical attention. IF...INHALATION THEN...Move victim to fresh air. If the victim has difficulty breathing or tightness of the chest, is dizzy, vomiting or unresponsive, give 100% oxygen with rescue breathing or CPR as appropriate and transport to the nearest medical facility. IF...INGESTION THEN...DO NOT GIVE LIQUIDS TO A DROWSY, CONVULSING OR UNCONSCIOUS PERSON. Otherwise, induce vomiting by administering 2 glasses of water and 2 tablespoons of Syrup of Ipecac. Keep head below hips while vomiting. Get medical attention. HELOXY TM MODIFIERS (Light colored, mobile liquids) Product Name HeloxyTM Modifiers

Flash Point (°F, SF)

Product Synonyms Aliphatic and aromatic glycidyl ethers

>200

HAZARD RATINGS Rating HMIS

Health 2 or 3

Fire 1

Reactivity 0

Shell Chemical - Online Literature NFPA

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1

0

SPECIAL STORAGE AND HANDLING GUIDELINES (Refer to the Introduction for general guidelines.) DO NOT reuse contaminated leather articles such as shoes as they cannot be decontaminated. RECOMMENDED PERSONAL PROTECTIVE EQUIPMENT (Refer to the Introduction for general PPE guidelines.) Respirator cartridge:

organic vapor

HEALTH HAZARDS (Refer to the MSDS for product specific health hazard information.) WARNING! Depending on the specific formulation, these products may be from irritating to corrosive to the eyes and skin. There is a potential for a skin allergy to develop after repeated skin exposure to these products. Inhalation of mists may produce irritation of the nose, throat and respiratory tract, and central nervous system depression (dizziness, light-headedness, headache, nausea and loss of coordination). Heloxy Modifiers may be toxic if swallowed. They may produce central nervous system depression and other toxic effects. Heloxy Modifiers 63, 69 and 163 are classified as possible human carcinogens (cancercausing agents). FIRE/EXPLOSION/REACTIVITY HAZARDS Products must be preheated to burn. Heloxy Modifier 61 is a combustible liquid with a flash point of 130 °F. EMERGENCY ACTION GUIDE IF...FIRE THEN...Use water fog, foam, dry chemical or CO2 to extinguish flames. Cool fire exposed containers with water spray. IF...SPILL OR LEAK THEN...Wear appropriate personal protective equipment. Shut off leak if safe to do so. Dike and contain spill. Remove with vacuum truck or pump to storage/recovery vessel. Soak up residue with suitable adsorbent material such as clay or sand. Flush area with water to remove trace residue. Contain run-off from water fog and residue flush and dispose of properly.

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IF...EYE CONTACT THEN...Flush eyes with large amounts of water for at least 15 minutes while holding eyelids open. Continue to rinse with water while on the way to get medical attention. IF...SKIN CONTACT THEN...Remove contaminated clothing and shoes. Flush affected area with large amounts of water for at least 15 minutes. Wash with soap if available. Get medical attention. IF...INHALATION THEN...Move victim to fresh air. If the victim has difficulty breathing or tightness of the chest, is dizzy, vomiting or unresponsive, give 100% oxygen with rescue breathing or CPR as required and transport to the nearest medical facility. IF...INGESTION THEN...DO NOT GIVE LIQUIDS TO A DROWSY, CONVULSING OR UNCONSCIOUS PERSON. Otherwise, induce vomiting by administering 2 glasses of water and 2 tablespoons of Syrup of Ipecac. Keep head below hips while vomiting. Get medical attention. HYDROCARBON SOLVENTS, AROMATIC (Clear, mobile liquids) Product Name Toluene 108-88-3 Xylene 1330-20-7 Cyclo-Sol

TM

28

Flash Point (°F, TCC)

Product Synonyms Methylbenzene; toluol; phenylmethane

40

Dimethylbenzene; xylol

76

Toluene/xylene blend with a light aliphatic solvent naphtha

21

HAZARD RATINGS Rating HMIS NFPA

Health 1 2

Fire 3 3

RECOMMENDED PERSONAL PROTECTIVE EQUIPMENT (Refer to the Introduction for general PPE guidelines.)

Reactivity 0 0

Shell Chemical - Online Literature Gloves: Respirator cartridge:

Page 46 of 84 polyvinyl alcohol, Viton organic vapor

HEALTH HAZARDS These products are moderate eye irritants and mild skin irritants. Prolonged or repeated skin contact can result in defatting or drying of the skin with resulting irritation and dermatitis. Breathing vapors may cause central nervous system depression (dizziness, lightheadedness, headache, nausea and loss of coordination). Liquid can directly enter the lungs (aspiration) when swallowed or vomited. Due to the low viscosity of these products, a serious and possibly fatal chemical pneumonia (aspiration pneumonitis) can develop if aspiration occurs. FIRE/EXPLOSION/REACTIVITY HAZARDS WARNING! FLAMMABLE. This product may be ignited by heat, sparks, open flames or contact with strong oxidizing agents. Vapors may accumulate and travel to ignition sources distant from the handling site resulting in a flash fire. EMERGENCY ACTION GUIDE IF...FIRE THEN...Use water fog, foam, dry chemical or CO2 to extinguish flames. DO NOT use a direct stream of water; product will float and can be reignited on surface of water. Cool fire exposed containers with water spray. IF...SPILL OR LEAK THEN...Eliminate ignition sources. Wear appropriate personal protective equipment. Shut off leak if safe to do so. If vapor cloud forms, use water fog to suppress or blanket spill area with foam. Dike and contain spill. Monitor area with a combustible gas indicator. Remove with vacuum truck or pump to storage/recovery vessel. Soak up residue with suitable adsorbent material such as clay or sand. Flush area with water to remove trace residue. Contain run-off from water fog and residue flush and dispose of properly. IF...EYE CONTACT THEN...Flush eyes with large amounts of water for at least 15 minutes while holding eyelids open. Rest eyes for 30 minutes. If redness, burning, blurred vision or swelling persists, get medical attention. IF...SKIN CONTACT THEN...Remove contaminated clothing and shoes. Flush affected area with large amounts of water. Wash with soap if file://C:\Shell%20Chemical%20-%20Online%20Literature.htm

21-04-01

Shell Chemical - Online Literature

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available. If blisters appear on the skin, or dizziness, vomiting or tightness of the chest occurs, get medical attention. IF...INHALATION THEN...Move victim to fresh air. If the victim has difficulty breathing or tightness of the chest, is dizzy, vomiting or unresponsive, give 100% oxygen with rescue breathing or CPR as required and transport to the nearest medical facility. IF...INGESTION THEN...Have victim rinse mouth out with water, then drink sips of water to help remove the taste from the mouth. If the victim is coughing, choking, has shortness of breath, or difficulty breathing, get medical attention. If any of the following delayed signs and symptoms appear within the next 6 hours, get medical attention: fever greater than 101 °F, shortness of breath, chest congestion, coughing or wheezing. HYDROCARBON SOLVENT BLENDS, COMBUSTIBLE (Clear, mobile liquids) Product Synonyms

Flash Point (°F, TCC)

Shell Mineral Spirits 135, 145-EC, 150 EC, 146 HT, 200 HT, 160, 1168, 1169, 147 EC; Shell Sol 71, 340-EC, 340 HT, 323-EC, 142 HT, 137, 130 B, WT, D60, FC

Medium aliphatic solvent naphtha and light aromatic solvent naphtha blends

102 to 147

Cyclo-Sol TM 42, 47, 53, 63; Shell TS-28, TS-28B, TS-28R

Medium aliphatic solvent naphtha, light and heavy aromatic solvent naphtha blends

103 to 133

Product Name

HAZARD RATINGS Rating HMIS NFPA

Health 1 0

Fire 2 2

RECOMMENDED PERSONAL PROTECTIVE EQUIPMENT (Refer to the Introduction for general PPE guidelines.) Respirator cartridge:

organic vapor

Reactivity 0 0

Shell Chemical - Online Literature

Page 48 of 84

HEALTH HAZARDS (Refer to the MSDS for product specific health hazard information.) These products are minimally irritating to the eyes and slightly irritating to the skin. Prolonged or repeated skin contact can result in defatting or drying of the skin with resulting irritation and dermatitis. Breathing vapors may cause central nervous system depression (dizziness, lightheadedness, headache, nausea and loss of coordination). Liquid can directly enter the lungs (aspiration) when swallowed or vomited. Due to the low viscosity of these products, a serious and possibly fatal chemical pneumonia (aspiration pneumonitis) can develop if aspiration occurs. FIRE/EXPLOSION/REACTIVITY HAZARDS CAUTION! COMBUSTIBLE. These products may be ignited by heat, sparks, open flames or contact with strong oxidizing agents. Vapors may accumulate and travel to ignition sources distant from the handling site resulting in a flash fire. EMERGENCY ACTION GUIDE IF...FIRE THEN...Use water fog, foam, dry chemical or CO2 to extinguish flames. DO NOT use a direct stream of water; product will float and can be reignited on surface of water. Cool fire exposed containers with water spray. IF...SPILL OR LEAK THEN...Eliminate ignition sources. Wear appropriate personal protective equipment. Shut off leak if safe to do so. If vapor cloud forms, use water fog to suppress or blanket spill area with foam. Dike and contain spill. Monitor area with a combustible gas indicator. Remove with vacuum truck or pump to storage/recovery vessel. Soak up residue with suitable adsorbent material such as clay or sand. Flush area with water to remove trace residue. Contain run-off from water fog and residue flush and dispose of properly. IF...EYE CONTACT THEN...Flush eyes with large amounts of water for at least 15 minutes while holding eyelids open. Rest eyes for 30 minutes. If redness, burning, blurred vision or swelling persists, get medical attention. IF...SKIN CONTACT THEN...Remove contaminated clothing and shoes. Flush affected area with large amounts of water. Wash with soap if available. If blisters appear on the skin, or dizziness, vomiting or tightness of the chest occurs, get medical attention.

Shell Chemical - Online Literature

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IF...INHALATION THEN...Move victim to fresh air. If the victim has difficulty breathing or tightness of the chest, is dizzy, vomiting or unresponsive, give 100% oxygen with rescue breathing or CPR as required and transport to the nearest medical facility. IF...INGESTION THEN...Have victim rinse mouth out with water, then drink sips of water to help remove the taste from the mouth. If the victim is coughing, choking, has shortness of breath, or difficulty breathing, get medical attention. If any of the following delayed signs or symptoms appear within the next 6 hours, get medical attention: fever greater than 101 °F, shortness of breath, chest congestion, coughing or wheezing. HYDROCARBON SOLVENT BLENDS, EXTREMELY FLAMMABLE (Clear, mobile liquids; hydrocarbon odor) Product Name Shell Sol BT -9, BT-67 EC, BT-4, BT 67W EC, B-8, Rubber Solvent, all Tolu-Sol TM products Shell Sol B HT, BJ-8, BT-20, BT-40, BT-60, BT-75, BT-83, WRG, Rubber Solvent 332

Flash Point (°F, TCC)

Product Synonyms Light aliphatic solvent naphtha and toluene blends

0 to 20

Light aliphatic solvent naphtha

<0

HAZARD RATINGS Rating HMIS NFPA

Health 1 1

Fire 3 3

RECOMMENDED PERSONAL PROTECTIVE EQUIPMENT (Refer to the Introduction for general PPE guidelines.) Respirator cartridge:

organic vapor

HEALTH HAZARDS (Refer to the MSDS for product specific health hazard information.)

Reactivity 0 0

Shell Chemical - Online Literature

Page 50 of 84

These products are from minimally to moderately irritating to the eyes. They are mildly irritating to the skin. Prolonged or repeated skin contact can result in defatting or drying of the skin with resulting irritation and dermatitis. Breathing vapors may cause central nervous system depression (dizziness, lightheadedness, headache, nausea and loss of coordination). Prolonged and repeated inhalation of n-hexane (a component of these products) may produce peripheral and central nervous system damage. Liquid can directly enter the lungs (aspiration) when swallowed or vomited. Due to the low viscosity of these products, a serious and possibly fatal chemical pneumonia (aspiration pneumonitis) can develop if aspiration occurs. FIRE/EXPLOSION/REACTIVITY HAZARDS DANGER! EXTREMELY FLAMMABLE. These products may be ignited by heat, sparks, open flames or contact with strong oxidizing agents. Vapors may accumulate and travel to ignition sources distant from the handling site resulting in a flash fire. EMERGENCY ACTION GUIDE IF...FIRE THEN...Use water fog, foam, dry chemical or CO2 to extinguish flames. DO NOT use a direct stream of water; product will float and can be reignited on surface of water. Cool fire exposed containers with water spray. IF...SPILL OR LEAK THEN...Eliminate ignition sources. Wear appropriate personal protective equipment. Shut off leak if safe to do so. If vapor cloud forms, use water fog to suppress or blanket spill area with foam. Dike and contain spill. Monitor area with a combustible gas indicator. Remove with vacuum truck or pump to storage/recovery vessel. Soak up residue with suitable adsorbent material such as clay or sand. Flush area with water to remove trace residue. Contain run-off from water fog and residue flush and dispose of properly. IF...EYE CONTACT THEN...Flush eyes with large amounts of water for at least 15 minutes while holding eyelids open. Rest eyes for 30 minutes. If redness, burning, blurred vision or swelling persists, get medical attention. IF...SKIN CONTACT THEN...Remove contaminated clothing and shoes. Flush affected area with large amounts of water. Wash with soap if available. If blisters appear on the skin, or dizziness, vomiting or tightness of the chest occurs, get medical attention.

Shell Chemical - Online Literature

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IF...INHALATION THEN...Move victim to fresh air. If the victim has difficulty breathing or tightness of the chest, is dizzy, vomiting or unresponsive, give 100% oxygen with rescue breathing or CPR as required and transport to the nearest medical facility. IF...INGESTION THEN...Have victim rinse mouth out with water, then drink sips of water to help remove the taste from the mouth. If the victim is coughing, choking, has shortness of breath, or difficulty breathing, get medical attention. If any of the the following delayed signs and symptoms appear within the next 6 hours, get medical attention: fever greater than 101 °F, shortness of breath, chest congestion or continued coughing or wheezing. HYDROCARBON SOLVENT BLENDS, FLAMMABLE (Clear, mobile liquids; hydrocarbon odor) Product Name

Product Synonyms

Flash Point (°F, TCC)

Shell VM&P Naphtha 3M, VM&P EC, High Flash VM&P, Sol LF, Sol M-95 EC, Sol 320 Shell Sol M-75 EC, Sol MS-14, Sol PS, Sol RB, Sol TV, Sol ZP, VM&P HB, VM&P HT, Mineral Spirits 6855

Light, medium aliphatic solvent naphtha and toluene blends

22 to 93

Light and medium aliphatic solvent naphtha blends

25 to 100

HAZARD RATINGS Rating HMIS NFPA

Health 1 1

Fire 3 3

RECOMMENDED PERSONAL PROTECTIVE EQUIPMENT (Refer to the Introduction for general PPE guidelines.) Respirator cartridge:

organic vapor

HEALTH HAZARDS (Refer to the MSDS for product specific health hazard information.)

Reactivity 0 0

Shell Chemical - Online Literature

Page 52 of 84

These products are from minimally to moderately irritating to the eyes. They are mildly irritating to the skin. Prolonged or repeated skin contact can result in defatting or drying of the skin with resulting irritation and dermatitis. Breathing vapors may cause central nervous system depression (dizziness, lightheadedness, headache, nausea and loss of coordination). Prolonged and repeated inhalation of n-hexane (a component of these products) may produce peripheral and central nervous system damage. Liquid can directly enter the lungs (aspiration) when swallowed or vomited. Due to the low viscosity of these products, a serious and possibly fatal chemical pneumonia (aspiration pneumonitis) can develop if aspiration occurs. FIRE/EXPLOSION/REACTIVITY HAZARDS WARNING! FLAMMABLE. These products may be ignited by heat, sparks, open flames or contact with strong oxidizing agents. Vapors may accumulate and travel to ignition sources distant from the handling site resulting in a flash fire. EMERGENCY ACTION GUIDE IF...FIRE THEN...Use water fog, foam, dry chemical or CO2 to extinguish flames. DO NOT use a direct stream of water; product will float and can be reignited on surface of water. Cool fire exposed containers with water spray. IF...SPILL OR LEAK THEN...Eliminate ignition sources. Wear appropriate personal protective equipment. Shut off leak if safe to do so. If vapor cloud forms, use water fog to suppress or blanket spill area with foam. Dike and contain spill. Monitor area with a combustible gas indicator. Remove with vacuum truck or pump to storage/recovery vessel. Soak up residue with suitable adsorbent material such as clay or sand. Flush area with water to remove trace residue. Contain run-off from water fog and residue flush and dispose of properly. IF...EYE CONTACT THEN...Flush eyes with large amounts of water for at least 15 minutes while holding eyelids open. Rest eyes for 30 minutes. If redness, burning, blurred vision or swelling persists, get medical attention. IF...SKIN CONTACT THEN...Remove contaminated clothing and shoes. Flush affected area with large amounts of water. Wash with soap if available. If blisters appear on the skin, or dizziness, vomiting or tightness of the chest occurs, get medical attention. IF...INHALATION

Shell Chemical - Online Literature

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THEN...Move victim to fresh air. If the victim has difficulty breathing or tightness of the chest, is dizzy, vomiting or unresponsive, give 100% oxygen with rescue breathing or CPR as required and transport to the nearest medical facility. IF...INGESTION THEN...Have victim rinse mouth out with water, then drink sips of water to help remove the taste from the mouth. If the victim is coughing, choking, has shortness of breath, or difficulty breathing, get medical attention. If any of the following delayed signs and symptoms appear within the next 6 hours, get medical attention: fever greater than 101 °F, shortness of breath, chest congestion, coughing or wheezing. HYDROCHLORIC ACID (Clear liquid; pungent odor) Product Synonyms

Flash Point (°F, TCC)

HCl; Muriatic Acid; aqueous hydrogen chloride

N/A

Product Name Hydrochloric acid 7647-01-0

HAZARD RATINGS Rating HMIS NFPA

Health 3 3

Fire 0 0

Reactivity 0 0

SPECIAL STORAGE AND HANDLING GUIDELINES (Refer to the Introduction for general guidelines.) Keep away from oxidizing agents. RECOMMENDED PERSONAL PROTECTIVE EQUIPMENT (Refer to the Introduction for general PPE guidelines.) Gloves: Respirator cartridge:

chlorinated polyethylene acid gas

HEALTH HAZARDS Hydrochloric acid is severely irritating to the eyes and skin. Breathing HCl gas is severely irritating (possibly corrosive) to the nose, throat and respiratory tract and may cause permanent damage and may be fatal. If swallowed, HCl is severely irritating to the mouth, throat and stomach and

Shell Chemical - Online Literature

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may cause permanent damage and may be fatal. FIRE/EXPLOSION/REACTIVITY HAZARDS Reacts with oxidizers to release toxic chlorine gas; may react with metals to release flammable hydrogen gas. EMERGENCY ACTION GUIDE IF...FIRE THEN...Cool fire exposed containers with water spray. IF...SPILL OR LEAK THEN...Wear appropriate personal protective equipment. Shut off leak if safe to do so. If vapor cloud forms, use water fog to suppress. Dike and contain spill. Remove with vacuum truck or pump to storage/recovery vessel. Neutralize contaminated area with soda ash or lime. Flush area with water to remove trace residue. Contain run-off from water fog and residue flush and dispose of properly. IF...EYE CONTACT THEN...Flush eyes with large amounts of water for at least 15 minutes while holding eyelids open. Continue to rinse with water while on the way to get medical attention. IF...SKIN CONTACT THEN...Remove contaminated clothing and shoes. Flush affected area with large amounts of water for at least 15 minutes. Wash with soap if available. Get medical attention. IF...INHALATION THEN...Move victim to fresh air. If the victim has difficulty breathing or tightness of the chest, is dizzy, vomiting or unresponsive give 100% oxygen with rescue breathing or CPR as required and transport to the nearest medical facility. IF...INGESTION THEN...DO NOT INDUCE VOMITING. DO NOT GIVE LIQUIDS TO A DROWSY, CONVULSING OR UNCONSCIOUS PERSON. Otherwise, give large amounts of water. If vomiting occurs spontaneously, keep head below hips while vomiting; then repeat water administration. Get medical attention. ISOPROPYL ETHER (Clear, mobile liquid) Product Name

Product Synonyms

Flash Point

Shell Chemical - Online Literature

Page 55 of 84 (°F, TCC) IPE; DIPE; diisopropyl ether; 2,2'oxybis-propane

Isopropyl Ether 108-20-3

-18

HAZARD RATINGS Rating HMIS NFPA

Health 1 2

Fire 3 3

Reactivity 1 1

SPECIAL STORAGE AND HANDLING GUIDELINES (Refer to the Introduction for general guidelines.) IPE should not be stored in glass containers or in containers which have been opened to the air. Unopened containers should not be stored for greater than six months. These storage precautions are taken to prevent the formation of explosive peroxides. RECOMMENDED PERSONAL PROTECTIVE EQUIPMENT (Refer to the Introduction for general PPE guidelines.) Gloves: Respirator cartridge:

polyvinyl alcohol, nitrile rubber, neoprene organic vapor

HEALTH HAZARDS Liquid and high vapor concentrations of isopropyl ether may be slightly irritating to the eyes. Prolonged or repeated skin contact with isopropyl ether can result in defatting or drying of the skin with resulting irritation and dermatitis. Prolonged breathing of high isopropyl ether vapor concentrations may cause dizziness, light-headedness, headache, nausea and loss of coordination (central nervous system depression). Isopropyl ether is considered practically non-toxic if swallowed. FIRE/EXPLOSION/REACTIVITY HAZARDS DANGER! EXTREMELY FLAMMABLE. Isopropyl ether may be ignited by heat, sparks, open flames or contact with strong oxidizing agents. Isopropyl ether in the presence of air may form explosive peroxides. Vapors may accumulate and travel to ignition sources distant from the handling site resulting in a flash fire. EMERGENCY ACTION GUIDE

Shell Chemical - Online Literature

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IF...FIRE THEN...Use water fog, "alcohol" foam, dry chemical or CO2 to extinguish flames. Cool fire exposed containers with water spray. IF...SPILL OR LEAK THEN...Eliminate ignition sources. Wear appropriate personal protective equipment. Shut off leak if safe to do so. If vapor cloud forms, use water fog to suppress or blanket spill area with foam. Dike and contain spill. Monitor area with a combustible gas indicator. Remove with vacuum truck or pump to storage/recovery vessel. Soak up residue with suitable adsorbent material such as sand; DO NOT use clay. Flush area with water to remove trace residue. Contain run-off from water fog and residue flush and dispose of properly. IF...EYE CONTACT THEN...Flush eyes with large amounts of water while holding eyelids open. Rest eyes for 30 minutes. If redness, burning, blurred vision or swelling persists, get medical attention. IF...SKIN CONTACT THEN...Remove contaminated clothing and shoes. Flush affected area with large amounts of water. Wash with soap if available. If blisters appear on the skin, get medical attention. IF...INHALATION THEN...Move victim to fresh air. If the victim has difficulty breathing or tightness of the chest, is dizzy, vomiting or nonresponsive give 100% oxygen with rescue breathing or CPR as required and transport to the nearest medical facility. IF...INGESTION THEN...Have victim drink sips of water to help remove the taste from the mouth. In general no treatment is necessary unless large quantities are ingested. However, get medical advice. KETONES (Clear, mobile liquids) Product Name

Product Synonyms

Flash Point (°F, TCC)

Acetone 67-64-1 Methyl Ethyl Ketone 78-93-3

DMK; dimethyl ketone; 2-propanone

0

MEK; 2-butanone

23

Methyl Isobutyl

MIBK; 4-methyl-

60

Shell Chemical - Online Literature Ketone* 108-10-1

Page 57 of 84 2-pentanone

HAZARD RATINGS Rating HMIS NFPA

Health 2 1, 2*

Fire 3 3

Reactivity 0 0

RECOMMENDED PERSONAL PROTECTIVE EQUIPMENT (Refer to the Introduction for general PPE guidelines.) Gloves: acetone methyl ethyl ketone methyl isobutyl ketone Respirator cartridge:

butyl rubber butyl rubber polyvinyl alcohol, butyl rubber organic vapor

HEALTH HAZARDS Ketones may be from slightly irritating to severely irritating to the eyes. Prolonged or repeated skin contact with ketones can result in defatting or drying of the skin with resulting irritation and dermatitis. Prolonged breathing of high ketone vapor concentrations may cause dizziness, lightheadedness, headache, nausea and loss of coordination (central nervous system depression). Ketones may be toxic if swallowed, producing central nervous system depression. FIRE/EXPLOSION/REACTIVITY HAZARDS WARNING! FLAMMABLE (ACETONE IS EXTREMELY FLAMMABLE). Ketones may be ignited by heat, sparks, open flames or contact with strong oxidizing agents. Vapors may accumulate and travel to ignition sources distant from the handling site resulting in a flash fire. EMERGENCY ACTION GUIDE IF...FIRE THEN...Use water fog, "alcohol" foam, dry chemical or CO2 to extinguish flames. Cool fire exposed containers with water spray. IF...SPILL OR LEAK THEN...Eliminate ignition sources. Wear appropriate personal

Shell Chemical - Online Literature

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protective equipment. Shut off leak if safe to do so. If vapor cloud forms, use water fog to suppress or blanket spill with "alcohol" foam. Dike and contain spill. Monitor area with a combustible gas indicator. Remove with vacuum truck or pump to storage/recovery vessel. Soak up residue with suitable adsorbent material such as clay or sand. Flush area with water to remove trace residue. Contain run-off from water fog and residue flush and dispose of properly. IF...EYE CONTACT THEN...Flush eyes with large amounts of water for at least 15 minutes while holding eyelids open. Rest eyes for 30 minutes. If redness, burning, blurred vision or swelling persists, get medical attention. IF...SKIN CONTACT THEN...Remove contaminated clothing and shoes. Flush affected area with large amounts of water. Wash with soap if available. If blisters appear on the skin, get medical attention. IF...INHALATION THEN...Move victim to fresh air. If the victim has difficulty breathing or tightness of the chest, is dizzy, vomiting or unresponsive, give 100% oxygen with rescue breathing or CPR as required and transport to the nearest medical facility. IF...INGESTION THEN...DO NOT GIVE LIQUIDS TO A DROWSY, CONVULSING OR UNCONSCIOUS PERSON. Otherwise, induce vomiting by administering 2 glasses of water and 2 tablespoons of Syrup of Ipecac. Keep head below hips while vomiting. Get medical attention. KRATON TM THERMOPLASTIC RUBBER (Solid pellets or crumbs of various colors; no odor) Product Name KratonTM Thermoplastic Rubber (multiple grades)

Flash Point (°F, TCC)

Product Synonyms Hydrogenated isoprene and styrene-isoprene block copolymers

Not applicable

HAZARD RATINGS Rating HMIS NFPA

Health 0 0

Fire 1 1

Reactivity 0 0

Shell Chemical - Online Literature

Page 59 of 84

SPECIAL STORAGE AND HANDLING GUIDELINES (Refer to the Introduction for general guidelines.) Do not store at elevated temperatures for prolonged periods. RECOMMENDED PERSONAL PROTECTIVE EQUIPMENT (Refer to the Introduction for general PPE guidelines.) Respirator cartridge:

particulate

HEALTH HAZARDS Dust from products may cause minor irritation to the eyes and respiratory tract. Contact with molten product can result in thermal burns. At processing temperatures some degree of thermal degradation will occur which could result in the release of potentially irritating, flammable or toxic gases in small quantities. FIRE/EXPLOSION/REACTIVITY HAZARDS Products have a tendency to accumulate static charge during transport, handling and processing which can be a potential fire hazard when used in the presence of volatile or flammable mixtures. EMERGENCY ACTION GUIDE IF...FIRE THEN...Use water fog, "alcohol" foam, dry chemical or CO2 to extinguish flames. Cool fire exposed containers with water spray. IF...SPILL OR LEAK THEN...Shovel and sweep up or use industrial vacuum cleaner; avoid creating dust clouds. Place in containers for recycling or disposal. IF...EYE CONTACT THEN...Flush eyes with water while holding eyelids open. Rest eyes for 30 minutes. If redness, burning, blurred vision or swelling persists, get medical attention. IF...SKIN CONTACT THEN...For contact with molten product, cool with water and treat as a thermal burn. DO NOT attempt to remove solidified material from the burn area. Get medical attention. IF...INHALATION

Shell Chemical - Online Literature

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THEN...Move victim to fresh air.

IF...INGESTION THEN...No action is ordinarily required. NEODENETM ALPHA AND INTERNAL OLEFINS (Clear, mobile liquids) Product Name Neodene TM Alpha and Internal Olefins 13, 1314, 14, 14/16, 14/16/18,1518, 16, 16/18, 18, 20, 2024, 2024+

Flash Point (°F, SF)

Product Synonyms C12 to C24 alpha and internal olefins

>200

HAZARD RATINGS Rating HMIS NFPA

Health 2 0

Fire 1 1

Reactivity 0 0

SPECIAL STORAGE AND HANDLING GUIDELINES (Refer to the Introduction for general guidelines.) High quality (low oxygen content) nitrogen blanketing of tanks for short and long-term storage is essential to fully protect product quality and minimize product degradation. RECOMMENDED PERSONAL PROTECTIVE EQUIPMENT (Refer to the Introduction for general PPE guidelines.) Respirator cartridge:

organic vapor

HEALTH HAZARDS Neodenes are irritating to the eyes and irritating to the skin. Prolonged or repeated skin contact can result in defatting or drying of the skin with resulting irritation and dermatitis. Liquid can directly enter the lungs (aspiration) when swallowed or vomited. Due to the low viscosity of these products, a serious and possibly fatal chemical pneumonia (aspiration pneumonitis) can develop if aspiration occurs. FIRE/EXPLOSION/REACTIVITY HAZARDS

Shell Chemical - Online Literature

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These products will not burn unless preheated. Olefins react readily with air or oxygen to form trace peroxides and other oxygenates. EMERGENCY ACTION GUIDE IF...FIRE THEN...Use water fog, foam, dry chemical or CO2 to extinguish flames. Cool fire exposed containers with water spray. IF...SPILL OR LEAK THEN...Wear appropriate personal protective equipment. Shut off leak if safe to do so. Dike and contain spill. Remove with vacuum truck or pump to storage/recovery vessel. Soak up residue with suitable adsorbent material such as clay or sand. Flush area with water to remove trace residue. Contain run-off from residue flush and dispose of properly. IF...EYE CONTACT THEN...Flush eyes with water with large quantities of water for at least 15 minutes while holding eyelids open. Rest eyes for 30 minutes. If redness, burning, blurred vision or swelling persists, get medical attention. IF...SKIN CONTACT THEN...Remove contaminated clothing and shoes. Flush affected area with large amounts of water. Wash with soap if available. If blisters appear on the skin, get medical attention. IF...INHALATION THEN...Move victim to fresh air.

IF...INGESTION THEN...Have victim rinse mouth out with water, then drink sips of water to help remove the taste from the mouth. If the victim is coughing, choking, has shortness of breath, or difficulty breathing, get medical attention. If any of the following signs and symptoms appear within the next 6 hours, get medical attention: fever greater than 101 °F, shortness of breath, chest congestion, coughing or wheezing. NEODENETM ALPHA AND INTERNAL OLEFINS, COMBUSTIBLES (Clear, mobile liquids) Product Name

Product Synonyms

Flash Point

Shell Chemical - Online Literature

Page 62 of 84 (°F, SF)

Neodene Alpha and Internal Olefins 10, 10-11/12-13, 10-11/12-13/14, 10/1112, 10/1112/ 1314/12, 10/12/1314, 1012, 1014, 1112, 1112/12, 12, 12/14, 12/14/16, 1213, 12/1314, 1112/1314 TM

C10 to C16 alpha and internal olefins

114 to 158

HAZARD RATINGS Rating HMIS NFPA

Health 2 0

Fire 2 2

Reactivity 0 0

SPECIAL STORAGE AND HANDLING GUIDELINES (Refer to the Introduction for general guidelines.) High quality (low oxygen content) nitrogen blanketing of tanks for short and long-term storage is essential to fully protect product quality and minimize product degradation. RECOMMENDED PERSONAL PROTECTIVE EQUIPMENT (Refer to the Introduction for general PPE guidelines.) Respirator cartridge:

organic vapor

HEALTH HAZARDS Neodenes are irritating to the eyes and irritating to the skin. Prolonged or repeated skin contact can result in defatting or drying of the skin with resulting irritation and dermatitis. Liquid can directly enter the lungs (aspiration) when swallowed or vomited. Due to the low viscosity of these products, a serious and possibly fatal chemical pneumonia (aspiration pneumonitis) can develop if aspiration occurs. FIRE/EXPLOSION/REACTIVITY HAZARDS CAUTION! COMBUSTIBLE. These products may be ignited by heat, sparks, open flames or contact with strong oxidizing agents. Vapors may accumulate and travel to ignition sources distant from the handling site resulting in a flash fire. Olefins react readily with air or oxygen to form trace peroxides and other oxygenates.

Shell Chemical - Online Literature

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EMERGENCY ACTION GUIDE IF...FIRE THEN...Use water fog, foam, dry chemical or CO2 to extinguish flames. DO NOT use a direct stream of water; product will float and can be reignited on surface of water. Cool fire exposed containers with water spray. IF...SPILL OR LEAK THEN...Eliminate ignition sources. Wear appropriate personal protective equipment. Shut off leak if safe to do so. If vapor cloud forms, use water fog to suppress or blanket spill area with foam. Dike and contain spill. Monitor area with a combustible gas indicator. Remove with vacuum truck or pump to storage/recovery vessel. Soak up residue with suitable adsorbent material such as clay or sand. Flush area with water to remove trace residue. Contain run-off from water fog and residue flush and dispose of properly. IF...EYE CONTACT THEN...Flush eyes with large amounts of water for at least 15 minutes while holding eyelids open. Rest eyes for 30 minutes. If redness, burning, blurred vision or swelling persists, get medical attention. IF...SKIN CONTACT THEN...Remove contaminated clothing and shoes. Flush affected area with large amounts of water. Wash with soap if available. If blisters appear on the skin, or dizziness, vomiting or tightness of the chest occurs, get medical attention. IF...INHALATION THEN...Move victim to fresh air. If the victim has difficulty breathing or tightness of the chest, is dizzy, vomiting or unresponsive, give 100% oxygen with rescue breathing or CPR as required and transport to the nearest medical facility. IF...INGESTION THEN...Have victim rinse mouth out with water, then drink sips of water to help remove the taste from the mouth. If the victim is coughing, choking, has shortness of breath, or difficulty breathing, get medical attention. If any of the following delayed signs and symptoms appear within the next 6 hours, get medical attention: fever greater than 101 °F, shortness of breath, chest congestion, coughing or wheezing. NEODENETM ALPHA AND INTERNAL OLEFINS, FLAMMABLE (Clear, mobile liquids)

Shell Chemical - Online Literature

Page 64 of 84

Product Synonyms

Flash Point (°F, SF)

C6 to C12 alpha and internal olefins

-14 to 95

Product Name Neodene TM Alpha and Internal Olefins 6 HP, 6/8, 6/8/10, 6/12, 8, 810 IO, 810 AO

HAZARD RATINGS Rating HMIS NFPA

Health 2 1

Fire 3 3

Reactivity 0 0

SPECIAL STORAGE AND HANDLING GUIDELINES (Refer to the Introduction for general guidelines.) High quality (low oxygen content) nitrogen blanketing of tanks for short and long-term storage is essential to fully protect product quality and minimize product degradation. RECOMMENDED PERSONAL PROTECTIVE EQUIPMENT (Refer to the Introduction for general PPE guidelines.) Respirator cartridge:

organic vapor

HEALTH HAZARDS Neodenes are irritating to the eyes and irritating to the skin. Prolonged or repeated skin contact can result in defatting or drying of the skin with resulting irritation and dermatitis. Liquid can directly enter the lungs (aspiration) when swallowed or vomited. Due to the low viscosity of these products, a serious and possibly fatal chemical pneumonia (aspiration pneumonitis) can develop if aspiration occurs. FIRE/EXPLOSION/REACTIVITY HAZARDS WARNING! FLAMMABLE. (Neodene 6HP is Extremely Flammable). These products may be ignited by heat, sparks, open flames or contact with strong oxidizing agents. Vapors may accumulate and travel to ignition sources distant from the handling site resulting in a flash fire. Olefins react readily with air or oxygen to form tract peroxides and other oxygenates. EMERGENCY ACTION GUIDE IF...FIRE

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THEN...Use water fog, foam, dry chemical or CO2 to extinguish flames. DO NOT use a direct stream of water; product will float and can be reignited on surface of water. Cool fire exposed containers with water spray. IF...SPILL OR LEAK THEN...Eliminate ignition sources. Wear appropriate personal protective equipment. Shut off leak if safe to do so. If vapor cloud forms, use water fog to suppress or blanket spill area with foam. Dike and contain spill. Monitor area with a combustible gas indicator. Remove with vacuum truck or pump to storage/recovery vessel. Soak up residue with suitable adsorbent material such as clay or sand. Flush area with water to remove trace residue. Contain run-off from water fog and residue flush and dispose of properly. IF...EYE CONTACT THEN...Flush eyes with large amounts of water for at least 15 minutes while holding eyelids open. Rest eyes for 30 minutes. If redness, burning, blurred vision or swelling persists, get medical attention. IF...SKIN CONTACT THEN...Remove contaminated clothing and shoes. Flush affected area with large amounts of water. Wash with soap if available. If blisters appear on the skin, or dizziness, vomiting or tightness of the chest occurs, get medical attention. IF...INHALATION THEN...Move victim to fresh air. If the victim has difficulty breathing or tightness of the chest, is dizzy, vomiting or unresponsive, give 100% oxygen with rescue breathing or CPR as required and transport to the nearest medical facility. IF...INGESTION THEN...Have victim rinse mouth out with water, then drink sips of water to help remove the taste from the mouth. If the victim is coughing, choking, has shortness of breath, or difficulty breathing, get medical attention. If any of the following delayed signs and symptoms appear within the next 6 hours, get medical attention: fever greater than 101 °F, shortness of breath, chest congestion, coughing or wheezing. NEODOL TM DETERGENT ALCOHOLS (C11 AND LESS) (Water-white liquids; slight odor) Product Name Neodol TM 1 112-42-5

Product Synonyms C11 linear primary alcohol; 1-undecanol

Flash Point (°F, TCC) 250

Shell Chemical - Online Literature Neodol 91 66455-17-2

Page 66 of 84 C9-C11 linear primary alcohol

228

HAZARD RATINGS Rating HMIS NFPA

Health 2 1

Fire 1 1

Reactivity 0 0

RECOMMENDED PERSONAL PROTECTIVE EQUIPMENT (Refer to the Introduction for general PPE guidelines.) Respirator cartridge:

organic vapor

HEALTH HAZARDS These products are severely irritating to the eyes and moderate skin irritants. Liquid Neodol can directly enter the lungs (aspiration) when swallowed or vomited. Due to the low viscosity of these products a serious and possibly fatal chemical pneumonia (aspiration pneumonitis) can develop if aspiration occurs. FIRE/EXPLOSION/REACTIVITY HAZARDS Products will not burn unless preheated. EMERGENCY ACTION GUIDE IF...FIRE THEN...Use water fog, foam, dry chemical or CO2 to extinguish flames. Cool fire exposed containers with water spray. IF...SPILL OR LEAK THEN...Wear appropriate personal protective equipment. Shut off leak if safe to do so. Dike and contain spill. Remove with vacuum truck or pump to storage/recovery vessel. Soak up residue with suitable adsorbent material such as clay or sand. Flush area with water to remove trace residue. Contain run-off from residue flush and dispose of properly. IF...EYE CONTACT THEN...Flush eyes with water with large quantities of water for at least 15 minutes while holding eyelids open. Rest eyes for 30 minutes. If redness, burning, blurred vision or swelling persists, get medical attention.

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IF...SKIN CONTACT THEN...Remove contaminated clothing and shoes. Flush affected area with large amounts of water. Wash with soap if available. If blisters appear on the skin, get medical attention. IF...INHALATION THEN...Move victim to fresh air.

IF...INGESTION THEN...Have victim rinse mouth out with water, then drink sips of water to help remove the taste from the mouth. If the victim is coughing, choking, has shortness of breath, or difficulty breathing, get medical attention. If any of the following delayed signs and symptoms appear within the next 6 hours, get medical attention: fever greater than 101 °F, shortness of breath, chest congestion, coughing or wheezing. NEODOL TM DETERGENT ALCOHOLS (C12 AND GREATER) (Viscous liquids to white solid; slight odor) Product Name Neodol TM 23 75782-86-4 Neodol 25 63393-82-8 Neodol 45 75782-87-5

Flash Point (°F, TCC)

Product Synonyms C12-C13 linear primary alcohol C12-C15 linear primary alcohol C14-C15 linear primary alcohol

279 286 315

HAZARD RATINGS Rating HMIS NFPA

Health 1 0

Fire 1 1

RECOMMENDED PERSONAL PROTECTIVE EQUIPMENT (Refer to the Introduction for general PPE guidelines.) Respirator cartridge:

HEALTH HAZARDS

organic vapor

Reactivity 0 0

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Page 68 of 84

(Refer to the MSDS for product specific health hazard information.) These products are from non-irritating to minimally irritating to the eyes. These products are mild to moderate skin irritants; slightly toxic if absorbed through skin. FIRE/EXPLOSION/REACTIVITY HAZARDS Products will not burn unless preheated. EMERGENCY ACTION GUIDE IF...FIRE THEN...Use water fog, foam, dry chemical or CO2 to extinguish flames. DO NOT use a direct stream of water; product will float and can be reignited on surface of water. Cool fire exposed containers with water spray. IF...SPILL OR LEAK THEN...Wear appropriate personal protective equipment. Shut off leak if safe to do so. Dike and contain spill. Remove with vacuum truck or pump to storage/recovery vessel. Soak up residue with suitable adsorbent material such as clay or sand. Flush area with water to remove trace residue. Contain run-off from residue flush and dispose of properly. IF...EYE CONTACT THEN...Flush eyes with water while holding eyelids open. Rest eyes for 30 minutes. If redness, burning, blurred vision or swelling persists, get medical attention. IF...SKIN CONTACT THEN...Remove contaminated clothing and shoes. Flush affected area with large amounts of water. Wash with soap if available. If blisters appear on the skin, get medical attention. IF...INHALATION THEN...Remove victim to fresh air.

IF...INGESTION THEN...Have victim drink sips of water to help remove the taste from the mouth. In general no treatment is necessary unless large quantities are ingested. However, get medical advice. NEODOL TM ALCOHOL ETHOXYLATES

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(Viscous liquids to white solids; slight odor) Product Name Neodol TM Alcohol Ethoxylates

Flash Point (°F, TCC)

Product Synonyms Linear primary alcohol ethoxylates

>200

HAZARD RATINGS Rating HMIS NFPA

Health 1, 2 1, 2

Fire 1 1

Reactivity 0 0

RECOMMENDED PERSONAL PROTECTIVE EQUIPMENT (Refer to the Introduction for general PPE guidelines.) Respirator cartridge:

organic vapor

HEALTH HAZARDS (Refer to the MSDS for product specific health hazard information.) Neodol ethoxylates may be from mildly to severely irritating to the eyes and skin depending on the product. Breathing Neodol ethoxylate vapors may be mildly irritating to the nose, throat and lungs. Neodol ethoxylates may be moderately toxic if swallowed. FIRE/EXPLOSION/REACTIVITY HAZARDS Products will not burn unless preheated. EMERGENCY ACTION GUIDE IF...FIRE THEN...Use water fog, "alcohol" foam, dry chemical or CO2 to extinguish flames. Cool fire exposed containers with water spray. IF...SPILL OR LEAK THEN...Wear appropriate personal protective equipment. Shut off leak if safe to do so. Dike and contain spill. Remove with vacuum truck or pump to storage/recovery vessel. Soak up residue with suitable adsorbent material such as clay or sand. Flush area with water to remove trace residue. Contain run-off from residue flush and dispose of properly.

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Page 70 of 84

IF...EYE CONTACT THEN...Flush eyes with large amounts of water for at least 15 minutes while holding eyelids open. Rest eyes for 30 minutes. If redness, burning, blurred vision or swelling persists, get medical attention. IF...SKIN CONTACT THEN...Remove contaminated clothing and shoes. Flush affected area with large amounts of water. Wash with soap if available. If blisters appear on the skin, get medical attention. IF...INHALATION THEN...Move victim to fresh air. If the victim has difficulty breathing or tightness of the chest, is dizzy, vomiting or unresponsive, give 100% oxygen with rescue breathing or CPR as appropriate and transport to the nearest medical facility. IF...INGESTION THEN...DO NOT GIVE LIQUIDS TO A DROWSY, CONVULSING OR UNCONSCIOUS PERSON. Otherwise, induce vomiting by administering 2 glasses of water and 2 tablespoons of Syrup of Ipecac. Keep head below hips while vomiting. Get medical attention. NEOSOLV TM SOLVENTS (Clear, mobile liquids) Product Name NeosolvTM Solvents 150, 2, 24, 4*, 6* and 8*

Flash Point (°F, SF)

Product Synonyms >C10 linear unsaturated hydrocarbons

168 to 272

HAZARD RATINGS Rating HMIS NFPA

Health 1 0

Fire 1*, 2 1*, 2

Reactivity 0 0

SPECIAL STORAGE AND HANDLING GUIDELINES (Refer to the Introduction for general guidelines.) High quality (low oxygen content) nitrogen blanketing of tanks for short and long-term storage is essential to fully protect product quality and minimize product degradation. RECOMMENDED PERSONAL PROTECTIVE EQUIPMENT

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(Refer to the Introduction for general PPE guidelines.) Respirator cartridge:

organic vapor

HEALTH HAZARDS (Refer to the MSDS for product specific hazard information.) Neosolvs range from non-irritating to irritating to the eyes and are moderate skin irritants. Prolonged or repeated skin contact can result in defatting or drying of the skin with resulting irritation and dermatitis. Liquid can directly enter the lungs (aspiration) when swallowed or vomited. Due to the low viscosity of these products, a serious and possibly fatal chemical pneumonia (aspiration pneumonitis) can develop if aspiration occurs. FIRE/EXPLOSION/REACTIVITY HAZARDS (Refer to the MSDS for product specific hazard information.) CAUTION! Some of these products are COMBUSTIBLE. These products may be ignited by heat, sparks, open flames or contact with strong oxidizing agents. Vapors may accumulate and travel to ignition sources distant from the handling site resulting in a flash fire. Olefins react readily with air or oxygen to form trace peroxides and other oxygenates. EMERGENCY ACTION GUIDE IF...FIRE THEN...Use water fog, foam, dry chemical or CO2 to extinguish flames. DO NOT use a direct stream of water; product will float and can be reignited on surface of water. Cool fire exposed containers with water spray. IF...SPILL OR LEAK THEN...Eliminate ignition sources. Wear appropriate personal protective equipment. Shut off leak if safe to do so. If vapor cloud forms, use water fog to suppress or blanket spill area with foam. Dike and contain spill. Monitor area with a combustible gas indicator. Remove with vacuum truck or pump to storage/recovery vessel. Soak up residue with suitable adsorbent material such as clay or sand. Flush area with water to remove trace residue. Contain run-off from water fog and residue flush and dispose of properly. IF...EYE CONTACT THEN....Flush eyes with large amounts of water for at least 15 minutes while holding eyelids open. Rest eyes for 30 minutes. If redness, burning, blurred vision or swelling persists, get medical attention. IF...SKIN CONTACT

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THEN...Remove contaminated clothing and shoes. Flush affected area with large amounts of water. Wash with soap if available. If blisters appear on the skin, or dizziness, vomiting or tightness of the chest occurs, get medical attention. IF...INHALATION THEN...Move victim to fresh air. If the victim has difficulty breathing or tightness of the chest, is dizzy, vomiting or unresponsive, give 100% oxygen with rescue breathing or CPR as required and transport to the nearest medical facility. IF...INGESTION THEN...Have victim rinse mouth out with water, then drink sips of water to help remove the taste from the mouth. If the victim is coughing, choking, has shortness of breath, or difficulty breathing, get medical attention. If any of the following delayed signs and symptoms appear within the next 6 hours, get medical attention: fever greater than 101 °F, shortness of breath, chest congestion, coughing or wheezing. OLEFINS (Colorless gases; liquids under pressure) Product Name Ethylene 74-85-1

Flash Point (°F, TCC)

Product Synonyms Ethene; olifiant gas 1-propene; methylethylene; methylethene 1-butene; butylene; ethylethylene

Propylene* 115-07-1 Butene-1** 106-98-9

-278 -162 -112

HAZARD RATINGS Rating HMIS NFPA

Health 1 1

Fire 4 4

Reactivity 1 2, 1*, 0**

SPECIAL STORAGE AND HANDLING GUIDELINES (Refer to the Introduction for general guidelines.) Storage and handling of these products should be in approved pressure vessels. RECOMMENDED PERSONAL PROTECTIVE EQUIPMENT

Shell Chemical - Online Literature

Page 73 of 84

(Refer to the Introduction for general PPE guidelines.) Respirator cartridge:

organic vapor

HEALTH HAZARDS Direct contact with liquified gas can result in burns (frostbite) to the eyes and skin due to rapid evaporation of the liquid. Breathing high gas concentrations can produce headache, light-headedness, dizziness, nausea, and possibly loss of consciousness and death. FIRE/EXPLOSION/REACTIVITY HAZARDS DANGER! EXTREMELY FLAMMABLE. These products may be ignited by heat, sparks, open flames or contact with strong oxidizing agents. Gas is explosive. Gas may accumulate and travel to ignition sources distant from the handling site resulting in a flash fire and explosion. Ethylene may undergo spontaneous hazardous polymerization. Avoid catalysts and conditions that promote polymerization. Exposure of containers to elevated temperatures will result in a build-up of pressure which may cause the container to rupture. EMERGENCY ACTION GUIDE IF...FIRE THEN...Allow gas to burn if flow cannot be shut off safely. Cool fire exposed containers with water spray from a safe distance. If a sufficient water supply cannot be sustained, withdraw from the area. IF...SPILL OR LEAK THEN...Evacuate surrounding areas. Eliminate ignition sources. Wear appropriate personal protective equipment. Disperse vapor cloud with sustained water spray until leak is stopped. If a sufficient water supply cannot be sustained, withdraw from the area. IF...EYE CONTACT THEN...Flush eyes with large amounts of water for at least 15 minutes while holding eyelids open. Rest eyes for 30 minutes. If redness, burning, blurred vision or swelling persists, get medical attention. IF...SKIN CONTACT THEN...Remove contaminated clothing and shoes. Flush affected area with large amounts of water for at least 15 minutes. Wash with soap if available. If blisters appear on the skin, get medical attention. IF...INHALATION

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THEN...Move victim to fresh air. If the victim has difficulty breathing or tightness of the chest, is dizzy, vomiting or unresponsive, give 100% oxygen with rescue breathing or CPR as required and transport to the nearest medical facility. IF...INGESTION THEN...No action required. Not applicable for gases or extremely volatile liquids. PHENOL (White crystals) Product Name

Flash Point (°F, TCC)

Product Synonyms Carbolic acid; monohydroxybenzene; benzenol; phenylic acid

Phenol 108-95-2

175

HAZARD RATINGS Rating HMIS NFPA

Health 3 3

Fire 2 2

Reactivity 0 0

RECOMMENDED PERSONAL PROTECTIVE EQUIPMENT (Refer to the Introduction for general PPE guidelines.) Full protective clothing should be worn when handling phenol. Protective clothing/gloves: Respirator cartridge:

neoprene, natural rubber, Teflon/Nomex organic vapor

HEALTH HAZARDS WARNING!

Phenol is extremely irritating to the eyes and may cause blindness. Phenol is extremely irritating to the skin and may be fatal if absorbed through skin even in small quantities. Breathing of phenol vapors is severely irritating to the nose, throat and respiratory tract and may be fatal. Phenol is highly toxic and may be fatal if swallowed. Phenol exposure can cause liver, kidney and heart damage.

Shell Chemical - Online Literature FIRE/EXPLOSION/REACTIVITY HAZARDS CAUTION! COMBUSTIBLE. This product may be ignited by heat, sparks, open flames or contact with strong oxidizing agents. Corrosive to carbon steel. EMERGENCY ACTION GUIDE IF...FIRE THEN...Use water fog, "alcohol" foam, dry chemical or CO2 to extinguish flames. Cool fire exposed containers with water spray. IF...SPILL OR LEAK THEN...Eliminate ignition sources. Wear appropriate personal protective equipment. Shut off leak if safe to do so. If vapor cloud forms, use water fog to suppress or blanket spill area with "alcohol" foam. Dike and contain spill. Monitor area with combustible gas indicator. Remove with vacuum truck or pump to storage/recovery vessel. Soak up residue with suitable adsorbent material such as clay or sand. Flush area with water to remove trace residue. Contain run-off from water fog and residue flush and dispose of properly. IF...EYE CONTACT THEN...Flush eyes with large amounts of water for at least 15 minutes while holding eyelids open. Continue to rinse with water while on the way to get medical attention. IF...SKIN CONTACT THEN...Remove contaminated clothing and shoes. Flush affected area with large amounts of water for at least 15 minutes. Get medical attention. IF...INHALATION THEN...Move victim to fresh air. If the victim has difficulty breathing or tightness of the chest, is dizzy, vomiting or unresponsive, give 100% oxygen with rescue breathing or CPR as required and transport to the nearest medical facility. IF...INGESTION THEN...DO NOT INDUCE VOMITING. DO NOT GIVE LIQUIDS TO A DROWSY, CONVULSING OR UNCONSCIOUS PERSON. Otherwise, give large amounts of water. If vomiting occurs spontaneously, keep head below hips while vomiting. Get medical attention.

Page 75 of 84

Shell Chemical - Online Literature

Page 76 of 84

PLASTICS (Solid pellets or crumbs of various colors) Product Name Polybutylene (multiple grades) CleartufTM /TraytufTM Resins (multiple grades)

Flash Point (°F, TCC)

Product Synonyms Butene polymer; butylene polymer; butene-ethylene copolymer Polyethylene terephthalate; polyester resin

Not applicable Not applicable

HAZARD RATINGS Rating HMIS NFPA

Health 0 0

Fire 1 1

Reactivity 0 0

SPECIAL STORAGE AND HANDLING GUIDELINES (Refer to the Introduction for general guidelines.) Do not store at elevated temperatures for prolonged periods. RECOMMENDED PERSONAL PROTECTIVE EQUIPMENT (Refer to the Introduction for general PPE guidelines.) Respirator cartridge:

particulate

HEALTH HAZARDS Dust from products may cause minor irritation to the eyes and respiratory tract. Contact with molten product can result in thermal burns. At processing temperatures some degree of thermal degradation will occur which could result in the release of potentially irritating, flammable or toxic gases in small quantities. FIRE/EXPLOSION/REACTIVITY HAZARDS Products have a tendency to accumulate static charge during transport, handling and processing which can be a potential fire hazard when used in the presence of volatile or flammable mixtures. EMERGENCY ACTION GUIDE IF...FIRE THEN...Use water fog, "alcohol" foam, dry chemical or CO2 to

Shell Chemical - Online Literature

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extinguish flames. Cool fire exposed containers with water spray. IF...SPILL OR LEAK THEN...Shovel and sweep up or use industrial vacuum cleaner; avoid creating dust clouds. Place in containers for recycling or disposal. IF...EYE CONTACT THEN...Flush eyes with water while holding eyelids open. Rest eyes for 30 minutes. If redness, burning, blurred vision or swelling persists, get medical attention. IF...SKIN CONTACT THEN...For contact with molten product, cool with water and treat as a thermal burn. DO NOT attempt to remove solidified material from the burn area. Get medical attention. IF...INHALATION THEN...Move victim to fresh air.

IF...INGESTION THEN...No action is ordinarily required. SHELLVISTM VISCOSITY INDEX IMPROVERS (Viscous amber liquids; slight odor) Product Name ShellvisTM Viscosity Index Improvers

Flash Point (°F, TCC)

Product Synonyms Viscosity improvers

279

HAZARD RATINGS Rating HMIS NFPA

Health 0 0

Fire 1 1

RECOMMENDED PERSONAL PROTECTIVE EQUIPMENT (Refer to the Introduction for general PPE guidelines.)

Reactivity 0 0

Shell Chemical - Online Literature Respirator cartridge:

Page 78 of 84 particulate (mist)

HEALTH HAZARDS (Refer to the MSDS for product specific health hazard information.) These products are from non-irritating to minimally irritating to the eyes. Prolonged or repeated skin contact may result in various skin disorders, such as dermatitis, oil acne or folliculitis. FIRE/EXPLOSION/REACTIVITY HAZARDS Products will not burn unless preheated. EMERGENCY ACTION GUIDE IF...FIRE THEN...Use water fog, foam, dry chemical or CO2 to extinguish flames. DO NOT use a direct stream of water; product will float and can be reignited on surface of water. Cool fire exposed containers with water spray. IF...SPILL OR LEAK THEN...Wear appropriate personal protective equipment. Shut off leak if safe to do so. Dike and contain spill. Remove with vacuum truck or pump to storage/recovery vessel. Soak up residue with suitable adsorbent material such as clay or sand. Flush area with water to remove trace residue. Contain run-off from residue flush and dispose of properly. IF...EYE CONTACT THEN...Flush eyes with water while holding eyelids open. Rest eyes for 30 minutes. If redness, burning, blurred vision or swelling persists, get medical attention. IF...SKIN CONTACT THEN...Remove contaminated clothing and shoes. Flush affected area with large amounts of water. Wash with soap if available. If blisters appear on the skin, get medical attention. IF...INHALATION THEN...Remove victim to fresh air.

IF...INGESTION THEN...Have victim drink sips of water to help remove the taste

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from the mouth. In general no treatment is necessary unless large quantities are ingested. However, get medical advice. SULFOLANE (White solid to viscous liquid; sulfide odor) Product Name Sulfolane Sulfolane W 126-33-0

Flash Point (°F, TCC)

Product Synonyms Thiophan sulfone; tetrahydrothiophene-1, 1- dioxide

320

HAZARD RATINGS Rating HMIS NFPA

Health 1 2

Fire 1 1

Reactivity 0 0

RECOMMENDED PERSONAL PROTECTIVE EQUIPMENT (Refer to the Introduction for general PPE guidelines.) Respirator cartridge:

organic vapor, particulate

HEALTH HAZARDS Sulfolane is mildly irritating to the eyes and skin. Breathing sulfolane vapors may produce lung damage at high doses. Under fire conditions, toxic and irritating sulfur oxides are formed. Sulfolane is moderately toxic if swallowed producing initial central nervous system stimulation followed by depression. FIRE/EXPLOSION/REACTIVITY HAZARDS Will not burn unless preheated. Decomposes at temperatures above 420 °F releasing sulfur oxides and carbon monoxide. EMERGENCY ACTION GUIDE IF...FIRE THEN...Use water fog, foam, dry chemical or CO2 to extinguish flames. Cool fire exposed containers with water spray. IF...SPILL OR LEAK THEN...Wear appropriate personal protective equipment. Place

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in containers for recycling or disposal. IF...EYE CONTACT THEN...Flush eyes with water with large quantities of water for at least 15 minutes while holding eyelids open. Rest eyes for 30 minutes. If redness, burning, blurred vision or swelling persists, get medical attention. IF...SKIN CONTACT THEN..Remove contaminated clothing and shoes. Flush affected area with large amounts of water. Wash with soap if available. If blisters appear on the skin, get medical attention. IF...INHALATION THEN...Move victim to fresh air.

IF...INGESTION THEN...DO NOT GIVE LIQUIDS TO A DROWSY, CONVULSING OR UNCONSCIOUS PERSON. Otherwise, induce vomiting by administering 2 glasses of water followed by 2 tablespoons of Syrup of Ipecac. Keep head below hips while vomiting. Get medical attention. THERMOPLASTIC POLYAMIDE RESINS (Amber solids) Product Name

Flash Point (°F, SF)

Product Synonyms

Azamide TM Resins

Thermoplastic polyamide resins

>200

Surfam TM Resin 58 112-84-5

1,3-docosenamide

>200

HAZARD RATINGS Rating HMIS NFPA

Health 0 0

Fire 1 1

RECOMMENDED PERSONAL PROTECTIVE EQUIPMENT (Refer to the Introduction for general PPE guidelines.)

Reactivity 0 0

Shell Chemical - Online Literature Respirator cartridge:

Page 81 of 84 organic vapor

HEALTH HAZARDS Vapors from heated product may cause irritation to eyes and respiratory system. Generally considered to be non-irritating to skin and non-toxic if swallowed. FIRE/EXPLOSION/REACTIVITY HAZARDS Products will not burn unless preheated. These products can react vigorously with strong oxidizing agents and strong acids and bases. Reaction with some curing agents can produce considerable heat. EMERGENCY ACTION GUIDE IF...FIRE THEN...Use water fog, "alcohol" foam, dry chemical or CO2 to extinguish flames. Cool fire exposed containers with water spray. IF...SPILL OR LEAK THEN...Shovel and sweep up or use industrial vacuum cleaner; avoid creating dust clouds. Place in containers for recycling or disposal. IF...EYE CONTACT THEN...Flush eyes with water while holding eyelids open. Rest eyes for 30 minutes. If redness, burning, blurred vision or swelling persists, get medical attention. IF...SKIN CONTACT THEN...For contact with molten product, cool with water and treat as a thermal burn. DO NOT attempt to remove solidified material from the burn area. Get medical attention. IF...INHALATION THEN...Move victim to fresh air.

IF...INGESTION THEN...No action is ordinarily required. WATER DISPERSIBLE RESINS (Opaque white viscous liquids)

Shell Chemical - Online Literature

Product Name Epi-RezTM Epoxy Resins, W and WD series Epi-Rez Epoxy Resins, WY series and 5510

Epi-Rez 3350 Epoxy Resin* Eco-Cryl(TM) Acrylic Resin Dispersion 9790

Page 82 of 84

Product Synonyms Bisphenol Aepichlorohydrin water dispersible epoxy resins Modified bisphenol A-epichlorohydrin or urethane modified epoxy resins containing 2propoxyethanol Bisphenol Aepichlorohydrin water dispersible epoxy resin solution containing alcohol diluents Acrylic copolymer mixture containing 2-propoxyethanol

Flash Point (°F, SF) >200

>200

123

>200

HAZARD RATINGS Rating HMIS NFPA

Health 2 2

Fire 1, 2* 1, 2*

Reactivity 0 0

SPECIAL STORAGE AND HANDLING GUIDELINES (Refer to the Introduction for general guidelines.) Do not reuse contaminated leather articles such as shoes as they cannot be decontaminated. HEALTH HAZARDS (Refer to the MSDS for product specific health hazard information.) Epi-Rez and Eco-Cryl resins may be irritating to eyes, skin and respiratory system. EpiRez WY and 5510 are severely irritating to the eyes; may cause corneal damage. There is a potential for skin sensitization (allergic reaction) after repeated skin contact with these products. The solvent in Epi-Rez WY series and 5510 may be toxic if absorbed through skin, producing damage to red blood cells. Inhalation of mists of these products may be irritating. These products may be toxic if swallowed. Ingestion of the solvent in Epi-Rez WY series and 5510 may cause damage to red blood cells. FIRE/EXPLOSION/REACTIVITY HAZARDS These products will not burn unless preheated. These products can react vigorously with strong oxidizing agents, strong Lewis or

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mineral acids and strong mineral and organic bases (especially primary and secondary aliphatic amines). Reaction with some curing agents can produce considerable heat. EMERGENCY ACTION GUIDE IF...FIRE THEN...Use water fog, foam, dry chemical or CO2 to extinguish flames. Cool fire exposed containers with water spray. IF...SPILL OR LEAK THEN...Wear appropriate personal protective equipment. Shut off leak if safe to do so. Dike and contain spill. Remove with vacuum truck or pump to storage/recovery vessel. Soak up residue with suitable adsorbent material such as clay or sand. Flush area with water to remove trace residue. Contain run-off from residue flush and dispose of properly. IF...EYE CONTACT THEN...Flush eyes with large amounts of water for at least 15 minutes while holding eyelids open. Rest eyes for 30 minutes. If redness, burning, blurred vision or swelling persists, get medical attention. IF...SKIN CONTACT THEN...Remove contaminated clothing and shoes. Flush affected area with large amounts of water. Wash with soap if available. If blisters appear on the skin, get medical attention. IF...INHALATION THEN...Move victim to fresh air. If the victim has difficulty breathing or tightness of the chest, is dizzy, vomiting or unresponsive, give 100% oxygen with rescue breathing or CPR as appropriate and transport to the nearest medical facility. IF...INGESTION THEN...DO NOT induce vomiting. If vomiting occurs spontaneously, keep head below hips while vomiting. Get medical attention. NOTE: An asterisk(s) following a product name indicates that different hazard ratings are noted in the next section, "Hazard Ratings". SC: 436-95 (Supersedes SC: 436-79)

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*CARDURA, ECO-CRYL, EPI -CURE, EPIKOTE, EPI -REZ, EPON, EPONEX, HELOXY & VeoVa are trademarks of the Royal Dutch/Shell group of companies All Rights Reserved. © 1998, Shell Chemicals Ltd. Last Modified:09/21/98 19:00:00 PM

Acquired by MYG for PPTSB Maint. Dept.

Centrifugal Pump Troubleshooting You have four opportunities to trouble shoot centrifugal pumps and each opportunity can offer you a clue as to what is wrong with the pump. Let's take a look at each of these conditions: The pump is hooked up to the piping and it is running : • • • • • • •

•

You can observe leakage from the stuffing box or some other area. This would include gaskets, bearing seals and cracks or holes in the piping. You can hear an abnormal noise. You can probably "pin point" the source if you try. There is evidence of excessive heat in one or more of the components. You can detect excessive vibration either from the use of instruments or one of your senses You can check if stuffing box environmental controls are hooked up properly, and in many cases tell if they are functioning correctly. You can check the position of control and isolation valves throughout the system. This is especially important to check while the pump is running. If there are meters available you can check : o Flow o Pressure o Power consumption o Temperature o Speed You can estimate if the foundation is too weak. It should be five times the mass of the hardware sitting on it.

The pump is still hooked up to the piping, but it is not running. You will be present during the removal process : • • •

You can check alignment between the pump and driver. During the removal process you can check for excessive pipe strain. You can check if the piping has been installed according to good engineering practices. This is a major factor in many cavitation problems.

The pump has been taken into the workshop, but has not yet been disassembled and you will be present at the disassembly. • • •

You can check the seal installation dimension. You can feel if there is restricted movement of any of the rotating parts. You can see if there is any loose hardware in the assembly

The pump has been disassembled. You were not present, but the parts are available for your inspection.

• • • •

You can see evidence of wear, rubbing or discoloration of the components. You can see evidence of corrosion. You can see if any parts are missing. You can see if any material or coating has attached its self to one of the components. As an example, calcium can build up on the inside of pipes and restrict flow, or magnetite (Fe304) build up on the seal components.

In this paper we will address the last condition. The pump has been disassembled. You were not present, but the parts are available for your inspection. When a rotating part such as a shaft seal, impeller, etc. comes in contact with a stationary part such as the inside of the stuffing box, a wear ring, stationary bushing etc., there will be evidence of this contact in the form of rubbing, wear, discoloration or damage to one or both of the components. There are four possibilities that we will be able to see : • • • •

A rubbing mark, or evidence of wear all around the rotating part and one place on the stationary part. All around the stationary part and one place on the rotating part. Evidence of rubbing or wear all around both the rotating and stationary parts. One mark on both the rotating and stationary component.

The cause could be the result of a problem in design, operation or maintenance. I will attempt to isolate these three areas as we look into the problems. All around the rotating part, one spot on the stationary part. • •

•

Design Problems: o The pump is pulley driven and the shaft L3/D4 is too high. Maintenance Problems: o The pump and driver are not aligned properly o The shaft is not centered in the stuffing box. o A gasket or fitting is protruding in, and touching the rotating part. o Excessive pipe strain. This is a common problem when a Centerline Design is not specified for applications over 200 F (100 C) Operation problems: o A major cause of this problem is the fact that the pump is operating too far from its best efficiency point (B.E.P.) and the shaft is not large enough to resist the bending.

All around the stationary, one spot on the rotary. •

•

Design Problems o You have converted the pump to a mechanical seal. The unit was originally designed for the packing to act as a bearing and stabilize the shaft. This is a very big problem with mixers and agitators Maintenance Problems

o o o o o o o o o o o

The rotating assembly is out of balance. Normal wear Damage Corrosion of the impeller Foreign material attached The impeller was trimmed and not re balanced. The seal, sleeve, or impeller is not concentric with the shaft. The unit never was balanced. The shaft is bent. Excessive heat or force was used during sleeve, seal, or bearing removal. The rotating unit is dragging something around with it.

All around both the rotating and stationary units. This problem could be caused by a combination ofthe first and second examples or: •

• •

Design Problems: o High temperature application. The shaft is expanding and a restriction bushing is growing in towards the shaft/sleeve. o The pump is operating at a critical speed. This can happen with variable speed motors. Operation Problems: o The pump is cavitating Maintenance Problems o Bad bearings. o The oil is contaminated with water, product, dirt, rust, casting leaching, etc.. o Incorrect oil level. o Poor fit because of shaft tolerances or the installation technique. o Excessive load due to a variety of reasons. o Oil temperature too high. Be sure to cool the oil not the bearings. Cooling the housing will cause it to shrink and thereby increase the squeeze on the bearing.

One mark on both the rotating and stationary component. •

I have only seen this one time and that was when the pump fell off the back of a pick up truck.

Corrosion problems associated with stainless steel The rotating equipment business uses a great deal of 300 series stainless steel, and as a result we often experience corrosion that is described in a variety of technical terms that include: • • • • • • • • • •

General corrosion Galvanic corrosion Pitting Inter granular corrosion Stress corrosion cracking Erosion- corrosion Fretting Concentrated cell or crevice corrosion Selective leaching Micro organisms

The last page of this report is a list titled "The Galvanic Series Of Metals and alloys". I will be referring to this chart during our discussion. The basic resistance of stainless steel occurs because of its ability to form a protective coating on the metal surface. This coating is a "passive" film which is resistant to further "oxidation" or rusting. The formation of this film is instantaneous in an oxidizing atmosphere such as air, water, or many other fluids that contain oxygen. Once the layer has formed we say that the metal has become "passivated" and the oxidation or "rusting" rate will slow down to less than 0.002" per year (0,05 mm. per year). Unlike aluminum or silver this passive film is invisible in stainless steel. It is due to the combining of oxygen with the chrome in the stainless to form chrome oxide which is more commonly called "ceramic". This protective oxide or ceramic coating is common to most corrosion resistant materials. Halogen salts, especially chlorides easily penetrate this passive film and will allow corrosive attack to occur. The halogens are easy to recognize because they end in the letters "ine". Listed in order of their activity they are: fluorine, chlorine, bromine, iodine and astatine. These are the same chemicals that will penetrate Teflon and cause trouble with Teflon coated or encapsulated O-Rings and/ or similar coated materials. Chlorides are one of the most common elements in nature and if that isn't bad enough they are also soluble, active ions; the basis for good electrolytes, the best conditions for corrosion or chemical attack. GENERAL OR OVERALL CORROSION.

This type of corrosion occurs when there is an overall breakdown of the passive film formed on the stainless steel. It is the easiest to recognize as the entire surface of the metal shows a uniform "sponge like" appearance. The rate of attack is affected by the fluid concentration, temperature, fluid velocity and stress in the metal parts subject to attack. As a general rule the rate of attack will double with an eighteen degree Fahrenheit rise in temperature (10° C.) of either the product or the metal part. If the rotating portion of the seal is rubbing against some stationary component, such as a protruding gasket or fitting the protective oxide layer will be polished off and the heat generated will increase the corrosion as noted above. This explains why corrosion is often limited to only one portion of the metal case. There are many good publications available to help you select the proper metal for any given mechanical seal application. As a general rule, if the wetted parts of the equipment are manufactured from iron, steel, stainless steel or bronze, and they are showing no signs of corrosion, grade 316 stainless is acceptable as long as you do not use stainless steel springs. (see chloride stress corrosion) GALVANIC CORROSION If you put two dissimilar metals or alloys in a common electrolyte, and connect them with a voltmeter, it will show an electric current flowing between the two. (This is how the battery in your automobile works). When the current flows, material will be removed from one of the metals or alloys ( the ANODIC one) and dissolve into the electrolyte. The other metal (the CATHODIC one) will be protected. Now let's take a look at the Galvanic Series chart that is attached to this report. The further apart the materials are located on this chart the more likely that the one on the ANODIC end will corrode if they are both immersed in a common fluid considered to be an electrolyte. water, containing chlorides, is one of the best. Example #1. A ship has lots of bronze fittings and a steel hull. Note that steel is located seven lines from the ANODIC end, and bronze is listed at twenty seven rows from the same end. Sea water is a perfect electrolyte so the bronze fittings would immediately attack the steel hull unless something could be done to either protect the steel or give the bronze something else to attack. The classic way to solve this problem is to attach sacrificial zinc pieces to the hull and let the bronze go after them. Again, looking at the chart, you will note that zinc is found on line three from the top of the chart. In other words the zinc is further away from the bronze than the iron is, so the galvanic action takes place between the zinc and the bronze, rather than between the steel and the bronze. Zinc paint is used for the same reason in many applications. Example #2

Nickel base tungsten carbide contains active nickel. When this face material is used in dual seal it is common to circulate water or antifreeze containing water between the seals (as mentioned in the beginning of this report, water can be an excellent electrolyte because of the addition of chlorine and fluorine). You will note that active nickel is located twenty one rows from the top of the chart. Passivated 316 stainless steel is positioned nine rows from the bottom. This means that the stainless steel can attack the nickel in the tungsten carbide causing it to corrode. Many of you have run into this problem already. The rate at which corrosion takes place is determined by : • •

• •

The distance separating the metals on the galvanic series chart The temperature and concentration of the electrolyte. The higher the temperature, the faster it happens. Any stray electrical currents in the electrolyte will increase the corrosion also. The relative size of the metal pieces. A large cross section piece will not be affected as much as a smaller one. Many metal seal components are isolated from each other by the use of rubber ORings or similar materials and designs. Shaft movement that causes fretting of the 316 stainless steel rubs off the passivated layer and exposes the active stainless to the electrolyte until the metal part becomes passivated once more. This is one of the reasons we see corrosion under O-rings, Teflon, and similar materials. In the next paragraph I will be discussing another cause of corrosion under rubber parts.

PITTING This is an accelerated form of chemical attack in which the rate of corrosion is greater in some areas than others. It occurs when the corrosive environment penetrates the passivated film in only a few areas as opposed to the overall surface. As stated earlier the halogens will penetrate passivated stainless steel. Referring to the galvanic chart you will note that passivated 316 stainless steel is located nine lines from the bottom and active 316 stainless steel is located thirteen lines from the top. Pit type corrosion is therefore simple galvanic corrosion, as the small active area is being attacked by the large passivated area. This difference in relative areas accelerates the corrosion causing the pits to penetrate deeper. The electrolyte fills the pits and prevents the oxygen from passivating the active metal so the problem gets even worse. This type of corrosion is often called "concentrated cell corrosion". You will also see it under rubber parts that tend to keep oxygen away from the active metal parts, retarding its ability to form the passivated layer. INTERGRANULAR CORROSION All Austenitic stainless steels (the 300 series, the types that "work harden", is one of them) contain a small amount of carbon in solution in the austenite. Carbon is precipitated out at the grain boundaries, of the steel, in the temperature range of 1050° F.

(565° C) to 1600° F. (870° C.). This is a normal temperature range during the welding of stainless steel. This carbon combines with the chrome in the stainless steel to form chromium carbide starving the adjacent areas of the chrome they need for corrosion protection. In the presence of some strong corrosives an electrochemical action is initiated between the chrome rich and chrome poor areas with the areas being low in chrome becoming attacked. The grain boundaries are then dissolved and become non existent. There are three ways to combat this: •

•

•

Anneal the stainless after it has been heated to this sensitive range. This means bringing it up to the proper annealing temperature and then quickly cooling it down through the sensitive temperature range to prevent the carbides from forming. When possible use low carbon content stainless if you intend to do any welding on it. A carbon content of less than 0.3% will not precipitate into a continuous film of chrome carbide at the grain boundaries. 316L is as good example of a low carbon stainless steel. Alloy the metal with a strong carbide former. The best is columbium, but sometimes titanium is used. The carbon will now form columbium carbide rather than going after the chrome to form chrome carbide. The material is now said to be "stabilized"

CHLORIDE STRESS CORROSION. If the metal piece is under tensile stress, either because of operation or residual stress left during manufacture, the pits mentioned in a previous paragraph will deepen even more. Since the piece is under tensile stress cracking will occur in the stressed piece. Usually there will be more than one crack present that causes the pattern to resemble a spider's web. Chloride stress cracking is a common problem in industry and not often recognized by the people involved. In the seal business it is a serious problem if you use stainless steel springs or stainless steel bellows material. This is the main reason that Hastelloy C is recommended for spring material. Here are some additional thoughts about chloride stress cracking that you will want to consider: •

•

Chlorides are the big problem when using the 300 series grades of stainless steel. The 300 series is the one most commonly used in the process industry because of its good corrosion resistant proprieties. Outside of water chloride is the most common chemical found in nature and remember that the most common water treatment is the addition of chlorine. Beware of insulating or painting stainless steel pipe. Most insulation contains plenty of chlorides and piping is frequently under tensile stress. The worst condition would be insulated steam traced, stainless steel piping.

If it is necessary to insulate stainless steel pipe a special chloride free insulation can be purchased or the pipe can be coated with a protective film prior to insulating.

• • •

• •

Stress cracking can be minimized by annealing the metal, after manufacture to remove residual manufactured stresses. Never replace a carbon steel bolt with a stainless steel one unless you are sure there are no chlorides present. Bolts can be under severe tensile stress. No one knows the threshold values for stress cracking to occur. We only know that you need tensile stress, chlorides, temperature and the 300 series of stainless steel. We do not know how much chloride, stress or temperature. Until I figured out what was happening I had trouble breaking stainless steel fishing hooks in the warm water we have in Florida. Many cleaning solutions and solvents contain chlorinated hydrocarbons. Be careful using them on or near stainless steel. Sodium hypochlorite, chlorethene. methylene chloride and trichlorethane are just a few in common use. The most common cleaner used with dye checking material is trichloroethane accounting for the reason we sometimes experience cracks after we weld stainless steel and die check it to check the quality of the weld.

EROSION CORROSION This is an accelerated attack resulting from the combination of mechanical and chemical wear. The liquid velocities in some pumps prevents the protective oxide passive layer from forming on the metal surface. The suspended solids also remove some of the passivated layer increasing the galvanic action. You see this type of corrosion very frequently at the eye of the pump impeller. FRETTING CORROSION This type of corrosion is easily seen on the pump shaft or sleeve. You will see the damage beneath: • • • • •

The grease or lip seal that is supposed to protect the bearings. The packing used to seal the fluid. The dynamic Teflon or elastomer used in most original equipment seals. The vibration damper used in rotating metal bellows seals. Under the rubber boot used in low cost seals, if they did not attach them selves to the shaft properly.

As mentioned earlier, 300 series stainless steel passivates its self by forming a protective chrome oxide layer when ever it is exposed to free oxygen. This oxide layer is very hard and when it imbeds into a soft elastomer it will cut and damage the shaft or sleeve rubbing against it. The mechanism works like this: • • • •

Oxygen passivates the active stainless steel forming a protective ceramic layer. The seal or packing removes the oxide layer as the shaft or sleeve rubs against it. The ceramic sticks into the soft elastomer turning it into a "grinding surface". The oxide reforms when the active metal is exposed and the process starts all over again.

•

A visible groove is cut into the shaft or sleeve that will cause seal leakage and "hang up".

CONCENTRATED CELL OR CREVICE CORROSION This corrosion occurs any time liquid flow is kept away from the attacked surface. It is common between nut and bolt surfaces, under O-rings and gaskets, and between the clamps and stainless steel shafts we find in many split seal applications. Salt water applications are the most severe problem because of the salt water low PH (8.0&endash;9.0). Here is the mechanism: • • • • •

Chlorides pit the passivated stainless steel surface. The low PH salt water attacks the active layer that is exposed Because of the lack of fluid flow over the attacked surface oxygen is not available to re passivate the stainless steel. Corrosion continues unhampered under the rubber and tight fitting clamp. The inside of the O-ring groove experiences the same corrosion as the shaft or sleeve.

SELECTIVE LEACHING The process fluid selectively removes elements from the piping or any other part that might be exposed to the liquid flow. The mechanism is: • • • •

Metals are removed from the liquid during a de-ionization or de-mineralizing process. The liquid tries to replace the missing elements as it flows through the system. The un-dissolved metals often coat them selves on the mechanical seal faces or the sliding components and cause a premature seal failure. Heat accelerates the process.

MICRO ORGANISMS These organisms are commonly used in sewage treatment, oil spills and other cleaning processes. Although there are many different uses for these "bugs", one common one is for them to eat the carbon you find in waste and other hydrocarbons, and convert it to carbon dioxide. The "bugs" fall into three categories: • • •

Aerobic, the kind that need oxygen. Anaerobic, the kind that do not need oxygen. Facultative, the type that goes both ways.

If the protective oxide layer is removed from stainless steel because of rubbing or damage, the "bugs" can penetrate through the damaged area and attack the carbon in the metal. Once in, the attack can continue on in a manner similar to that which happens when rust starts to spread under the paint on an automobile.

GALVANIC SERIES OF METALS AND ALLOYS CORRODED END ( ANODIC OR LEAST NOBLE) MAGNESIUM MAGNESIUM ALLOYS ZINC ALUMINUM 5052, 3004, 3003, 1100, 6053 CADMIUM ALUMINUM 2117, 2017, 2024 MILD STEEL (1018), WROUGHT IRON CAST IRON, LOW ALLOY HIGH STRENGTH STEEL CHROME IRON (ACTIVE) STAINLESS STEEL, 430 SERIES (ACTIVE) 302, 303, 304, 321, 347, 410,416, STAINLESS STEEL (ACTIVE) NI - RESIST 316, 317, STAINLESS STEEL (ACTIVE) CARPENTER 20CB-3 STAINLESS (ACTIVE) ALUMINUM BRONZE (CA 687) HASTELLOY C (ACTIVE) INCONEL 625 (ACTIVE) TITANIUM (ACTIVE) LEAD-TIN SOLDERS LEAD TIN INCONEL 600 (ACTIVE) NICKEL (ACTIVE) 60 NI-15 CR (ACTIVE) 80 NI-20 CR (ACTIVE) HASTELLOY B (ACTIVE) BRASSES COPPER (CA102) MANGANESE BRONZE (CA 675), TIN BRONZE (CA903, 905) SILICONE BRONZE NICKEL SILVER COPPER - NICKEL ALLOY 90-10 COPPER - NICKEL ALLOY 80-20 430 STAINLESS STEEL NICKEL, ALUMINUM, BRONZE (CA 630, 632) MONEL 400, K500 SILVER SOLDER NICKEL (PASSIVE) 60 NI- 15 CR (PASSIVE) INCONEL 600 (PASSIVE) 80 NI- 20 CR (PASSIVE) CHROME IRON (PASSIVE) 302, 303, 304, 321, 347, STAINLESS STEEL (PASSIVE) 316, 317, STAINLESS STEEL (PASSIVE)

CARPENTER 20 CB-3 STAINLESS (PASSIVE), INCOLOY 825 NICKEL - MOLYBDEUM - CHROMIUM - IRON ALLOY (PASSIVE) SILVER TITANIUM (PASS.) HASTELLOY C & C276 (PASSIVE), INCONEL 625(PASS.) GRAPHITE ZIRCONIUM GOLD PLATINUM PROTECTED END (CATHODIC OR MOST NOBLE)

How To Measure Mechanical Seals

Page 1 of 2

HOW TO MEASURE MECHANICAL SEALS Section (1) In some cases, we are not able to directly cross reference a seal in our cross reference to pump manufacturers. In order to assure you receive the correct seal for your pump, it is very important that you supply us with accurate information. Below, you will see a sequence of photos which will help us identify the seal assembly you have.

The major rotating components are the "seal head" and the "spring", remove the "seal head from the "spring".

http://www.epm.com/measure.htm

On the back side of the seal head (the end which fits towards the spring), measure the inside diameter. This will determine the seal size diameter - which is either the shaft size or the sleeve size that the seal fits over. The front side of the seal usually has a black carbon face. The carbon face runs against the "stationary" portion or "seat" of the seal assembly.

Now, measure the "spring" length while it is separated from the "seal head".

09-04-01

How To Measure Mechanical Seals

Page 2 of 2

The next step is to measure the outside diameter of the "stationary" or "seat".

Then measure the thickness of the "stationary" or "seat".

Section (2) Once you have determined, the sizing - then we need to know the materials the seal is made from. Please also supply us with the following additional information: 1. What is the rotating seal face made from - is it a black ridgid material - most likely carbon? 2. What are the rubber or elastomer portions of the seal made from? Unless you are pumping chemicals or high temperature liquids, then it is most like Nitrile; also called NBR or Buna-N. 3. What is the "stationary" or "seat" made from? Is it a white, hard material? Or, is it a metal? 4. What shape is the "stationary" or "seat"? Does it have an o-ring around the outside, which is in a groove (like the photo above)? Or, does the "stationary" or "seat" have a rubber piece it fits into which is like a cup? So, the "stationary" or "seat" is either an o-ring mount or a cup mount type. Once you have given us the information above, you will be assured of receiving a seal assembly that will fit and function correctly in your pump. The alternative is to send us the entire seal assembly and we will size it for and identify the materials, even if the parts are broken or worn. information click here to contact us. Join Our Affiliate Program | Terms & Conditions Home | Services | Online Store | EPM In The News | FAQs | Contact Us | E-mail

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09-04-01

Improved Gear Bearings NASA Goddard Space Flight Center (GSFC) offers potential partners a superior evolution of its patented gear bearing technology. The new design incorporates rifle true anti-backlash, improved thrust bearing performance, and phase-tuning techniques for superior low speed reduction. The gear bearing technology combines gear and bearing functions to reduce weight, number of parts, size, and cost, while also increasing capacity and performance.

Benefits •

Precise control: Rifle true anti-backlash produces a planetary transmission with zero backlash, resulting in smoother operation and superior control.

•

Improved thrust bearing: Gear teeth give superior thrust bearing performance.

•

Improved speed reduction: Through phase tuning, a one-tooth difference between ground and output rings is possible, creating opportunities for significant reduction ratios at both low and high speeds.

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Less noise and vibration: More evenly distributed planet loading reduces cyclic loading and rough spots, reducing noise and vibration.

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Low cost, simple design: Gear bearings combine gear and bearing functions to reduce materials and cost, while also reducing weight and simplifying the design.

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Fewer fatigue failures: Reduced cyclic loading reduces susceptibility to fatigue failures.

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High strength: Gear bearings are more structurally rigid and provide higher overall load capacity compared to fixed planetary designs.

Applications •

Transportation (including automotive, aircraft, marine, and rail): Transmissions, electric windows, windshield wipers, steering mechanisms, alternators/generators, engines and propellers, control systems, landing gear, door openers, rudders/steering/leveling controls, winches, rail switching systems

•

Power tools: Garden equipment, hand tools, lawn-mowers, chain saws, log splitters

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Industrial machinery: Power presses, lathes and grinders, slitting and rolling equipment, construction equipment, lifting and handling equipment

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Farm equipment: Tractors, harvesters, hay rollers

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Medical equipment: MRI, CT, PET scanners

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Toys: Electric robots, cars, and other motorized toys

Technology Details Gear bearings function both as gears and bearings, providing superior speed reduction in a small package. Smoothness and precision control are created through reduced micro chatter and elimination of rotational wobble. The technology provides for tighter mesh, more even gear loading, and reduced friction and wear. NASA’s gear bearing technology has been further enhanced with the following: •

To increase thrust bearing point contact loads, helical gear teeth forms (including herringbone) were developed. These provide outstanding thrust bearing performance.

•

To provide unprecedented high and low speed reduction, phase tuning was implemented into the design, achieving successful reduction ratios of 325:1. Phase tuning allows differentiation in the number of teeth that must be engaged between input and output rings in a planetary gearset.

•

To provide smooth and accurate control, rifle true anti-backlash was incorporated into the gear bearing design to produce a planetary transmission with zero backlash.

Gear Bearings are compatible with most gear types, including spur, helical, elliptical, and bevel gears. In planetary gearsets, Gear Bearings can eliminate planet carriers or planet bearings, substantially reducing parts count and cost. The ring gear can be mated with a motor housing to eliminate the motor's front bearing and further reduce cost. In addition, Gear Bearings eliminate concerns that the center of the carrier is coincident with the center of the sun gear and equalizes the loading for the planet gears. Gear Bearings are also very compact. The prototype two-stage planetary gearset shown in the header of this document is 1.25 inches in diameter and can produce speed reductions from 2:1 to 2,000:1. By selecting the appropriate manufacturing method, Gear Bearings can be tailored to benefit any application, from toys to aircraft. For tight tolerance applications such as in helicopters, machined Gear Bearings could provide very high performance. For medium range applications such as in power tools, cast Gear Bearings could reduce cost and size over conventional gearsets. For low tolerance applications such as in toys, injection molded plastic Gear Bearings could substantially reduce cost. Beyond reducing parts, these examples illustrate how the high load capacity of Gear Bearings could enable cost reductions by use of lower price, lower strength materials. Alternatively, Gear Bearings with higher price, high strength materials could enable lighter and more compact gearsets for the same cost. NASA's Gear Bearing technology is based on two key concepts, the Roller Gear Bearing and the Phase-Shifted Gear Bearing. All designs are capable of efficiently carrying large thrust loads.

Roller Gear Bearing The Roller Gear Bearing has crowned roller ends with the roller radius equaling the gear pitch radius. The gear can mesh with another roller gear such that the crowned ends will be in rolling contact -- no sliding will occur.

Phase-Shifted Gear Bearing The Phase-Shifted Gear Bearing has teeth rotated with respect to each other such that the top land of one side intersects with the bottom land of the other side. In the area where the teeth intersect, the teeth are beveled from the top of the tooth down to the root circle. When meshed with another phase-shifted gear, the beveled tooth surfaces contact each other as with four-way thrust bearings.

Potential Gear Bearing technology is being developed for use in two NASA missions. For the Next Generation Space Telescope (NGST), electric motors turn 325:1 Gear Bearings to drive jack screws that position telescope mirrors. Gear Bearings are also being considered for high performance speed reduction in the next generation Pistol Grip Tool (PGT), the power tool used by astronauts for space station assembly and servicing. The annual U.S. planetary gearbox market is estimated at $700 million. Other gearset markets may also have applications for this technology, including use in linear actuators, speed reducers, gearheads, and wheel and track drives. In addition, Gear Bearings could play a key role in the transition from hydraulics to electric actuators.

Prototypes NASA has prototype hardware to show how the base Gear Bearing technology works. If a company requires

prototype hardware directly related to an application, the company(s) listed below is/are licensed to provide technical consulting and prototype hardware as needed. No companies are currently licensed.

Prototyping Opportunities If your company is interested in becoming a licensed supplier of gear bearing prototypes, please refer to Partnering Options.

Partnering Opportunities For information and forms related to the technology licensing and partnering process, please visit the Licensing and Partnering page. (Link opens new browser window.)

For More Information If you are interested in more information or want to pursue transfer of this technology, please contact: Office of Technology Transfer NASA Goddard Space Flight Center Greenbelt, MD 20771 Phone: (301) 286-5810 E-mail: [email protected]

Maintenance practices that cause high seal and bearing maintenance problems Maintenance departments seldom return savings to the company management. They fear that if they do not spend this year's budget, next year's will be reduced. Management views maintenance savings as bottom line money and works at reducing maintenance man power and inventory costs. Here are some of the maintenance practices that increase the pump failure rate: Problems with pump maintenance that can cause excessive shaft movement and deflection. This shaft displacement is a major cause of premature seal and bearing failure. •

• • •

•

•

•

• •

•

•

•

Failure to align the pump and driver. Misalignment will cause the mechanical seal to move excessively, increasing the chance for the seal faces to open and fail the seal. Pipe strain is another cause of misalignment between the seal's stationary and rotating faces. Wear ring damage is common if pipe strain is present. Failure to dynamically balance the rotating assembly can result in "whip, wobble, and run-out problems." Damage to the shaft and bearings during the sleeve removal process. Banging on the sleeve with a large hammer or heating the shaft with a torch are common methods used to remove sleeves. Needless to say the seal and bearings stand a good chance of being destroyed in the process, along with the shaft that will be bent or warped. Using the coupling to compensate for misalignment. A coupling is used to transmit torque and compensate for axial growth of the shaft, nothing else! It cannot compensate for misalignment between the pump and its' driver. Trimming the impeller without dynamically re-balancing it. The impeller casting is not homogeneous, it must be re-balanced after any machining operation has taken place. Throttling the pump discharge to stop a cavitation problem. The more you pump the more N.P.S.H. you need, so throttling does work, but you may be now operating off the pumps' B.E.P. resulting in shaft deflection. Failure to machine the stuffing box square to the shaft will result in excessive seal movement. Repairing the cutwater to the wrong length can cause a cavitation problem known as the "Vane Passing Syndrome" that will damage the tips of the impeller blades and damage the volute just beyond the discharge nozzle. Failure to properly adjust the open impeller clearance or letting the closed impeller wear ring clearance become excessive can make the pump run inefficiently and vibrate. Turning down a shaft and repairing it with a polymer material will weaken the shaft making it more sensitive to deflection forces. That practice was common with packed pumps, but should be avoided when mechanical seals are being used. Substituting a globe valve for a gate valve will throw the pump off of its B.E.P. causing shaft deflection.

• •

Any alteration in the piping system or failure to prevent solids "build up" in the lines will have the same affect. Mounting the pump and motor on too light a foundation. The foundation should be at least five times the mass of the equipment sitting on it or vibration will become a problem. Proper grouting is also necessary to mate the base of the pump to the foundation.

Seal handling practices can also lead to premature seal failure. •

•

Installation problems: o The seal is installed at the wrong length. There are a lot of ways to do this from reading the print wrong to the sleeve moving after the seal was attached to the sleeve. o The wrong lubricant was used on the dynamic rubber part causing it to be chemically attacked. Petroleum grease on Ethylene Propylene O-rings is a good example of this problem. In salt water applications zinc oxide should be used on all rubber parts and metal components that clamp together. o The seal was installed before the impeller setting was made or an impeller adjustment was made without resetting the mechanical seal. In most cases this will cause the seal faces to open prematurely. o The shaft or sleeve is out of tolerance. This can cause serious problems with those seal designs that have a dynamic elastomer sliding on the shaft (most original equipment seals fit into this category). o The sleeve was hardened to resist packing wear causing the seal set screws to slip and the faces to open. o The elastomer (rubber part) exceeded its' shelf life. This is a real problem with the Buna "N" material found in most rubber bellows seals. o Installing a stationary seal on a cartridge will cause the rotating face to "cock" when the set screws are attached to the shaft. An environmental control was lost while the seal was installed in the pump. Typical environmental controls include: o Clean flushing liquid to keep solids away from the moving seal parts. o Controlling stuffing box temperature with a cooling or heating jacket. If the circulating water is "hard" condensate may have to be substituted to prevent the cooling jacket from becoming coated with calcium and other solids that will interfere with the heat transfer. o Barrier fluid is used to circulate between two mechanical seals. Sometimes the circulation is done by simple convection, but pumping rings and forced circulation are common also. Check to see if your convection tank has to be pressurized. This is a common problem with many original equipment seals. Feel the convection lines to make sure the convection is taking place in the right direction. o A steam quench is often used to remove dangerous vapors and to keep the seal area warm when the pump is shut down. Metal bellows applications use the steam quench to cool down hot oil to prevent unwanted "coking".

A stuffing box vent should be connected from the area of the seal faces to the suction side of the pump, or some other low pressure area to prevent air from being trapped at the seal faces. o A discharge recirculation line, and a bushing in the bottom of the stuffing box are often used to pressurize the stuffing box to prevent the product from vaporizing at or between the lapped seal faces. Is there enough clearance between the seal outside diameter and the inside of the stuffing box? Solids build up in the stuffing box can interfere with the free movement of the seal. The seal was installed with unidentified materials, making troubleshooting almost impossible. Which carbon is being used? There are a hundred available and they are not all alike. Which elastomer? Do you know both the material and the grade? What material are the metal components manufactured from? Not all stainless steel grades are alike, and stainless steel springs or metal bellows should never be used because of potential problems with chloride stress corrosion. There are many hard seal faces in use. All ceramics, silicone carbides and tungsten carbides are not alike. The outside springs were painted on a double seal when the pump area was refurbished. The pump discharge recirculation line is handling abrasive solids. They are being directed at the lapped seal faces or at the thin metal bellows. If the open impeller is adjusted backwards (this is a common problem if a facility has both Duriron and Goulds pumps), it can create a vacuum in the stuffing box as the "pump out vanes" are running too close to the back plate. Do not shut off the stuffing box cooling jacket when a metal bellows seal is installed. The stuffing box is cooling down the shaft as well a the seal area. Shaft cooling is necessary to prevent heat conduction to the bearings. o

•

• • • •

• • • •

•

Poor bearing maintenance practices are a major cause of premature bearing failure. • •

•

•

If the oil level is too high or the bearings are over greased, the low specific heat of the lubrication and its poor conductivity will cause the bearing area to over heat. The inside of the bearing case must be protected against rust when it is stored as a spare. The bearings should be coated with an appropriate grease because they can rust also. During storage, or while in a standby condition, nearby equipment that is vibrating can induce vibration into the static bearings causing "false brinneling" or hardening of the bearing balls and races. If the oil becomes contaminated with water you will experience a very rapid bearing failure. The water can enter through the grease or lip seals from several sources: o leakage through the packing or mechanical seal. o From the water hose that is used to wash down the base plates and pump area.

From moisture in the air as the moisture enters the bearing casing through "aspiration". o From the quench gland on a mechanical seal. The bearing was installed improperly: o The shaft outside diameter was the wrong tolerance. Remember that the tolerance is given in tenths of thousands of an inch or thousands of a millimeter. o Too much pressure was put on the arbor press during the assembly sequence. o The bearing was heated in contaminated oil that has deposited the contaminates in the races o The oil was over heated and varnish particles are now in the race ways. o The bearing was pushed too far up a tapered shaft. The thrust bearing is being retained by a simple snap ring. During operation the shaft thrust is usually toward the volute and this thin ring. o

•

•

Mechanical Seal design, operation, and maintenance problems In my presentations I teach that mechanical seals fail prematurely because: • •

The lapped faces open. A seal component becomes damaged.

In the following paragraphs we will learn how these failures can be separated into: • • •

Design problems. Operation problems. Maintenance problems.

These individual subjects have been discussed in other sections of this Technical Series. The purpose of this paper is to give you an overview of the subject, and assist you in your troubleshooting function. MECHANICAL SEAL DESIGN PROBLEMS Problems with the Seal Faces: • • • • • • • • • • • • • • • • •

Wrong carbon or hard face selected. The material is not compatible with the fluid you are sealing, and the cleaner or solvent used to clean or flush the system. Face flatness problems: The face cross section is too narrow causing temperature or pressure distortion problems. The material modulus of elasticity is too low. The face is not hard enough. All clamping forces must be "equal and opposite" to prevent face distortion. In many designs they are not. The differential expansion between the seal face and its holder can cause the face to go out of flat. The faces were not lapped at a cryogenic temperature and the seal is being specified for cryogenic service. Bad packaging. Poor heat conductivity: Carbon is a poor conductor of heat compared to most hard faces. Many ceramics are not good conductors of heat . Plated or coated faces can "heat check" due to a differential expansion rate between the coating and the base material. The seal face is sometimes insulated by a gasket or elastomer. Low expansion steel face holders are not usually corrosion resistant. No vibration damping has been provided to prevent "slip stick" vibration problems. This is a major problem with metal bellows seals. Unbalanced seal designs require excessive flushing or cooling to remove unwanted heat.

•

The carbon must be dense enough to prevent entrained air pockets from expanding and causing pits in the carbon face. An "unfilled carbon" with four impregnates is the best.

The Springs or bellows. • • • • • • •

Springs in the fluid can clog easily, especially the small springs. Stainless steel springs and bellows are sensitive to chloride stress corrosion problems. A single spring can be wound in the wrong direction. Thin bellows plates and small cross section springs are sensitive to abrasive wear. Rubber bellows experience a catastrophic failure mode when the bellows ruptures. Stressed metal corrodes faster. Springs and metal bellows are subjected to high stress. Too much spring or bellows movement will cause an early fatigue of the metal.

The Dynamic Elastomer (the one that moves) • • • •

Some elastomers do not move to a clean surface as the face wears. Spring loaded elastomers stick to the shaft or sleeve and are sensitive to the shaft diameter and finish. Elastomers positioned in the seal face are subject to the heat generated between the seal faces. Dynamic elastomers are very sensitive to the shaft tolerance and finish.

Operating conditions too severe for the design. • • • • • • •

Elastomers and some seal faces are sensitive to temperature extremes. Excessive pressure can distort seal faces causing them to go out of flat. Excessive pressure can cause elastomer extrusion. High speed can separate the seal faces in rotating seal designs. High speed can cause excessive heat at the seal faces. Excessive shaft movement separates faces also. Hard vacuum can "out gas" an elastomer causing it to leak.

Dual seals • • • • •

Rotating "back to back" designs: Centrifugal force throws solids into the inner faces. Inner seal blows open if barrier fluid pressure is lost. Inner stationary face is not positively retained to prevent movement if the pressure is lost between the faces. When the outboard seal fails the inboard will fail also due to the pressure drop between the faces.

• •

The inner seal has to move into the sealing fluid as the face wears. This is a major problem if the fluid contains solids. Failure to use "two way" hydraulic balance causes the inner faces to open with a reversal in barrier fluid pressure.

Design problems that cause excessive shaft movement • • • • • •

An elbow is installed too close to the pump suction inlet. The mass of the foundation is not five times the mass of the pump and its driver. Wrong size pump was specified because of safety factors and, as a result, the pump is operating off the B.E.P. The pump was selected oversize in anticipation of a future need. A "centerline" design should have been selected when the operating temperate exceeded 200°F (100°C). The shaft L3/D4 is too high.

The pump is cavitating due to a design problem. • • • • • • •

Too high a N.P.S.H. is required. You need a double suction pump. The suction specific speed number is too high. You are using too low a specific speed impeller. A reducer has been installed up side down, letting an air pocket into the suction. The impeller to cutwater clearance is too low. There is too much suction resistance due to excessive piping. Too much suction lift for the fluid temperature.

Other design problems • • • • • • • • • • • • • •

Some seal designs cannot compensate for thermal shaft growth or impeller adjustment. Cartridge versions are needed for this feature. The pumping fluid is located at the inside diameter of the seal faces. Solids will be thrown into the lapped faces destroying some face materials. Solids will pile up in front of the movable faces, preventing them from compensating for wear. Most seal faces are weak in tension. Hysteresis (delay) problems caused by the seal mass and sliding elastomers. Poor packaging that allows face damage during shipment and storage. Designs that frett (damage or groove) the shaft or sleeve. High speed requires the use of stationary seal designs. Centrifugal force can open rotating designs above 5000 fpm. (25 m/sec.) The seal is positioned too far from the bearing housing. Lack of a self-aligning feature is causing excessive face movement. A tapered stuffing box can cause face damage. No vent has been provided to vent the stuffing box in a vertical application. Hardened shafts and sleeves can cause the seal set screws to slip.

•

A discharge recirculation line is aimed at the lapped faces, causing them to wear, and interfering with the seal movement.

Problems caused by the product you are sealing. • • • • • • • • • • • • • • • • •

The fluid can flash or vaporize between the faces. Viscous fluids open seal faces as they restrict seal movement. Products that solidify will open and damage seal faces. Crystallizing products restrict seal movement and open the faces. Film building products cause the faces to open. Hot oil is typical. The fluid can attack one of the seal components, especially the elastomer. All chemicals have the potential for corroding a seal component. It is just a mater of time. Some fluids are poor lubricants: This can cause excessive wear. Color contamination problems as the carbon wears. "Slip stick" vibration problems. Slurries clog up the sliding seal components and open the faces. Cryogenic fluids can attack some carbon faces and most elastomers. High temperature fluids attack elastomers and change the state of the fluid you are sealing. Some fluids can cause the formation of ice outboard the seal, restricting seal movement as the face wears. Agitation can cause some fluids to change their viscosity. Cleaners or solvents are attacking a seal component.

OPERATION PROBLEMS Operations that cause excessive shaft movement that will open or damage the seal faces • • • • • • • • • • • • • •

Opening and closing valves in the suction and/or discharge causing the pump to operate off the B.E.P, and the shaft to deflect. Pumping the supply tank dry, causing excessive vibration and heat. Series or parallel pump operation can cause shaft deflection. Running at a critical speed will cause the shaft to defect. Cavitation problems: Low N.P.S.H. Air getting into the system through packing. A stuffing box, suction recirculation line is heating the incoming fluid. A discharge bypass line is heating the suction fluid A discharge recirculation line is aimed at the seal face restricting its movement. Water hammer is opening or damaging the lapped faces. The piping system has been altered since the pump was installed. The pump is being started with the discharge valve shut or severely throttled. Starting a pump with the discharge valve open is just as bad.

Operations that cause excessive heat and corrosion problems • • • •

Cleaners or solvents used in the lines can attack a seal component, especially the elastomer. A product concentration change will affect corrosion. A change of product. Either a temperature or pressure change in the system will affect both.

Operations that cause the seal faces to open • • • • • • • • • • •

The seal is seeing frequent reversing pressures. Loss or lack of an environmental control. Flush not working. Quench is shut off. Barrier fluid not circulating. Loss of heating or cooling. Heating jacket clogged. Pressure drop in the stuffing box. Flushing with a dirty product. Quenching with shop water leaves solids outboard Of the seal that will cause a hang-up as the seal moves forward to compensate for wear. The quenching steam pressure is too high. It is getting into the bearings.

MAINTENANCE PROBLEMS • • • • • • • • • • • • • • • • • • •

The pump and driver are not aligned&emdash;causing excessive seal movement. Pipe strain. Thermal growth. Bad installation techniques that can injure a seal component. Wrong lubricant put on the dynamic elastomer. The impeller clearance was set after the seal installation. The face is inserted backwards, only one side is lapped. The seal is set at the wrong installation length. The sleeve moved when the impeller was tightened to the shaft. A lubricant was put on the seal face that froze when the product evaporated across the lapped faces. The rotating assembly is not dynamically balanced. The shaft is bent. The sleeve is not concentric to the shaft. Impeller clearance is not being maintained, causing vibration problems. The impeller is positioned too close to the cutwater. The seal has been set screwed to a hardened shaft. No seal or gasket between the shaft sleeve and the solid shaft. This is a big problem with double ended pumps. The seal environmental control is not being maintained. Flushing fluid is being restricted or shut off.

• • • • • • • • •

Quenching steam is shut off. The barrier fluid tank level is too low The convection tank is running backwards. The cooling jacket is restricted due to a calcium build up. You are running both a discharge recirculation line and a cooling jacket. Out of tolerance shaft dimensions will restrict seal movement. The impeller clearance was made without re-adjusting the seal face load. The shaft sleeve was removed to accommodate a smaller diameter seal. The sleeve was providing corrosion resistance. A gasket is protruding into the stuffing box restricting the seal movement.

Pump and seal problems with no apparent cause These problems are the ones that drive you crazy. No matter how hard you look the solution keeps evading you. Over the years I have collected quite a few examples. I offer some of them for your enjoyment and maybe, in the process, they will help you solve the "un-solvable" CAVITATION The pump cavitated every time it rained. •

Solution: The product temperature would cause it to vaporize very close to ambient pressure, and when it rained atmospheric pressure dropped enough to cause the problem.

The pump never cavitated in the summer months, only during the winter when everything was cooler. •

Solution: The tank vent froze during the winter months causing the pump to pull a partial vacuum in the tank.

The cavitation started suddenly. •

Solution: A plastic pipe liner collapsed at the suction side of the pump or the gate fell off a gate valve.

The cavitation started after the packing was converted to a mechanical seal. A careful inspection showed that the seal was not leaking air into the suction. •

Solution: The pump had speeded up (increased the rpm) when the packing was removed. This increase in speed and capacity caused the cavitation.

The cavitation kept getting worse with time, nothing obvious had changed in the system. •

Solution: The product had formed a coating on the inside of the suction pipe increasing the pressure drop and resulting in a loss of suction head.

The cavitation only occurred when there was a higher head at the suction of the pump and stopped cavitating when the level fell in the tank - just the opposite of what should have happened. •

Solution: The pump was pumping to a fixed discharge head. The capacity of the pump increased when the suction level was higher, because the pump delivers the difference between the suction and discharge head. When the differential went down, the capacity increased.

Two pumps were installed in parallel, one cavitated the other did not. They had separate suction lines so that was not the problem. •

Solution: Some one had installed an oversized section of pipe on the discharge side of the pump that was cavitating. The lower discharge resistance caused an increase in capacity which caused the cavitation. When the proper sized pipe was installed the cavitation stopped.

The pump had been cavitating for some time, but after a visual check everything appeared normal. •

Solution: A globe valve had been substituted for a gate valve on the suction of the pump. A globe valve can add the equivalent of another 100 feet (30,5 meters) of pipe to the system.

The pump started to cavitate when a flange gasket was replaced on the suction side of the pump. •

Solution: The inside diameter of the gasket was too small. It was acting as an orifice, and restricting the flow.

The pump cavitated about one third of the time it was running. •

Solution: A close inspection of the system revealed that there was no surge tank installed between the pump discharge and the multiple outlets that were using the product. The pump was acting like an accumulator and started to cavitate when the demand went up and the discharge head dropped.

The pump cavitated although here was excessive suction head available. •

Solution: There was too much velocity on the suction side of the pump. I saw this problem in Scandanavia in an application where the pump was taking a suction on a flow of water coming off of a mountain.

THE SEAL WAS GETTING HOT The seal was showing evidence of running dry, but the fluid level was never lost in the pump. •

Solution: Air was trapped in the stuffing box of a vertical pump after it was converted from packing to a mechanical seal. Most seal designs have no facility for venting the stuffing box in a vertical application

The seal showed evidence of running dry.

•

Solution: The open impeller had been adjusted backwards and the "pump out vanes" on the rear of the impeller were pumping the stuffing box dry. This happens if you are using several brands of pumps and the maintenance mechanics confuse the impeller adjustment method. Some pumps adjust towards the volute (Goulds), some adjust towards the back plate (Duriron). It is easy to mix them up.

There was little to no fluid circulating between the two seals. •

Solution: The pipe fitting had bottomed out in a gland inlet elbow shutting off the flow. This sometimes happens after the seal has been repaired several times and the pipe thread shows some wear letting it protrude further into the elbow fitting.

The mechanic had marked the seal location on the shaft sleeve before the impeller was installed. When the impeller was tightened against the shaft shoulder the sleeve moved and over compressed the seal. Durco pump impellers adjust to the pump back plate. When you make impeller adjustments you over compress the mechanical seal. A cooling jacket was being used, but the seal continued to get hot. I have seen multiple reasons for this: •

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• • •

A discharge recirculation line had also been installed, but it was hidden by some insulation. The cooling jacket could not keep up with the heat being added by the recirculation line. The inside of the cooling jacket had become coated with a layer of calcium because hard water was being used as the cooling medium. Condensate should have been substituted. A thermal bushing had not been installed in the bottom of the stuffing box. The cooling jacket flow changed with fluctuations in shop water pressure. The inner seal of some double seal applications can get hot if the mechanic installs the cartridge seal by pushing on the gland and fails to reset the seal compression with the installation clips. The interference from the cartridge sleeve elastomer can cause enough resistance to compress the inner seal and unload the outer seal.

THE SEAL WOULD LEAK FOR NO APPARENT REASON The open impeller was being adjusted without resetting the seal. Many operators make their own impeller adjustments. The seal faces were opening because the equipment's sliding foot had been bolted to the floor allowing the shaft to grow through the stuffing box when the unit came up to temperature.

The cartridge seal had been hydrostatically tested with water and then put into a hot oil application. It leaked almost immediately. •

Solution: The trapped water vaporized when the unit was started. This could be a dangerous condition because water trapped in a gasket and then flashed to steam could blow the equipment apart.

The seal would start leaking about thirty minutes after the pump started. •

Solution: The carbon insert would come loose in its holder when the seal came up to temperature. At shut down the metal holder would shrink and everything appeared normal.

The seal was tested in the shop, but leaked when it was installed in the pump that was operating at cryogenic (cold) temperature. •

Solution: The faces had to be lapped at cryogenic temperature to keep them flat at the seal operating temperature. The cryogenic temperature can also harden the Oring and freeze any lubricant that was put on the seal face.

The seal was found to be leaking every Monday morning. •

Solution: A utility man did not know about seals. He would loosen the gland on the weekend so that what he thought was packing would drip a little. The leak was found by the regular maintenance people every Monday morning.

The leakage occurred during the winter months. •

Solution: Someone circulated commercial anti freeze between two seals to act as a barrier fluid. The brand they selected contained a chemical to plug up radiator leaks and it kept plugging up the seal.

The seal would fail only during the winter months. The problem was traced to swelling of the dynamic O-ring but no logical reason could be found for its failure. •

Solution: During the winter months a worker decided to oil the bed of his dump truck to make the mined, raw product slip off easier. The petroleum oil he used attacked the Ethylene Propylene (EPR) O-ring in the mechanical seal, installed downstream in the system.

The seal area was wet, but no visible leakage could be seen. •

Solution: It turned out that there was a flange leaking above the pump and dripping the product next to the shaft.

The problem was traced to the fact that the mechanic was installing the seal at the wrong dimension. The written instructions were clear and placed in the box and yet the mechanic continued to do the installation incorrectly. •

Solution: The mechanic could not read. He had been faking it for many years and was quite good at it. The same problem occurs with older mechanics that refuse to wear glasses and as a result cannot see the funny little lines between the numbers on their measuring scale.

The centrifugal pump discharge was connected to the bottom of a surge tank. As the tank filled, the pump operating point shifted from too much capacity to too much head, deflecting the shaft in two directions. The outside seal in a double seal application failed suddenly. Nothing had changed in the system. •

Solution: Routine maintenance included repainting the pump. The paint spray got into the outside seal springs and stopped them from moving.

The seal ran great for several days and then started to leak. It tested all right on the test bench after it had been removed from the pump. •

Solution: It had been set screwed to a hardened sleeve and the set screws gradually loosened.

The seal was changed several times, but the steady leak persisted. •

Solution: The leak was occurring between the pump sleeve and the shaft. This is a common problem in double ended pumps that have been converted to a mechanical seal. You often have to devise a method of sealing the sleeve to the shaft or the sleeve to the impeller because the manufacture has not provided one.

The seal started to leak after many months of service. A bench vacuum test showed that the seal was all right. •

Solution: The seal was fretting the shaft below the Teflon wedge allowing the leak to come through this groove.

The seal ran approximately six months and then failed. •

Solution: The lines were steam cleaned and the wrong grade of Viton® was in the seal. Most Viton® compounds will be attacked by steam, caustic or other water based solutions.

The seal was installed correctly, but it leaked immediately.

•

Solution: The solid, hard face is usually lapped on only one side. The face had been installed backwards and the rotary unit was running on a non lapped surface.

SEAL COMPONENT DAMAGE IS VISIBLE, BUT WHAT IS THE CAUSE? It looked like a seal part had come loose in the stuffing box, but all of the parts were there. •

Solution: During a previous installation a small spring had been lost when it fell into a drain hole in the bottom of the seal gland. It came loose after a later installation. This is a problem when several people work on the same pump.

The bellows plates were breaking but there was no evidence of corrosion, excessive wear, or vibration. •

Solution: A discharge recirculation line was directing high velocity abrasive particles at the thin metal section of a metal bellow seal.

The inner seal of a dual, rotating "Back to back" seal was showing excessive face wear in a short period of time. •

Solution: The inner seal stationary face was not locked in the bottom of the stuffing box and when the system pressure overcame the barrier fluid pressure, the stationary face was pushed into the inside rotating face. When the pump was stopped the spring pushed the stationary face back to its normal position.

The carbon showed massive damage in a cryogenic (cold) application. •

Solution: The carbon had been lubricated at assembly and the lubricating oil froze in the cryogenic atmosphere.

The bellows plates showed massive wear. •

Solution: The seal was rotating in an abrasive slurry. Metal bellows seals should be designed to rotate the fluid inside of the stuffing box, instead of rotating through the fluid.

OTHER PROBLEMS The pump had been recently overhauled and at start up the pump was reading high amperage, but low flow. •

Solution: One of the wear rings had been left off of the suction side of the impeller and the fluid was recirculating to the pump suction.

The pump made a terrific racket during start up. It produced the proper head, but the capacity was less than anticipated. •

Solution: It was a two speed pump and the second speed had been wired backwards.

In an acid application, a stationary seal showed localized corrosion only on the gland. •

Solution: This was an older pump with a bolted on stuffing box that would slip because the bolts were worn. This caused the shaft to run against the gland causing it to overheat and, in an acid application, the corrosion rate of the acid doubles with an 18° F. (10° C) rise in temperature. It doesn't make any difference if the acid or the part gets hot, the affect is the same.

The dual seal convection tank was running backwards. •

Solution: The seal was not centered in its gland, and as the shaft turned, the close tolerance between the seal and the gland outlet increased the velocity of the liquid enough to drop the pressure and cause the tank to convect backwards.

The pump was converted from packing to a mechanical seal and then started to break shafts. •

Solution: The pump was operating way off of its best efficiency point, causing major shaft deflection. The packing was acting as a bearing and supported the shaft during this deflection.

The product was solidifying in the stuffing box. Steam was being used to heat a jacket around the pump. The header gauge showed adequate pressure. •

Solution: The gage was located too far away from the pump jacket. The line was not insulated and this allowed the steam to experience a pressure drop between the header and the stuffing box heating jacket. The result was that the steam cooled down below the necessary heating temperature. The problem was only visible when the pump was stopped for a period of time.

The nickel base tungsten carbide face was being chemically attacked. •

Solution: A galvanic action occurred between the passivated stainless steel and the active nickel contained in the tungsten carbide face.

Shaft displacement cheat sheet This paper is the first in a series of "Cheat Sheets" that will summarize the information you have been learning through out this series. These sheets will not substitute for your learning of the subject. They are intended to outline the subject in a logical way so that you will be able to make an effective presentation on any of the subject material. We know that seals fail for only two reasons: • •

One of the seal components becomes damaged. The lapped faces open.

Common sense dictates that the more the shaft deflects from the center of the stuffing box, the more likely the lapped faces are to separate. Rotating seals (the spring loaded face rotates with the shaft) are very sensitive to this type of shaft displacement or any other form of misalignment between the stationary and rotating faces. THE PUMP IS OPERATING OFF THE BEST EFFICIENCY POINT (BEP) • • • • • • • •

• • • • • • • • • •

The wrong size pump was selected because safety factors were added to the computations. A discharge valve is throttled to decrease the excessive capacity . An orifice has been installed in the discharge piping to limit flow. Two pumps are being operated in a parallel mode with different diameter impellers. Two pumps are being operated in series with different width impellers. The suction tank level is increasing or decreasing dramatically. The centrifugal pump is discharging into the bottom of the tank instead of the top. The head is changing. The system pressure is being maintained by a head tank. The centrifugal pump is acting like an accumulator because it stars when the head tank pressure falls and stops when the pressure tank pressure is reached. You are using a variable speed motor, trying to maintain a flat system curve. The impeller diameter has been changed. The specific speed of the impeller is too high or too low for the application. The piping system has changed: There have been piping additions and deletions since the pump was originally sized. Extra pumps have been installed into the system. The piping inside diameter is reduced because of product build up. A globe valve has been substituted for a gate valve in the system. An oversized impeller was installed to satisfy a system requirement. The pipe was damaged when a truck ran over it.

THE PUMP IS CAVITATING

• • • • • •

Suction vaporization. The suction temperature is too high or the suction pressure is too low. Vane passing syndrome. There is not enough clearance between the tip of the impeller and the pump cut water. The suction specific speed number is in excess of 8500 (5000 in the metric system) Air ingestion. The fluid is vortexing at the pump suction or air is entering the system through packing, valves above the water line, flange gaskets, etc... Turbulence&emdash;there is an elbow too close to the suction. A discharge bypass line is recirculating to the pump suction.

THE PUMP IS VIBRATING • • • • • • • • • • • • •

Dynamic unbalance of the rotating assembly caused by erosion, corrosion, or damage. Harmonic vibration. The shaft is vibrating in harmony with something close by. Slip stick. The seal faces are slipping and sticking due to poor lubrication. Water hammer. The pump is hitting a critical speed. Bent shaft Bad bearings Poor lubrication. Contamination of the lubricant. Poor quality. Bad installation. Over lubrication. The bearing is being retained by a snap ring..

OTHER CAUSES OF SHAFT DISPLACEMENT • • • • • • • • •

Pipe strain caused by either mechanical or thermal expansion. Misalignment between the pump and driver. Pulley driven designs. Start up thrust. Water hammer High L3/D4 number Thermal growth, both axial and radial. Impeller adjustment. The pump pedestal is not five times the mass of the hardware sitting on it.

THE STUFFING BOX IS CAUSING THE MISALIGNMENT PROBLEM • • •

The face is not machined square to the shaft. The stuffing box is not concentric with the shaft. Some bolted on stuffing boxes can slip with vibration.

Some common misconceptions about mechanical seals 1. Two hard faces are a sensible choice if there are dirt or solids in the product you are pumping. o

ans. Seal faces are lapped to less than three light bands (less than one micron) of flatness. Dirt and solids cannot penetrate these faces unless they open. The trick to sealing solids and slurry is to keep the lapped seal faces together.

2. Dual seals are a good choice for a slurry application. o

ans. Putting a clean liquid between two seals is not going to stop solids from clogging the inner seal. Since the barrier fluid is at a higher pressure than the stuffing box pressure you will probably end up diluting your product.

3. Putting the seal outside the stuffing box can keep the springs and other parts from clogging in an abrasive slurry. o

ans. As the seal faces wear the seal is going to have to move into the slurry that will restrict its movement. It is the same problem you face with many of the dual seal applications used to seal dirt and solids..

4. You should not use ceramic seal faces in a mechanical seal. They will crack when subjected to temperature transients. o

ans. Space vehicles are covered with ceramic so they can take temperature transients, its just a matter of which ceramic you are using.

5. Modern seal designs are made to fit A.N.S.I. and I.S.O. pump designs without having to make any modifications to the seal or pump. o

ans. The seals will fit, but they will not have enough outside diameter clearance for proper operation. The stuffing box bore should be enlarged.

6. Seal faces have to be lubricated. o

ans. Not necessarily. Carbon graphite is a natural lubricant. Electric motors have used carbon/graphite brushes for years that do not use any external lubricating source.

7. Vibration analysis is a good technique for predicting seal failure. o

ans. Vibration analysis requires that you know the frequency of the piece of hardware you are analyzing. This is easy for bearings that are always

made out of the same material, always in the same basic medium and vary little in shape. Seals come in a variety of shapes and materials and run in all sorts of mediums. 8. Oil is a good barrier fluid to use between dual mechanical seals. o

ans. Actually it is one of the worse. It has too low a specific heat number and it is not a very good conductor of heat compared to other liquids.

9. Teflon is a universal elastomer. It makes sense to use it in mechanical seals. o

ans. Teflon is not an elastomer because it does not have a memory. To use it in a mechanical seal you must spring load it to the shaft and that is never a good idea because you will end up with expensive shaft damage (fretting). O.E.M. suppliers use Teflon because they are not sure where the pump is going to be used.

10. Shrinking a carbon seal face into a metal holder is an acceptable manufacturing technique. o

ans. It really is a bad one. The out of roundness tolerance of the metal holder will clash with the out of roundness tolerance of the carbon, causing high loading at several points on the carbon outside diameter. The carbon should be pressed into the metal holder allowing it to shear and conform to the metal out of roundness.

11. It is a good engineering practice to glue the O-rings in a split mechanical seal design. o

ans. The glue will create a hard spot that will give you a leakage problem.

12. You should connect the flush connection to the top of gland. o

ans. It should be connected to the bottom of the gland or stuffing box. This will allow the flushing fluid to fill the box prior to spilling over the end restriction in the stuffing box. American prints show the top half of the drawing, that is why this error is so frequently made.

13. The elastomer Viton is acceptable in water. o

ans. It is a worse choice. The proper material for water is ethylene propylene. Some specific grades of Viton can be used in cold water , but none of them are good for hot water. Viton is cured in sulfur and what ever attacks the cure attacks the compound. Needless to say sulfur and water are not a good combination.

14. Split seals leak o

ans. It all depends upon your definition of leakage. If you are talking "fugitive emissions" that are measured at parts per million you can build a case for leakage, but if you mean "no visible leakage" then split seals should be as leak free as any other mechanical seal manufactured from the same materials.

15. You should put a lubricant on seal faces when you install them. o

ans. It's not a good idea to put anything on the lapped faces. The trick is to keep the lapped faces together.

16. No one can predict seal life. o

ans. That is a fact, but we know how long seals should last. They should run leak free until the sacrificial carbon wears down. (90% of mechanical seals fail long before that happens).

17. In most seal applications the carbon is running on a hard face. o

ans. The graphite comes out of the carbon /graphite face and deposits on the hard face. You can easily see the black mark made by the graphite. The seal face you are actually running is carbon on graphite. The hard face is just some place to put the graphite. This is the reason the seal faces can run dry.

18. It is good engineering practice to put a stationary seal ( the type where the springs do not rotate with the shaft) on a cartridge. o

ans. Tightening the cartridge sleeve set screws will pull the cartridge sleeve to one side, causing the rotating face to no longer be perpendicular or square to the rotating shaft. This squareness to the shaft is essential to the performance of any stationary seal design.

19. If you are installing a mechanical seal in a vertical centrifugal pump, you use the same procedure as installing a seal in the horizontal version. o

ans. Vertical pumps trap air in the stuffing box. You will have to install some type of vent above the seal faces and dynamic elastomer to avoid "dry running" in these locations.

20. PV factors are a legitimate way of predicting seal performance. o

ans. Carbon/graphite seal faces are sensitive to pressure(P), but not to velocity(V) so PV has limited value.

21. The three hundred series of stainless steel is a good choice for seal metal components. o

ans. That is true for most of the metal components, but not for springs or metal bellows. The three hundred series is sensitive to chloride corrosion problems in these locations.

22. You must not use ceramic as a seal hard face because it will crack with a rapid temperature change. o

ans. Some ceramics have this problem, and you should not use them in your designs. The ceramic called "silicone carbide" is a good choice as a hard face and does not have the "cold shocking" problem.

23. The metal bellows seal should be your first choice for a hot application. o

ans. the metal bellows seal is always a good choice in hot fluid to eliminate the temperature sensitive elastomer or O-ring, but it is not effective in hot petroleum applications because of "coking problems. In these applications you have to cool the stuffing box area to prevent the oil from forming coke solids on the seal moving parts and faces.

24. The mechanical seal should be positioned against a shaft or sleeve shoulder to insure the correct face load and then set screwed to the shaft or sleeve.. o

ans. To compensate for shaft axial growth or open impeller adjustment, the seal must be positioned on an adjustable cartridge sleeve.

Troubleshooting mechanical seals at equipment disassembly All seals fail for the same reasons: • •

The faces open up and allow dirt or solids to penetrate. One of the seal components has been damaged by either the product, heat, or a cleaner used to flush the system.

After the failure has occurred you will frequently get a chance to analyze the failed components. You are going to be looking for several things: • • • • • • •

Evidence of corrosion. Wear patterns on those parts that should be rubbing. Evidence of rubbing or wear on those components that should not be in contact. Discoloration of any of the seal components, especially the metal parts. Parts that are missing. Springs, set screws and drive lugs as an example. Loose hardware. Either a seal component or a foreign object. Product attaching to a rotating component. Carefully inspect the impeller and rotating part of the seal.

In the following paragraphs we will be inspecting the individual components and looking for evidence of the above. THE CARBON FACE Chipping on the O.D. of the carbon. Indicating vibration. • • • • • • •

This can be caused by harmonic vibration, or if the rotating equipment hits a critical speed. Slipstick can occur if you are pumping a fluid with poor lubricating qualities. Mishandling is a common problem. Look for evidence of drive lug wear to eliminate this as a possibility. Vaporization of the liquid causing the faces to rapidly open and then close as the leaking fluid cools the faces. A discharge recirculation line is aimed at the carbon seal face. The pump is cavitating. Remember there are five types of cavitation. Water hammer is a another possibility.

Pits in the carbon face. This problem is usually associated with poor grades of carbon/ graphite. •

•

Exploded carbon. Air trapped in the pores of the carbon expands and expels pieces of the carbon when the seal faces get hot. Prior to ejection polished patches will be visible, usually with small cracks visible in the center. If the product solidifies between the faces it will tear out pieces of the carbon at start up. This is a common occurrence with ammonia compressor seals because

•

petroleum oil is mixed with the ammonia and it can coke at the elevated temperature. Most petroleum products will "coke" because of the higher face temperature, and pull out small pieces of the carbon as the faces rotate. You will see evidence of these small pits if you inspect the carbon face under a magnifying glass.

Chips at the I.D. of the carbon •

•

•

Solids, or a foreign object of some type from outside of the pump are getting under the gland and are being thrown into the seal faces. This can occur if the seal leaked at some time and the product solidified on the outboard side of the seal. It can also occur if liquid, containing solids, is used in the quench connection of an A.P.I. type gland. If the seal was installed outside of the stuffing box, as is the case with non metallic seals, solid particles in the fluid can be centrifuged into the rotating carbon face. If the stationary face is manufactured from carbon it can be chipped if it comes into contact with the rotating shaft. This is a common problem at pump start up, or if the pump is operating off of its B.E.P.

Phonograph finish on the carbon face. •

A solid product was blown across the seal face. This happens in boiler feed water applications.

Chemical attack of the carbon. •

•

•

You are using the wrong carbon. Something in the product or the flush is attacking the carbon filler. Switch to an unfilled carbon such as Pure grade 658 RC or C.T.I. grade CNFJ. You are trying to seal an oxidizing agent. Oxidizers attack all forms of carbon including the unfilled type. The carbon combines with the oxygen to form either carbon monoxide or carbon dioxide. Some forms of de ionized water will pit and corrode carbon faces

Cracked or damaged carbon face. •

• • • • •

The product is solidifying between the faces. Carbons are strong in compression but weak in tension or shear. This problem is common with intermittent pumps each time they start up. Excessive vibration can bang the carbon against a metal drive lug. A cryogenic fluid is freezing a lubricant that was put on the face. The elastomer is swelling up under a carbon or hard face. The shaft is hitting the stationary face or the rotating seal face is hitting a stationary object. Mishandling.

• •

Poor packaging. The lapped seal faces should be able to survive a 39" (one meter) drop. Ice is forming on the outboard side of the seal and preventing the seal from moving to compensate for face wear.

A coating is forming on the carbon face: • • • • •

A change in temperature. Many products solidify at temperature extremes. The product is taking a pressure drop across the seal faces and solidifying. Selective leaching is picking up an element from the system and depositing it on the seal face. The stuffing box is running under a vacuum because the impeller was adjusted backwards and the impeller "pump out vanes" are causing the vacuum. The system protective oxide is depositing at the faces. In hot water systems we experience this problem with magnetite (Fe3O4) until the system stabilizes.

Coking • •

This is a problem with all oils, and petroleum products in particular. Coking is caused by the combination of high temperature and time. Contrary to popular belief the presence of air or oxygen is not necessary.

Shiny spots, cracks and raised portions of carbon. • •

The carbon is not dense enough, causing the expanding gases trapped beneath the surface of the carbon to explode through the face. Product is solidifying between the faces and pulling out pieces of the carbon as the seal revolves.

Excessive carbon wear in a short period of time. Evidence of excessive heat is usually present. •

• •

• •

•

Heat checking of the hard face. It shows up as a cracking of the hard face. This is a problem with coated or plated hard faces. Cobalt base tungsten carbide is a typical example. The shaft is moving in an axial direction because of thrust. This can cause an over compression and heating of the seal faces The impeller is being adjusted towards the back plate. This is problem with seals installed in Duriron pumps or any other pump that adjusts the open impeller against the back plate. Any installation problem: The inner face of a "back to back" double seal application is not positively locked in position. A snap ring must be installed to prevent the inboard stationary face from moving towards the rotating face when the high pressure barrier fluid pressure is lost or overcome by system pressure. The seal was installed at the wrong dimension.

• • • • • • •

A cartridge double seal was installed by pushing on the gland. Friction, between the shaft and the sleeve O-Ring is compressing the inner seal. A vertical pump was not vented. Solids have penetrated between the faces. The faces are not flat. The movable face is sluggish. The product is vaporizing between the faces because of either high temperature or low stuffing box pressure . Non lubricants will cause rapid face wear. A non lubricant is any fluid with a film thickness less than one micron at its load and operating temperature..

The carbon has a concave or convex wear pattern • • • •

High pressure distortion. The stationary face is not perpendicular to the shaft. Some companies lap a concave pattern as standard. Check with your manufacturer. The shaft is bending because the pump is running off of its B.E.P.

The carbon is not flat. • • • • • • • •

Mishandling. Poor packaging. The hard face has been installed backwards and you are running on a non lapped surface. The seal was shipped out of flat. The metal/ carbon composite has not been stress relieved and it is distorting the carbon. When the carbon was lapped the lapping plate was too hot and as a result, not flat. The carbon was lapped at room temperature and the seal is running at cryogenic temperatures. Solids are imbedded in the carbon. The faces have opened. o The seal was set screwed to a hard shaft. o The elastomer (rubber part) is spring loaded to the shaft causing the faces to open as the shaft moves due to end play, vibration or carbon wear. The shaft/ sleeve is over sized causing an excessive interference between the elastomer and the shaft/ sleeve. o The sleeve finish is too rough. o The product has changed from a liquid to a solid. o Dirt or solids are interfering with the seal movement. o Some one put the wrong compression on the faces. o Shaft fretting is hanging up the face. o The face has been distorted for some reason allowing solid particles to enter. o The sliding elastomer has swollen up causing too much interference on the shaft/ sleeve.

Poor centering is causing the rotating face to run off the stationary face. Keep in mind the gland bolts are not always concentric with the shaft. o The single spring was wound in the wrong direction. o An out of balance rotating assembly or bent shaft is causing the rotating face to "run off" of the stationary face. o

THE HARD FACE. Chemical attack. •

Some ceramics and silicone carbides are attacked by caustic. Check to see if your seal face contains silica. As an example: both reaction bonded silicone carbide and 85% ceramic have this high silica content.

Cracked or broken. • • • • • • •

The product is solidifying between the faces. Most hard faces have poor tensile or shear strength. Excessive vibration will cause cracking at the drive lug location.. A cryogenic fluid is freezing a lubricant that was put on the face. The elastomer is swelling up under an outside seal face. This problem can also occur if the seal design allows a spring to contact the I.D. of the hard face. The shaft is hitting the stationary face or the rotating seal face is hitting a stationary object. Mishandling. Poor packaging.

Heat check (a common problem with coated or plated faces) •

Caused by a high heat differential across the face. Most hard coating have only one third the expansion rate of the stainless steel base material.

Hard coating coming off of the face. •

•

The base material not compatible with the sealed product. These coating are very porous so if the product attacks the base material the coating will come off in sheets. The plating process was not applied correctly.

Analysis of the wear track on the hard face. Deep grooves&emdash;excessive wear. Solids imbedded in the carbon are causing the problem. The solids were trapped between the faces when the seal faces opened. •

The seal was set screwed to a hard shaft.

• • • • • • • • •

• •

The elastomer is spring loaded to the shaft preventing it from flexing as the shaft vibrates.. The shaft/ sleeve is over sized causing the dynamic elastomer or bellows vibration damper to hang up.. The shaft/ sleeve finish is too rough The product has solidified in the seal components. Dirt or solids are interfering with seal movement. Not enough spring compression on the faces. Fretting of the shaft/ sleeve is hanging up the face. The face has been distorted by either excessive temperature or pressure. The sliding elastomer has swollen up due to chemical attack of the product or a cleaner that was flushed through the lines. The wrong choice of rubber lubricant, at installation, can also cause the problem Poor centering is causing the rotating face to run off of the stationary face.. The single spring was wound in the wrong direction.

The wear track is wider than the carbon. • • • • • •

Worn bearings. Bent shaft. Unbalanced impeller. Sleeve not concentric with the shaft. Seal not concentric with the sleeve. In a stationary seal, the stationary carbon is often not centered to the shaft, causing a wiping action.

The wear track is narrower than the carbon. • • • •

The soft face (carbon) was distorted by pressure. The hard face was over tightened against an uneven surface. The hard face clamping forces are not "equal and opposite". The face never was flat, or it was damaged during shipment.

Non Concentric pattern. The wear track is not in the center of the hard face. • • • • • •

The shaft is bending because the pump is running off of its best efficiency point. Poor bearing fit. Pipe strain. Temperature growth is distorting the stuffing box. The stationary face is not centered to the shaft. Misalignment between the pump[ and its driver.

Uneven face wear. The hard face is distorted: • •

High pressure. Excessive temperature.

• • • • •

Over tightening of the stationary face against the stuffing box. The clamping forces are not equal and opposite. The hard face is not wide enough. You are using a two bolt gland and the gland is too thin causing it to distort. You are using a pump seal in a motion seal application.

The product is sticking to the seal face. The product is changing state and becoming a solid. Most products solidify for the following reasons: • • • •

A change in temperature. A change in pressure. Dilatants will solidify with agitation. As an example: cream becomes butter. Some products solidify when two or more chemicals are mixed together.

The hard face is not flat. • • • •

Mishandling. Poor packaging. The hard face has been installed backwards and you are running on a non lapped surface. It was shipped out of flat.

THE ELASTOMER. Compression set. The O-ring has changed shape. •

… High heat is almost always the cause unless you are dealing with Kalrez, Chemraz, or a similar material where a certain amount of compression set is normal.

Shrinking, hardening or cracking. • • • • •

High heat. The shelf life was exceeded. This is a big problem with "Buna N" that has a shelf life of only twelve months. Cryogenics will freeze just about any elastomer. Chemical attack normally causes swelling, but in rare cases can harden an elastomer. Oxidizing liquids can attack the carbon that is used to color most elastomers black.

Torn nibbled, or extruded. • • •

Mishandling. Sliding over a rough surface. Forced out of the O-Ring groove by high pressure.

• •

The liquid has penetrated the elastomer, vaporizing inside and blowing out pieces. This is a problem with Ethylene Oxide. Halogenated fluids can penetrate the Teflon coating on an elastomer and cause the base material to swell up, splitting the Teflon jacket.

Swelling, changing color, weight or size. Almost always caused by: • • • • •

Chemical attack. Be careful of the lubricant used to install the elastomer. Solvents or cleaners used in the system may not be compatible with the elastomer. Some compounds are sensitive to steam. Most Vitons are a good example of this problem. The elastomer is not compatible with something in the fluid you are sealing.

Torn rubber bellows. • • • •

The bellows did not vulcanize to the shaft because you used the wrong lubricant. The shelf life was exceeded. The seal faces stuck together and the shaft spun inside the bellows. The pump discharge recirculation line was aimed at the rubber bellows. Solids entrained in the high velocity liquid are abrading the bellows.

THE METAL CASE OR BODY OF THE SEAL. Corrosion. • •

• •

•

•

General or overall. This is the easiest to see and predict. The metal has a "sponge like" appearance. It always increases with temperature. Concentrated cell or crevice corrosion. Caused by a difference in concentration of ions, or oxygen in stagnant areas causing an electric current to flow. Common around gaskets, set screws, threads, and small crevices. Pitting corrosion. Found in other than stagnant areas. Extremely localized. Chlorides are a common cause. Can be recognized by pits and holes in the metal. Stress corrosion cracking. Threshold values are not known. A combination of chloride, tensile stress, and heat are necessary. Chloride stress corrosion is a serious problem with the 300 series of stainless steels used in industry. This is the reason you should never use stainless steel springs or stainless metal bellows in mechanical seals. Inter granular corrosion. Forms at the grain boundaries. Occurs in stainless steel at 800-1600 F. (412-825 C.), unless it has been stress relieved. A common problem with welded pieces. Stabilizers such as columbium are added to the stainless steel to prevent this. Rapid cooling of the welds, the use of 316L and stress relieving after the welding are the common solutions. Galvanic corrosion. Occurs with dissimilar materials in contact with and connected by an electrical current. Common in brine, caustic, and salt water applications.

•

• •

Erosion / Corrosion. An accelerated attack caused by a combination of corrosion and mechanical wear. Vaporization, liquid turbulence, vane passing syndrome, and suction recirculation are special cases often called cavitation. Solids in the liquid and high velocity increase the problem. Selective leaching. Involves the removal of one or more elements from an alloy. Common with demineralized or de ionized water applications. Micro organisms, that will attack the carbon in active stainless steel.

Rubbing--All around the metal body. • • • • • •

A gasket or fitting is protruding into the stuffing box and rubbing against the seal. The pump discharge recirculation line is aimed at the seal body. The shaft is bending due to the pump operating off of its best efficiency point. Pipe strain. Misalignment between the pump and its driver. A bolted on stuffing box has slipped.

Partial rubbing -- On the metal body. • • • • • • •

Bent shaft. An unbalanced impeller or rotating assembly. Excessively worn or damaged by corrosion or solids in the product. The product has attached its self to the impeller. The impeller never was balanced. The impeller was trimmed, and not re balanced. The seal is not concentric with the shaft, and is hitting the stuffing box I.D..

Discoloration. Caused by high heat. Stainless steel changes color at various temperatures. FAHRENHEIT

COLOR OF THE METAL

CENTIGRADE

700 - 800

Straw Yellow

370 - 425

900 - 1000

Brown

480 - 540

1100 - 1200

Blue

600 -650

> 1200

Black

> 650

NOTE: To tell the difference between discoloration caused high heat and product attaching to the metal part, try to erase the color with a common pencil eraser. Discoloration will not erase off. Product sticking to the metal surfaces. •

Heat is the main cause.

• • • • • • • • • •

The product pressure has dropped. Air or oxygen is getting into the system. Valves above the water line. Through the stuffing box. The product was not deaerated. The pump suction is not completely submerged. The bypass return is too close to the pump suction. The liquid is vortexing in the suction line. A non O-Ring elastomer is being used in the seal allowing air to enter the stuffing box when you are sealing a vacuum application. The system protective oxide coating is depositing on the sliding metal components.

The following applications cause a vacuum to be present in the pump stuffing box. • • • • • •

Pumps that lift liquid. Heater drain pumps. Pumping from an evaporator. Pumping from the hot well of a condenser. Pumps that prime other pumps. The open impeller was adjusted in the wrong direction and the impeller pump out vanes are causing the vacuum.

The Teflon coating is coming off some of the metal parts. •

Coatings are very porous. They do not provide corrosion resistance. The base material is being attacked by the product.

DRIVE LUGS, PINS, SLOTS, etc. Broken. • • • •

Chemical attack. Excessive side load. The seal faces are glued together because the product has solidified. A cryogenic fluid is sticking the faces together.

Wear on one side of the drive lug or slot. • • •

Vibration. Slipstick. The stationary is not perpendicular to the shaft.

The drive pins are falling out of the holder. •

Corrosion.

• • •

Improper fit. Bad part. Excessive vibration.

THE SPRINGS. Broken or cracked. •

•

The stationary face is not perpendicular to the shaft causing excessive spring flexing in the metal "plastic range". The spring material has "work hardened" and fatigued. Chloride stress corrosion problems with 300 series stainless steel.

Corroded. •

Stressed material corrodes much faster than unstressed material. The springs are always under severe stress.

Clogged. • • •

Be sure to distinguish between "cause and effect". If the springs are located outside the liquid, it happened after the failure. If the product solidifies or crystallizes it can clog springs exposed to the pumped fluid. Dirt or solids in the fluid can clog exposed springs.

Twisted. •

Almost always an assembly problem. The lugs were not engaged in the slots. This is a problem with many seal designs. Check to see if your seals can come apart easily or if the drive lugs can change position when the seal is not compressed.

The drive lugs or slots are worn on both sides. • • •

Excessive vibration. The single spring, rubber bellows seal, was not vulcanized to the shaft. The stationary is not perpendicular to the shaft, causing excessive spring movement.

Broken Metal Bellows. •

• •

Fatigue caused by over flexing in the plastic range of the metal o Harmonic vibration. o Slipstick. The discharge recirculation line is aimed at the thin bellows plates. Excessive wear from solids in the stuffing box.

• •

Faces sticking together as the product solidifies. Chloride stress corrosion with 300 series stainless steel.

Because these seals do not have a dynamic elastomer to provide vibration damping some other means must be provided or vibration will always be a problem. THE SLEEVE, OR SHAFT. Grooves or pits at the seal dynamic elastomer location. • • • • •

Fretting. Concentrated cell corrosion. The rubber bellows did not vulcanize to the shaft/ sleeve. The set screws slipped on a hardened shaft or were not tightened properly. The seal faces stuck together causing the shaft to rotate inside the static elastomer. Salt water applications are particularly troublesome when a static elastomer or clamp is attached to the shaft. Pitting caused by the chlorides and the low PH of salt water are the main problems.

Rubbing at the wear ring location. • • • • • • • • • •

The pump is running off of its best efficiency point. The shaft is bending. Bad bearings. Excessive temperature. Sleeve is not concentric with the shaft, or the seal with the sleeve. Bent shaft. Unbalanced impeller or rotating assembly. Pipe strain. Misalignment between the pump and its driver High temperature applications require a "center line: pump design.

Corrosion. See above description under metal corrosion .THE SET SCREWS. • • • •

Stripped from over tightening. Corroded. Check to see if you are using hardened set screws. This type is normally supplied with most cartridge seals and can corrode easily. Rounded Allen Head. Alan wrenches wear rapidly. They are an expendable tool. Loose. o Sleeve too hard. They are not biting in. o Sleeve too soft. They are vibrating loose.

THE GLAND.

Rubbing at the I.D. • • • • • • • • • • • • •

Partial rubbing. The gland has slipped. Improper installation. It was not centered to the shaft. The shaft is bending. Pipe strain. Rubbing all around. The shaft is not concentric with the sleeve. The seal is not concentric with the sleeve. Bad bearings. Bent shaft. Unbalanced impeller or rotating assembly. Solids attached to the shaft, or caught between the shaft, and the gland. Cavitation.

Corrosion. •

If there is evidence of rubbing the corrosion will be accelerated.

Passages clogged or not connected properly. •

A.P.I Gland. o Hooked up wrong. o Flushing connection clogged. o Quench connection clogged.

BUSHINGS Rubbing at the I. D. • • • • • • • • • • • • • •

Partial rubbing. The A.P.I. gland has slipped. Improper installation. It was not centered to the shaft. The shaft is bending. The gland bolt holes are often not concentric with the shaft/ sleeve. Misalignment between the pump and its driver. Excessive pipe strain. Rubbing all around. The shaft is not concentric with the sleeve. The seal is not concentric with the sleeve. Bad bearings. Bent shaft. Unbalanced impeller. Cavitation

Erosion. •

Dirt and solids are present in the discharge or suction recirculating fluid.

Troubleshooting mechanical seals at the pump site

Leakage can occur at any time throughout the life of the mechanical seal. To troubleshoot seals effectively it is helpful to know just when the leakage starts. This is the advantage of being able to troubleshoot a running pump or one that is still hooked up to its piping. By noting the type of leakage and when the leakage occurs we can do a more thorough job of analyzing any seal failure. In addition to leakage we will be looking for other symptoms that are visible to the trained troubleshooter. We will start with the different types of leakage. Please look at the following diagram.

The leakage occurs while the pump is both running and stopped. The leakage can be detectable visually, by odor, or by instrumentation. A strobe light can sometimes be used to determine its location. As you can see in the above diagram there are several leak paths possible. You must determine which ones you have. The seal can leak : At the lapped faces. Since they are a wearable surface the leak will probably get either better or worse. It should never remain constant. The leak started because: • • • •

The outside springs in a dual cartridge seal were painted during routine maintenance. The spring load has been reduced because of thermal growth, axial thrust, or impeller adjustment. The seal was set screwed to a hardened shaft and has vibrated loose. One or both of the seal faces is not flat. Solid tungsten carbide and silicone carbide faces are often lapped flat on only one side. Check to see if the face has been installed backwards.

• • • • • •

The dynamic elastomer has swollen up and seized the spring loaded face, preventing it from remaining in contact with the stationary face. The product prevented the lapped seal faces from remaining in contact. Dirt has gotten into the sliding components. The product has crystallized. The product solidified or became very viscous. The product is vaporizing across the seal faces expanding and blowing them open.

At the static and dynamic elastomer locations. •

This type of leak tends to remain constant and will often stop when the small opening clogs up with solids. The leak can be caused by a damaged elastomer or damage on the surface where the elastomer seals. In some instances the elastomer is not seated properly. It is twisted because of either poor installation, excessive shaft movement, or high pressure extrusion.

At the gland gasket. •

This is the easiest leak to detect because it is very visible and does not change with shaft rotation.

Between the shaft sleeve and the shaft. •

This is a common problem with double ended pumps, where the sleeve is used to position the impeller and there is no method of sealing the sleeve against the impeller.

Between the seal face and its metal holder. •

The leakage frequently increases, as the product temperature increases, because the metal face holder has an expansion rate three times that of the carbon or hard faces.

Through fretting damage • •

The damage is caused by spring loaded dynamic O-rings, Teflon wedges, chevrons, U- cups etc. You can't miss the frett marks. They will be located on the pump shaft, pump sleeve, or inner sleeve of the mechanical seal.

The seal leaks only when the pump is running. •

The stationary face has been over tightened against the stuffing box face causing it to go out of flat. Statically the carbon will readjust to the distorted hard face.

•

•

• • •

• • •

• • • • • • • •

•

The clamping is not equal and opposite across the static seal face. Look for different width gaskets at the front and rear of the static face. Again, the carbon will readjust when the shaft is not turning. Between face and holder. The holder heats up and expands faster than the pressed in face. The leak will begin when the metal holder comes up to temperature. Remember that metal expands three times faster than a seal face. Cryogenic (cold) service will harden the elastomer. Be sure to check the lower temperature limit of the elastomer that was selected. Misalignment between the pump and the driver. The shaft is bending and not allowing the seal to move freely. This occurs if the pump is operating off of its best efficiency point and the shaft L3/D4 is not small enough to resist the bending. The product is vaporizing across the seal faces. Cavitation, slip stick, harmonic, or some other type of vibration is bouncing the faces open, check the lugs or drive pins for sign of excessive wear. The seal was installed without enough compression or the impeller was adjusted after the seal was installed and thermal expansion of the shaft is opening the faces. A discharge recirculation line is aimed at the seal faces or some other critical point and the faces are being forced open. A non- concentric seal, bad sleeve installation, or an out of balance rotating assembly, is causing the rotating portion of the seal to run off the stationary face. A bent shaft can cause the rotating portion of the seal to run off the stationary face. The rotating portion of the seal is hitting a stationary object. Look for: A protruding gasket or fitting. A foreign object that has worked its way into the stuffing box area. A stationary portion of the rotating equipment, such as a close fitting bushing. At elevated temperature the product thins out (the viscosity decreases) and is leaking through an elastomer. It will not leak at the cooler temperature when the product viscosity is higher. High temperature is causing the lapped seal face to go out of flat.

The seal leaks only when the pump is not running. • •

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The seal is also leaking while running, but the leak is vaporizing and not visible. Hold a piece of white paper over the seal area and see if the paper becomes damp. A meniscus caused by centrifugal force and liquid surface tension had formed at the inside diameter of the seal faces. This prevented a leaking seal from dripping while the shaft was turning. You are using high temperature grade Kalrez. It is too hard at ambient temperature and will soften at operating temperature. The pump is running under vacuum and while it is running air is being pulled into the system. The fluid leaks out when the shaft is static. This can occur if an open impeller that was designed to be adjusted against the volute has accidentally been adjusted backwards against the back plate. The impeller "pump out vanes" can

then pull a vacuum in the stuffing box. This is a common problem if you use a lot of Duriron pumps and then bring in a few of another brand. The leak occurs only at start up and then stops after a short time. • • •

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Face distortion. Caused by a high pressure surge that was created when the pump was started with the discharge valve shut. The shaft is bending, and interfering with the seal movement. Occurs because the pump is running with the discharge valve throttled or shut. Operators shut the discharge valve at start up to save electricity and prevent cavitation. The same problem can occur if the pump is started with the discharge valve wide open and because of the lack of discharge resistance the pump will run to the right hand side of its curve. In some cases you could also burn out the electric motor. The product has changed state, and becomes a liquid again when the pump comes up to operating temperature. The product had : Crystallized Solidified Became viscous Excessive axial shaft movement at start up. This is a common problem with sleeve bearing equipment.

The seal leaks intermittently or after the pump has run for a fixed period of time Look for reoccurring events that initiate the leakage. They can include: • • • • • •

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Flushing the lines at the end of a batch or season. Alternating pumps in a multiple pump arrangement. An additive is being put into the product. Batch operations are beginning or ending. The cooling water is passing through temperature cycles. The outside ambient temperature has changed dramatically. I ran into a situation where a supplier was oiling the bed of his truck to prevent solids from sticking in the winter and this oil attacked the elastomer in the seal. Hard water is being used as a flush and it is gradually restricting the flush lines or cooling jacket. A filter or strainer is clogged in a flush line. The flushing water pressure drops at certain times of the day because of demand. The boiler or cooling tower is being blown down. There is a control valve in the pump discharge that is causing the pump to occasionally operate too far off of the B.E.P. The stuffing box is cycling between a positive and negative pressure. Vortexing can occur if the pump suction falls too low. This also occurs in mixers and agitators..

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You are quenching a high temperature application with water. As the quench water vaporizes it leaves dissolved solids outboard of the seal restricting axial movement as the seal faces wear. The pump is cavitating on a regular or intermittent basis. Here are a few possibilities: The suction level falls too low The tank vent freezes. The velocity is too high on the suction side of the pump. A suction strainer is plugged up. A stuck or broken check valve in the pump suction piping. A temporary loss of discharge head. A booster pump has shut off. A suction eccentric reducer was installed up side down allowing slugs of air into the suction of the pump. The fluid is vortexing in the supply tank. The level is too low for the pump capacity. The pump is lifting liquid and the foot valve is sticking.. The impeller is too close to the cutwater. Air is entering the system through the pump packing. A lower "specific speed" impeller as been substituted. The pump was specified with too low a "suction specific speed" number. The pump is running at a higher speed or a larger impeller was installed after the system heads were calculated. In some parallel pump installations, a stronger pump can throttle the weaker one causing shaft deflection. The wrong lubricant was used on the dynamic elastomer, causing it to swell up. Reaction bonded type, Silicone Carbide can crack if the lines were flushed with Caustic solution.

TYPES OF LEAKAGE. The leak rate is changing, It gets better or worse. • • • • • • •

This type of leak is usually associated with seal face leakage because the seal face is a wearable surface. The carbon seal face is not flat. The seal face was damaged at the time of assembly. Dirt or solids are imbedded into one of the faces Coke (over heated oil) or some other solid has formed on the seal faces causing them to separate. The rotating face is hung up on the shaft. Outside seal springs have been painted during routine maintenance.

The faces spit liquid.

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The product is vaporizing at the faces - check the fluid vapor point. When using balanced seals the stuffing box pressure must be at least one atmosphere higher than the product vapor point. Unbalanced seals require a much higher differential pressure. The rotating face is running off of the stationary face. The stationary was not centered to the shaft - a common problem. The seal is not concentric with the shaft. The rotating assembly is out of balance. The shaft is bent.

Fire hose type leakage. The leak is following shaft rotation. • • •

Product has solidified on the seal face and a piece has broken off. This is usually initiated by a high temperature between the faces. The rotating face is cracked. The hard surfacing, or coating, is lifting off of the rotating face.

Intermittent leakage. • • • • • •

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Temperature changes or pressure surges are altering the face flatness within the elastic range of the material. The stuffing box is alternating between vacuum and pressure The movable face is sluggish and not able to follow run out. The product is viscous. The product has started to solidify. The shaft/ sleeve is too large in diameter restricting movement of the seal. Spring loaded, dynamic elastomers such as Teflon wedges, U- cups, Chevrons and spring loaded O-ring designs are very sensitive to this problem Dirt or solids are clogging the seal and preventing it from following shaft run out. In a non O-Ring version, the spring load is too high causing the elastomer or Teflon to stick to the shaft. The product is occasionally vaporizing between the faces. There is a leak between the face and the holder that becomes visible only when the unit comes up to operating temperature. A bending, or bent shaft is causing the seal outside diameter to contact the inside diameter of the stuffing box, or some other stationary object. The pump is running with too high or too low a head. Check the pump curve against actual operating conditions. The application is cycling between ambient and cryogenic temperatures causing the elastomer to harden on the cold cycle and the faces to go out of flat.

The seal area is damp. There is no visible leakage. •

There is a leaking flange or fitting above the seal that is dripping close to the seal location.

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The product is vaporizing. Hold a clean piece of white paper over the running seal, and check for leakage. The paper will become damp. Any condition that could cause intermittent leakage will cause this problem.

Constant dripping. It gets neither better nor worse. This cannot be a damaged seal face leak because seal faces are a wearable surface and the leak rate would have to change. • • • • • • • • • •

The elastomer is cut or nicked. The shaft/sleeve is damaged at the elastomer location. There is damage in the O-Ring groove. Maybe the O-ring was removed with a sharp metal instrument and this has caused a scratch in the O-ring groove. There is a leak path between the carbon and the holder. Leaking at the cartridge sleeve location. Leaking between the sleeve and the shaft. Leaking between the gland and the stuffing box. Leaking between the stationary face and the seal gland. Seal faces are stuck open. The elastomer has swollen up due to chemical attack by either the product, the flush, what ever is being used to clean the lines, or by the lubricant that was put on the elastomer to help the installation. This attack usually takes place within one week of exposure to the non compatible lubricating fluid.

THE STUFFING BOX AREA IS GETTING HOT. Heat is being generated at the seal faces. Unbalanced seals generate more heat than balanced seals. • • • • • • • • • • •

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The carbon is being insulated by an elastomer and cannot dissipate the heat. High friction face materials. Two hard faces usually generate more heat than carbon vs. a hard face. The faces are running dry. The stuffing box has not been vented. This is especially important in vertical applications. You do not have a barrier fluid between the seals in a dual seal application. You have lost an environmental control. Flushing. Quenching. The cooling jacket is clogged or not functioning for some reason. The discharge or suction recirculation line is clogged. The barrier fluid has stopped circulating in a dual seal application or you are using oil as a barrier fluid. Oil has a low specific heat and poor conductivity, making it a poor choice as a heat transfer medium. If you must use oil as the barrier fluid you may have to forsake convection and go to a forced circulation system or a pumping ring. An A.P.I. type gland has been piped incorrectly Poor conductivity of the hard face. Silicone carbide is better than 99.5 ceramic.

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There is too much spring load on the seal faces: A wrong installation measurement. The impeller was adjusted after the seal was installed. Any pump impeller that adjusts against the back plate has this problem. Durco pumps are a good example. Excessive axial movement of the shaft. Thermal expansion

A seal component is rubbing the inside diameter of the stuffing box, or against a product that has attached its self to the inside of the stuffing box. • • •

The seal is not concentric with the shaft. The shaft is out of balance. The shaft is bent.

The sleeve, shaft or rotating seal is hitting a stationary component. • •

A protruding gasket or fitting. A bushing in the bottom of the stuffing box.

A foreign object is loose in the stuffing box. A suction recirculation line was used to lower stuffing box pressure. The high velocity recirculation is heating up the return line. NOISE IN THE STUFFING BOX. •

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The product is cavitating in the pump. There are five types of cavitation: o Vaporization. o Internal recirculation o The Vane Passing Syndrome o Turbulence o Air ingestion A component is rubbing. The bearings are bad. The seal has come loose from the shaft. A foreign object has entered into the stuffing box. The sleeve is hitting an A.P.I. disaster bushing. The seal faces are running dry. They will make a whistling noise. You have hit a critical speed. Coupling misalignment. The noise is coming from the motor or some near by equipment. "Slip stick" at the seal faces.

AUXILIARY EQUIPMENT FAILURE. The convection tank

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It is running backwards. The seal faces are off center causing a pumping action across the faces. The inlet and outlet ports are not drilled properly. A cartridge double seal has not been centered properly The pressure or level in the tank changes. One of the seals is leaking. The pressure or level change should tell you which one. Temperature change. No air pocket in the tank. Not convecting. It was installed incorrectly. The minimum and maximum dimensions were ignored.

Flow meter not indicating. • • • •

Meter broke. Line clogged. The flow is not high enough. The gage graduations are too large.

No flow through the quench and drain connections. • • •

You are piped to the wrong connection. Most glands that have been drilled for a quench connection, have a flush connection also. Valve not open Line clogged

Loss of jacket cooling. The incoming and out going lines are at the same temperature. • • •

A layer of calcium has built up on the inside of the cooling jacket. A discharge recirculation line is connected to the stuffing box (it may be hidden inside the insulation). Some one has shut off the cooling water or steam.

VIBRATION. • • • • • • • • • • •

Cavitation. Remember there are five types. The pump is operating off of its best efficiency point. Unbalanced impeller or rotating assembly. Look for wear or product is attached. Bent shaft. Bad bearings. Misalignment between the pump and driver. Pipe strain. Maybe you need a center line design pump Rotating component hitting a stationary component. The pump is running at a critical speed Harmonic vibration induced by nearby equipment. Loose hold down bolts.

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Pipe hangers are spaced improperly. The mass of the pump base is not five times the mass of the pump, and motor. The base is too narrow. Imaginary lines extended downward thirty degrees to either side of a vertical through the pump shaft must pass through the bottom of the foundation, not the sides. Seal "slip stick" that can occur when pumping non lubricants such as hot water and most solvents.

Troubleshooting rub marks in a centrifugal pump When a centrifugal pump is disassembled there are a couple of things visible to the trained trouble shooter. He can see either corrosion or evidence of rubbing, damage or wear. When ever a rotating piece of hardware hits a stationary piece it leaves a mark that is clearly visible and capable of being analyzed for cause. This type of rub mark should never be confused with the dull appearance we see on a piece of metal that has been rotating in an abrasive slurry. In strong corrosive applications the rub mark may not be visible. The contact will cause an increase in the metal temperature causing rapid chemical attack. This condition is easy to identify because the corrosion is localized at the rubbing location. Shaft fretting is another common rub mark that should not be confused with the rub marks we will be discussing in the following paragraphs. Fretting is visible between the dynamic elastomer in the mechanical seal and the sleeve or shaft that the elastomer is sealing against. You will also observe this type of damage immediately under the grease or lip seals that we find being used to seal most bearing applications. There are five possible rubbing combinations that can be seen: 1. All around the rotary and one spot on the stationary. 2. All around the stationary and one spot on the rotary. 3. All around both the rotary and stationary. 4. One spot on both the rotary and stationary. 5. One spot on the rotating component. You should look for the rub marks on those pieces that normally come in close contact. Common sense will dictate that the further the hardware is located from the bearings, the more likely the contact will occur. Here are some likely candidates for rubbing when the pump experiences shaft deflection, or any other type of radial displacement. Look for contact between : • • • • •

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The stationary and rotary parts of the wear rings that are installed in most closed impeller pump designs. The shaft&endash;sleeve and the mechanical seal stationary face inside diameter. The shaft&endash;sleeve and the bottom of the pump stuffing box, or stuffing box restrictive bushing. The shaft&endash;sleeve and the A.P.I. gland disaster bushing. The outside diameter of the mechanical seal rotating element and the inside diameter of the stuffing box. You will need a mirror and flashlight to see the stuffing box inside diameter. The impeller and the volute casing or the pump back plate. The outside diameter of the rotating seal, and a protruding gasket or fitting.

In the following paragraphs I will list the observations, explain the causes and where practical list some of the conditions that can initiate the problem with centrifugal pumps. If you would like to learn more about how to trouble shoot the rubbing marks we normally find in ball bearings, please refer to another paper in this series Observation - All around the rotary, one spot on the stationary. The shaft is being deflected from its true position or the hardware surrounding the rotating piece is being forced into the rotary unit. •

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The pump is operating off of its B.E.P. The stationary mark will be visible at either 240° or 60° from the discharge "cut-water" as measured in the direction of shaft rotation. o Some one has throttled the pump discharge valve. o The capacity has increased. o The discharge lines have a solids build up on the I.D. or there is a restriction in the discharge piping. o The tank is being filled from the bottom. The head is increasing as the tank fills. o The discharge by-pass line is not functioning. o You have the wrong size pump. o Two pumps are piped in parallel. The larger pump is shutting the discharge check valve of the smaller pump. o The pump speed has changed. o The system has been altered. Piping and fittings have been added or removed. o The pump was started with the discharge valve fully open or shut. o The viscosity of the liquid has changed. o The impeller has been trimmed. o The discharge piping or a fitting on the discharge has been damaged. o The motor is running at the wrong speed. This could be caused by a change in the specific gravity of the pumped fluid. o The suction head has changed & the discharge head changed to compensate. o An in-line filter is clogged. The shaft is pulley driven. The off-set driver is causing the deflection. Misalignment between the pump and the driver. o They never were aligned. o Thermal growth. o Vibration has loosened the hold down bolts. o The seal was changed and the pump was not realigned. o A universal joint has been installed between the pump and the driver. Pipe strain o Thermal growth - no expansion joints. o During the installation process the piping was forced to the pump suction&emdash; instead of piping from the suction to the pipe rack. o A center line design pump was not specified for elevated temperatures.

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A protruding piece of stationary hardware is contacting the rotating part. o A fitting is protruding into the stuffing box through the lantern ring connection. o A gasket on the gland face is extruding into the stuffing box. A recirculation line aimed at the seal will give the appearance of rubbing marks if there is a lot of abrasives in the re-circulating fluid. The mechanical seal gland has slipped and is now contacting the rotating shaft. A bad foot bearing on a mixer. The stationary seal face was not centered on the shaft and now the inside diameter of the seal face is rubbing on the shaft. A severe cocking of the seal face can cause the same problem.

Observation - All around the stationary, one spot on the rotary. •

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The shaft is bent. o It never was straight. o The shaft was damaged when it was dropped. o The shaft was overheated and warped when the sleeve was removed. The rotary unit is out of balance. You must balance everything that rotates with the shaft such as the impeller, sleeve, sleeve gasket, drive key, seal, bearings, coupling, motor etc. o It never was balanced. o Cavitation damage. o Some of the product has attached it self to the rotating assembly. o The impeller is the most logical place to look for un balance problems, especially in the balancing holes. o Erosion can remove metal from the rotating parts. o Corrosion can do the same. o Temperature distortion. o A non concentric sleeve, seal, impeller, coupling, etc. o The impeller was trimmed and not re&endash;balanced. o A piece was damaged during the installation process. The rotary unit is dragging something around with it. o A piece left over from the last seal change. No one notices that one of the springs has fallen out and is resting in the bottom of the stuffing box, getting ready to be picked up by the new seal. o A piece of the seal has come loose. Look for set screws, springs, drive lugs and all of the obvious seal parts. The pump is running at a critical speed, or it has passed through a critical speed. The seal or sleeve is not concentric with the shaft.

Observation - The mark is all around both units. • •

Look for a combination of the first two discussed. This is a very common condition. Thermal expansion.

The shaft usually expands faster than restriction bushings placed in the bottom of the stuffing box. o Hot oil applications use a thermal bushing in the bottom of the stuffing box to gain more efficiency from the cooling jacket. o Steam is often used as a quench with an A.P.I. gland. This gland has a close fitting disaster bushing that can be overheated by the quench temperature. Excessive vibration. o Bad bearings or a loose bearing fit. o Cavitation - there are five types&emdash; o Harmonic, from nearby equipment. o Seal "slip stick". o

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Observation - One spot on both the stationary and rotary units. •

This is caused by a momentary deflection of the rotary unit. Just about the only time it happens is when some one drops the pump while it is being transported.

Observation &endash; One spot on the rotating unit. •

Someone has hit the piece with a hammer.

Why do not good seals wear out? We know that a mechanical seal is supposed to run until the carbon wears down, but our experience shows us this never happens with the original equipment seal that came installed in the pump. We buy an expensive new mechanical seal and that one doesn't wear out either. What is wrong? Was the new seal a waste of money? Not really. You are doing something that appears logical, you are trying to solve the seal problem by purchasing a different seal, but that is like trying to get a good paint job on an automobile by buying a good brand of paint. If you wanted to get a good paint job on an automobile you would have to do four things and purchasing a good brand of paint is only one of them. Here are the things you would have to do in no particular order: • • •

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Prepare the body. This is the most costly part involving metal repair, rust removal, sanding, masking etc. Buy a good brand of paint. All paint is not the same, and like anything good it will cost more money than many other brands. Apply the paint correctly. This means exactly the right amount of air pressure and a technique that guarantees no drips or runs. It also means a super clean paint room and frequent sanding between primer and finish coats. Needless to say the paint job can be ruined in this step. Take care of the paint after it has been applied. This means that you have to keep the car washed and waxed and garaged in bad weather. It also means frequent touch ups and paying attention to small details.

If you did those four things correctly, how long can a paint job last on an automobile? Obviously for years. Step outside and watch the cars go by and you will see evidence of people that are not doing those four things. In fact it is so rare that when we see an older car that looks good, we stare at it. Getting good seal life involves four steps also. They should be obvious, but let's look at them any way.: • • • •

Prepare the pump for the seal, that's the body work Purchase a good seal, the good paint. Install the seal correctly, apply the paint correctly Apply the correct environmental control if necessary (and it probably is),washing and waxing.

We will look at each of these subjects in detail and hopefully begin to increase the life of our mechanical seals to the point where most of them wear out. We will be discussing

seals for centrifugal pumps in this paper, but the information applies to just about any kind of rotating equipment including mixers and agitators. Prepare the pump for the seal • • • • • • • • • • • • • • • • •

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Do an alignment between the pump and driver. Use a laser aligner. A "C or D" frame adapter is an even better choice. Dynamically balance the rotating assembly. You can use most vibration analysis equipment to do this. Check with your supplier if you do not have the program. Make sure the shaft is not bent. Rotate it between centers. Avoid shaft sleeves. A solid shaft is less likely to deflect and is much better for a mechanical seal. Reduce pipe strain where ever possible. Use a "center line" design pump if the product temperature is greater than 200°F (100°C). This will reduce some pipe strain problems at the pump. Use pumps with a low ratio. This is extremely important with intermittent service pumps. Use an oversize stuffing box. Avoid tapered designs. Give the seal lots of room. Try to get the stuffing box face as square to the shaft as possible. There are facing tools available to do this. Reduce vibration by any techniques you know or can learn. Do not let the pump cavitate. The seal faces will bounce open and possibly become damaged. Water hammer can occur if power is lost to the pump while it is running. Maybe you can take some preventative action to avoid water hammer problems. Be sure the mass of the pump/motor pedestal is at least five times the mass of the hardware sitting on it. Be sure there are ten diameters of pipe between the pump suction and the first elbow. Be sure the base plate is level and grouted in place. Keep the open impeller adjusted to lessen vibration and internal recirculation problems. Make sure the bearings have the proper amount of lubrication and that water and solids are not penetrating into the bearing cavity. Replace the grease or lip seals with labyrinth or face seals. Avoid discharge recirculation lines connected to the stuffing box. In most instances suction recirculation will be better. If the pump has wear rings, check their clearance. Make sure the wetted parts of the pump are manufactured from corrosion resistant materials. Cleaners and solvents in the lines sometimes cause problems that the designer never anticipated. Seal off any air that might be leaking into the suction side of the pump and remove any that might be trapped in the volute.

Purchase a good seal

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Use hydraulically balanced designs that seal both pressure and vacuum. If you are going to use an elastomer in the seal try to use an o-ring. They are the best shape for lots of reasons, but don't let any one spring load the o-ring or it will not flex or roll as it should. Use non fretting seal designs. Shaft fretting is a major cause of premature seal failure. Stationary seals (the springs do not rotate with the shaft) are better than rotating seals (the springs rotate) for sealing fugitive emissions and any other fluids. If the seal has small springs keep them out of the fluid or they will clog easily. There are plenty of seal designs that have this non-clogging feature. A wide hard face is excellent for the radial movement we see in mixer applications and those seals that are physically positioned a long way from the bearings. You will need some sort of vibration damping for high temperature metal bellows seals. They lack the elastomer that normally performs that function. Use designs that keep the sealing fluid at the seal outside diameter, or centrifugal force will throw solids into the lapped faces and restrict their movement when the carbon wears. Use unfilled carbons for the seal faces. They are the best kinds and the cost is not excessive. Be sure you can identify all of the seal materials. It is impossible to troubleshoot a "mystery material". Do not let the supplier tell you that his material is proprietary. If that is his attitude find another supplier or manufacturer, otherwise you deserve all of the problems you are going to have. Try to keep elastomers away from the seal face. The elastomer is the one part of the seal that is the most sensitive to heat, and the temperature is hottest at the faces. Any dangerous or expensive product should be sealed with dual seals. Be sure the hydraulic balance is in both directions or you are gambling that one of the faces might open in a pressure reversal or surge. If the design has a carbon pressed into a metal holder, be sure the carbon was pressed and not "shrunk in". Pressed carbon will shear to conform to irregularities in the metal holder&emdash;helping to keep the lapped faces flat.

Install the seal correctly •

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Cartridge seals are the only design that makes sense if you want to make impeller adjustments and they are a lot easier to install because you do not need a print, or take any measurements to get the correct face load. Cartridge dual seals should have a pumping ring built in. Use buffer fluid (lower pressure) between the seals when ever possible to avoid product dilution problems. Avoid any type of oil as a buffer fluid because of oil's low specific heat and poor conductivity. Keep the seal as close to the bearings as possible. There is usually room to move the seal out of the stuffing box and then use the stuffing box area for a support

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bushing to help stabilize the rotating shaft. Depending upon the application you will have to decide if this support bushing has to be retained axially. Split seals make sense in just about any application that does not require dual seals or fugitive emission sealing (leakage measured in parts per million) Split seals are the only design to use on double ended pumps or otherwise you will have to replace both seals when only one seal has failed. They also allow you to change seals without having to do a re-alignment with the pump driver. Do not lubricate seal faces at installation. Keep solids off the lapped faces. If there is a protective coating on the seal faces be sure to remove it prior to installation Rubber bellows seals require a special lubricant that will cause the bellows to stick to the shaft. It is normally a petroleum based fluid, but you can check with your supplier to be sure. Rubber bellows seals require a shaft finish of no better than 40 RMS, or the rubber will have difficulty sticking to the shaft. In a vertical application, be sure to vent the stuffing box at the seal faces. You may have to install this vent because the pump manufacturer never provided it. Many cartridge seals have a vent built in that you can connect to the pump suction or some other low pressure point in the system.

Take care of the seal •

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The seal would prefer to be sealing a cool, clean, lubricating liquid. We seldom have one of those to seal so maybe you can apply an environmental control in the stuffing box area to change your product into a cool, clean, lubricating liquid: If you are using a jacketed stuffing box, be sure the jacket is clean. Condensate or steam are the best fluids to circulate through the jacket. Install a carbon bushing in the end of the stuffing box to act as a thermal barrier that will help to stabilize the stuffing box temperature. Flushing is the ultimate environmental control. It causes product dilution, but if you are using the correct seal you won't need much flush. Four or five gallons per hour (notice I said hour not minute) should be enough for that type of seal. Keep the fluid moving in the stuffing box to prevent a build up of heat. Suction recirculation will remove solids that are heavier than the product you are sealing. Since that is the most common slurry condition, use suction recirculation as your standard. Learn where not to use it also. Discharge recirculation will allow you to raise the pressure in the stuffing box to prevent a fluid from vaporizing between the lapped faces. Try not to aim the recirculation line at the lapped faces, it could injure them. If you are using a metal bellows the recirculation line can act as a sand blaster and cut the thin bellows plates. If the product is too hot, cool the stuffing box area, There are lots of ways to do this. Check other sections of the Technical Series for ideas. It is important to remember that these environmental controls are often more important when the pump is stopped because soak temperatures and shut down cooling can change the stuffing box temperature drastically, causing the product to change state.

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Dangerous products will need an A.P.I. type gland if you elect not to use dual seals. The disaster bushing that is part of the A.P.I. configuration should protect the seal from physical damage if you should lose a bearing when the pump is running. Do not put too much steam or water through the quench connection or it will get into the bearing case. Leakage out the drain connection is often perceived as a seal failure by operators. Be sure they know the difference. Be sure the connections are made correctly. It is easy to mix up the four ports and get the flush or recirculation line into the quench port.

Does any one ever do all of these four things? Unfortunately not. If we did eighty five or ninety percent of our seals would be wearing out rather than the ten or fifteen percent that wear out now. The prematurely failed seal with plenty of carbon face left continues to be the rule. The most common excuse we hear to explain our lack of good seal life is that there is never time to do it right, followed by the cliché, "but there is always time to fix it". Most of us do one or two of the necessary steps and experience an increase in our seal life. There is nothing wrong with an increase in seal life, but that is a long way from wearing out seals. Think about it for a minute. If the seal is lasting a year, how big can the problem be? The temperature cannot be too high or the pressure too severe. If that were true it wouldn't take a year to fail the seal. The product can't be too dirty for the same reason. We often find the problem is as simple as a seal design that is fretting the shaft, causing a leak path through the damaged sleeve or shaft. Other times we find that the flush that is used to clean the lines once a year is the culprit, and no one is changing the seal materials to reflect this threat to the seal components.

Why Mechanical Seals fail A mechanical seal can either wear out or fail. To determine which your seals are doing, look at the wearable face. In most instances this will be the face manufactured from some grade of carbon/ graphite. Since the seal face is the only sacrificial part of the mechanical seal, a worn out seal is identified as one that has no carbon nose piece left at the time it started to leak. A failed seal is identified by the fact that it has substantial carbon remaining at the time it started to leak.

The above illustrations show the difference between a worn out and a new mechanical seal. Most consumers experience seal failure rates in excess of 85%, and for the most part these seal failures are easily correctable. Seal failures fall into only two broad categories, either the seal faces opened, or one of the seal components was damaged by contact, heat or corrosion. Whenever we try to troubleshoot any mechanical seal it is wise to remember that only three things are visible: • •

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Evidence of rubbing. Evidence of damage including corrosion, physical damage, or discoloration of one of the seal component materials. Most mechanical seals are constructed of three materials: o Metal parts o A face combination o Some rubber like parts (called elastomers) The product is attaching to a sliding component causing sticking, or coating on the face causing face separation.

Here are some reasons why a mechanical seal face would open: The dynamic elastomer is not free to slide or move on the rotating shaft or sleeve. • •

The shaft is oversize. A tolerance of + 0.000 - 0.002 inches (+ 0,00 - 0,05 mm) would be typical. The shaft finish is too rough. Most seal companies want at least a 32 R.M.S. (0,8 micro meters) surface finish in the area of the dynamic (sliding) elastomer.

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The fluid we are pumping is causing the elastomer to stick to the shaft. The dynamic O-ring can generate a lot of heat if there is misalignment between the shaft and the stuffing box face. The rapid movement of the elastomer will generate localized heat causing the following to occur at a faster rate: o The product is solidifying (glue and paint will do this) o It is crystallizing (sugar syrup and caustic are good examples) o It is building a coating on the shaft (petroleum products will form varnish or coke at elevated temperatures, or hard water will form a layer of calcium. etc.) Dirt or solids are restricting the elastomer from moving. Chemicals added to treat water or impurities in the water can collect on the seal sliding surfaces A chemical has attacked the elastomer causing it to swell up and restrict the movement of the seal. In some instances a swollen elastomer has been known to open seal faces while the pump was not running in a standby mode. The shaft or sleeve has been hardened and the set screws have slipped. Many sleeves were hard coated to resist packing wear. Stock rooms are full of these sleeves. The seal has lost its compression. o It was installed with the wrong compression. o The impeller was adjusted after the seal was attached to the shaft. This is a very common problem with A.N.S.I. or other back pull out pumps. o A temperature change has altered the location of the seal. Remember that each inch of stainless steel shaft will grow one thousandth of an inch for each one hundred degree Fahrenheit rise in temperature or 0.001"/1"/100°F . Metric grows 0,001 mm/1 mm of shaft for each 50°C rise in temperature. o The open impeller was adjusted to compensate for normal wear. Typical pump specifications allow the impeller and the casing each to wear as much as 0.125 inch (3 mm) and still be adjusted back to the correct pump efficiency. This is important when you realize that the average mechanical seal has a carbon nose that extends only 0.125 inch (3 mm). The springs, spring or bellows are not operating properly. o A single spring has been installed backwards allowing the faces to stay in contact while the shaft or sleeve rotates within the dynamic elastomer or end fitting. o Excessive misalignment is causing rapid flexing of the spring or bellows causing them to fatigue. o The drive lugs have failed and the multiple springs are twisted in their holder. o The product has clogged the springs. o Many times the outside springs of a dual seal have been painted either at the pump company or as part of a normal maintenance routine.

Something is restricting the free movement of the seal.

•

• • •

The product is viscous. Remember that some products become more viscous with agitation. These products are called dilatants (cream becomes butter with agitation) A recirculation line from the discharge of the pump is aimed at the seal and interfering with its movement. A foreign object is in the stuffing box. A protruding gasket is touching the movable part of the seal

The shaft is being displaced causing the seal to hit something as it rotates or to cause the rotating face to run off of the stationary face. • • • • • • • • •

The pump is operating off of its best efficiency point (B.E.P.) causing the shaft to bend. The rotating assembly is out of dynamic balance. The shaft is bent. There is misalignment between the motor and the pump. Pipe Strain is twisting the pump stuffing box. Heat causes expansion and that always opens the possibility for rubbing or wear. Cavitation, slip stick, harmonic vibration, bad bearings or some other form of vibration is causing excessive movement of the shaft. The shaft sleeve is not concentric with the shaft causing it to run "off center". The pump designed with sleeve or babbitted bearings and shaft movement is excessive.

The seal face is being distorted by either temperature or pressure. The product is vaporizing between the seal faces causing the faces to blow apart. • •

If boiler feed water vaporizes it leaves behind all of the chemicals that were added to the water to prevent hardness, adjust PH, soften boiler scale etc.... In cryogenic (cold) applications the vaporizing fluid can freeze any lubricant that might have been placed on the seal faces. This frozen lubricant can damage the carbon/ graphite seal face.

An environmental control has failed. There are many types used with Mechanical Seals, here are a few of the common environmental controls: • • • • •

Flushing is used for cooling and to wash away solids. Quenching is used for temperature control and vapor removal. Barrier fluids are used to keep air away from a fluid and to provide temperature control. Cooling/ heating jackets are used to keep products in a liquid state and at the proper temperature. A suction recirculation line is installed from the bottom of the stuffing box to the suction side of the pump. This is done to remove stuffing box solids in the pumping fluid and to provide cooling to the seal components.

•

•

A line can be installed from the discharge of the pump to the stuffing box to increase stuffing box pressure whenever you pump a fluid close to its vapor point. It is also wise to install a carbon restriction bushing in the bottom of the stuffing box with a clearance of approximately 0.005" to 0.007" (0,13 mm to 0,018 mm) on the inside diameter. Dual seals can be installed to prevent a pressure drop across the inside seal face and to control the temperature at the seal face.

Unbalanced seals and some split seals can open their lapped faces in vacuum applications. •

•

Those pumps that run under vacuum include: condensate pumps, heater drain pumps, pumps that lift liquid and any pump that takes its suction from a condenser or evaporator. Remember to use O-ring elastomers in vacuum applications as this shape elastomer will seal either vacuum or pressure. The product has built up on one of the seal faces causing the faces to separate. This is a common problem with petroleum products or any product that can build a film on a surface. Since this coating is not dense enough to provide good sealing, it can cause the faces to leak at shutdown.

When a seal face opens it allows solids to penetrate between the lapped surfaces. The solids imbed themselves into the softer carbon/graphite face causing it to act like a grinding wheel. This grinding action will cause severe wear in the hard face. It should be noted that seal face opening accounts for the largest majority of mechanical seal failures. The second major cause of seal failure is when one of the seal components is attacked by the sealing fluid or a chemical being used to clean or flush the lines. Chemical attack is easy to see: • •

•

•

The Carbon will appear to have a sponge like appearance Plated materials will have their hard coating peel off when the base material is attacked. This same thing happens when you allow rust to penetrate behind automobile paint and you then notice that the paint is peeling off in sheets. The elastomer will usually swell up and get soft. When an elastomer shrinks and gets hard it is almost always evidence of excessive heat. Prior to failure caused by excessive heat, most elastomers will take a compression set ( the round O-Ring becomes square) Metal components will develop pits and an overall dull appearance. The color of the metal is often an indication of the amount of heat it was subjected to:

FAHRENHEIT

COLOR OF THE METAL

CENTIGRADE

700 - 800

Straw Yellow

370 - 425

900 - 1000

Brown

480 - 540

1100 - 1200

Blue

600 -650

> 1200

Black

> 650

Here are a few things to consider when you suspect corrosion is the problem : The corrosion rate of almost all chemicals doubles with each 18 degree Fahrenheit (10 C.) rise in temperature. • •

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•

•

•

• •

• • •

•

Be sure to vent vertical pumps. Air trapped in the stuffing box is a good insulator. See if the operator is running the pump with a restricted discharge. In addition to deflecting the shaft it can cause a severe heat rise in the pump. The control valve may be stuck in the throttled position. Try to use a recirculating line from the bottom of the stuffing box to the suction side of the pump. This is practical in almost any application other than when we are pumping a product close to its vapor point and there would be a danger of vaporizing the product in the stuffing box. When ever possible bore out the packing stuffing box or install a large seal chamber in place of the packing stuffing box. This extra room will allow centrifugal force to centrifuge and clean the fluid in the seal chamber as well as provide extra cooling in the seal area. It is normal to dead end the fluid in the stuffing box when a cooling or heating jacket is being used. If a recirculation line is installed in the stuffing box along with the cooling jacket, the jacket will become inoperative because the circulating hot fluid will not be in the stuffing box long enough to be cooled by the jacket. Be sure to check that the cooling jacket is functioning. A layer of calcium inside the jacket, can just about stop heat transfer. If the water is too hard in your area, consider condensate as an alternative cooling fluid. More than one stuffing box jacket has frozen in cold weather, be sure to use non freezing cooling fluids at lower temperatures If a convection tank is being used with dual seals make sure it is operating. Every design has limits, make sure you are not exceeding them. Also check that the fluid is flowing from the top of the stuffing box to the convection tank and returning to the bottom of the stuffing box. I have seen many of these applications running backwards. Use only balanced seals. They generate less heat than unbalanced seals. If there is a bypass line installed from the discharge piping to the suction side of the pump, it may be heating up the incoming fluid. Check to see if the cooling jacket has been isolated and drained. This often occurs when a metal bellows seal is used in hot oil applications. An empty cooling jacket will act as an insulation to the stuffing box fluid. Remember that the cooling jacket is also there to cool down the shaft and protect the bearings. Do not disconnect it.

When you look for corrosion be sure to check out any cleaners or solvents that are used to flush out the system or clean the lines. Many grades of Viton® can be attacked by cleaning the lines with steam or caustic. It is important to identify all of the materials used in the seal components. •

Carbon fillers can be attacked by heat and chemicals

• • • •

Plated materials can crack due to differential expansion. Stainless Steel springs can break due to Chloride Stress Corrosion. Hardened set screws can corrode and vibrate loose. Some elastomers can be attacked by steam. Be careful of using petroleum grease on elastomers as some compounds can be attacked by any petroleum product.

Some hard coatings have very little flexibility and will crack with a small differential temperature. Be careful of tungsten carbide with a cobalt binder; nickel binder would be a much better choice. ®DuPont Dow elastomer

FAQ ON PUMPS AND FLUID SYSTEMS

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1. System 1.1 What is a system? 1.2 What is a siphon and how does it work? 1.3 Is a siphon reliable in an industrial application? 1.4 What is the difference between head and pressure? 1.5 How is the effect of nozzle(s) taken into account in a pumping system? 1.6 How is the pressure drop established for a control valve? 1.7 What is the relevance of the highest point of the system, assuming it is higher than the discharge point? 1.8 What is the pressure drop for several pieces of equipment in the same line? 1 9 What are fittings? 1.10 Why is the term head drop or pressure drop used when describing the effect of equipment on a system? 1.11 How can the same pump satisfy different flow requirements of a system? 1.12 Is the head at the suction side of a pump equal to the N.P.S.H. available? 1.13 Is the head at the discharge side of the pump equal to the Total Head? 1.14 What is the difference between the N.P.S.H. available and the N.P.S.H. required? 1.15 How is the pressure head at any location in a piping system determined and why bother? 1.16 What is the purpose of a variable speed drive?

2. Pump or Performance curve 2.1 What is Total Head? 2.2 What is Friction Head? 2.3 What is Velocity Head? 2.4 What is Static Head or Total Static Head? 2 5 What is N.P.S.H.? 2.6 What information is required to determine the Total Head of a pump? 2.7 What information do I need to order a pump? 2.8 What is the best way to start a pump? 2.9 What is a performance curve? 2.10 What does "centrifugal" refer to in centrifugal pump? 2.11 What is the Best Efficiency Point (B.E.P.)?

3. Calculations 3.1 What is barometric pressure and why should I care? 3.2 What is my elevation above sea level and why should I care? 3.3 What is the best equation to use for calculating the friction head of a Newtonian fluid? 3.4 What is the Moody diagram? 3.5 What is the Newton-Raphson iteration technique? 3.6 What is the Reynolds number?

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3.7 What is the effect of pipe roughness on Friction Head? 3.8 What is the effect of pipe fittings on the total pipe friction loss? 3.9 How can the Total Head of a system that has more then one outlet be determined and what is the effect compared to a system with one outlet? 3.10 How do you calculate pressure drop due to fluid friction?

4. General 4.1 What are some good sources of information (i.e. references) on pumps and pumping systems? 4.2 Does it take longer to cook a 4 min egg in Mexico city than on the beach ? Are you kidding 4 minutes is 4 minutes? I mean to get an egg to the same consistency as a typical 4 min egg, how long would it take in Mexico city vs. somewhere at sea level? 4.3 What is negative pressure? 4.4 What is relative and absolute pressure? 4.5 What is a control volume and how is it used? 4.6 What is an energy balance? 4.7 What is the system equation and how is it developed? 4.8 Does a fluid system with no pump have a Total Head? 4.9 What other devices can create pressure in such a way as to move fluid through a system? 4.10 In a multiple and identical pump system, if one pump is in poor running order what is the effect on the discharge header head and the flow to the system? 4.11 What happens if the damaged pump's performance curve has all points at a lower head than the good pump's performance curve?

5.Fluid 5.1 What is specific gravity? 5.2 What is viscosity? 5.3 What is the difference between Newtonian and non-Newtonian fluids and why should I care? 5.4 What is laminar and turbulent flow?

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1. System 1.1 What is a system? A system most specifically a fluid (i.e. liquid) system consists of a continuous fluid body and the devices that contain it. The system starts at the inlet(s) and ends at the outlet(s) including the pipes and tanks (I am referring to open vs. closed systems). The system can generate heat (i.e. friction) or loose heat (through a heat exchanger for example), it can also do work (i.e. pump, inductor, etc.). It is critical to clearly identify the inlet and outlet points of the system. In a typical pumping system where fluid is pumped from one tank to another, the inlet point is the surface of the suction tank and the outlet point is the surface of the discharge tank Sometimes the discharge point can be difficult to identify. An example is the case where the fluid is pumped up to an elevation, say z2, and is then transferred to a trough which takes it down to an elevation z3.. Which is the exit point of the system z2 or z3? The answer is z2. When the fluid hits the troth which is an open pipe, no more energy is required from the pump to push the fluid further, gravity takes over. Let's make this a little trickier, what happens if the trough is a pipe, is the discharge point now z3? Depends. If the pipe is full all the way to z3 then yes, if the pipe is not full, then no. The discharge point is the point where the fluid fails to completely fill the pipe. If the discharge pipe has a long portion which is sloping downward but nearly horizontal, how is it possible to ensure that the pipe is full at all times? I'll keep the answer for later.

1.2 What is a siphon and how does it work? A siphon is a system of pipe or tubing with the fluid inlet - the surface of the inlet reservoir at a higher elevation than the outlet and where some portion of the fluid path is higher than the inlet point. The potential energy of the fluid at the inlet of the siphon is higher than the potential energy at the outlet, this difference in energy drives the fluid through the system. How high can the siphon piping go above the inlet? This depends on the local barometric pressure. If the barometric pressure or head is 34 feet of water, which is the value at sea level, then the maximum rise after the inlet point is 34 feet when the fluid is water.

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1.3 Is a siphon reliable in an industrial application? No, not really, sometimes it might be handy to run a pipe from an elevated tank (starting from the fluid surface moving upwards and then downwards) to a lower location without installing a pump. The system should work as long as the line was full. Of course, at some time or other the line will be emptied and then the problem would be how to refill the line. A higher tank could be used to refill the line but this adds more complication. The practical approach would be to pump the fluid downwards to ensure that fluid could always be transferred. Or alternatively, run a line from directly underneath the inlet tank in a path that brings it continuously downwards to the discharge point. In this case the piping must be of sufficient size to deliver the flow and head required at the discharge point. 1.4 What is the difference between head and pressure? To start with, head is not equivalent to pressure. Head is a term which has units of a length or feet. In the following equation (Bernouilli 's equation) each of the terms is a head term: elevation head h, pressure head p/y and velocity head v2/2g. Head is equal to specific energy, of which the units are Ibf-ft/lbf. Therefore the elevation head is actually the specific potential energy, the pressure head, the specific pressure energy and the velocity head is the specific kinetic energy (specific means per unit weight).

So what is the difference? Head is energy per unit mass whereas pressure is a force per unit area. 1.5 How is the effect of nozzle(s) taken into account in a pumping system? The nozzle manufacturers will normally give the Ap (pressure drop) vs. flow for their nozzles. Since the purpose of a nozzle is to accelerate the fluid, we might expect that the velocity head to be mentioned. Two factors need to be considered when the fluid goes through the nozzle:

1. friction is increased due to the high velocity through the restriction and 2. the fluid velocity is increased which requires additional energy (kinetic energy). Why does the higher velocity require additional energy? Well consider this situation, if we had a restriction (such as a valve) and the fluid went back up to its initial velocity, no additional energy would be required other then the friction energy. However, nozzles are usually positioned at the discharge point of the system which means that a velocity increase for the fluid as it leaves the system affects the energy balance causing the kinetic energy to increase. The kinetic energy increase which is the velocity head must be supplied by the source of work in the system, the pump. The manufacturers don't really want to complicate matters by giving two pressure losses, it is simpler to give the one pressure loss required to run the nozzle at a given flow rate. Their philosophy is, if you supply the required pressure ahead of the nozzle, the nozzle will produce the required effect and don't bug me about the velocity. In other words, make sure that you have enough pressure ahead of the nozzle.

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1.6 How is the pressure drop established for a control valve? When designing a new system, if we assume a pressure drop across the valve of 10 ft of fluid, then it mil be generally possible to select a valve that mil give this pressure drop at a reasonable opening of say 90%. In other words, by using a Delta p of 10 ft for the pressure drop, we have fixed one of the parameters required to size a valve, without unduly restricting the task. 10 ft of pressure drop is a common value used in designing systems with control valves. This criteria will generally result in a valve size one size smaller than the line (i.e. if the line is 8", the valve is 6"). In the case of existing systems where the control valve is in place, we should be more careful. While the system is operating, the position of the valve should be noted. The manufacturers tables for this valve will give the pressure drop corresponding to the flow rate and valve opening. This pressure drop should be used in the calculations for Total Head. 1.7 What is the relevance of the high point of the system, assuming it is higher than the discharge point? After the initial start of a pump, the high point of a system will have to be reached before the system is entirely filled. When the system is filled the high point is no longer relevant since the static head required is equal to the elevation of the discharge point minus the elevation of the inlet point. During the initial phase, the discharge point's elevation is continually changing as it moves towards the outlet. If the high point of the line is unusually high, then during the start-up it may require more head than is available. To avoid this, make sure that the shut-off head is greater than the static head required to reach the high point. 1.8 What is the total pressure drop for several pieces of equipment in the same line? The pressure drop associated with each piece of equipment is additive. 1.9 What are fittings? Fittings are all the miscellaneous pipe connections (tees, elbows, Ys, etc.) sometimes known as hardware, required to run pipes and their branches in various directions to their destination. Manual valves are also considered fittings. 1.10 Why is the term pressure drop used when describing the effect of equipment on a system? To drive fluid through apiece of equipment there must be a force at the inlet greater than the force at the outlet. These forces are converted to pressure, which is more convenient in a fluid system. The difference (or drop) in pressure between the inlet and outlet is proportional to the overall force pushing the liquid forwards. If we convert pressure drop to head then we obtain the pressure drop value in terms of head (i.e. fluid column height) or pressure head. 1.11 How can the same pump satisfy different flow requirements of a system? If a pump is sized for a greater flow and head that is required for the present conditions, then a manual valve at the outlet of the pump can be used to throttle the flow down to the present

requirements. Therefore, at a future date the flow can be increased by simply opening a valve. This however is wasteful of energy and a variable speed drive should be considered. 1.12 Is the head at the suction side of a pump equal to the N.P.S.H. available? No, the N.P.S.H. available is the head in absolute fluid column height minus the vapor pressure (in terms of fluid column height) of the fluid. Close, but no cigar.

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1.13 Is the head at the discharge side of the pump equal to the Total Head? No, the Total Head is the difference in head between the discharge and the suction. 1.14 What is the difference between the N.P.S.H. available and the N.P.S.H. required? The N.P.S.H available can be calculated for a specific situation and depends on the barometric pressure, the friction loss between the system inlet and the pump suction flange, and other factors (see book). The N.P.S.H. required is given by the pump manufacturer and depends on the head, flow and type of pump. The N.P.S.H. available must always be greater than the N.P.S.H. required for the pump to operate properly. 1.15 How is the pressure head at any location in a piping system determined and why bother? First, calculate the Total Head of the system. Then, using a control volume, set one limit at the point where the pressure head is required and the other at the inlet or outlet of the system. Apply an energy balance and convert all energy terms to head. The resulting equation gives the pressure head at the point required (see Pump Handbook). Why bother? The most common reason for this calculation is to establish the pressure ahead of a control valve which is required to size the valve. 1.16 What is the purpose of a variable speed drive? All systems require a means of flow control. The plant's output requirements may change causing flow demand to vary and therefore the various systems throughout the process must be able to modify their output flow rate. To achieve this, pumps are sized for the maximum anticipated flow rate. The most frequent means of reducing the output flow rate is to have a line which re-circulates flow back to the suction tank. Another method is to have a valve in the discharge line which reduces the output flow rate when throttled. Either method works well, but there is a penalty to be paid in consumption of extra power for running a system which is oversized for the normal demand flow rate. A solution to this power waste is to use an electronic variable speed drive. For a new installation this alternative should be considered. This provides the same flow control as a valved system without energy waste. How does a variable speed drive work? The head and flow produced by a pump is the result of centrifugal force imparted to the liquid by the impeller. Centrifugal force is directly proportional to impeller diameter and rotational speed. We can affect the centrifugal force by either changing the impeller diameter, which is difficult, or varying the impeller speed, which of course is what a variable speed drive does. The family of curves shown on pump performance charts corresponds to the performance of a pump at constant speed with various impeller sizes. If we keep the impeller size constant and vary the speed of the pump, a similar set of curves for different pump speeds is produced. Therefore, when a variable speed drive is used, only the required pump head and flow is produced resulting in an appropriate power consumption.

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2. Pump or Performance curve 2.1 What is Total Head? Total Head is the difference between the head at the discharge vs. the head at the inlet of the pump. Total head is a measure of a pump's ability to push fluid through a system. This parameter (with the flow) is a more useful term than the pump discharge head since it is independent of a specific system. Also Total Head, just as any head at any location in the system, is independent of the fluid density. 2.2 What is Friction Head? Fluid layers move at different speeds depending on their position with respect to the pipe axis. The velocity is zero at the pipe wall and maximum at the pipe center. This difference in velocity between fluid layers is a source of friction. Another source of friction is the interaction between the fluid layers close to the pipe wall and the pipe roughness or the small peaks and valleys on the wall (for turbulent flow only). The sum of these two sources of friction is the total friction due to fluid movement. Friction head is the energy loss due to fluid movement and is proportional to the flow rate, pipe diameter and viscosity. Tables of values for friction head are available in many references. The Colebrook and Darcy equations provide a method of calculating friction head for Newtonian fluids. Another component of friction head is the pressure drop due to fittings. Many references supply the data for determining the friction loss due to fittings. The 2K method is recommended (see Pump Handbook). 2.3 What is Velocity Head? Velocity head is the kinetic energy of the fluid particles. Velocity head difference is the difference in kinetic energy between the inlet and outlet of the system.

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2.4 What is Static Head or Total Static Head? The static head or total static head is the potential energy of the system. It is the difference between the elevation of the outlet vs. the inlet point of the system. 2.5 What is N.P.S.H.? The Net Positive Suction Head (N.P.S.H.) is the head at the suction flange of the pump less the vapour pressure converted to fluid column height of the fluid. The N.P.S.H. is always positive since it is expressed in terms of absolute fluid column height. The term "Net" refers to the actual head at the pump suction flange and not the static head. The N.P.S.H. is independent of the fluid density as are all head terms. 2.6 What information is required to determine the Total Head of a pump? 1. Flow rate through the pump and everywhere throughout the system. 2. Physical parameters of the system: length and size of pipe, no. of fittings and type, elevation of inlet and outlet. 3. Equipment in the system: control valves, filters. 4. Fluid properties: temperature, viscosity and specific gravity. 2.7 What information do I need to order a pump? Total head, flow and fluid properties (i.e. temperature,, pH, composition etc). 2.8 What is the best way to start a pump? Start the pump with a closed discharge valve. 2.9 What is a performance curve? A performance curve is a plot of Total Head vs. flow rate for a specific impeller diameter. The plot starts at zero flow, The head at this point corresponds to the shut-off head of the pump, or point A. The curve then decreases to a point where the flow is maximum and the head minimum, point B. This point is sometimes called the run-out point. Beyond this, the pump cannot operate. The pump's range of operation is from point A to B.

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2.10 What does "centrifugal" refer to in centrifugal pump? A centrifugal pump consist of an impeller rotating within a fixed casing or volute. Because the impeller blades are curved, the fluid is pushed in a tangential and radial direction. A force which acts in a radial direction is known as a centrifugal force. This force is the same one that keeps water inside a bucket which is rotating at the end of a string. 2.11 What is the Best Efficiency Point (B.E.P.)? The B.E.P. (best efficiency point) is the point of highest efficiency of the pump. All points to the right or left of B.E.P have a lower efficiency. The impeller is subject to non-symmetrical forces when operating to the right or left of the B.E.P.. These forces manifest themselves as vibration depending on the speed and construction of the pump. The most stable area is near or at the B.E.P..

3. Calculations 3.1 What is barometric pressure and why should I care? Barometric pressure is the air pressure in absolute terms in the local environment. The air pressure is highest at sea level and gradually diminishes with elevation. Barometric pressure is often expressed inpsia (pound per square inch absolute) or feet of water absolute. The barometric pressure at sea level is 14.7 psia or 34 feet of water absolute. Barometric pressure is used to calculate the N.P.S.H. available, which is required to determine if the pump will operate properly as designed. 3.2 What is my elevation above sea level and why should I care? Your elevation above sea level varies with your location. Your local airport can give you their elevation and barometric pressure. The relationship between elevation and barometric pressure is well documented and available in many reference books as charts or tables. You can find your

local elevation on a topographic map and determine the barometric pressure at your location. For example, the air pressure at sea level is 14.7 psia, at 10,000 feet it is 10.2

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psia, and at 35,000 feet (the cruising altitude of most passenger jets) 3.5 psia. The local barometric pressure is required to calculate the N.P.S.H. available at the pump suction. Ever see a movie where people and things are sucked out of an airplane after the bad guy shoots a hole through a window. Well at a 35,000 feet altitude, an object located over a 12" diameter hole (approximate size of a window) will be subject to a force of 1270 pounds, frightening isn 't it? 3.3 What is the best equation to use for calculating the friction head of a Newtonian fluid? For turbulent flow, the Colebrook equation to calculate the friction factor followed by the Darcy equation to get the friction head. For laminar flow, the laminar flow equation followed by the Darcy equation. 3.4 What is the Moody diagram? A graphical representation of the Colebrook and laminar flow equation. 3.5 What is the Newton-Raphson iteration technique? A technique used to solve for a non-explicit variable in an equation. An example is the friction factor in the Colebrook equation. The technique can resolve a complex equation very quickly, usually converging to a solution within 4 iterations (see Pump Handbook). 3.6 What is the Reynolds number? The Reynolds number is proportional to the kinematic viscosity, the average velocity and the pipe inside diameter. The kinematic viscosity (v) is the ratio of the absolute viscosity to the fluid density.

The Reynolds number is a non-dimensional number (i. e. has no units). It combines 3 important characteristics of the system and the fluid, velocity, viscosity and density. The diameter is termed the characteristic length. One of the many uses for this number is to establish if the flow is laminar or turbulent. A Reynolds number below 2000 indicates laminar flow and above 4000 turbulent flow. 3.7 What is the effect of pipe roughness on Friction Head? The Colebrook equation gives the value of the friction parameter f with respect to the Reynolds number and the pipe roughness. When the Reynolds number is small, below 2,000 (laminar flow region), pipe roughness has no effect at all. When the Reynolds number is between 4,000 and 50,000, that is low velocity and/or high viscosity, then the influence of pipe roughness is as equally important as the effect of velocity. When the

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Reynolds number is large, above 50,000, that is high velocity and/or low viscosity, then the friction is entirely dependent on pipe roughness. 3.8 What is the effect of pipe fittings on the total pipe friction loss? Any fitting inserted into a pipe run has an effect since it either obstructs the flow or re-directs it or both. Most common fittings have been studied and their effect quantified, the results are available in many reference books. 3.9 How. can the Total Head of a system that has more then one outlet be determined and what is the effect compared to a system with one outlet? One fluid path from inlet to a selected outlet is used for the calculation of Total Head. This path is assumed to require the highest Total Head, if there is a doubt about the head required for the other path then the calculation is done on the other path and a comparison is made. Also the velocity head input difference to the two separate branches needs to be added to the Total Head. This however is normally a small and negligible term (see book for a detailed explanation). 3.10 How do you calculate pressure drop due to fluid friction? The Colebrook equation is the most accepted formula for calculating the pressure or head drop due to friction in pipes for Newtonian fluids. This equation relates the friction factor to the Reynolds number and the pipe roughness. The friction factor is then used in the Darcy formula (see book) to calculate head drop. For non-Newtonian fluids, which is mostly slurries of one kind or another, the process is much more complicated and many factors are taken into account. Some of these factors are: particle size and distribution, settling velocity of the particles in the mixture, viscosity variation of the mixture, solids transportation mode, etc..

4. General 4.1 What are some good sources of information (i.e. references) on pumps and pumping systems? I haven't found yet any one comprehensive source whose theme matched my interests. At first, I found this discouraging, However, I consulted the following books:

1. Hydraulic Institute Engineering Data Book, Cleveland, Ohio, 1979 2. Goulds Pump Manual, Seneca Falls, New York, 1972 3. The Chemical Engineering Guide to Pumps, Ed. by K. McNaughton, McGraw-Hill Publications Co., New York, 1984 4. Durco Pump Engineering Manual, The Duriron Co., 1960 5. Principles of Unit Operations, A. Foust, L.A. Wenzel, C.W. Clump, L. Maus, L.B. Anderson, John Wiley & Sons, New York, 1960 6. The Piping Handbook, edit. Reno C. King, 5th Edition, McGraw Hill, New York, NY 1973

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7. Slurry Transport Using Centrifugal Pumps, K.C. Wilson, G.R. Addle, R. Clift, Elsevier Science Publishers Ltd., Crown House, LintonRoad, Barking, Essex 1 Gil 8JU, England 8. Cameron Hydraulic Data, Ed. by C.R. Westaway & A.W. Loomis, 16th edition, Ingersoll-Rand, Wooddiff, New Jersey, NJ 07675 9. Some Pipe Characteristics of Engineering Interest, L.F. Moody, Houille Blanche, June 1950 10. Turbulent Flow in Pipes with Particular Reference to the Transition Region between the Smooth and Rough pipe Laws, C.F. Colebrook, J. Inst. Civil Engrs. (London), February 1959 11. Fluid Mechanics with Engineering Applications, R.L. Daugherty & J.B. Franzini, 7th edition, McGraw-Hill Book Company, New York, NY 12. Esso, Product Information, Lubricants and Specialties, 1990 13. Van Nostrand Reinhold Encyclopedia of Chemistry, ed. D.M. Considine, 4th edition, Van Nostrand Reinhold Company, 1984, New York 14. Chemical Engineering, William B. Hooper, August 24, 1981 15. The Pump Handbook, McGraw Hill 4.2 Does it take longer to cook a 4 min egg in Mexico city than on the beach ? Are you kidding 4 minutes is 4 minutes? I mean to get an egg to the same consistency as a typical 4 min egg, how long would it take in Mexico city vs. somewhere at sea level? Different liquids boil at different temperatures for a given air pressure. For example, water boils at a temperature of 212 °F at an air pressure of 14.7 psia (the pressure at sea level). However, a temperature of 189 °F is required to boil water at a pressure of 11 psia which is the air pressure at 8,500 feet above sea level, the altitude of Mexico city. Just because water boils at a lower temperature in Mexico city doesn 't mean that it takes a shorter time to boil an egg. The same amount of heat transfer is required to get the egg to the right consistency regardless of water temperature. It will take longer to transfer enough heat to cook the egg if the water is boiling at a lower temperature than a higher one. We are so used to water boiling at the same temperature that it is very surprising to find that it takes longer than 4 minutes to boil a 4 minutes egg in Mexico city. How much longer I don't know, I'd have to go to Mexico city. If there are any Mexico city residents out there on the web, please try the experiment and let me know.

4.3 What is negative pressure ? Pressure is said to be negative when it is less than the local barometric or atmospheric pressure. 4.4 What is relative and absolute pressure? A pressure measurement that is absolute is not related to any other. The atmospheric pressure at sea level is 14.7psia (pounds per square inch absolute), that is, 14.7psi above zero absolute. Relative pressure is always related to the local atmospheric pressure. For example, 10 psig (pound per square inch gauge) is 10 psi above the local atmospheric pressure. Most pressure measurements are taken in psig which is relative to the local pressure. Pressure measurements do not normally have to be corrected for altitude since all the measurements you might do on a system are relative to the same atmospheric pressure

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therefore the effect of elevation is not a factor. An important exception to this is when taking a pressure measurement at the pump suction to determine the N.P.S.H. available. This pressure measurements is converted to absolute pressure which should be corrected for altitude. 4.5 What is a control volume and how is it used? A control volume is a theoretical boundary which helps delimit the extent of a system, particularly all its inputs and outputs. The principles of conservation of mass and energy can then be applied within this region. 4.6 What is an energy balance? Because of the principle of conservation of energy, any energy gain or loss in a system must be accounted for. Therefore, making an energy balance is the process of identifying all the sources of energy gain or loss and adding them up. The result must be equal to zero. 4.7 What is the system equation and how is it developed? The system equation has on the left hand side the Total Head (difference between the pump discharge head and suction head), and on the right hand side, all the terms which impede fluid flow such as: friction, velocity, elevation difference, etc. An energy balance is used to derive the system equation. 4.8 Does a fluid system with no pump have a Total Head? No, Total Head is a term that is used only for a pump. 4.9 What other devices can create pressure in such a way as to move fluid through a system? An inductor can raise the pressure of a fluid by using another fluid at a higher pressure. The company Schute & Koerting manufacture these devices. 4.10 In a multiple and identical pump system, if one pump is in poor running order what is the effect on the discharge header head and the flow to the system? Consider a two pump system where one of the pumps is in poor running condition as compared to the other. This could be due to: worn or damaged impeller, worn casing, worn bearings and shaft, wrong impeller, etc. Any or a combinations of these factors will have an effect on the pump's performance. The efficiency of the pump will be affected as well as the head and flow. It is difficult to predict the resulting performance curve without doing tests. However, unless the pump has gaping holes, the performance curve should look similar to that of the good pump but with a lower capacity and head. Let's assume that there is negligible friction loss between the discharges of pumps A or B and the header, also the head at the inlet of both pumps is the same. The operating point is point 1 on curve A which corresponds to 500 USGPM and 96 ft. The curve for the bad pump B, being slightly lower will contribute 265 USGPM at 96ft since it must operate at the same head as pump A. Therefore, the total flow will be 765 USGPM. If we had two good pumps the total flow would be 1000 USGPM instead of 765 USGPM. The head is not affected since it is the pump with the higher head which will control the pressure head in the discharge header by forcing the other pump to reduce its flow to match the higher pressure head. This what is meant when people say that one pump is fighting the other. Improperly designed suction or discharge piping can have this effect also. Page 13 of 15

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4.11 What happens if the damaged pump's performance curve has all points at a lower head than the good pump's performance curve? The best that the damaged pump can do is to produce the head corresponding to its shut-off head AHC (point 2) at 0 flow. Since the head produced by the good pump is higher, there will be flow through the damaged pump in the reverse direction. The flow however will be impeded since the pump can produce some head. The system behaves as a branch system. The branch flow sees a head drop which is the sum of the shut-off head of the damaged pump, plus any friction loss, plus the static head of the suction tank on the inlet of the damaged pump. 4.12 Ever wonder how it occurs that you can get partially full pipes in what appears to be a pressurized system? What would the difference in Total Head be for the same flow rate for a system with partially full pipes and a system with full pipes? (See the Maintaining a Full Pipe,Piping Handbook) 4.13 What is head actually? (See 1.4)

5.Fluid 5.1 What is specific gravity? By definition, the specific gravity of a fluid is:

where PF is the fluid density and PW is water density at standard conditions.

5.2 What is viscosity? Viscosity is the fluid property from which the resistance to movement can be evaluated (see Pump Handbook). The higher the viscosity the more difficult it is to move the fluid. 5.3 What is the difference between Newtonian and non-Newtonian fluids and why should I care? It is the relationship between the tangential stress or shear within the fluid (i.e. the friction force between layers per unit surface) and the velocity gradient or shear rate (i.e. the difference in speed between fluid layers divided by the distance between them) which defines whether a fluid is Newtonian or not. If the relationship is linear and the fluid has zero stress at zero velocity gradient then it is Newtonian (see Table A2) for a graphic representation of stress vs. velocity gradient. Many fluids do not behave in the well ordered fashion of Newtonian fluids. These are known as non-Newtonian fluids. They fall in several categories (see Table A2) depending on what shape the stress vs. velocity gradient takes. For these fluids, the velocity gradient is dependent on the viscosity. That is, the velocity affects the viscosity resulting in a much higher (or in some cases lower) stress than for a Newtonian fluid. A typical household product will help illustrate this point, try the following experiment. Make a solution of corn starch and water, approximately 1 part water to 2 parts cornstarch. Mix well into a large shallow bowl. Try moving this fluid

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rapidly around with yourflngers. The faster you try to move through the fluid, the higher the resistance. If you move your fingers fast enough they will skip over the surface. At that rate of shear, the solution almost behaves as a solid, when the fingers are moved slowly, the solution behaves more as expected offering little resistance. Compare this behaviour to another fluid that seems equally thick, such as molasses (molasses is not considered a Newtonian fluid, however it is much closer to being Newtonian than a starch solution). The molasses flows readily no matter how fast the movement. This is what is meant by viscosity being dependant on rate of shear. Why do I care? Many fluids that we deal with are Newtonian, but not as many as you might think. (see Pump Handbook), the data was not easy to find. I would appreciate references for any good sources on this subject.

Table A2 5.4 What is laminar and turbulent flow? Laminar flow is a very well behave flow usually occurring at low speeds for most fluids. In the laminar flow regime it is possible to determine theoretically the speed of any particle between the center of a pipe and the wall. Most fluids have to be carried at a much higher velocity which puts them in the turbulent flow regime. For turbulent flow, the fluid particles move in many directions, each particle reacts with its neighbor in an unpredictable fashion creating much higher internal friction than is present in the laminar flow situation. If you put dye in a laminar flow system, you will observe nice long streams of dye undisturbed by the surrounding liquid. The same dye inserted in a turbulent flow will immediately be dispersed through out the liquid.

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PUMP INFORMATION 1. Main characteristics to be considered 2. Impeller Diameter and RPM 3. Pump Characteristic Performance curve 4. Brake Horsepower 5. Cavitation 6. Time, Temperature and Pressure effects 1. There are two main characteristics to be considered when determining the correct pump: 1. The right type of pump for the application and liquid. 2. The pump performance or characteristic requirements. First, in order to select the right type of pump you need to know the fundamentals. Pumps can be broken into two major categories. The first category is the positive displacement (PD) pump in which the liquid is moved by physically displacing that liquid from a confined space. Positive displacement pumps come in several configurations such as vane, gear, screw, progressive cavity, Archimedes' screw, piston, plunger, and others. Some of the very first pumps used by man for primitive irrigation were positive displacement pumps. PumpU will discuss PD pumps second. The next major category is the centrifugal pump. The centrifugal pump adds energy to the liquid by first increasing the velocity of the liquid (adding kinetic energy) and then converting the kinetic energy to potential energy by directing the liquid into a progressively increasing volume, which in turn slows the velocity. Centrifugal pumps come in several configurations such as axial flow or propeller pumps, mixed flow or turbine pumps, and Francis vane and radial vane pumps. The last two categories represent the type of pumps dealt with most. A centrifugal pump is a device, which converts driver energy to kinetic energy in a liquid by accelerating it to the outer rim of a revolving device known as an impeller. The impeller, driven by the pump shaft adds the velocity component to the liquid by centrifugally casting the liquid away from the impeller vane tips. The key idea here is that the energy created is kinetic energy. The amount of energy given to the liquid corresponds to the velocity at the edge or vane tip of the impeller. The faster the impeller revolves or the bigger the impeller is, then the higher will be the velocity of the liquid at the vane tip and the greater the energy imparted to the liquid. The kinetic energy of a liquid coming out of an impeller is harnessed by creating a resistance to the flow. The pump volute (casing) which catches the liquid and slows it down creates the first resistance. The case is the pressure vessel required to confine the liquid at the pressures generated by the system and the pump itself. The volute is the cavity cast into the case that progressively slows the liquid after the liquid has exited at high velocity from the impeller. When the liquid slows down in the pump casing some of the kinetic energy is converted to pressure energy. It is the resistance to the pump's flow that is read on a pressure gauge attached to the discharge line. Note!!! A pump does not create pressure, it only creates flow! Pressure is a measurement of the resistance to flow. In Newtonian fluids (non-viscous liquids like water or gasoline) we use the term head to measure the kinetic energy which a pump creates. Head is a measurement of the height of a liquid column which the pump could create resulting from the kinetic energy the pump gives to the liquid (imagine a pipe shooting a jet of water straight up into the air, the height the water goes up would be the head). The main reason for using head instead of pressure to measure a centrifugal pump's energy is that the pressure from a pump will change if the specific gravity (weight) of the liquid changes, but the head will not change. So we can always describe a pump's performance on any

Newtonian fluid, whether it's heavy (sulfuric acid) or light (gasoline) by using the term head. Remember, head is related to the velocity, which the liquid gains when going through the pump. To convert head to pressure (psi) the following formula applies: PSI = Head (FT) x S. G. 2.31 Newtonian liquids have specific gravities typically ranging from 0.5 (light, like light hydrocarbons) to 1.8 (heavy, like concentrated sulfuric acid). Water is a benchmark, having a specific gravity of 1.0. 2. Impeller Diameter and RPM The two main factors in determining how much head a pump creates are: The Impeller Diameter The RPM of the Impeller (revolutions per minute,) Impeller Diameter If the speed (revolutions per minute) of the impeller remains the same then the larger the impeller diameter the higher the generated head. Note that as you increase the diameter of the impeller the tip speed at the outer edge of the impeller increases commensurately. However, the total energy imparted to the liquid as the diameter increases goes up by the square of the diameter increase. This can be understood by the fact that the liquid's energy is a function of its velocity and the velocity accelerates as the liquid passes through the impeller. A wider diameter impeller accelerates the liquid to a final exit velocity greater than the proportional increase in the diameter. RMPs (Revolutions Per Minute) As the number of revolutions per minute of an impeller increases, the velocity (and head) imparted to the liquid passing through it increases as well. As the impeller revolves more rapidly the rate of increase in the liquid velocity is higher than the rate of rpm increase. In other words, an impeller spinning at 2000 RPMs generates more than twice the head of the same impeller spinning at 1000 RPMs. A version of the single centrifugal is the multiple stage pumps that either have volutes cast into the case as in the horizontally split design or can have a diffuser as part of the stationary components at the discharge of each impeller. The diffuser provides the same function as the volute. The diffuser provides an expanding cavity for the high velocity liquid to slow and build pressure. In the multiple stage pump, the volutes or diffuser for each stage is the beginning of the passage that redirect the liquid back to the center of the pump to enter the eye of the next stage. Specific Speed (Ns) is used to relate the hydraulic performance of a centrifugal pump to the shape and physical proportions of its impeller. Most pumps in the industry today range from a specific speed of 3000 down to 500. The 3000 index indicates a characteristic of higher flow rates at lower differential heads. The index of 500 indicates a characteristic of high differential head at relatively low flow rates. 3. Pump performance A pump's performance is shown in its characteristics performance curve where its capacity (GPM) is plotted against its total developed head (FT), efficiency (%), required input power (BHP), and NPSHr (FT) The pump curve also shows its speed (in RPM) and other information such as pump size and type, impeller size, etc.

Pump Speed in RPMs Also in the notice that most curves indicates performance at the speed of 3450 RPM (a common electric motor speed in 60 Hz countries). All the information given in the curve is valid only for 3450 RPM. Generally speaking, curves that indicate RPM to be between 3400 and 3600 RPM are used for all two pole (3600 RPM nominal speed) motors applications. Flow The pump's flow range is shown along the bottom of the performance curve. Note that the pump, when operating at one speed, 3450 RPM, can provide various flows. The amount of flow varies with the amount of head generated. As a general rule with centrifugal pumps, an increase in flow causes a decrease in head. Head The left side of the performance curve indicates the amount of total head a pump is capable of generating. Trimmed Impeller Curves Notice that on some there are several curves, which slope generally downward as they move from left to right on the curve. These curves show that actual performance of the pump at

various impeller diameters. Duty Point The point on the curve where the flow and head match the application's requirement is known as the duly point. A centrifugal pump always operates at the point on its performance curve where its head matches the resistance in the pipeline. The term head refers to the differential head developed by a pump expressed in feet of liquid: H = [Pd-Ps] x 2.31 / SG where: H = pump head, FT of liquid Pd = pump discharge pressure, PSIG Ps = pump suction pressure, PSIG SG = liquid specific gravity It is important to understand that a centrifugal pump is not limited to a single flow at a given speed. Its flow depends on the amount of resistance it encounters in the pipeline. To control the flow of a centrifugal pump it is normally necessary to restrict the discharge pipeline, usually with a valve, and thus set the flow at the desired rate. Note: Generally speaking, do not restrict a pump’s flow by putting a valve on the suction line. This can cause damage to the pump! 4. Brake Horsepower (BHP) Along the bottom of the typical performance curve are brake horsepower lines sloping upward from left to right. These lines correspond to the performance curves above them (the top performance curve corresponds to the top BHP line and so on). These lines indicate the amount of driver horsepower, which is required at different points of the performance curve. The lines correspond to a BHP horsepower scale on the lower right hand corner of the page. In our example operating point at 120 gpm and 150 feet of head we observe that the corresponding BHP line equals about 6.8 horsepower. See the chart below. End of Curve Horsepower When sizing a motor driver to fit an application it is necessary to consider whether the pump will ever be required to operate at a flow higher than the duty point. The motor will need to be sized accordingly. If the pump may flow out to the end of the curve (if someone opens the restriction valve all the way, for example) it is important that the motor does not become overloaded as a result. Therefore it is normal practice to size the motor not for the duty point, but for the end of curve (EOC) horsepower requirements. 5. Cavitation Cavitation is a phenomenon, which occurs when a liquid vaporizes as it passes through a pump and then quickly turns back into a liquid. The collapse of the vapor bubbles creates destructive microjets of liquid strong enough to damage the pump. Vaporization occurs if the pumped liquid drops below its vapor pressure. As a liquid accelerates through a pump it loses pressure (Bernoulli's Principle). If the pressure drops below the vapor pressure of the liquid then gas bubbles will instantly form as the liquid vaporizes. These bubbles just as quickly collapse, causing cavitation to occur. To prevent cavitation the pressure (more correctly the head) of the liquid entering the pump must be high enough to prevent the subsequent liquid pressure drop from reaching liquid vapor pressure. Cavitation is the vaporization of the liquid caused by the pressure dropping below the vapor pressure of the liquid at the flowing temperature. NPSH

A minimum amount of suction pressure (head) is needed for a pump to operate without cavitating. The term used to describe this suction pressure is Net Positive Suction Head (NPSH). The amount of NPSH the pump requires to avoid cavitation is called NPSHr. The amount of NPSH available to the pump from the suction line is termed NPSHa. When selecting a pump it is necessary to see how much NPSH it requires at the duty point and make sure the NPSH available exceeds that amount. It is normal practice to have at least 2 feet of extra NPSH available at the suction flange to avoid any problems at the duty point. Also, if the pump were inadvertently operated at a flow higher than the rating point then a higher NPSH would be required to avoid cavitation. NET POSITIVE SUCTION HEAD Flooded suction: NPSH = ha - hv + hs - hf Suction lift: NPSH = ha – hv - hs - hf ha = the absolute pressure in feet of liquid on the surface of the supply liquid. hv = the vapor pressure of the liquid being pumped expressed in feet of head. Hs = the height in feet of the supply liquid surface with respect to the pump inlet. Hf = suction line friction losses expressed in feet of head.

These calculations yield the available net positive suction head for a given system. This must be compared to the required net positive suction head NPSHr calculated by the manufacturer. NPSHa must exceed NPSHr. Specific speed Specific speed (NS) is calculated from: NS = [N x FAQ^0.50] / [H^0.75] N = pump speed, RPM FAQ = capacity at best efficiency point (BEP) at maximum impeller diameter, GPM H = head at BEP at maximum impeller diameter, FT Specific speed identifies the type of pump according to its design and flow pattern. According to this criteria a pump can be classified as radial flow, mixed flow, or axial flow type. A radial flow pump is one where the impeller discharges the liquid in the radial direction from the pump shaft centerline, an axial flow pump discharges the liquid in the axial direction and a mixed flow pump is one that is a cross between a radial and an axial flow pump design. Specific speed identifies the approximate acceptable ratio of the impeller eye diameter (D1) to the impeller maximum diameter (D2) in designing a good impeller. NS: 500 to 5000; D1/D2 > 1.5 - radial flow pump NS: 5000 to 10000; D1/D2 < 1.5 - mixed flow pump NS:10000 to 15000; D1/D2 = 1 - axial flow pump 6. Time, Temperature and Pressure effects. A general rule of thumb for engineering plastics, thermoplastics, is that their mechanical properties are impacted, some rather dramatically. These factors affect the creep of the plastic, much differently than is the case with metals. Due to this fact and for certain processing benefits, many thermoplastics are reinforced with fiber (glass, carbon). Be sure to understand these relationships for the material of the product you are using. UHMW PE (ultra high molecular weight polyethylene) is nearly chemically inert, has great impact strength, abrasion resistance (sliding), toughness, lubricity. It is a highbred of the standard PE. PP (polypropylene) is a crystalline polymer, light weight (sg=0.91), excellent in chemical resistance with caustics, solvents and acids, and other organic chemicals, it is not recommended for use with oxidizing type acids, detergents, low-boiling point hydrocarbons, alcohol’s and some chlorinated organic materials. Upper practical use temperature limit is 160 – 180 F, depending on the load. Natural PP is white.

PVDF (polyvinylidene fluoride) is a thermoplastic fluoropolymer and is similar to PTFE. It is has high chemical resistance like the fluoropolymer group with the exception of not be being suitable with strong acids, fuming acids, polar solvents, amines, ketones, and esters. It has a high tensile strength and heat distortion temp (HDT) of around 300F. It is sold under the trade names of Kynar (Atochem) and Solef (Solvay). PVDF is naturally an off white / cream color. PFA (perfluoroalkoxy) is a thermoplastic fluoropolymer and is very similar to PTFE but lacks some of the chemical resistance and mechanical properties. However, due to its better processibility and near PTFE chemical resistance it is a PTFE alternative often. PFA is often used in roto-molding of metal parts. PTFE (polytetrafluorethylene) is resistant to practically every known chemical or solvent and highest use temperature of all the thermoplastic fluoropolymers. PTFE has a low coefficient of friction. Due to its physical properties it is not readily processed and there for does not lend itself to wide use in injection molded products. Components tend to be machined only. It is sold under the trade name Teflon (DuPont). PTFE is bright white in color. ETFE (ethylene tetrafluoroetheylene) is a tough plastic with good tensile strength, high impact resistance. It lacks the full chemical resistance of PTFE, but again is easier to process. It has us useful temperature limit around 350F. It has excellent resistance to solvents, caustics, chlorides, and most corrosive chemicals. It is sold under the trade name Tefzel (DuPont). It is a creamy-white color. ECTFE (ethylene chlorotrifluoroethyhlene) is a tough plastic with good tensile strength, high impact resistance. It lacks the full chemical resistance of PTFE, but again is easier to process. It has excellent resistance to solvents, caustics, chlorides, and most corrosive chemicals. It has a useful temperature limit of around 300F. It is sold under the trade name Halar (Ausimont). It is a shade of white in color. PVC (polyvinyl chloride) is the most popular of the vinyl thermoplastics, making its way into almost every day life in some form of a product. It is easily processed and joined together, stock shapes and extrusions are common. It has a useful temperature of around 140F, high impact strength and is amorphous. It has broad general chemical resistance. It is gray in color CPVC (chlorinated PVC) is very similar to PVC but it has a higher operating temperature of around 212F. Chemical resistance and impact properties are the same as PVC. It is gray to gray-blue in color. Santoprene is a competitive diaphragm material. It is similar to EPDM. Geolast is similar to that of Buna-N (nitrile). Non-aggressive fluids only FLOW RATES IN PIPES - Normal to peak 1" = 16-30 GPM 3" = 120-270 GPM 1-1/4" = 30-35 GPM 4" = 250-500 GPM 1-1/2" = 40-70 GPM 6" = 500-1100 GPM 2" = 65-120 GPM 8" = 1000-2000 GPM 2-1/2" = 80-170 GPM 10" = 1500-3000 GPM QUICK CALCULATE - FLOW RATE Normal Flow Rate of a Pipe = D 2 x 20 (Diameter Squared x 20) Twice the Diameter = 4 times the flow

Shaft fretting The next time you remove a grease or lip seal (the rubber seal located next to the bearing) you will note that the shaft is grooved and damaged under the rubber lip. You will see this same damage in a few other locations also: • • • •

On the sleeve under the stuffing box packing, if you are still using packing in your pumps. On the sleeve under the Teflon® wedge, "U' cup, or "V" rings if you are using original equipment type mechanical seals. Underneath the spring loaded o-ring found on many popular single and double mechanical seals. Underneath the rubber bellows of the type #1 seal if the rubber bellows did not vulcanize to the shaft.

This shaft or sleeve damage is called fretting and it will cause you several problems: • • •

• •

Sleeve replacement is costly. The pump bearings are often destroyed in the process of removing the damaged shaft sleeve. The shaft diameter was reduced to accommodate the wear sleeve. This reduction weakened the shaft, raised the L3/D4 number, and increased shaft deflection problems. The seal can "hang up" in the fretted groove, opening the lapped seal faces. The fretted grove becomes an additional leak path for the fluid. This is a major cause of seal failure

What causes this fretting problem? How can a soft piece of rubber or a slick wedge of Teflon cut a hard shaft? It doesn't seem to make any sense. Surprisingly it has nothing to do with dirt in the air or abrasives in the fluid. The problem will occur even if you are pumping a filtered, clean lubricant in a sterile atmosphere. To understand fretting you must first understand the term "corrosion resistant". Some materials resistant corrosion others do not. What is the difference? We say that iron rusts, but aluminum oxidizes. A look at any dictionary will verify that these terms mean the same thing. So why do we use different terms to describe the same problem?

The answer lies in the way a metal rusts or oxidizes. If the oxide layer is protective we say that the material is corrosion resistant. Take aluminum as an example: •

•

Aluminum protects its self by forming a layer of aluminum oxide on the surface when it is exposed to oxygen. It is very visible and looks almost white in color. A more common name for aluminum oxide is ceramic,a dense, hard, corrosion resistant material.

After this dense layer is formed on the surface of the aluminum the oxidation or rusting rate is slowed down to less than 0.002 inches (0,05 mm) per year, and this is the definition of corrosion resistant. If this protective oxide layer is rubbed or polished off by the packing, lip seal or Teflon wedge the oxide will immediately reform to protect the base material. It is this constant oxide removal and reforming that causing the shaft grooving that is so visible. We get the same reaction when we polish silver. The "tarnish" replaces its self to protect the silver. Shaft vibration and end play causes a constant axial movement of the shaft through the mechanical seal dynamic rubber or Teflon® part. Bearing grease seals and stuffing box packing are stationary, so the rotating shaft is constantly being polished by these materials when the pump is running There is a second problem associated with fretting. The ceramic oxide that is removed imbeds its self into the rubber part causing a wear or grinding action on the base metal. Stainless steel protects its self by forming a protective oxide called chrome oxide, one of the hardest ceramics. When this oxide forms we say that the active stainless steel is now "passivated". It is this chrome oxide imbedded into the packing, Teflon®, or rubber lip that does so much damage to the shaft sleeve. So now we have two causes of fretting: • •

The removal of the passivated layer by the rubbing action of the rubber or Teflon®. The hard ceramic that we removed sticking into the rubber or Teflon causing a grinding action.

Now that we know the causes of fretting what is the solution? • • •

Replace bearing lip or grease seals with labyrinth or the newer positive face seals. Face seals are the better choice. Stop putting packing into pumps. You don't need that kind of leakage any more. Do not use mechanical seals that are designed with a dynamic elastomer positioned on the pump shaft or sleeve. Most original equipment seals are

designed this way. Stationary cartridge seals, most balanced O-ring seals and all bellows seals eliminate the shaft dynamic elastomer and the fretting associated with it. ® DuPont Dow elastomer

Pump and seal problems with no apparent cause These problems are the ones that drive you crazy. No matter how hard you look the solution keeps evading you. Over the years I have collected quite a few examples. I offer some of them for your enjoyment and maybe, in the process, they will help you solve the "un-solvable" CAVITATION The pump cavitated every time it rained. •

Solution: The product temperature would cause it to vaporize very close to ambient pressure, and when it rained atmospheric pressure dropped enough to cause the problem.

The pump never cavitated in the summer months, only during the winter when everything was cooler. •

Solution: The tank vent froze during the winter months causing the pump to pull a partial vacuum in the tank.

The cavitation started suddenly. •

Solution: A plastic pipe liner collapsed at the suction side of the pump or the gate fell off a gate valve.

The cavitation started after the packing was converted to a mechanical seal. A careful inspection showed that the seal was not leaking air into the suction. •

Solution: The pump had speeded up (increased the rpm) when the packing was removed. This increase in speed and capacity caused the cavitation.

The cavitation kept getting worse with time, nothing obvious had changed in the system. •

Solution: The product had formed a coating on the inside of the suction pipe increasing the pressure drop and resulting in a loss of suction head.

The cavitation only occurred when there was a higher head at the suction of the pump and stopped cavitating when the level fell in the tank - just the opposite of what should have happened. •

Solution: The pump was pumping to a fixed discharge head. The capacity of the pump increased when the suction level was higher, because the

pump delivers the difference between the suction and discharge head. When the differential went down, the capacity increased. Two pumps were installed in parallel, one cavitated the other did not. They had separate suction lines so that was not the problem. •

Solution: Some one had installed an oversized section of pipe on the discharge side of the pump that was cavitating. The lower discharge resistance caused an increase in capacity which caused the cavitation. When the proper sized pipe was installed the cavitation stopped.

The pump had been cavitating for some time, but after a visual check everything appeared normal. •

Solution: A globe valve had been substituted for a gate valve on the suction of the pump. A globe valve can add the equivalent of another 100 feet (30,5 meters) of pipe to the system.

The pump started to cavitate when a flange gasket was replaced on the suction side of the pump. •

Solution: The inside diameter of the gasket was too small. It was acting as an orifice, and restricting the flow.

The pump cavitated about one third of the time it was running. •

Solution: A close inspection of the system revealed that there was no surge tank installed between the pump discharge and the multiple outlets that were using the product. The pump was acting like an accumulator and started to cavitate when the demand went up and the discharge head dropped.

The pump cavitated although here was excessive suction head available. •

Solution: There was too much velocity on the suction side of the pump. I saw this problem in Scandanavia in an application where the pump was taking a suction on a flow of water coming off of a mountain.

THE SEAL WAS GETTING HOT The seal was showing evidence of running dry, but the fluid level was never lost in the pump. •

Solution: Air was trapped in the stuffing box of a vertical pump after it was converted from packing to a mechanical seal. Most seal designs have no facility for venting the stuffing box in a vertical application

The seal showed evidence of running dry. •

Solution: The open impeller had been adjusted backwards and the "pump out vanes" on the rear of the impeller were pumping the stuffing box dry. This happens if you are using several brands of pumps and the maintenance mechanics confuse the impeller adjustment method. Some pumps adjust towards the volute (Goulds), some adjust towards the back plate (Duriron). It is easy to mix them up.

There was little to no fluid circulating between the two seals. •

Solution: The pipe fitting had bottomed out in a gland inlet elbow shutting off the flow. This sometimes happens after the seal has been repaired several times and the pipe thread shows some wear letting it protrude further into the elbow fitting.

The mechanic had marked the seal location on the shaft sleeve before the impeller was installed. When the impeller was tightened against the shaft shoulder the sleeve moved and over compressed the seal. Durco pump impellers adjust to the pump back plate. When you make impeller adjustments you over compress the mechanical seal. A cooling jacket was being used, but the seal continued to get hot. I have seen multiple reasons for this: •

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A discharge recirculation line had also been installed, but it was hidden by some insulation. The cooling jacket could not keep up with the heat being added by the recirculation line. The inside of the cooling jacket had become coated with a layer of calcium because hard water was being used as the cooling medium. Condensate should have been substituted. A thermal bushing had not been installed in the bottom of the stuffing box. The cooling jacket flow changed with fluctuations in shop water pressure. The inner seal of some double seal applications can get hot if the mechanic installs the cartridge seal by pushing on the gland and fails to reset the seal compression with the installation clips. The interference from the cartridge sleeve elastomer can cause enough resistance to compress the inner seal and unload the outer seal.

THE SEAL WOULD LEAK FOR NO APPARENT REASON The open impeller was being adjusted without resetting the seal. Many operators make their own impeller adjustments.

The seal faces were opening because the equipment's sliding foot had been bolted to the floor allowing the shaft to grow through the stuffing box when the unit came up to temperature. The cartridge seal had been hydrostatically tested with water and then put into a hot oil application. It leaked almost immediately. •

Solution: The trapped water vaporized when the unit was started. This could be a dangerous condition because water trapped in a gasket and then flashed to steam could blow the equipment apart.

The seal would start leaking about thirty minutes after the pump started. •

Solution: The carbon insert would come loose in its holder when the seal came up to temperature. At shut down the metal holder would shrink and everything appeared normal.

The seal was tested in the shop, but leaked when it was installed in the pump that was operating at cryogenic (cold) temperature. •

Solution: The faces had to be lapped at cryogenic temperature to keep them flat at the seal operating temperature. The cryogenic temperature can also harden the O-ring and freeze any lubricant that was put on the seal face.

The seal was found to be leaking every Monday morning. •

Solution: A utility man did not know about seals. He would loosen the gland on the weekend so that what he thought was packing would drip a little. The leak was found by the regular maintenance people every Monday morning.

The leakage occurred during the winter months. •

Solution: Someone circulated commercial anti freeze between two seals to act as a barrier fluid. The brand they selected contained a chemical to plug up radiator leaks and it kept plugging up the seal.

The seal would fail only during the winter months. The problem was traced to swelling of the dynamic O-ring but no logical reason could be found for its failure. •

Solution: During the winter months a worker decided to oil the bed of his dump truck to make the mined, raw product slip off easier. The petroleum oil he used attacked the Ethylene Propylene (EPR) O-ring in the mechanical seal, installed downstream in the system.

The seal area was wet, but no visible leakage could be seen. •

Solution: It turned out that there was a flange leaking above the pump and dripping the product next to the shaft.

The problem was traced to the fact that the mechanic was installing the seal at the wrong dimension. The written instructions were clear and placed in the box and yet the mechanic continued to do the installation incorrectly. •

Solution: The mechanic could not read. He had been faking it for many years and was quite good at it. The same problem occurs with older mechanics that refuse to wear glasses and as a result cannot see the funny little lines between the numbers on their measuring scale.

The centrifugal pump discharge was connected to the bottom of a surge tank. As the tank filled, the pump operating point shifted from too much capacity to too much head, deflecting the shaft in two directions. The outside seal in a double seal application failed suddenly. Nothing had changed in the system. •

Solution: Routine maintenance included repainting the pump. The paint spray got into the outside seal springs and stopped them from moving.

The seal ran great for several days and then started to leak. It tested all right on the test bench after it had been removed from the pump. •

Solution: It had been set screwed to a hardened sleeve and the set screws gradually loosened.

The seal was changed several times, but the steady leak persisted. •

Solution: The leak was occurring between the pump sleeve and the shaft. This is a common problem in double ended pumps that have been converted to a mechanical seal. You often have to devise a method of sealing the sleeve to the shaft or the sleeve to the impeller because the manufacture has not provided one.

The seal started to leak after many months of service. A bench vacuum test showed that the seal was all right. •

Solution: The seal was fretting the shaft below the Teflon wedge allowing the leak to come through this groove.

The seal ran approximately six months and then failed.

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Solution: The lines were steam cleaned and the wrong grade of Viton® was in the seal. Most Viton® compounds will be attacked by steam, caustic or other water based solutions.

The seal was installed correctly, but it leaked immediately. •

Solution: The solid, hard face is usually lapped on only one side. The face had been installed backwards and the rotary unit was running on a non lapped surface.

SEAL COMPONENT DAMAGE IS VISIBLE, BUT WHAT IS THE CAUSE? It looked like a seal part had come loose in the stuffing box, but all of the parts were there. •

Solution: During a previous installation a small spring had been lost when it fell into a drain hole in the bottom of the seal gland. It came loose after a later installation. This is a problem when several people work on the same pump.

The bellows plates were breaking but there was no evidence of corrosion, excessive wear, or vibration. •

Solution: A discharge recirculation line was directing high velocity abrasive particles at the thin metal section of a metal bellow seal.

The inner seal of a dual, rotating "Back to back" seal was showing excessive face wear in a short period of time. •

Solution: The inner seal stationary face was not locked in the bottom of the stuffing box and when the system pressure overcame the barrier fluid pressure, the stationary face was pushed into the inside rotating face. When the pump was stopped the spring pushed the stationary face back to its normal position.

The carbon showed massive damage in a cryogenic (cold) application. •

Solution: The carbon had been lubricated at assembly and the lubricating oil froze in the cryogenic atmosphere.

The bellows plates showed massive wear. •

Solution: The seal was rotating in an abrasive slurry. Metal bellows seals should be designed to rotate the fluid inside of the stuffing box, instead of rotating through the fluid.

OTHER PROBLEMS The pump had been recently overhauled and at start up the pump was reading high amperage, but low flow. •

Solution: One of the wear rings had been left off of the suction side of the impeller and the fluid was recirculating to the pump suction.

The pump made a terrific racket during start up. It produced the proper head, but the capacity was less than anticipated. •

Solution: It was a two speed pump and the second speed had been wired backwards.

In an acid application, a stationary seal showed localized corrosion only on the gland. •

Solution: This was an older pump with a bolted on stuffing box that would slip because the bolts were worn. This caused the shaft to run against the gland causing it to overheat and, in an acid application, the corrosion rate of the acid doubles with an 18° F. (10° C) rise in temperature. It doesn't make any difference if the acid or the part gets hot, the affect is the same.

The dual seal convection tank was running backwards. •

Solution: The seal was not centered in its gland, and as the shaft turned, the close tolerance between the seal and the gland outlet increased the velocity of the liquid enough to drop the pressure and cause the tank to convect backwards.

The pump was converted from packing to a mechanical seal and then started to break shafts. •

Solution: The pump was operating way off of its best efficiency point, causing major shaft deflection. The packing was acting as a bearing and supported the shaft during this deflection.

The product was solidifying in the stuffing box. Steam was being used to heat a jacket around the pump. The header gauge showed adequate pressure. •

Solution: The gage was located too far away from the pump jacket. The line was not insulated and this allowed the steam to experience a pressure drop between the header and the stuffing box heating jacket. The result was that the steam cooled down below the necessary heating temperature. The problem was only visible when the pump was stopped for a period of time.

The nickel base tungsten carbide face was being chemically attacked. •

Solution: A galvanic action occurred between the passivated stainless steel and the active nickel contained in the tungsten carbide face.

Operation practices that cause frequent mechseal and bearing maintenance problems Wouldn't it be wonderful if the plant operation and maintenance departments could work independently? The fact of the matter is that there are three types of problems we encounter with centrifugal pumps and poor operation is one of them. If you are curious, the other two are design problems and poor maintenance practices. Seals and bearings account for over eighty five percent (85%) of premature centrifugal pump failure. In the following paragraphs we will be looking at only those operation practices that can, and will cause premature seal and bearing failure. Design and maintenance practices will be discussed in other papers in this series. When pumps were supplied with jam packing, the soft packing stabilized the shaft to prevent too much deflection. In an effort to save flushing water and to conserve power, many of these same pumps have since been converted to a mechanical seal and the radial stabilization the packing provided has been lost. The bad operating practices include: Running the pump dry will cause over-heating and excessive vibration problems that will shorten seal life. Here are some of the common reasons why a pump is run dry: • • • • •

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Failing to vent the pump prior to start-up. Running the tank dry at the end of the operation cycle. Emptying the tank for steaming or introduction of the next product. Running on the steam that is being used to flush the tank. Starting the standby pump without venting it. Venting a hazardous product can cause a lot of problems with the liquid disposal. Many operators have stopped venting for that reason. Tank vents sometimes freeze during cold weather. This will cause a vacuum in the suction tank, and in some cases could collapse the tank. Sump fluids are often dirty, corrosive or both. The control rods for the float switch will often "gum up" or corrode and give a false reading to the operator. He may think that there is an adequate level, when in fact, the tank is empty.

Dead heading the pump can cause severe shaft deflection as the pump moves off of its best efficiency point (B.E.P.). This translates to excessive heat that will affect both the seal and the bearings as well as causing the seal faces to open, and the possibility of the impeller contacting the volute when the shaft deflects.

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Starting the centrifugal pump with a shut discharge valve is standard practice with many operation departments. The concern is to save power without realizing the damage that is being done to the mechanical seal, impeller, wear rings and bearings. Some pumps are equipped with a recirculation valve that must be opened to lessen the problem, but many times the valve is not opened, or the bypass line is clogged or not of the correct diameter to prevent the excessive head. Another point to remember is that if the bypass line is discharged to the suction side of the pump the increased temperature can cause cavitation. After a system has been blocked out the pump is started with one or more valves not opened. Discharge valves are shut before the pump has been stopped.

Operating off of the best efficiency point (B.E.P.). Changing the flow rate of the liquid causes shaft deflection that can fail the mechanical seal and over-load the bearings. • •

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Starting the pump with the discharge valve closed to save power. The level in the suction tank is changing. Remember that the pump pumps the difference between the discharge and suction heads. If the suction head varies, the pump moves to a different point on its curve. Any upset in the system such as closing, throttling or opening a valve will cause the pump to move to a new point on the curve as the tank fills. Pumping to the bottom of a tank will cause the pump to move to a different point on the curve as the tank fills. Some systems were designed for a low capacity positive displacement pump and have since been converted to a centrifugal design because of a need for higher capacity. Centrifugal pumps must discharge to the top of the tank to prevent this problem. If the discharge piping is restricted because of product build up on the inside walls, the pump will run throttled. This is one of the reasons that it is important to take periodic flow and amperage readings. Increasing the flow will often cause cavitation problems.

Seal environmental controls are necessary to insure long mechanical seal life. It is important that operations understand their function and need because many times we find the controls installed, but not functioning. •

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Cooling-heating jackets should show a differential temperature between the inlet and outlet lines. If the jacket clogs up, this differential will be lost and seal failure will shortly follow. Barrier fluid is circulated between two mechanical seals. There may or may not be a differential temperature depending upon the flow rate. If a convection tank is installed, there should be a temperature differential between the inlet and outlet lines. The line coming out of the top of the seal to the side of the tank should be warmer than the line from the bottom

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of the tank to the bottom of the seals, otherwise the system is running backwards and may fail completely. The level in the tank is also critical. It should be above the tank inlet line or no convection will occur. Some convection tanks are pressurized with a gas of some type. Many original equipment (O.E.M.) seal designs will fail if this differential pressure is lost. Some seal glands (A.P.I. type) are equipped with a quench connection that looks like the seal is leaking water or steam. If there is too much steam pressure on this quench connection, the excessive leakage will get into the bearings causing premature failure. The steam is often used to keep the product warm to prevent it from solidifying, crystallizing, getting too viscous, building a film on the faces etc. Operating people frequently shut off the quench to stop the condensate from leaking. Flushing fluids are used for a variety of purposes, but most of the time they are used to get rid of unwanted solids. The flush can be closely controlled with a flow meter or throttling valve. The amount of flush is determined by the seal design. As an example, those designs that have springs in the product require more flush. It is important to check that the stuffing box has been vented in vertical pumps. The vent should be coming out of the seal gland and not the stuffing box lantern ring connection.

There are some additional things that all operators should know to insure longer rotating equipment life. As an example : •

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Mechanical seals have an 85% or more failure rate that is normally correctable. This is causing unnecessary down time and excessive operating expense. Seals should run until the sacrificial carbon face is worn away, but in more that 85% of the cases the seal fails before this happens. There are five different causes of cavitation. You should know where the best efficiency point (B.E.P.) is on a particular pump, and how far it is safe to operate off the B.E.P. with a mechanical seal installed. You should be aware that washing down the pump area with a water hose will cause premature bearing failure when the water penetrates the bearing case. Learn about the affect of shaft L3/D4 on pump operation. Know how the pumped product affects the life of the mechanical seal and why environmental controls are necessary. If you are not using cartridge seals, adjusting the open impeller for efficiency will shorten the seal life. In most cases the seal will open as the impeller is being adjusted to the volute. Durco pumps are the best example of the exception to this rule. The popular Durco pumps adjust to the back plate causing a compression of the seal faces that can create mechanical seal "over heating" problems.

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Cycling pumps for test will often cause a mechanical seal failure unless an environmental control has been installed to prevent the failure. Mechanical seals should be positioned after the impeller has been adjusted for thermal growth. This is important on any pump that is operated above 200°F (100°C) or you will experience premature seal failure. Some elastomers will be affected by steaming the system. A great deal of caution must be exercised if a flushing fluid such as caustic is going to be circulated through the lines or used to clean a tank. Both the elastomer and some seal faces (reaction bonded silicone carbide is a good example) can be damaged. If the elastomer is attacked, the failure usually occurs within one week of the cleaning procedure. The stuffing box must be vented on all vertical centrifugal pumps or otherwise air will be trapped at the seal faces that can cause premature failure of many seal designs. Most original equipment seal designs cause shaft damage (fretting) necessitating the use of shaft sleeves that weaken the shaft and restrict pump operation to a narrow range at the B.E.P..

Here are a few common misconceptions that cause friction between maintenance and operation departments • • • • • • •

Shutting the pump discharge valve suddenly, will blow the seal open. All ceramics cold shock. High head, low capacity consumes a lot of power. The pump must come into the shop to change a mechanical seal. If you use two hard faces or dual mechanical seals in slurry applications, you will not need flushing water with its corresponding product dilution. If you use metal bellows seals for hot oil applications, you will not need the stuffing box cooling jacket operating. It is O.K. to use an oversized impeller because throttling back will save power.

A few more thoughts on the subject •

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Operators should receive proper schooling on the trouble shooting and maintenance of pumps. In the military and many modern plants, the operator and the maintenance mechanic are often the same person. If the operator knows how the pump works he will have no trouble figuring out the solution to his problem. Too often he is told to keep the flow gage at a certain point, or between two values without understanding what is actually happening with the equipment. If the operator recognizes cavitation he can tell the maintenance department and help them with their trouble shooting. As you wander around the plant look out for painters that paint the springs of outside and double mechanical seals. There is a trend to putting two

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seals in a pump for environmental reasons and the painting of springs is becoming a common problem. If someone is adjusting the impeller make sure he is resetting the seal spring tension at the same time. If the pump is getting hot or making excessive noises, report it immediately. After the failure, it does no good to tell maintenance that it was making noise for two weeks. If you are the floor operator it is common knowledge that taking temperature and pressure readings is very boring, especially on those gages that are located in hot or awkward locations. Avoid the temptation to "radio" these readings. From hot to failure is a very short trip. Maintenance's favorite expression is "there is never time to do it right, but there is always time to fix it." Try to keep this in mind when the pressure is on to get the equipment running again. Do not let cleaning people direct their "wash down" hoses directly at the pump. Water entering the bearings through the lip or grease seals is a major cause of premature bearing failure. Most water wash downs are used to dilute and wash away seal leakage. Stop the leak and you have eliminated the reason for the hose. A great many motor and electrical problems are caused by these same wash down hoses. Cooling a bearing outside diameter will cause it to shrink and the bearing will get hotter as the radial load increases. Keep the water hose and all other forms of cooling off of the bearing casing.

Pump selection practices that cause high seal and bearing maintenance problems.

Purchasing well designed hardware does not bring automatic trouble free performance with it. The very best equipment will cause problems if it was not designed for your particular application. Here are a few of the more common selection problems we find with centrifugal pumps: •

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Buying the same size pump as the one that came out of the application. That's O.K. If the old pump was the correct size, but the odds are that it was too big because of the safety factors that were added at the time of purchase. This will cause the pump to run off of its best efficiency point (B.E.P.) and you will spend a lot of production money for the additional power that is needed to run against a throttled discharge valve or orifice installed in the discharge piping. Buying to a standard, or making a decision based on efficiency, and believing that these two some how relate to quality. Standards were written for packed pumps. When a mechanical seal is being used the shaft L3/D4 number is almost always too large. Efficiency is always gained at the expense of maintenance. Efficiency means tight tolerances and smooth passages that will eliminate reliable double volute designs and keep the maintenance department busy adjusting tight tolerances to maintain the efficiency you paid for. Series and parallel installation problems. We often find pumps installed in parallel, but no one knows it because the second pump was installed at a much later date and no one has bothered to trace the piping. Pumps in parallel require that they have the same diameter impeller and that they run at the same speed, or the larger pump will throttle the smaller one causing it to run off the best efficiency point, deflecting the shaft. The capacity should be looked at if the higher capacity pump might exceed the N.P.S.H. available. When pumps are installed in series the impellers must be the same width and they must run at the same speed or the higher capacity pump will either cavitate because the smaller capacity pump can not feed liquid at the proper capacity, or it will run throttled if it is feeding the smaller pump. In either case the larger of the two pumps will be adversely affected. Purchasing a larger pump because it will be needed in the future. Will raise the operating cost to unacceptable levels (Power = head x capacity) as the pump is run against a throttled discharge valve. This inefficient use of power will translate to a higher heat environment for the seal along with all of the problems associated with shaft deflection. Using a variable speed motor to compensate for a pump curve that is not flat enough. Many boiler feed pumps require a flat curve so that the pump

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can put out varying capacities at a constant boiler pressure (head). We see this same need if we are pumping a varying amount of liquid to a very high constant height. Varying the speed of a pump is similar to changing the diameter of the impeller. If you look at a typical pump curve you will observe that the best efficiency point (B.E.P.) comes down with impeller size to form an angle with the base line (capacity line) of the graph. This means that if you vary the speed of the impeller, the pump always runs off the B.E.P. except in the case where the system curve intersects the pump curve, or in the case of an exponential system curve such as we find in a typical hot or cold water circulating system. Double ended pumps installed in a vertical position to save floor space. Makes seal replacement a nightmare unless you are using split or cartridge designs. Specifying a desired capacity without knowing the true system head. You can't guess with this one. Some one has to make the calculations and "walk the system". The present pump is not a reliable guide because we seldom know where it is pumping on its' curve. Chart recorders installed on both the suction and discharge side of the pump will give a more accurate reading of the present head if they are left on long enough to record the differences in flow. The trouble with this method is that it will also record a false head caused by a throttled valve, an orifice, or any other restriction that might be present in the piping. Requesting too low a required N.P.S.H. will cause you to end up with a different kind of cavitation problem. See another paper in this series for information about "Internal recirculation". Failure to request a "center line design" when pumping temperature exceeds 200°F (100°C) it will cause pipe strain that will translate to wear ring damage and excessive mechanical seal movement. The use of "inline" pumps to save floor space. Many of these designs are "close coupled" with the motor bearings carrying the radial and thrust loads. Because of typical L3/D4 numbers being very high, the wear rings act as "steady bearings" after the pump is converted to a mechanical seal. The pump should have been designed with a separate bearing case and a "C" or "D" frame adapter installed to connect a motor to the bearing case. Thrust bearings being retained by a simple snap ring. Beyond 65% of its rated efficiency most centrifugal pumps thrust towards the pump volute. The thin snap ring has to absorb all of this axial thrust and most of them can not do it very well . The mechanical seal has been installed in a packing stuffing box that is too narrow to allow free seal movement. If a mechanical seal was specified, the pump back plate should have been manufactured with a large diameter seal chamber. In most cases the stuffing box recirculation line should be installed from the bottom of this large seal chamber to the suction side of the pump or a low pressure point in the system. There are some exceptions to this, however:

If you are pumping at or close to vapor point. If the entrained solids have a low specific gravity. If you are using a Duriron pump that adjusts to the back plate. If you are using a double suction pump where the stuffing boxes are at suction pressure. High temperature applications have several special needs: o A jacketed stuffing box that isolates the pumpage from the stuffing box contents by a carbon bushing to retard heat transfer. o A centerline design to compensate for thermal expansion. o A cartridge seal design that allows open impeller adjustment after the pump has come up to operating temperature. o A stainless steel shaft to retard heat transfer to the bearings. o A method of cooling the bearing oil, but never the bearings. o A coupling that will compensate for axial expansion. o o o o

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Why do pump bearings fail? What do we mean by good bearing life? Most of us change the bearings every time we disassemble the equipment to replace the mechanical seal or the packing sleeve. Is this really a sensible thing to do? If you think about it for a minute there is nothing in a bearing to wear out, there are no sacrificial parts. Bearing life is determined by the number of hours it will take for the metal to "fatigue" and that is a function of the load on the bearing, the number of rotations, and the amount of lubrication that the bearing receives. Pump companies predict bearing life measured in years. As an example, the Duriron pump company anticipates a three hundred year life for the radial bearing on their 3 x 2 x 10 pump ( 75 mm. x 50 mm. x 250 mm.) when pumping a liquid with a specific gravity of "one" (fresh water). To understand the term "fatigue" we will conduct an experiment: • •

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Straighten out a standard paper clip. Flex it a little and then let it go. You will notice that it returns to the straightened position. You could repeat this cycle many times (many years actually) without breaking (fatiguing) the metal because you are cycling the metal in its "elastic range" ( it has a memory similar to piece of rubber). Now we will bend (stress) the paper clip a lot further and you will note that it did not return to the straightened position. This time you stressed the metal in its' "plastic range" where it did not have a memory. If you bend the metal back and forth in this plastic range it will crack and break in less than twenty cycles. The metal fatigued more quickly because it "work hardened" and became brittle. The more you stress the metal by flexing it the quicker it will work harden and break. You have just demonstrated that fatigue is a function of stress and cycles. When the bearing is pressed on a rotating shaft the load passes from the inner race( inside ring) through the balls to the bearing outer race (the outside ring). Each ball carries a portion of the stress as the balls roll under the load. It is this stress that will eventually fatigue the metal parts.

When a pump is operating at its best efficiency point (B.E.P.) the only load the bearing has to carry is: • • •

The weight of the rotating assembly. The stress caused by the interference fit on the shaft. Any bearing preload specified by the manufacturer.

The fact is that most bearings become overloaded because of:

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The wrong interference fit between the bearing and the shaft ( the shaft was out of tolerance). Misalignment between the pump and its' driver. Bent shafts. An unbalanced rotating element. Pushing the bearing too far up a tapered sleeve. Operating the pump off of its best efficiency point (B.E.P.). Shaft radial thermal expansion. A futile attempt to cool the bearings by cooling the bearing housing with a water hose or some other similar system. Cooling the outside diameter of a bearing causes it to shrink, increasing the interference and causing additional stress. Cavitation. Water hammer. Axial thrust. The bearing housing is sometimes out of round. Pulley driven designs. Vibration of almost any form. The impeller is located too far away from the bearing. This is a common problem in many mixer/ agitator applications. A bad bearing was supplied. This is becoming more of a problem with the increase in counterfeit parts we are finding in industry.

This overloading will cause heat to be generated, and heat is another common cause of premature bearing failure. Heat will cause the lubricant to: • •

Decrease in viscosity, causing more heat as it loses its ability to support the load. Form a "varnish" residue and then "coke" at the elevated temperature. This "coking" will destroy the ability of the grease or oil to lubricate the bearing. It will also introduce solid particles into the lubricant.

In addition to the heat generated by overloading we get additional heat from: • • •

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The oil level is too high or too low. Too often pumps are aligned but not leveled. The bearing was over greased. The shaft material is conducting heat from the pumpage back to the bearing housing. This is a common problem in heat transfer oil pumps, or any time a metal bellows seal is used in an application and the stuffing box cooling jacket is shut off or inoperative. A loss of barrier fluid between double seals causing a temperature rise that conducts heat back to the bearings.

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A failed cooling jacket in the bearing housing around the stuffing box or built into the seal gland. Grease or lip seal contact on the shaft, right next to the bearings. A failed cooling "quench" in an A.P.I. type seal gland.

A leading bearing manufacturer states that the life of bearing oil is directly related to heat. Non contaminated oil cannot wear out and has a useful life of about thirty years at thirty degrees centigrade (86 F.). They further state that the life of the bearing oil is cut in half for each ten degree centigrade rise (18 F.) in temperature of the oil. This means that oil temperature regulation is critical in any attempt to increase the useful life of anti friction bearings. Probably the major cause of premature bearing failure is the contamination of the bearing lubrication by moisture and solids. As little as 0.002% water in the lubricant can reduce bearing life by 48%. Six percent water can reduce bearing life by 83% percent. There are several methods used by pump companies to keep this water and moisture out of the bearing housing: • •

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A flinger ring to deflect packing or seal leakage away from the bearings. A silly arrangement at best. Keeping the bearing oil hot to prevent the forming of condensation inside the bearing case. A ridiculous system when you consider that bearing life is directly related to heat. The use of "so called" sealed bearings. You can call them any thing you want, but the seals will not seal anything, especially moisture or water. Grease or lip seals that have a useful life of about two thousand hours (84 days at 24 hours per day) and will cut the expensive shaft directly under the seal lip. Double lip seals will cut the shaft in two places. Labyrinth seals that are superior to lip seals but not totally effective because you are still trying to seal with non contacting surfaces that are useless Statically.

The moisture comes from multiple sources: • • • •

Packing leakage flows back to the bearing area. Because of packing leakage a water hose is used to wash down the area. This washing splashes on to the pump bearing case also. Aspiration, moist air enters through the lip or labyrinth seals when the bearing case cools down. A seal quench gland that often has steam, condensate or cooling water leaking out and directed at the radial bearing.

The moisture causes several problems: • • •

Pitting and corrosion of the bearing races and rolling elements that will increase the fatigue of the metal components. Free atomic hydrogen, in the water, appears to cause hydrogen embrittlement of the bearing metal accelerating the fatigue. A water and oil emulsion does not provide a good lubricating film.

We get solids into the lubricant from several sources: •

• • • • •

Metal seal cage wear. This is the part the separates the balls that are held between the bearing races. It is often manufactured from brass or a non metallic material. Abrasive particles leach out of the bearing housing casting. Often solid particles were already contaminating the grease or oil we are using for the lubricant. Solids were introduced into the system during the assembly process because of a lack of cleanliness. Airborne particles penetrate the bearing seals. Particles worn off of the grease or lip seals penetrate into the bearings.

How to keep solids and moisture out of the bearing housing. •

•

•

•

Seal the inside of the bearing housing with epoxy or some other suitable material to stop rusting and to prevent solids from leaching out of the metal case. If you do this be careful about using some of the new high detergent lubricants. They might be powerful enough to remove this protective coating. Replace the grease or labyrinth seals with positive face seals. In the future, you are going to need these seals to prevent hydrocarbon fugitive emissions. Install an expansion chamber outside of the bearing casing to accept the air (approximately 16 oz. or 475 ml. in a typical process pump) that expands as the bearing casing increases in temperature. Without this expansion chamber approximately one atmosphere of pressure will build up in the bearing housing. This is not a problem for a mechanical seal, but during long periods of shut down the pressure could be lost. Clean the oil in the bearing casing by installing a simple oil circulating and filtering system or change the oil frequently.

When do you go from anti-friction ball and roller bearings to hydrodynamic (sleeve) bearings in a centrifugal pump? •

Any time the DN number exceeds 300,000 (Bearing bore times rpm)

•

•

If the standard bearings fail to meet an L10 life of 25.000 hours in continuous operation or 16,000 hours at maximum axial and radial load and rated speed. If the product of the pump horsepower and speed in rpm is 2.7 million or greater.

The past several years have seen a decrease in the quality of the bearings available for rotating equipment. We find prepacked bearing being shipped with too much, or no grease at all. Stabilization temperatures have changed and overall quality has diminished. If you adopt the above suggestions you should not have to be changing your bearings as frequently as you are now.

Troubleshooting the positive displacement rotarypump. No liquid discharge. • • • • • • • • • • • • •

The pump is not primed. Prime it from the outlet side by keeping the outlet air vent open until liquid comes out the vent. The rotating unit is turning in the wrong direction. Valves are closed or there is an obstruction in the inlet or outlet line. Check that the flange gaskets have their center cut out. The end of the inlet pipe is not submerged. You can either increase the length of the inlet pipe into the liquid level or raise the level in the tank. The foot valve is stuck. A strainer or filter is clogged. The net inlet pressure is too low. A bypass valve is open. There is an air leak some where in the inlet line. Air can come in through gaskets or valves above the fluid line. The stuffing box is under negative pressure. Packing is allowing air to get into the system. You should convert the packing to a mechanical seal The pump is worn. The critical clearances have increased. Something is broken. Check the shaft, coupling, internal parts, etc. There is no power to the pump.

The pump is putting out a low capacity. • • • • • • • • • •

The pump's internal clearances have increased. It is time to change some parts. The net inlet pressure is too low; the pump is cavitating. A strainer or filter is partially clogged. The speed is too low. Check the voltage. The tank vent is partially frozen shut. A bypass line is partially open. A relief valve is stuck partially open. The inlet piping is damaged. Something ran over it A corrosion resistant liner has collapsed in the inlet piping. Air is leaking through the packing. You should go to a mechanical seal.

The pump looses its prime after it has been running for a while. • •

The liquid supply is exhausted. Check the tank level; sometimes the float is stuck, giving an incorrect level reading. The liquid velocity has increased dramatically.

• • •

The liquid is vaporizing at the pump inlet. A bypass line is heating the incoming fluid. An air leak has developed in the suction piping.

The pump is using too much power • • • • •

The speed is too high. The liquid viscosity is higher than expected. The discharge pressure is higher than calculated The packing has been over tightened. You should convert to a mechanical seal. A rotating element is binding. Misalignment could be the problem or something is stuck in a close clearance and binding the rotating element.

Excessive noise and vibration. • • • • •

Relief valve chatter. Foundation or anchor bolts have come loose. The pump and driver are misaligned. The piping is not supported properly. The liquid viscosity is too high. The pump is starving. Check the temperature of the incoming liquid. Check to see if the supply tank heater has failed.

Excessive noise or a loss of capacity is frequently caused by cavitation. Here is how the NPSH required was determined initially: With the pump initially operating with a 0 psig. inlet pressure and constant differential pressure, temperature, speed and viscosity; a valve in the inlet line is gradually closed until cavitation noise is clearly audible, there is a sudden drop off in capacity or there is a 5% overall reduction in output flow. Cavitation occurs with: • • •

A loss of suction pressure. An increase in fluid velocity. An increase in inlet temperature.

Here are some common causes of cavitation problems: • •

• • •

A foot valve or any valve in the suction piping is sticking. Something is occasionally plugging up the suction piping. If the pump suction is coming from a river, pond or the ocean, grass is a strong possibility. A loose rag is another common cause. A collapsed pipe liner. A filter or strainer is gradually clogging up.

• • • • • • • •

• •

• • • •

•

The tank vent partially freezes in cold weather. The sun is heating the suction piping, raising the product temperature close to its vapor point. The level in the open suction tank decreases causing vortex problems that allow air into the pump suction. Several pumps in the same sump are running, decreasing the level too much. The suction tank float is stuck. It will sometimes show a higher level than you really have. A discharge recirculation line, piped to the pump suction, opens and heats the incoming liquid. Sometimes the suction lift is too high. The increase in pipe friction will reduce the suction head. The vapor pressure of the product is very close to atmospheric pressure. The pump cavitates every time it rains because of a drop in atmospheric pressure. The tank is being heated to de-aerate the fluid. Sometimes it is being heated too much. The process fluid specific gravity is changing. This can happen with a change in product operating temperature or if a cleaner or solvent is being flushed through the lines. The source tank is changing from a positive pressure to a vacuum due to the process. A packed valve in the suction piping is at a negative pressure and air is leaking in through the packing. The tank is being pumped dry. The inlet piping has been moved or altered in some way. Has a foot valve, strainer, elbow, or some other type of hardware been installed in the suction piping? Has a layer of hard water calcium or some other type of solid formed on the inside of the suction piping reducing its inside diameter over some period of time?

You are experiencing rapid pump wear. •

• • • • • •

There are abrasives in the liquid you are pumping causing erosion problems. You may have to go to a larger pump running at a slower speed. There is some corrosion in one or more of the pump elements. There is a lack of lubrication. You have a severe pipe strain problem. It could have been caused by thermal growth of the hardware. Too much misalignment. The pump is running dry. When all else fails the best way to reduce NPSH required is to select a larger pump and run it at a slower speed.

Troubleshooting premature bearing failure As discussed in a another technical paper on this web site, bearings have no wearable surfaces, they are instead designed to fatigue after many hours of service. In a properly operating bearing the race ways and rolling elements will become dull in appearance. This dullness is not an indication of wear and has no affect on the life of the bearing. These dull surfaces form the visible paths that I will be referring to in the following paragraphs, so their appearance and location is important in analyzing any type of bearing failure. When we install a bearing into a piece of rotating equipment the general rule is to have the interference fit on the race that is rotating and, therefore, carrying the load. Almost all centrifugal pumps, motors, and a high percentage of other types of rotating equipment have the bearings installed with the inner race an interference fit and rotating with the shaft . The outer race remains stationary or in a fixed position. In the following paragraphs I will be discussing various load conditions and the resultant appearance of the raceways and rotating elements in this type of an installation: The radial load is rotating with the shaft, This is caused by an unbalanced rotating assembly or a bent shaft. •

•

The inner ring appearance. The load acts all of the time at the same place in the race way. Here the path pattern is at its widest, tapering off at the ends. If the load is only radial, the pattern will be in the center of the race way and will extend around slightly less the half the race way circumference. The outer ring appearance. The path will extend around the entire race way. It will be uniform in width and if the load is only radial, it will be in the center of the race way.

The radial load is unidirectional. This is what we would expect to find with a properly operating piece of equipment. If the equipment is operating off of its best efficiency point, is misaligned, or if there is excessive pipe strain the pattern will be the same, only more pronounced. •

The inner ring appearance. The path will be in the center of the race way, uniform in width and visible around the entire circumference of the race way.

•

The outer ring appearance. The pattern will be widest at the load point and tapering towards the ends. If the fit and clearances are normal the pattern will extend around to slightly less than one half of the raceway. It will be located in the center of the race way, if the load is only radial.

The radial load is multidirectional . Cavitation, too tight an interference fit, preloading, or cooling a bearing outside diameter are all common causes of this problem. • •

The inner ring appearance. All around the race way, widest where the load was the greatest. The outer ring appearance. All around the race way, widest where the load was the greatest.

The axial load is unidirectional. This is the normal condition of all end suction centrifugal pumps. •

Both the inner and outer rings. The pattern will extend around both raceways and is displaced axially from the center. A centrifugal pump thrusts towards the thrust bearing until it reaches 65% of its efficiency and then it thrusts towards the volute or wet end during normal operation.

An oval compression of the outer ring. Caused by an out of round housing. • •

The inner ring appearance. The path extends around the entire ring and is uniform in width. The outer ring appearance. Two wider paths where the ring was distorted to the oval shape.

The inner ring was misaligned. Normally happens during the installation process. • •

The inner ring appearance. The pattern extends around the entire ring and is uniform in appearance. The outer ring appearance. The ball path will be oval, extending from one side of the race way to the other, and wider in two diametrically opposite sections.

Now that we know what some typical wear paths look like, we will inspect the only two things that are visible to the trained trouble shooter. • •

Evidence of rubbing. Evidence of corrosion and damage.

Look for damage caused by solid particles. These particles will be rolled into the race ways and can:

• •

Score, or cause small indentations in the precision races and rolling elements. Interfere with the transfer of heat within the tight tolerances, causing discoloration, thermal expansion, seizing etc.… The particles come from: o Varnish and "coke" that forms where the lubricant overheated. o Parts of the ball cage that have broken loose due to a lack of lubrication. Brass cage parts will turn the lubricant green. o Pieces from a failed grease or lip seal. o A contaminated lubricant. o Lack of cleanliness during the installation process. The bearings are being installd next to the area where the mechanic is grinding a new edge on his lawnmower blade. o The bearing lubricant could have been over heated during the installation process. o Rust coming off the inside of the casting. o Silica leaching out of the casting o Particles of material flaking off of the protective coating put on the inside of the housing to prevent rust. o Airborne - through the bearing seals or housing vent.

Look for lack of lubrication that can eventually cause the bearing to seize: • •

•

• •

•

•

You will see" mirror like" surfaces on the metal parts that look like the piece was "lapped". The metal will become discolored and soften as it anneals. Annealing can occur any time the temperature exceeds 300°F (150°C): o Straw yellow 600° F. 315° C. o Brown 700° F. 370° C. o Blue 800° F. 425° C. o Black 900° F. 480° C. If a pre- lubricated bearing was heated by immersing it in a hot oil bath (200°F or 100°C), the hot oil will wash out the grease and leave the bearing with little to no lubrication. o Many pre-lubricated bearings actually have no lubricant at all installed. Check yours to be sure. Bearing quality is a serious maintenance problem. A clogged oil level gauge can give a false reading of lubrication level. If the bearing case has no expansion chamber installed, a build up of internal pressure, as the bearing case comes up to temperature, can blow out of the seals. At shut down, moisture laden air will return to the case through the same seals. A poorly designed labyrinth seal can pump hot oil out of the bearing case. The lubricating oil level should be at the middle of the lower bearing ball when the pump is at rest. Be sure the pump has been leveled to insure the correct lubrication height.

Look for smearing of the metal . When two non lubricated surfaces slide against each other, under load, the material transfers from one surface to the other. • • •

The metal melts and then re-hardens, causing localized stress that can produce cracks in the metal.. The load was too light for the speed. Centrifugal force threw the balls out. The outer race will smear on the outside diameter if it slides during operation due to an improper "slip fit". This slipping can also cause "fretting corrosion" as the protective oxide film is worn away from the metal surface.

Look for evidence of static vibration. You will see indents in the raceway that could be either shiny or rusted in the bottom. The frequency of the vibration has no affect, but greater energy causes greater damage. Roller bearings are more susceptible to this type of damage because the balls, in a ball bearing, can roll in many directions. Rollers, how ever, can roll in only one direction. Movement in the other directions takes the form of "sliding". • • •

The pump was located too close to another piece of equipment that was vibrating. This can be a big problem during storage. The shaft was not locked during shipment. In addition to vibration, equally spaced indents can be caused by: o An induction heater was used during assembly, causing "false Duriron". o The bearing was installed by pressing on the wrong race. o The bearing was driven too far up a tapered shaft.

Look for electric current damage. It will show up on both the races and the rolling element. The bottom of the depression will be dark in color. •

The pump was used as a ground for a welding rig.

Look for flaking or spalling of the metal race way. Since there is nothing in a bearing to wear out, flaking or spalling is a sign of normal fatigue. Overloading however, can cause premature fatigue. Look for the following causes of bearing overloading: • • • • • • • • •

The bearing housing is out of round. The shaft is over size. The bearing was driven up too far on a tapered shaft. Misalignment between the pump and its driver. The rotating assembly is out of balance. The shaft is bent. The pump is operating too far off of its best efficiency point (B.E.P.). Pipe strain. Water hammer in the lines.

• • • • • • • •

Cavitation. The bearing had a quality problem to start with. Shaft thermal expansion. The bearing housing is being cooled, causing the outer race to shrink, increasing the load. Excessive axial thrust. Pulley driven design. Hydrogen embrittlement of the metal caused by moisture entering the lubricant. Pumping a high specific gravity fluid such as sulfuric acid can almost double the radial load.

Overloading is often accompanied by a change in appearance of the lubricant. You will see varnish or coke as the lubricant is subjected to this high heat. In addition to overloading there are additional sources of heat that can destroy the lubricant : • • • • • • • • • • • • •

Soak temperatures through the shaft. This can be a big problem in either hot oil or hot water applications. Over lubrication of the bearing. Plugged oil return holes. Constant oil cups at the wrong level. Insufficient clearance in labyrinth seals. The oil gage breather hole is blocked and showing the wrong lubrication level. Bent lock washer prongs can rub against the bearing race. Grease or lip seals are too tight on the shaft. The pump stuffing box cooling jacket was shut off and drained when the metal bellows seal was installed in a high temperature oil application. Someone is cooling the power end case causing the bearing outer race to shrink. Friction with the seal cage. Sliding friction caused by small changes in the shaft speed. Inertia keeps the balls moving as the shaft slows down. The stuffing box packing has been over tightened.

Look for cracks in the metal. • • •

Mishandling. The bearing was driven too far up a tapered shaft. Any type of flaking or smearing can cause a fracture notch that will lead to cracking.

Look for signs of corrosion.

•

•

Moisture is in the lubricant. It came from: o Packing or seal leakage. o A water hose being used to wash down the area. o Normal aspiration as the pump cooled down, and the moisture ladened atmosphere entered the bearing case. o Steam or water from a seal quench gland. This is a common problem with the A.P.I. gland that is commonly used in oil refineries. Regardless of the protective coating put on the bearing races, (cadmium, chromium, zinc, etc.) the rolling elements are almost always fabricated from 52100 bearing steel, and it rusts.

The major bearing companies do a good job of providing the literature and photographs that you need to do effective "comparison troubleshooting". Check with your bearing supplier for the availability of this information.

The pump works for a while and then loses suction part:4 This is the fourth paper in a four part series about pump troubleshooting. As mentioned in paper number eleven, volume number ten, there are a couple of things you must keep in mind when troubleshooting centrifugal pump problems: •

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• •

•

The centrifugal pump always pumps the difference between the suction and discharge heads. If the suction head increases, the pump head will decrease to meet the system requirements. If the suction head decreases the pump head will increase to meet the system requirements. A centrifugal pump always pumps a combination of head and capacity. These two numbers multiplied together must remain a constant. In other words, if the head increases the capacity must decrease. Likewise if the head decreases, the capacity must increase. The pump will pump where the pump curve intersects the system curve. If the pump is not meeting the system curve requirements the problem could be in the pump, the suction side including the piping and source tank, or somewhere in the discharge system. Most pumps are oversized because of safety factors that were added at the time the pump was sized. This means that throttling is a normal condition in most plants, causing the pump to run on the left hand side of its curve.

Cavitation is a main cause of losing pump suction, but remember that there are several different types of cavitation: • • •

• •

Vaporization of the liquid within the pump caused by a loss of suction head or an increase in suction temperature. The "vane passing syndrome" caused by too small an impeller to cutwater clearance. Too high a suction specific speed number will cause internal recirculation problems resulting in cavitation. The suction specific speed number is obtained from a formula that can be found in paper 9-12 of this series. Air ingestion on the suction side of the pump allows air and bubbles into the suction of the pump. Turbulence of the fluid that releases entrained gases into the suction piping.

Each of these cavitations has been addressed in other papers in this Technical Series. In this paper we will be looking at only the intermittent loss of suction fluid. You will be looking at several possibilities: •

A recurring restriction in the suction piping that may or may not be causing a cavitation problem within the pump.

• •

Intermittent cavitation problems as opposed to a design or operation problem that causes a constant cavitation condition. A repetitive need for an increase in the pump's capacity.

Now we will take a look at each of these possibilities in detail: A re-occurring restriction in the suction piping that may or may not be causing a cavitation problem within the pump. • •

• • • • •

A foot valve or any valve in the suction piping is sticking. Something is occasionally plugging up the suction piping. If the pump suction is coming from a river, pond or the ocean, grass is a strong possibility. A loose rag is another common cause. A collapsed pipe liner will restrict the piping at higher velocities. The suction is being throttled to prevent heating of the process fluid. This can happen with some volatile fuel applications. A filter or strainer is gradually clogging up. Air is being introduced into the suction side of the pump to reduce the capacity. This is sometimes done with low specific gravity fluids to avoid throttling the discharge that might overheat and flash the product.

Intermittent incidents that cause cavitation problems • • •

• • • • • •

•

The tank vent partially freezes in cold weather. The sun is heating the suction piping, raising the product temperature close to its vapor point. The level in the suction tank increases, decreasing the differential head across the pump. This will increase the pump capacity until the level in the tank drops. The level in the open suction tank decreases causing vortex problems that allow air into the pump suction. Several pumps in the same sump are running, decreasing the level too much. The suction tank float is stuck. It will sometimes show a higher level than you really have. A discharge recirculation line, piped to the pump suction, opens and heats the incoming liquid. Sometimes the suction lift is too high. The increase in pipe friction will reduce the suction head. The vapor pressure of the product is very close to atmospheric pressure. The pump cavitates every time it rains because of a drop in atmospheric pressure. The tank is being heated to de-aerate the fluid. Sometimes it is being heated too much.

•

• • • •

The process fluid specific gravity is changing. This can happen with a change in product operating temperature, or if a cleaner or solvent is being flushed through the lines. A booster pump is malfunctioning or leaking excessively. The source tank is changing from a positive pressure to a vacuum due to the process. A packed valve in the suction piping is at a negative pressure and air is leaking in through the packing. The tank is being pumped dry.

A repetitive need for an increase in the pumps capacity. • • • •

A bypass line, or relief valve opens, decreasing the discharge resistance, increasing the capacity. A break or leak in the line down stream of the pump will increase the capacity of the pump as the head drops. The pump is supplying many sources and too many valves are open at one time. The pump discharge is being directed to several different tank farm locations. The changing piping resistance is changing the pump's head and capacity.

The centrifugal pump is drawing too much amperage part:3 This is the third paper in a four part series about pump troubleshooting. As mentioned in paper number ten, volume number ten, there are a couple of things you must keep in mind when troubleshooting centrifugal pump problems: •

•

• •

•

The centrifugal pump always pumps the difference between the suction and discharge heads. If the suction head increases, the pump head will decrease to meet the system requirements. If the suction head decreases the pump head will increase to meet the system requirements. A centrifugal pump always pumps a combination of head and capacity. These two numbers multiplied together must remain a constant. In other words, if the head increases the capacity must decrease. Likewise if the head decreases, the capacity must increase. The pump will pump where the pump curve intersects the system curve. If the pump is not meeting the system curve requirements the problem could be in the pump, the suction side including the piping and source tank, or somewhere in the discharge system. Most pumps are oversized because of safety factors that were added at the time the pump was selected. This means that throttling is a normal condition in most plants, causing the pump to run on the left hand side of its curve.

The increased amperage can be caused by a pump that is too large for the application. • • • • •

A large pump was specified in anticipation of future needs. The pump was sized for the maximum operating condition, but does not run anywhere near that point most of the time. The capacity requirement has been lowered and the pump is being throttled rather than cut back the impeller diameter. The pump was oversized because of safety factors that were added at the time the pump was sized. Increasing the speed of the pump causes a dramatic change in the amperage required. The amperage changes by the cube of the change in speed or impeller diameter. If you double the speed of a pump you will need eight times the amperage.

The increased amperage can be caused by a change in the product. • •

The motor was sized for a low specific gravity fluid, but the lines are being flushed or tested with water. The specific gravity of the fluid has increased for some reason.

•

•

The viscosity of the liquid is increasing with a change in temperature. Some viscosities increase with a lower temperature, some with a higher temperature. The viscosity of a liquid can increase with agitation. That is how cream becomes butter.

The increased amperage is caused by two part rubbing together as a result of shaft displacement. Here are some common causes of shaft displacement: • • • • • • • • • •

Pipe strain Misalignment between the pump and driver. A bent shaft. The rotating assembly is not dynamically balanced. Cavitation. Water hammer. Operating off the BEP. Thermal growth. Pulley driven pumps. Different types of vibration including harmonic, slipstick, induced, etc...

There are many parts that can come into contact when the shaft displaces. • • • • • • •

The impeller can contact the pump volute or back plate. This can also happen with an improper impeller adjustment or thermal growth. The end of the stuffing box can be hit by the shaft or sleeve. There is often a close fitting bushing installed in this location. The outside diameter of the rotating mechanical seal and the inside of the stuffing box. A gasket or fitting protruding into the stuffing box that rubs against the mechanical seal. The rotating shaft and the stationary seal face. The shaft and the API gland disaster bushing. The closed impeller wear rings are a common source of rubbing.

The increased amperage can be caused by an increase in bearing loading. • • •

Check the shaft and housing tolerances along with the installation method. Cooling a bearing outside diameter causes it to shrink and over compress. The wrong lubrication level. There is too much lubricant in the bearing

The starting procedure could be the problem. • •

The radial flow pump is being started with the discharge valve open. Radial flow pumps use the most horsepower at high capacity. The axial flow pump is being started with the discharge valve shut. Axial flow pumps use the most horsepower at high head.

Check to see if there is too much axial thrust. • •

•

See if the impeller balance holes are clogged. If there is an elbow too close to the suction of a double ended pump, and the piping is running parallel with the shaft, The change in velocity of the incoming fluid will cause axial thrust. Converting packing to a mechanical seal can increase the axial loading on the bearing

Here are a few more reasons why you might be using too much amperage. • • • • •

The stuffing box packing has been tightened too much. An unbalanced mechanical seal is being used in a high pressure application. There is too much face load The impeller has been installed backwards. The shaft is running in the wrong direction. The open impeller needs adjusting. You have too much clearance between the impeller and the volute, or back plate, depending upon the pump design.

The pump is not producing enough capacity to satisfy the application part:2 This is the second paper in a four part series about pump troubleshooting. Let me begin by pointing out that there are a couple of things you must keep in mind when troubleshooting centrifugal pump problems: •

•

• •

•

The centrifugal pump always pumps the difference between the suction and discharge heads. If the suction head increases, the pump head will decrease to meet the system requirements. If the suction head decreases the pump head will increase to meet the system requirements. A centrifugal pump always pumps a combination of head and capacity. These two numbers multiplied together must remain a constant. In other words, if the head increases the capacity must decrease. Likewise if the head decreases, the capacity must increase. The pump will pump where the pump curve intersects the system curve. If the pump is not meeting the system curve requirements the problem could be in the pump, the suction side including the piping and source tank, or somewhere in the discharge system. Most pumps are oversized because of safety factors that were added at the time the pump was sized. This means that throttling is a normal condition in most plants, causing the pump to run on the left hand side of its curve.

THE PROBLEM IS IN THE PUMP ITS SELF: • • • • • • • • • • • • • • • • •

The impeller diameter is too small The impeller width is too narrow The impeller speed is too slow. Check the voltage and frequency The impeller is damaged. The impeller is clogged. The open impeller clearance is too large. The impeller to cutwater clearance is too large. The impeller specific speed number is too low. The impeller has been installed backwards The shaft is running backwards. The wear ring clearance is too large. A wear ring is missing. The second stage of a two stage pump is wired backwards. A bubble is trapped in the eye of the impeller. A low suction tank level is increasing the differential pressure across the pump decreasing its capacity. Air is coming into the pump suction through the packing. Air is coming into the pump suction through an unbalanced mechanical seal.

• • • •

The pump was not primed prior to star up. You may need a concentric casing rather than the volute design. You are using a variable speed motor trying to produce a flat curve. Remember that both the head and capacity change with speed. The pump is the wrong size. Someone gave the pump distributor a wrong system curve

THE PROBLEM IS ON THE SUCTION SIDE OF THE PUMP • • • • • • • • • • • • • • • • • • • • • • • • • • •

There is too much piping between the pump suction and the source tank. There is an elbow too close to the pump suction. A filter or strainer is clogged. Intermittent plugging of the suction inlet. Loose rags can do this. A foot valve is stuck The tank float is stuck. Showing a higher tank level that does not exist. The tank vent is partially shut or frozen. A globe valve has been substituted for a gate valve. A check valve is stuck partially closed Solids have built up on the piping walls. A liner has broken away from the piping wall and has collapsed in the piping. The piping was collapsed by a heavy object that hit the outside of the piping. A foreign object is stuck in the piping It was left there when the piping was repaired. A small clam cleared the suction screen, but has now grown large on the pump side of the screen. The sun is heating the inlet piping. It should be insulated to prevent this problem. Piping was added on the inlet side of the pump to compensate for a piece of equipment that was installed in the shop. A reducer has been installed upside down. A discharge recirculation line is heating the incoming fluid. The pump capacity is too high for the tank volume. Multiple pump inlets are too close together. The suction lift is too high. There is not enough NPSH available for the fluid you are pumping. Maybe you can use an inducer to increase the suction pressure. Air is coming into the system through valves above the water line or gaskets in the piping. Air is being pumped into the suction piping to reduce cavitation problems Fluid returning to the sump is being aerated by too far a free fall. The fluid is vortexing at the pump inlet because the sump level is too low. The tank is being heated to deaerate the fluid, but it is heating the fluid up too much.

• • •

•

Two pumps are connected in series. The first pump is not sending enough capacity to the second pump. The operating temperature of the pumped fluid has increased. The vapor pressure of the fluid is too close to atmospheric pressure. When it rains the drop in atmospheric pressure causes the inlet fluid to vaporize. The suction is being throttled to prevent the heating of the process fluid.

PROBLEMS ON THE DISCHARGE SIDE OF THE PUMP INCLUDING THE PIPING • • • • • • • • • • • • •

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Extra piping has been added to the system to accommodate extra storage capacity. A bypass line has been installed in the pump discharge. Piping or fittings have been added to the discharge side of the pump. An orifice has been installed in the discharge piping to reduce the capacity or produce a false head. A gate valve has been substituted for a globe valve in the discharge piping. A check valve is stuck partially closed. An orifice has been installed into the piping to restrict flow. The piping was collapsed by a heavy object that hit the outside of the piping. The discharge valve is throttled too much. There is a restriction in the discharge piping. Extra pumps have been installed into the existing piping They are connected in parallel, but are not producing the same head. Two pumps are in parallel. The larger one is shutting the check valve of the smaller pump. Two pumps are in connected in series. The first pump does not have enough capacity for the second pump. They should be running at the same speed with the same width impeller The pump discharge is connected to the bottom of the tank. The head is increasing and the capacity is decreasing as the tank fills. The pump is acting as an accumulator&emdash;coming on when the tank level drops. The head is too high when the tank fills.

The pump is not producing enough head to satisfy the application This is the first paper in a four part series about pump troubleshooting. Let me begin by pointing out that there are a couple of things you must keep in mind when troubleshooting centrifugal pump problems: •

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The centrifugal pump always pumps the difference between the suction and discharge heads. If the suction head increases, the pump head will decrease to meet the system requirements. If the suction head decreases the pump head will increase to meet the system requirements. A centrifugal pump always pumps a combination of head and capacity. These two numbers multiplied together must remain a constant. In other words, if the head increases the capacity must decrease. Likewise if the head decreases, the capacity must increase. The pump will pump where the pump curve intersects the system curve. If the pump is not meeting the system curve requirements the problem could be in the pump, the suction side including the piping and source tank, or somewhere in the discharge system. Most pumps are oversized because of safety factors that were added at the time the pump was sized. This means that throttling is a normal condition in most plants, causing the pump to run on the left hand side of its curve.

THE PROBLEM COULD BE IN THE PUMP ITS SELF •

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The impeller diameter is too small. o The impeller is running at too slow a speed o You are running an induction motor. Their speed is different than synchronous motors. It's always slower. The pump curve was created using a variable frequency motor that ran at a constant speed. Put a tachometer on your motor to see its actual speed. o Your pulley driven pump is running on the wrong pulley diameter. o A variable frequency motor is running at the wrong speed. o Check the speed of the driver if the pump is driven by something other than an electric motor. There is something physically wrong with the motor. Check the bearings etc. Check the voltage of the electric motor. It may be too low. The impeller is damaged. The damage could be caused by excessive wear, erosion, corrosion or some type of physical damage. o Physical damage often occurs during the assembly process when the impeller is driven on or off the shaft with a wooden block and a

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mallet. Many impeller designs do not have a nut cast into the impeller hub to ease removal. o Erosion occurs when solids enter the eye of the impeller. The solids can chip off pieces of the ceramic that are passivating the impeller, causing localized corrosion. o Damage can occur if the impeller to volute, or back plate clearance is too small and the shaft experiences some type of deflection. The original clearance could have diminished with thermal growth of the shaft. Keep in mind that some open impellers adjust to the volute (Goulds) while other designs adjust to the back plate (Duriron). In an ANSI and similar design centrifugal pumps, the normal thrust towards the volute has bent the snap ring designed for bearing retention. This can allow the rotating impeller to hit the stationary volute. Here are some examples of shaft displacement: o Operating the pump too far off the BEP. o Pulley driven applications. o Pipe strain. o Misalignment between the pump and driver. o The shaft could be bent. o The rotating assembly was probably not dynamically balanced. The impeller is clogged. This is a major problem with closed impellers. With the exception of finished product, most of what you will be pumping contains entrained solids. Remember also that some products can solidify, or they can crystallize with a change in fluid temperature or pressure. Impeller balance holes have been drilled between the eye and the wear rings of a closed impeller. The reverse flow is interfering with the product entering the impeller eye. A discharge recirculation line should have been used in place of the balance holes to reduce the axial thrust. The double volute casting is clogged with solids or solids have built up on the surface of the casting. The open impeller to volute clearance is too large. 0.017" (0,5 mm) is typical. This excessive clearance will cause internal recirculation problems. A bad installation, thermal growth, or normal impeller wear could be the cause. o A large impeller to cutwater clearance can cause a problem called discharge recirculation. Wear is a common symptom of this condition. If the impeller is positioned too close to the cutwater you could have cavitation problems that will interfere with the head. The impeller specific speed number is too high. Lower specific speed numbered impellers are used to build higher heads. An impeller inducer was left off at the time of assembly. Inducers are almost always needed with high specific speed impellers. Leaving off the inducer can cause cavitation problems that will interfere with the head. The impeller is loose on the shaft. The impeller is running backwards

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The shaft is running backwards because of a wiring problem. The pump is running backwards because the discharge check valve is not holding and system pressure is causing the reverse rotation. This is a common problem with pumps installed in a parallel configuration. Check valves are notoriously unreliable. The impeller has been installed backwards. This can happen with closed impellers on double ended pumps The second stage of a two stage pump is wired backwards. The pump reverses when the second stage kicks in. You should have heard a loud noise when this happened. The wear ring clearance is too large. o This is a common problem if the shaft L3/D4 number is greater than 60 (2 in the metric system). o You should replace the rings when the original clearance doubles. Needless to say this can only be determined by inspection. If you are pumping a product at 200°F (100°C) or more you should use a centerline design volute to prevent excessive wear ring wear as the volute grows from the base straight up, engaging the wear rings. A wear ring is missing. It was probably left off during the installation process. A high suction tank level is reducing the differential pressure across the pump increasing its capacity. The pump pumps the difference between the suction and discharge heads. A bubble is trapped in the eye of the impeller. The eye is the lowest pressure area. When this bubble forms it shuts off all liquid coming into the pump suction. This could cause the pump to lose its prime. You cannot vent a running pump because centrifugal force will throw the liquid out the vent leaving the air trapped inside. Air is coming directly into the pump. This happens with a negative pressure at the suction side. Negative suction happens when the pump is lifting liquid, pumping from a condenser hot well etc. o Air is coming into the stuffing box through the pump packing. o Air is coming into the stuffing box through an unbalanced mechanical seal. As the carbon face wears the spring load holding the faces together diminishes. o If you are using mechanical seals in vacuum service, they should be of the O-ring design. Unlike other designs, O-rings are the only shape that seals both pressure and vacuum. o The pump was not primed prior to start up. With the exception of the self priming version, centrifugal pumps must be full of liquid at start up. o Air can enter the stuffing box if the gasket between the two halves of a double ended pump is defective or does not extend to the stuffing box face. Any small gaps between the face of the stuffing box and the split at the side of the stuffing box will allow either air in, or product out.

Air is coming into the suction side of the pump through a pin hole in the casing. o Air is entering the stuffing box between the sleeve and the shaft. This happens if you convert a double ended pump from packing to a mechanical seal and fail to install a gasket or o-ring between the impeller hub and the sleeve. The open impeller was adjusted backwards and now the close fitting "pump out vanes" are creating a vacuum in the stuffing box. You need a volute casing instead of a concentric casing. Volute casings are much better for producing head. You have the wrong size pump. It cannot meet the system curve requirements: The pump was not selected to meet the system curve requirements because no system curve was given to the pump supplier. At replacement time the same size pump was purchased because no one had calculated losses in the system. The pump was sized from a piping diagram that was thirty five years old. There have been numerous piping changes and additions since the original layout. In many instances additional pumps have been installed and this pump is running in parallel with them, but nobody knows it. o

• • • • • •

THE PROBLEM IS ON THE SUCTION SIDE OF THE PUMP. THE PUMP COULD BE CAVITATING. •

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Air is entering the suction piping at some point. o Air is being pumped into the suction piping to reduce cavitation problems o Fluid returning to the sump is being aerated by too far a free fall. The return line should terminate below the liquid level. o The fluid is vortexing at the pump inlet because the sump level is too low and the pump capacity is too high. o Air is coming into the system through valves above the water line or gaskets in the piping flanges. o The liquid source is being pumped dry. If this is a problem in your application you might want to consider a self priming pump in the future. The vapor pressure of the fluid is too close to atmospheric pressure. When it rains the drop in atmospheric pressure causes the inlet fluid to vaporize. There is a problem with the piping layout. It is reducing the head on the suction side of the pump. o There is too much piping between the pump suction and the source tank. You may need a booster pump or an inducer. The higher the pump speed the bigger the problem. o There is an elbow too close to the pump suction. There should be at least ten diameters of pipe between the elbow and the pump

• • • • • • • • •

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•

suction. Suction piping should never run parallel with the pump shaft in a double ended pump installation. This can cause unnecessary shaft thrusting. o A piece of pipe of reduced diameter has been installed in the suction piping. o Piping was added on the inlet side of the pump to by-pass a piece of equipment that was installed on the floor. o A piping to pump reducer has been installed upside down causing an air pocket. Concentric reducers can cause the same problem.. o Multiple pump inlets are too close together. The pump inlet is too close to the tank floor. The suction lift is too high. A gasket with too small an inside diameter has been installed in the suction piping restricting the liquid flow. A gasket in the suction piping is not centered and is protruding into the product stream. A globe valve has been substituted for a gate valve in the suction piping. The loss of head in a globe valve is many times that of a gate valve. Two pumps are connected in series. The first pump is not sending enough capacity to the second pump. The piping inlet is clogged. A filter or strainer is clogged or covered with something. Intermittent plugging of the suction inlet. o Loose rags can do this. o If the suction is from a pond, river, or the sea, grass can be pulled into the suction inlet. A foot valve is stuck. A check valve is stuck partially closed The foot valve is too small. A small clam or marine animal cleared the suction screen, but has now grown large on the pump side of the screen. The suction piping diameter has been reduced. o The suction piping collapsed when a heavy object either hit or ran over the piping. o Solids have built up on the piping walls. Hard water is a good example of this problem o A liner has broken away from the piping wall and has collapsed in the piping. Look for corrosion in the piping caused by a hole in the liner. o A foreign object is stuck in the piping It was left there when the piping was repaired. o The suction is being throttled to prevent the heating of the process fluid. This is a common operating procedure with fuel pumps where discharge throttling could cause a fire or explosion. The pump inlet temperature is too high.

The tank is being heated to deaerate the fluid, but it is heating the fluid up too much. Look for this problem in boiler feed pump applications. o The sun is heating the inlet piping. The piping should be insulated to prevent this problem. o The operating temperature of the pumped fluid has been increased to accommodate the process requirements. o A discharge recirculation line is heating the incoming fluid. You should direct this line to a reservoir rather than the pump suction. o Steam or some other hot cleaner is being circulated through the lines. The problem is in the tank connected to the suction of the pump. o The pump capacity is too high for the tank volume. o The tank float is stuck, showing a higher tank level that does not exist. o The tank vent is partially shut or frozen, lowering the suction pressure. o There is not enough NPSH available for the fluid you are pumping. Maybe you can use an inducer or booster pump to increase the suction pressure. o A high suction tank level is reducing the differential pressure across the pump, increasing its capacity and lowering the head. o

•

PROBLEMS ON THE DISCHARGE SIDE OF THE PUMP INCLUDING THE PIPING •

• • •

• • • •

Two pumps are in connected in series. The first pump does not have enough capacity for the second pump. They should be running at the same speed with the same width impeller. The pump discharge is connected to the bottom of the tank. The head is low until the level in the tank increases. Units in the discharge piping should not normally be shut off, they should be by-passed to prevent too much of a change in the pump's capacity. If too many units are being by-passed in the discharge system the head will decrease as the capacity increases. This can happen if an extra storage tank farm is being by-passed because the storage capacity is no longer needed. A bypass line has been installed in the pump discharge increasing the capacity and lowering the head. Piping or fittings have been removed from the discharge side of the pump reducing piping resistance. Connections have been installed in the discharge piping that have increased the demand that increases capacity. The pump is acting as an accumulator, coming on when the tank level drops. The head will be low until the accumulator is recharged.

•

•

Consider the possibility of a siphon affect in the discharge piping. This will occur if the pump discharge piping is entering into the top of a tank and discharging at a lower level The pump must build enough head initially to take advantage of the siphoning action. A discharge valve (manual or automatic) is opened too much.

OCCUPATIONAL MULTITASKING TRAINING FOR OPERATION TECHNICIAN

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EVALUATION KEY 1= Excellent, Outstanding Competence 2=Satisfactory, Acceptable Competence 3= Marginal, Needs Additional Training 4=Unsatisfactory, Needs Extensive Training

e e m m Ti i n tT g tio en e in l p p m g e t ip n om ke es ni qu C l e n T E s a ia t ce Pl ia er ou ls, n s t o es H at y it c a o l T es M a T a m n r y T f n f t e m l o t o A fo e e ra rit e b af er st se s O o y P S W J U U S

FA CT O RS

Name:.......................................................... Staff No:...................................................... Dates: Commence:.................. Completed:................

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1. BASIC WORKSHOP PRACTICES Good Workshop Floor Housekeeping: Able to safely use Bench Vices : Able to safely use and maintain Drill press: Able to safely use Bench Grinder: Able to safely Dress worn-out Grinding Wheel: Able to safely Cut Gasket using cutting tools: Able to use standard workshop Measuring tools: Able to use the different type of Hammers,Mallet etc Able to work with hacksaws and select blades: Able to work with various types of Files: Test and Evaluation No. 1

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e e im m T i n tT io n e et l ng p m pi g p i m e n st ni qu ke Co l e n E e a T a i s s, t Pl er ce ou sia ol at es n y it c t o H l a T M a T a es f n f m n y m r o T t o A te e l e fo st rit se ra se af ob er y U J U S W O P S

FA CT O RS

1= Excellent, Outstanding Competence 2=Satisfactory, Acceptable Competence 3= Marginal, Needs Additional Training 4=Unsatisfactory, Needs Extensive Training

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Rotating Element - Clearance, Limits, Fits, Tolerance: Bearings and Lubrications: Mechanical Seals and Packing Alignment and Couplings Preventive Maintenance and Troubleshooting Central Lubrication System (Pure Mist) : Test and Evaluation No. 2

Installation, Alignment Removal, Repair, Operation, Troubleshooting and Preventive Maintenance of : - Reciprocating Pumps - Rotary Pumps PETRONAS PENAPISAN (T) SDN. BHD. Proprietary Document

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ENGINEERING TECHNICIANS ORIENTATION (MECHANICAL)

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e e m m Ti i n tT io g en et in l p p m g e t ip m in ke es qu Co nn l e T E a s a i Pl ce st ia er ou ls, n s e o H ic at t y a t T a al To es n rm y fM n f m t T e o t o o A l e fe rf rit e e st b ra Sa Pe W O Us Jo Sy Us

FA CT O RS

EVALUATION KEY 1= Excellent, Outstanding Competence 2=Satisfactory, Acceptable Competence 3= Marginal, Needs Additional Training 4=Unsatisfactory, Needs Extensive Training

SKILL BLOCK ANALYSIS AND EVALUATION GUIDE (ROTATING AND RECIPROCATING EQUIPMENT)

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- Diaphragm Pumps - Variable Displacement - Proportionary Pumps - Magnetic Driven Centrifugal Pumps - High Speed Centrifugal Pumps Test and Evaluation No. 3

Installation, Repair, Operation, Troubleshooting and Preventive Maintenance of : - Centrifuge - Fans and Blowers - Tasnk Mixers Hoses and Rail Car Pullers, etc. Test and Evaluation No.4 PETRONAS PENAPISAN (T) SDN. BHD. Proprietary Document

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e e m m Ti i n tT tio en e ng pl m pi g p i m e n t u e ni Co es l ek Eq T an a l s i , e s P t ia er ou ol nc ic at ys es t H t o l a T a a M T es n f f n y m rm T t o o A e e o l t fe rf e e b st ra rit Jo Sa O Us Us Pe Sy W

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EVALUATION KEY 1= Excellent, Outstanding Competence 2=Satisfactory, Acceptable Competence 3= Marginal, Needs Additional Training 4=Unsatisfactory, Needs Extensive Training

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POWER TRANSMISSION COMPONENTS: Parts, Inspection, Operation, Troubleshooting and Repair : Gears and Gearing : Coupling: Conveyers: Drive Belts : Chains : Auxiliary Drive Units : Test and Evaluation No. 5

DRIVERS - Types, Characteristics : Steam Turbines - Construction : Combustion Gas Turbines - Construction: PETRONAS PENAPISAN (T) SDN. BHD. Proprietary Document

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e e m m Ti i n tT tio g en e in pl m g p p i m n e t u ni Co ke es l Eq e T an a l i , s e s P t ia er ou ol nc ic es at ys t t o l H a T a es na fM fT n m rm y T o o e A t e o t l rf e e b fe st rit ra Jo Us Us Pe W O Sy Sa

FA CT O RS

EVALUATION KEY 1= Excellent, Outstanding Competence 2=Satisfactory, Acceptable Competence 3= Marginal, Needs Additional Training 4=Unsatisfactory, Needs Extensive Training

SKILL BLOCK ANALYSIS AND EVALUATION GUIDE (ROTATING AND RECIPROCATING EQUIPMENT)

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Induction Motors (Low Voltage) : Large Induction Motors (High Voltage) : Internal Combustion Engines (Diesel, Dual-Fuel, Gasoline) - Principal and Operations Hydrodynamic Bearings and Lubrications : Central Lubrication System (Purge Mist) : Test and Evaluation No.6 DEISEL AND DUAL FUEL ENGINES : Description, Overhaul, Disassembly: Fuel Systems, Injectors, Cooling Systems: Ignition and Electrical Systems Air and Exhaust System : Turbo and Super Charging: Engine Disassembly - Crank End, Power End : PETRONAS PENAPISAN (T) SDN. BHD. Proprietary Document

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ENGINEERING TECHNICIANS ORIENTATION (MECHANICAL)

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e e m m Ti i n tT tio en e ng pl m pi g p i m e n t u o e ni l es Eq ek -C an ir a T l s , s P t e e ou sia ol at y ic es nc t o H l t a T M a T a es f n f m n y m r o T t o A e e t l e fe e b rfo st rit ra Us Jo Us Sa O W Sy Pe

EV AL UA TI O N

EVALUATION KEY 1= Excellent, Outstanding Competence 2=Satisfactory, Acceptable Competence 3= Marginal, Needs Additional Training 4=Unsatisfactory, Needs Extensive Training

SKILL BLOCK ANALYSIS AND EVALUATION GUIDE (ROTATING AND RECIPROCATING EQUIPMENT)

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Control Gear and Instruments : Fits, Tolerances, Clearances : Lubrication Systems : Periodic Services and Inspections: Test and Evaluation No: 7 :

AUXILLARY EQUIPMENT: Monitoring and Troubleshooting : Preventive Maintenance Lubrication : Lubricators : Test and Evaluation No. 8 : EVALUATOR's COMMENTS:

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ENGINEERING TECHNICIANS ORIENTATION (MECHANICAL)

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e e m m Ti i n tT tio en e ng pl m pi g p i m e n t u e ni Co es l ek Eq T an a l s i , e s P t ia er ou ol nc ic es at ys t H t o l a T a es na fM fT n y m rm T t o o e A e o t l fe rf e e b st rit ra Jo Sa O Us Us Pe W Sy

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EVALUATION KEY 1= Excellent, Outstanding Competence 2=Satisfactory, Acceptable Competence 3= Marginal, Needs Additional Training 4=Unsatisfactory, Needs Extensive Training

SKILL BLOCK ANALYSIS AND EVALUATION GUIDE (ROTATING AND RECIPROCATING EQUIPMENT)

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COMPRESSORS : Types, Characteristics, Terminology : Reciprocating Compressors Parts, Construction, Inspection, Troubleshooting, Repair, Testing : Rotary Compressors Parts, Construction, Inspection, Troubleshooting, Repair, Testing : Centrifugal Compressors Parts, Construction, Inspection, Troubleshooting, Repair, Testing : Compressor Valves : Compressor Controls: Test and Evaluation No.9 : PETRONAS PENAPISAN (T) SDN. BHD. Proprietary Document

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ENGINEERING TECHNICIANS ORIENTATION (MECHANICAL)

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e e m m Ti i n tT g tio en e in l p p m g e t ip n om ke es ni qu C l e n T E s a ia t ce Pl ia er ou ls, n s t o es H at y it c a o l T es M a T a m n r y T f n f t e m l o t o A fo e e ra rit e b af er st se s O o y S P W U J U S

FA CT O RS

EVALUATION KEY 1= Excellent, Outstanding Competence 2=Satisfactory, Acceptable Competence 3= Marginal, Needs Additional Training 4=Unsatisfactory, Needs Extensive Training

SKILL BLOCK ANALYSIS AND EVALUATION GUIDE (ROTATING AND RECIPROCATING EQUIPMENT)

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COMPRESSOR ACCESSORIES : Reciprocating to Rotary Movement Conversion Components: Test and Evaluation No.10:

UOP's ROTARY VALVES: Parts, Construction, Inspection, Troubleshooting, Repair, Testing : Test and Evaluation No.11:

MISCELLANEOUS / ADDITIONAL : Contract Supervision and Administration: Test and Evaluation No.12:

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ENGINEERINGTECHNICIANS ORIENTATION (MECHANICAL) INTEGRATED MACHINERY INSPECTION

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e e m m Ti i n tT g tio en e in l p p m g e t ip n om ke es ni qu C l e n T E s a ia t ce Pl ia er ou ls, n s t o es H at y it c a o l T es M a T a m n r y T f n f t e m l o t o A fo e e ra rit e b af er st se s O o y S P W U J U S

FA CT O RS

EVALUATION KEY 1= Excellent, Outstanding Competence 2=Satisfactory, Acceptable Competence 3= Marginal, Needs Additional Training 4=Unsatisfactory, Needs Extensive Training

SKILL BLOCK ANALYSIS AND EVALUATION GUIDE (ROTATING AND RECIPROCATING EQUIPMENT)

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PERMENENT MONITORING SYSTEM: Relative shaft vibration or absolute casing vibration: Test and Evaluation No.13:

INTERMITTEENT MONITORING SYSTEM: Analysis of velocity or acceleration of bearing or casing : Test and Evaluation No.14:

INTERMITTEN - SIMPLE PARAMETERS : RMS velocity Peak vs RMS acceleration: Test and Evaluation No.15:

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ENGINEERING TECHNICIANS ORIENTATION (MECHANICAL) INTEGRATED MACHINERY INSPECTION

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m Ti

e

e m Ti

n g tio e in l p p m e g t ip n om ke es ni qu l C e T n E a s i a e , t er ia ou Pl ls nc es s o H at st a y it c T o l e M a m T a n r T y t e l m of of An fe rfo itt te ra e r e b a e s s s O S P W U Jo U Sy

FA C TO R S

EVALUATION KEY 1= Excellent, Outstanding Competence 2=Satisfactory, Acceptable Competence 3= Marginal, Needs Additional Training 4=Unsatisfactory, Needs Extensive Training

SKILL BLOCK ANALYSIS AND EVALUATION GUIDE (ROTATING AND RECIPROCATING EQUIPMENT)

en

t

Name of Trainer/ Evaluator

INTERMITTEN - FIELD ANALYSIS: RMS velocity Peak vs RMS acceleration on motors, pumps fin-fans etc Test and Evaluation No.16:

INTERMITTEENT FIELD OR LAB ANALYSIS Normal spectrum comparison in laboratory and occasional trouble shooting in field: Test and Evaluation No.17:

PERMANENT MONITORING & DETAIL ANALYSIS : Utilisation of computer and analyser Test and Evaluation No.18:

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ENGINEERING TECHNICIANS ORIENTATION (MECHANICAL) INTEGRATED MACHINERY INSPECTION

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m Ti

e

e m Ti

n g tio e in l p p m e g t ip n om ke es ni qu l C e T n E a s i a e , t er ia ou Pl ls nc es s o H at st a y it c T o l e M a m T a n r T y t e l m of of An fe rfo itt te ra e r e b a e s s s O S P W U Jo U Sy

FA C TO R S

EVALUATION KEY 1= Excellent, Outstanding Competence 2=Satisfactory, Acceptable Competence 3= Marginal, Needs Additional Training 4=Unsatisfactory, Needs Extensive Training

SKILL BLOCK ANALYSIS AND EVALUATION GUIDE (ROTATING AND RECIPROCATING EQUIPMENT)

en

t

Name of Trainer/ Evaluator

VIBRATION TRANSDUCERS Eddy Current or Proximity Probe: Moving Element and Accelerometer Piezoelectric Materials Test and Evaluation No.19: Special Type Accelerometers: Compression Type Accelerometers Shear & Delta Shear Type Accelerometers Uni-Gain Accelerometers: Test and Evaluation No.20: Choosing Mounting Position for Accelerometer: Influence of Temperature Transients: Handling of Accelerometers: Test and Evaluation No.21: PETRONAS PENAPISAN (T) SDN. BHD. Proprietary Document

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ENGINEERING TECHNICIANS ORIENTATION (MECHANICAL) INTEGRATED MACHINERY INSPECTION

EV AL U AT IO N

PERFORMANCE FACTORS

i tT

m

e

e m Ti

n tio ng e pi pl pm i e g t u n om ke es ni l Eq C e T n a , s i a e s t a er ou Pl ol nc es si o H at st a y it c T l T e M a m f a n r T y o t e l m of An fe rfo itt te ra e r se b a e s s O U S P W U Jo Sy

FA C TO R S

EVALUATION KEY 1= Excellent, Outstanding Competence 2=Satisfactory, Acceptable Competence 3= Marginal, Needs Additional Training 4=Unsatisfactory, Needs Extensive Training

SKILL BLOCK ANALYSIS AND EVALUATION GUIDE (ROTATING AND RECIPROCATING EQUIPMENT)

en

Name of Trainer/ Evaluator

FREQUENCY ANALYSIS AND INSTRUMENTATION

Understanding of FFT / DFT Analyzers: Understanding of Flat and Hanning weighting: Understanding and analysis of Sine Wave Weighting anf Filter Characteristics Picket Fence effect and corrections Linear and Exponential Averaging: Scan Analysis, Scan Average and Zoom Analysis: Test and Evaluation No.22: BASIC FAULT DETECTION AND DIAGNOSIS Overall Measurement Techniques: Unbalance/Misalignment/Crack shafts detection:: Detecting Vibration from Gears and Roller Bearings:

Test and Evaluation No.23: PETRONAS PENAPISAN (T) SDN. BHD. Proprietary Document

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ENGINEERINGTECHNICIANS ORIENTATION (MECHANICAL) INTEGRATED MACHINERY INSPECTION

EV AL U AT IO N

PERFORMANCE FACTORS

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FA C TO R S

EVALUATION KEY 1= Excellent, Outstanding Competence 2=Satisfactory, Acceptable Competence 3= Marginal, Needs Additional Training 4=Unsatisfactory, Needs Extensive Training

SKILL BLOCK ANALYSIS AND EVALUATION GUIDE (ROTATING AND RECIPROCATING EQUIPMENT)

en

Name of Trainer/ Evaluator

COMPUTERISED FAULT DETECTION AND TREND ANALYSIS Logging Amplitude Scale for Fault Detection: Logging Frequency Scale for Fault Detection:

Test and Evaluation No.24:

DIAGNOSTICS OF BEARINGS AND GEARS: Faults in Rolling Element Bearings: Gears and tooth deflection analysis: Ghost components in gear vibration spectra: Establishing Toothmesh frequency: Test and Evaluation No.25: FREQ ANALYSIS OF SIGNALS FROM ROTATING MACHINES: CEPSTRUM ANALYSIS ( SPECTRUM OF A SPECTRUM)

Test and Evaluation No.26: PETRONAS PENAPISAN (T) SDN. BHD. Proprietary Document

MYG /imi / 82 - 98

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ENGINEERING TECHNICIANS ORIENTATION (MECHANICAL) INTEGRATED MACHINERY INSPECTION

PERFORMANCE FACTORS

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EV AL U AT IO N

EVALUATION KEY 1= Excellent, Outstanding Competence 2=Satisfactory, Acceptable Competence 3= Marginal, Needs Additional Training 4=Unsatisfactory, Needs Extensive Training

SKILL BLOCK ANALYSIS AND EVALUATION GUIDE (ROTATING AND RECIPROCATING EQUIPMENT)

en

Name of Trainer/ Evaluator

MEASUREMENTS OF RUN-UPS AND RUN-DOWNS (TRANCIENT) Electrical and mechanical Run-out:: Understanding the Bode Plots: Understanding the Polar Plots: Understanding the Cascade Plots: Understanding the Campbell Diagrams:

Test and Evaluation No.27: RECIPROCATING MACHINE FAULT DETECTION: Required instrumentations: Analysis Procedures of Recip-Machine:

Test and Evaluation No.28:

PETRONAS PENAPISAN (T) SDN. BHD. Proprietary Document

MYG /imi / 82 - 98

Page: 14 of 14

Basic Overhaul and Inspection Procedures for Centrifugal Pumps All measurements in imperial units, (metric in brackets)

Parts Needing Attention. Centrifugal pumps—whether horizontal or vertical—can be considered to have two basic types of parts—rotating and stationary. Rotating parts include the impeller, shaft, wearing rings, shaft sleeves, bearings, and mechanical seals. Stationary parts include the casing, bearing housing, suction and discharge flanges, packing, leakoff tubing, and baseplate. Most maintenance work on centrifugal pumps is concerned with the rotating pans, but some work is also performed on the stationary parts. Rotating parts will be considered first in this procedural paper because they generally require more care. ROTATING PARTS Impellers. Immediately after removing the impeller from a pump (Fig.x) inspect its eye, vanes, shrouds, wearing rings, passages, hub, and other parts. Wear may occur at the eye, vanes, shrouds, and other impeller parts. Corrosion, cavitation, and erosion are generally accompanied by a wasting away of the impeller or vane surfaces. Where the attack is severe, the thinned sections may have holes through them or may warp and deflect. To reduce wear from liquids containing abrasives, wearing pads (Fig.x ) are sometimes welded to the impeller or coating. Since the procedures for coating an impeller with special protective materials are specialised, it is generally best to have this work done by an organisation specialising in it. If cavitation is severe, it may be necessary to change the suction conditions or install an impeller suitable for the existing suction conditions. If the impeller is dirty when it is removed from the pump, clean it carefully before making an inspection. Use a soft wire brush or a steam lance to remove thick gummy residues. Scale, coke, and other deposits can be removed by chemical cleaning or sandblasting. In either case; precautions must be taken to see that the impeller material is not damaged by the cleaning method chosen. Pitting of the impeller may be caused by cavitation, which can occur without audible noise. While it is possible to recondition an impeller that is worn or corroded, it is often better to replace it with a new one suitably protected to resist wear or corrosion. Pump manufacturers are always willing to advise users as to which maintenance procedures should be followed for best results in service. Often, before the impeller can be removed for inspection, scale and burrs must be removed from the shaft with a file (Fig.x). To prevent damage of the packing and fittings, clean the shaft thoroughly. After unbolting and removing the case (Fig.x), use a screw driver or an allen key to loosen the impeller setscrew. Clean the shaft as the work progresses. If a bronze impeller is used and was shrunk fitted on the shaft, as many are, slip a metal sleeve over shaft and against the impeller hub, to protect the shaft while the impeller is heated. Start heating the impeller with an oxy-acetylene torch from the outside of the shroud, working toward the hub. Revolve the impeller while heating it so its temperature will be equalized. When the impeller is loose pry it off the shaft, being careful to press only against the shroud.Check with the manufacturer for removal procedures for impellers made of materials other than bronze. Impeller Runout. With pumps having bearings at each end of the shaft, mount the impellers, wearing rings, spacer, and shaft sleeves on the shaft and support the assembly between centers (Fig.x). Set a dial gage at zero and take readings near each end and at the center of each shaft sleeve. Also take similar readings at each impeller wearing ring. As a rule of thumb, for most pumps, if the runout is not more than 0.0015 in (0.0381mm)., the assembly can be considered accurate and the shaft installed as is. If the reading is greater, check for a bent shaft, out-of-square, dirty, or burred impeller-end of a shaft or spacer sleeve. Check the runout on single-stage cantilevered-shaft pumps as shown in Fig.x. If the shaft binds or the dial gage shows a runout greater than 0.0015 in (0.0381mm)., loosen the ball-bearing lock nut and check

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Basic Overhaul and Inspection Procedures for Centrifugal Pumps All measurements in imperial units, (metric in brackets)

the runout again. If the impeller and shaft assembly runs true, either the lock nut washer is burred or the faces of it and the bearing are not parallel. Smooth out the burrs, tighten the lock nut, and check the runout. If there are no burrs, install a new washer and check the runout. If the assembly runs out more than 0.0015 thousandth of an inch (0.0381mm). after the lock nut is loosened, dismantle and check for a bent shaft, cocked or loose ball bearing, or out-of-square adjoining faces of parts on the shaft. All adjacent faces should be square with the shaft center line and parallel to each other. True up the faces in a lathe, if necessary (Fig. x). Figure x shows a gage being used to check the eye of an impeller. Balance. Badly worn or corroded impellers may or will vibrate excessively. While the presence of vibration is usually easy to detect, a special balancing machine (Fig.x) is needed to detect how much unbalance exists. It is sometimes (usually) necessary to return the impeller and shaft to the manufacturer for a check of this type if a balancing machine is not available or in not working order. To balance an impeller by hand, press it on an arbor, the ends of which rest on two parallel and level knife-edges. If out of balance, the impeller will turn and come to rest with its heavy side down. To balance the impeller, metal must be removed from the heavy side. This must be done without impairing the pump’s performance or accelerating erosion. For this reason, drilling holes in the heavy side is unacceptable. The best practice is to mount a shrouded impeller off center in a lathe and take a cut from the shroud, deepest at the rim (Fig.x ). This may be done on one or both shrouds by a proficient machinist, depending on their thickness and the amount of metal to be removed. In semi open impellers, remove metal from the shroud if the design permits, or from the underside of the vanes of open impellers. To check the diametral clearance of an impeller, place it in its stage piece, and move it laterally against a dial gage. Compare this reading with the manufacturer’s recommendation or required dimensions. Figure x shows the steps necessary to locate the impeller on the shaft of a pump fitted with ball bearings. Shafts. Check for a bent shaft by means of a dial gage, as described above. Badly bent shafts should be either be replaced or returned to the pump manufacturer for straightening because the average plant does not have the necessary facilities. A shaft may also be checked for trueness by swinging between lathe or other centers and checking the runout with a dial gage. Tap the impeller shaft key (Fig.x ) to see that it is tight. Twist of the shaft under load, expansion, or corrosion will progressively loosen the impeller. Reconditioning a Shaft. Centrifugal pump shafts wear while in use. Typical wear points are at the packing box (Fig.x ) and other places where the friction load is high. Keep the friction wear low by using a good grade of packing and adjusting the glands evenly. Be sure that the gland follower does not ride on the shaft. As soon as the packing becomes dry, replace or relubricate it, If a new shaft is costly, or wear is rapid, it may pay to add a tougher wear-resisting surface to the shaft at the points of sliding or rotating contact. This process is known as hard surfacing and can increase the life of some parts from four to thirty times. Center the shaft in a lathe (Fig.x), supporting it in a steady rest located near the point of work, if necessary. Use calipers to measure any unevenness to find the depth of cut needed to give the necessary thickness of metal overlay. When the buildup is done by welding (Fig.) keep the heat low and lay the bead longitudinally along the shaft. For 2-in (50mm), and smaller shafts, alternate the beads at 180-deg positions. On larger shafts, lay the beads at 90-deg intervals until the shaft is completely covered at its worn section. Before spraying a shaft with metal, take a rough thread cut of 24 threads per inch (per 25.4mm) Then break the apex or tip of the thread with a knurling tool (Fig.). Mount the gun on the tool post with the nozzle about 6 in.(150mm) from the work (Fig.). Hold the gun by hand when spraying the undercut or dovetail section at each end of the machined area. Take a rough finishing cut (Fig.), leaving enough metal for filing and lapping of the shaft to its PP(T)SB PROPRIETARY DOCUMENT MYGAZEE ‘99 Pg:2 of 9

Basic Overhaul and Inspection Procedures for Centrifugal Pumps All measurements in imperial units, (metric in brackets)

final size. File lightly' (Fig. ), moving the file back and forth as the shaft revolves. When the shaft is correctly filed, polish it with No. 0 and 00 emery cloth ( metric dimensional codes may vary slightly) (Fig.). The shaft is now ready to be replaced in the pump. There are other methods of bard-surfacing shafts. For best results, consult the manufacturer of the equipment for this work. Wearing Rings. These (Fig. ) are installed in the casing or impeller, or both, to take the wear resulting from rotation of the impeller, grit, and other abrasives incause. They are replaceable at a peller or casing, whose wear they the liquid handled, and any other far lower cost than either the imprevent. Although wearing rings are designed for uniform clearance around their circumference, certain conditions may cause them to rub during pump operation. Pump designers recognise this and try to choose wearing-ring materials that will run as a bearing while lubricated by the liquid pumped. Wearing-ring. Wearing ring clearance is of extreme importance because, as the clearance increases in a given pump, leakage of liquid past the rings becomes greater, reducing the efficiency of the pump. In one design of single-stage double-suction pump, tripling the wearing-ring clearance reduces the pump efficiency by, more than 5%. For this and other reasons, designers try to choose nongalling materials for wearing rings. Typical nongall combinations are: (1) bronze with a dissimilar bronze, (2) cast iron with bronze, (3) steel with bronze, (4) monel metal with bronze, and (5) cast iron with cast iron. There is less leakage after accidental contact between casing and impeller wearing rings made of these materials. Figure xa shows typical wearing-ring clearances recommended by one manufacturer. One method of measuring wearing-ring clearance is shown in Fig. xb.Installation. Centrifugal pumps fitted with wearing rings come supplied with the rings. So it is not necessary to install rings on a new pump. Once the rings wear, they’ must be replaced. - To do this, first secure suitable replacements for the rings in the pump from the manufacturer. Remove worn impeller rings which are threaded or shrunk in place, by heating the ring with a torch, being careful not to heat the impeller. Or insert a few pieces of dry ice in the impeller’s eye to shrink the impeller away from the ring. The casing wearing rings can be removed using similar methods. Since many impeller rings are shrink fits (Fig.x ), heat the ring before slipping into place and pinning. For ring diameters of 2.5 to 6 in.(63.5 to 152.4mm), interference between the ring and the impeller is 0.001 to 0.0015in.(0.0254 to 0.0381mm) Between 6 and 12 in (154.5 to 304.8mm), interference is 0.002 (0.0508mm) and 0.0025 in (0.0635mm). Insert the pins or locking with set screws after the ring is in place. Figure x shows the methods and tools used in taking measurements of the impeller and casing wearing rings. to restore the clearance in a pump having a single wearing ring, obtain a new ring bored undersize and true up the impeller hub by taming/trimming it dawn in a lathe. Sometimes it may be necessary to build up either the casing ring or the impeller hub, machining both to give the correct clearance. This is difficult and feasible only if the pump is large and the equipment to do the work is in good working condition. Pumps with double wearing rings can have their clearance renewed in three ways: (1) Obtain a new oversized impeller ring and use the old casing ring, bored true to the larger diameter. (2) Obtain a new casing ring bored undersize and use the old impeller ring turned to a smaller diameter. (3) Renew both rings, if necessary. When a new wearing ring is put on an impeller, its surface is often off center with the shaft. So, after mounting a new ring, check its wearing surface, Machine, if necessary. Do this whether the ring is pressed, bolted, or screwed on the impeller. Measure the clearance of flat wearing rings with a feeler gage between the stationary and rotating parts. In multistage pumps, where the wearing ring may be L-shaped, the lip of the L prevents using the gage. A fairly close check of the clearance may be made by mounting a dial indicator on the impeller (Fig. xb) and setting it to zero with the casing ring resting on the impeller wearing-ring hub. Without moving the impeller or PP(T)SB PROPRIETARY DOCUMENT MYGAZEE ‘99 Pg:3 of 9

Basic Overhaul and Inspection Procedures for Centrifugal Pumps All measurements in imperial units, (metric in brackets)

dial indicator, push up on the stationary ring from below and record the maximum dial reading. This is the diametral clearance. Divide by 2 to get radial clearance. In Fig. xa, tolerances are always plus for casing wearing rings, minus for impeller wearing rings. Shaft Sleeves. These (Fig. xb) wear when packed too tightly. They may be reconditioned by welding or metallising in the same manner as described above for shafts. Where wear is extreme, replacement of the worn sleeve with a new one is often recommended. Use a sleeve puller to remove old sleeves from the shaft. When the sleeve is rusted to the shaft, use the impeller nut to help loosen the sleeve. In extreme cases a hammer and chisel may be needed to split the sleeve before removal. Be careful not to damage the shaft with the hammer or chisel After installing a new shaft sleeve, check its concentricity on the shaft. Bearings. Figure x shows the high-pressure end of a barrel-type feed pump and the ring-oiled sleeve bearing and pressure-lubricated Kingsbury thrust bearing opened for inspection and maintenance. Check the journal and thrust bearings as shown in Fig. xc. Ball bearings may be checked as shown in Fig. x. Clearance between the shaft and the babbitt of sleeve bearings should not exceed 150 per cent of the original value. Sleeve Bearings. These can be rebabbitted in the field, if desired, but this practice is generally confined to larger plants. To replace the metal in babbitt-lined bearings: (1) Machine or melt out the old metal and place it with new metal in a melting pot. (2) Wash the shell in a weak solution of lye and water. Rinse in clean running water. Do not touch or wipe the area to be metalled. Dry with compressed aft. (3) Dip the shell in a bath of hydrochloric acid (1 part acid to 5 parts water) for 10 minutes, or paint with this solution, using an old but clean brush. Rinse in water. (4) Plug oil holes with heat-resistant plastic or other material (fire clay, etc.). Coat all the surfaces not to be tinned with a fire-clay wash (made up like a cement wash). (5) Swab the surface to be tinned with a flux. This can be made by dissolving zinc in muriatic acid until no more will dissolve and the acid is neutralized or “killed.” (6) Place the bearing shell in a molten solder bath. Keep it there until it is as hot as the solder—575 to 625 F. (300 to 330 C.) Remove and inspect the tinning. Any bare patches must be scraped bright, fluxed, and retinned. (7) Place the bearing shell in a previously prepared cylindrical jig of suitable size and center an undersized length of shaft in the shell. (8) Pour the babbitt at 800 to 850 F (425 to 455 C) into the space be. tween the jig and a mandrel placed inside it. The diameter of the mandrel should be slightly less than that of the pump shaft Use a ladle large enough to complete the pouring in one filling. the molten babbitt with an iron wire, working the wire up and down. This prevents cavities fonning during cooling. (9) If the metal shrinks from the edges of the bearing shell,. correct by peening it outward. Clean all the hammer marks when fitting and scraping the bearing. Set up large bearings in a lathe chuck and bore to the correct size after the babbitt is cool. Bolt both halves of the bearing together before the boring is begun. Make the final fit by scraping and checking as in Fig. x. Small bearings can be bored in a drill press. Use an adjustable double-edge culling tool. Take fine cuts to prevent the tool from digging into the metal. Scrape to size. Rolling-contact Bearings. The Anti-Friction Bearing Manufacturers Association (AFBMA) has published a number of recommendations pertaining to maintenance, overhaul and care of ball, roller, needle, and various types of thrust bearings. Since they are extremely useful to follow, several of them are given below.

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Basic Overhaul and Inspection Procedures for Centrifugal Pumps All measurements in imperial units, (metric in brackets)

Keep all rolling-contact bearings clean at all times - dirt means damage. Work on the bearings with clean tools in clean surroundings. Remove all outside dirt from the housing before exposing the bearing. Handle bearings only with clean dry hands. Treat a used bearing as carefully as a new one. Use approved cleaning solvents and flushing oils. Lay and wrap the bearings on only clean or special fibrous-free paper. Protect disassembled bearings from dirt; use only clean lint-free rags if the bearings are wiped. Keep the bearings in oil proof VPI (vapor phased inhibitor) paper when not in use. Clean the inside of the housing before replacing a bearing. Install new bearings as removed from their packages, without washing. Keep bearing lubricants clean when applying and cover containers when not in use. An arbor press is usually the best tool for removing a bearing. But most field work is done with some form of bearing puller. To remove a bearing, press or pull only on the ring which is tight. Press or pull straight and square to keep the ring from cocking. Never press or pull against the bearing shields or separators. Mechanical Seals. These (Fig. x) are extremely popular, and will find more use in all types of pumps in the future. The useful life of a mechanical seal varies with its operating conditions—some last for years, while others have a life of only a few months. While seals differ in construction, the following recommendations by NSO Eagle-Sealol are helpful in establishing general maintenance methods. Remove the seal cover containing the stationary face and pull the face from the cover by hand. Using two wire hooks, engage them in the holes provided and remove the rotating face. Although normally only the faces require repair, it is a good idea to remove the spring assembly for cleaning and inspection. Use the same wire hooks as before. if the stationary face is scored, it probably has dirt or scale embedded in it Take a light clean-np cut to remove the metal to below any possible dirt. If the face is made of carbon, machine it from the outside, working inward, to prevent chipping. It is best to use a carbide-tipped tool with a 5-deg lead angle and 7-deg rake. Zero rake can be used for bronze faces. If the rotating face is only slightly scored, lap it without machining, But if the scoring is deep, take a fine cut in a lathe, if the face is stellite or hardened chrome steel of 500 brinell or softer, machine it with a carbide-tipped tool having zero rake and 7-deg lead angle. If possible, hold it in a four-jaw chuck to avoid distortion. Next, the faces must be lapped. If good lapping plates are not available in the plant, it may be possible to purchase them from the pump or seal manufacturer. The plate should be made of soft close grained cast iron and properly charged with diamond powder in paste form. Loose lapping compound should not be used for carbon or bronze faces. Both are porous and the loose compound fills the pores, and in effect makes a lap of the seal face, Washing the face is not always effective. Use two lapping plates, one charged with 750-mesh powder for roughing and one with 1,600-mesh for finishing. Apply diamond paste from its tube, and roll it into the plate surface with a hard roller. The plate is ready for lapping when the diamond particles are embedded in the iron with the points protruding. Keep the plate wet with solvent at all time to wash the residue into the plate grooves. Use a figure-eight stroke, covering as much of the plate as possible to obtain a flat face, When finished, a lapped carbon face should have a high polish. Do not leave loose particles of carbon on the face; they are abrasive and will cause wear. The rotating face of a seal can be lapped using the same technique, but with a light pressure, to avoid damaging the plate. Do not use loose compound on diamond-charged laps. If an optical flat is available, check the faces for flatness. Remember, it is impossible to obtain a flat seal face using a lapping plate that is not flat. When reinstalling the seal, be sure that all parts are thoroughly clean. Use new gaskets.

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Basic Overhaul and Inspection Procedures for Centrifugal Pumps All measurements in imperial units, (metric in brackets)

Seal Tips. Handle all mechanical seals with care. Rough handling can distort the faces. Keep the pads clean; rubbing, off the lubricant applied by the manufacturer may give operating trouble with new seals. Check the shaft for runout and bearing end play—both should be as small as possible. Never operate a seal dry unless it is a dry-type unit; faces will gall quickly in other types. When handling hot liquids in a pump equipped with mechanical seals, keep the seal faces flooded to prevent vapor binding. It is a good practice to rotate pumps equipped with mechanical seals once a day when stopped. If there is any danger of losing the pump suction, keep the vent or equalizing line open during starting and initial operation of the pump. (Note: This should also be done with standard-packed pumps handling volatile liquids. Close vent-line valve after the pump starts. If the pump has water-cooled glands, be sure the cooling water to the glands is turned on before starting the pump. If ice forms on seals of pumps installed outdoors, use steam or hot water on the seal boxes to melt the ice. Most plants today are replacing conventional packing with mechanical seals. Outside seals fit in the pump stuffing box, usually without extensive alterations. Inside seals go in the pump and some special work may be required to adapt them to the pump. For specific instructions on installing mechanical seals in an existing pump, consult both the pump and seal manufacturer. The exact procedure to follow will differ from one seal and pump to another. STATIONARY PARTS Casing. When the pump casing is open (Fig. xa), examine the waterways for corrosion and erosion. If the casing is pitted or eroded, this can often be corrected by welding, brazing, or metal spraying, depending on the material from which the casing is made and the facilities available for repair. If the signs of wear are severe, consult the manufacturer on the possibility of using more resistant materials. Clean and paint the waterways before closing the pump casing. Most split-casing pumps have one or more gaskets between the upper and lower halves. These should be replaced whenever the casing is opened for inspection or repairs. To prevent delays during pump reassembly, be sure to have a new gasket on ‘hand before the casing is opened. New gaskets should be the same thickness as the original. Trim the inner edges of the gasket accurately along the inside of the pump waterways. Have the gasket mounting surfaces absolutely clean before applying the gasket. Ingersoll-Rand Co. recommends the following for parting-flange gaskets. Since paper and asbestos gaskets may dry out while a pump is stored or idle, check and tighten, if necessary, the flange bolts before starting the pump. Since the gaskets may be further compressed by differential expansion during pump operation, check the bolts for tightness after the first shutdown and during periodic inspections. Do not make this check on hot-liquid pumps until they cool to room temperature. If the upper casing half is removed for a lengthy period, moisten the gasket to prevent drying and cracking. When installing a new gasket, shellac the lower half of the casing. Do not use oil, grease, or varnish. Place the gasket in position, replace the upper half of the casing, and tighten the bolts. Allow at least 8 hr for the gasket to set before removing the upper half of the casing. With the pump open, finish cutting the gaskets with a knife at the cored passages and machined fits. Some horizontally split pumps have metal shims between the two casing halves. The shim thickness is stamped on the lower half of the casing. Cut four small gaps near the outer edge of the gasket for inserting feelers. Place the rotor in the lower half of the casing, sprinkle the gasket evenly with powdered graphite, and replace the upper half of the casing. Pull down the nuts evenly until the gasket is compressed to the value marked on the shim, as measured at the feeler gaps. Worthington Corporation recommends the bolting-tightening sequence shown in Fig.x. Use a similar sequence when tightening the casing bolts of all split-casing pumps. PP(T)SB PROPRIETARY DOCUMENT MYGAZEE ‘99 Pg:6 of 9

Basic Overhaul and Inspection Procedures for Centrifugal Pumps All measurements in imperial units, (metric in brackets)

Bedplate. Clean the bedplate of grease and oil at regular intervals. Snake out the drain lines so drainage is not impaired. Paint the bed-plate regularly—a clean bedplate makes a neater-looking pump and encourages maintenance personnel to give greater attention to their duties. Piping. Check the piping regularly for leaks, damaged insulation, water hammer, defective valves, and other defects. When trouble is found, make the necessary repairs immediately. Defective piping can cripple the best pumps. Foundation. Keep the foundation clean, and freshly painted. Repair cracks before they become so wide that the entire block must be replaced. Where the cracking is excessive, determine the cause and eliminate it, if possible. Poor construction of the foundation, incorrect mix, insufficient setting time, and wrong bolting procedures are common causes of cracks. But where a foundation was correctly built and cracks develop, check for excessive vibration of the pump or piping, settling of the floor or ground, or piping movement that transmits a strain (push or pull) to the pump foundation. OPEN-IMPELLER (PROPELLER) PUMPS The wide use of vertical propeller and axial-flow pumps for circulating-water and plant-water services gives plant maintenance men some new problems. In the higher capacities these pumps may be bulky (Fig.30), but their parts are relatively simple and require no special tools for maintenance. Vibration. When a propeller pump vibrates it indicates poor mechanical condition that may be caused by a motor out of balance, or worn bearings. If the complete unit vibrates the propeller may be out of balance. Propeller blading damaged by a solid object drawn into the pump can cause sudden out of balance. Shaft wobble indicates the need for an overhaul at the stuffing box. Wobble or whip of the shaft at this box makes it hard to keep the packing tight. Because of long shaft, spans and liberal bearing clearance, propeller-pump shafts never run dead true. Water in Oil. On oil-lubricated inner-column pumps, if water backs up into the oiler, wearing rings at the top of the propeller are worn. On water-lubricated-bearing column pumps, wear may cause lubricating-water Bow to increase so much that the column pressure is reduced;

Shaft Wear. Depending on the pump design, shaft-journal wear takes place on the shaft or on the journal sleeve when used. Where bearings run on the shaft, it may be possible to turn shaft sections end for end to get a new journal surface. When this cannot be done, journal areas may be built up by metal spraying. After spraying, grind the journal surface to a fine finish and check the shaft to make sure that it is straight Replace badly worn shafts. Shaft Couplings. How straight a shaft is, when made up of sections screwed together, depends on how square the shaft-section end faces are and the condition of the shaft and coupling threads. Reface shaft ends that are scuffed or dented. Rechase damaged threads and renew shaft couplings if they are in poor condition.

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Basic Overhaul and Inspection Procedures for Centrifugal Pumps All measurements in imperial units, (metric in brackets)

Bearing Wear. Check metal-bearing wear with micrometers. Make measurements at both ends of the bearing and in two planes, as wear is often bell-mouthed or elliptical. Replace bearings if wear is more than 2 or 5 mils.(0.0508 to 0.1270mm) Rubber Bearings. Because of the fluted surface and their resiliency, rubber bearings cannot be checked with micrometers. A shaft-size plug gage or a test on an undamaged section of the shaft will indicate the amount of wear. Loose Bearings. A whipping shaft hammering rubber bearings may loosen the bond between the rubber and metal backing. Discard bearings in this condition. Silt getting behind the lining may force it against the shaft. Or the lining may come loose from its metal backing and drop out of its holder, and this is serious. Metal Backings. Do not tack-weld metal backings of rubber bearings as the welding heat will separate the lining from the backing. Press-fit the bearings in their housings or hold them with setscrews. Water Lubrication. Where parts of the rubber lining stick to the shaft journal and the lining bore is torn, the bearing has run without lubricating water. Water-lubricated column bearings require a continuous supply of clean water at a pressure high enough to push it down the column, Grease or Oil. On grease-packed suction-head bearings, run a supply line from the surface and regularly give the bearings a shot of grease. Once a day is not too often in dirty-water service, as the grease pushes out the abrasive dirt. Give oil-lubricated column bearings several drops a minute from the oiler. Turn on the oiler 15 minutes before starting the pump. Propeller Blades. Bent or damaged propeller blades or pitted areas on them cause unbalance and, as a result, rapid journal and bearing wear. The kind of wear on the journal above and below the propeller gives a clue to the amount of balance. Propeller Balance. If the journals are worn evenly all around, the propeller is in good balance. If the wear is eccentric, but both journals are worn in the same plane, the unbalance is mainly static. If the journals are worn out of round and the wear planes are 90 to 180 deg from each other, the unbalance is dynamic. Wearing Rings. Where the propeller has a wearing ring on its hub, check the clearance between this ring and the stationary casing ring. Replace them when they are worn more than 0.010 in (0.254mm), Throttle Bushing. High-head multistage pumps may have a throttle bushing and bleed-off port above the last impeller. If this bushing wears, replace it to prevent high-pressure water entering the column. Suction Bell. When the bottom bearings and journals are badly worn, the propeller-blade edges may contact and groove or scuff the suction bell. This type of damage can be corrected by a light cut to restore the original bevel. Pitting and Erosion. Cavitation may cause pitting of the suction bell near the leading edges of the propeller blades. Erosion caused by abrasive action of solids in the pumped liquid is a third type of wear. Figure x shows a suction bell which has been attacked by cavitation (deep grooves) and erosion (general scoring). Cast iron, being a brittle material, is most subject to this type of attack. Bronze and stainless steel have high resistance to the effect of cavitation because of their better fatigue qualities. Metal Spraying. Pitted areas in cast-iron bells may be repaired by metal spray if care is taken to ensure proper bonding. Replace suction bells that are badly eroded.

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Basic Overhaul and Inspection Procedures for Centrifugal Pumps All measurements in imperial units, (metric in brackets)

Check Propeller. After the pump is in operation and has been started and stopped several times, check the propeller setting because the turning effort of the motor may further tighten the shaft couplings. If the propeller sets too high, it can cause turbulence at the blade ends which will speed up suction-bell erosion.Pump Runs Roughly. If, after overhauling a vertical pump, the unit is still rough, particularly if this happens off and on, check the water surface for vortex formation. When the bottom of the suction bell is not far enough under water, air may be pulled into the impeller. Wood rafts on the water’s surface or baffles attached to the pit’s floor may break up the whirls where air enters. Inspector’s Responsibilities. Typical line of responsibility and accountability in attaining decisions should be made clear upfront, It is the responsibility of a Rotating Specialist to determine if parts needs replacement and it is the duty of a Metallurgical Specialist to put forth and recommends necessary action pertaining to critical repairs and revivification of metallurgical parts.

The preceding article above are approaches and practices seldom discussed in both the instruction and/or maintenance manual since they are assumed as common knowledge to the operator Extreme care has been exploited in ensuring that non of the above supplement would contradict manufacturer’s procedures and good practices. It is again stressed that the Operation, Maintenance and Installation Manual should be stringently observed or manufacturer of the respective machines be consulted for any abnormalities that may emerge as uncontrollable or unsure by the operator.

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Initial Start. Starting a centrifugal pump for the first time can be a troublesome experience unless our plant crew has made a thorough check of the unit during and after installation. There are probably more unusual accidents and troubles during initial starting of a pump than at almost any other time. Factors to be considered in starting any centrifugal pump include pipe cleanliness, pump alignment, rotation, lubrication, position of valves, stuffing-box and mechanical seals leakage, effect of changes, bypass quantities, throttling of the discharge, and performance checks. Pipe Cleanliness. Multistage pumps, and many single-stage units, have close-clearance running parts which must be protected from Abrasive particles often found in new piping systems. In one subsidiary, Rust particles which lodge in the running clearance of a large barrel type feed pump caused the unit to seize, requiring expensive repairs and loss of pumping capacity during the ensuing shutdown.

Fig. 1. Temporary screen for pump inlet

This is but one example of damage caused by materials lodged in new piping. Many others could be cited because this problem is rather common. To reduce the possibility of larger abrasive particles from reaching the pump, install a flat or conical strainer made up of No. 20 or 30 mesh window screen backed with ¼ inch mesh screen. Place the screen in the suction line as close as possible to the pump suction nozzle. Install pressure gauges on each side of the screen so the pressure drop across it can be measured. When starting the pump for the first time, watch both gages and when a pressure difference occurs between the two, indicating the screen is clogged, stop the pump and remove the dirt and scale from it. When the screen is clean, replace it. PP(T)SB PROPRIETARY DOCUMENT

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Continue cleaning and replacing, as necessary, until no further pressure drop occurs. For most installations, the screen should be used for at least 24 hrs before being removed. Alignment. Bring the pump to operating temperature by admitting liquid to the casing. Check the alignment,. Turn the pump over by hand. It should turn freely , without binding, scraping, or making any noise. Inspect the pump footings to see that any device for expension of the casing are free and in good working condition. Rotation. Check the driver and pump rotation. When a polyphase driving motor is used, touch the starter button just long enough to make the motor turn a few revolutions. The pump shaft should turn in the direction of the arrow on the casing. Figure 2a shows how to determine the direction of rotation of a horizontal centrifugal pump. Stand at the driver end facing the pump. If the top of the shaft revolves from left to right when viewed from this position the pump is said to rotate clockwise. A counter-clockwise pump turns in the opposite direction when viewed from the same point. For a vertical pump (Fig. 2b) look down at the top of the pump. If a point on the shaft revolves from left to right, when viewed from this position, the rotation is clockwise. When the shaft turns in the opposite direction its rotation is counterclockwise.

Fig.2. Method of determining the direction of rotation as specified by the Hydraulic Institute (a) Horizontal pumps and (b) Vertical pumps.

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Pump Bearings. Before any pump is started its bearings must be carefully inspected, cleaned and lubricated. With Oil-lubricated shoes sleeve bearings, thrust shoes (if used), and drain plug. Flush out the housing, oil piping, cooler, pump, and sump tank with kerosene, carbon-tetrachloride or safety solvent such as Klenco. Wash the bearing parts throughly and reassemble them the housing. Flush entire system with lubricating oil and allow it to drain to waste. This will ensure removal of any dirt, metallic, or waste particles present in the bearings or lube sysem. Replace the drain plug, caps, and other parts and fill the bearing as directed by the manufacturer or as given later in this article. Grease-lubricated ball, roller, and needle bearings are usually packed with grease at the factory before the pump is shipped. So no lubrication may be necessary before starting the pump. Check the condition of the grease by removing the bearing housing cover. See that there is grease in the bearing. In general, do not add any grease unless the pump manufacturer gives specific direction to do so. An over-greased bearing may overheat soon after the pump is started. If it has, replace as directed later in this article. Never start a pump equipped with Kingsbury-type thrust bearings without first pouring enough oil into the bushing to protect the thrust shoes. Extreme care must be exercised with all types of bearings on pumps to see that they have enough clean lubricant, Watch for contamination and for dirt picked up during shipment. Thrust Bearings. Figure 3 shows the steps in priming two designs of Kingsbury thrust bearings commonly used in centrifugal pumps. In the bearing in Fig.3a, fill the housing with the correct. grade of oil until the level reaches the ring on the oil-level gage. This resembles the gage shown in Fig. 4a. Remove the locking screw in the top of the thrust-bearing cap and pour a liberal amount of oil into the bushing (Fig.3a). Replace the locking screw, making sure it fits into the bushing, to prevent it from rotating. Remove the vent plug at the top of the oil-pump body and pour in enough oil to wet the parts of the oil pump. Watch the oil level in the gage glass after the main pump starts. If the level stabilises too far below the marker, add oil until the level is ½ (12.5mm) to ¼ in. (6.5mm) below the ring. The bearing in Fig. 3b and 3c does not have an integral oil pump. It is, however, filled in a manner similar to that described above for Fig. 3a. Be sure to replace the locking screw tightly, to prevent the bushing from rotating. The procedures given here apply to one make of Kingsbury thrust bearing. To be completely safe, check the pump instruction manual before flushing and filling thrust bearings of this type. The exact procedure may differ from that given here.

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Fig. 3. Filling thrust bearings

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Fig.4. Two types of commonly used oil level indicators.

Fig.5. Typical forced-feed lubrication system.

Sight Gage Glasses. Figure 4a shows a bearing housing equipped with a typical sight gage glass. The bearing in Fig.4b has a constant-level oiler. This device maintains a constant oil level in the bearing as long as there is oil in its reservoir (oil bottle).

Lube-oil Systems. Figure 5 shows a typical forced-feed lube-oil system for a horizontal multistage centrifugal pump. It consists of an integral gear-type oil pump mounted within the thrust-bearing housing and taking its suction from an oil reservoir mounted below the pump shaft, as shown, a tubular oil cooler, piping, and pressure and level gages. With this type of system, which is often used with large pumps having sleeve bearings for the main shaft, clean and flush the bearings as outlined above, including

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the sump-tank reservoir and the oil cooler. Pour oil into each bearing bushing and fill the sump tank until it is ¾ full, or more. Important: Fill the supply line to the bearings so it will be full when the pump starts and there will be no delay in the oil reaching the bearings. The system supplies oil to both the bearings of the main pump and its driver. Some pumps have a centrifugal-type impeller on the end of their shaft, instead of a gear pump.

Fig.5b. An advanced (Pure) oil mist lubrication system console

Ball Bearings. Almost all modern pumps use ball bearings to carry radial or thrust loads, or both. (These are usually lubricated by an advanced pure or purge mist systems- (fig.5b). The types of ball bearings used for centrifugal pumps include single-row deep-groove, double-row deep groove, double-row self-aligning, and angular contact types made in double-and single-row designs. They may be greaseor oil-lubricated, with oil being popular for the al rger sizes. Figure 6a shows a typical double-type out-board bearing designed to take small axial thrust loads during pump starts and stops. It consists of two angular-contract ball thrust bearings mounted back to back. A locknut and washer hold this bearing and the rotor assembly in correct position with respect to the rest of the pump. Cooled Ball Bearings. Where high temperatures are expected in bearing operation the race may be surrounded by a cooling jacket (Fig.6b). Water or another suitable liquid is circulated through the jackets during pump operation.

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Sleeve Bearings. In horizontal pumps these resemble the bearing shown in Fig. 3b, except that a thrust collar is not fitted. A spherical seat permits the bearing bushing to align itself with the shaft journal. Bearing linings are removable. Vertical deepwell turbine pumps use either water-lubricated bearings (Fig.6b) or oil-lubricated bearings (Fig. 6d). With the first type the liquid pumped serves as the lubricant and no supply from the surface is needed. Oil-lubricated bearings are inside a shaft-enclosing tube and are fed oil from the ground surface by a lubricator mounted on or near the pump drive. Lubricants. Many manufacturers recommend a straight correctly refined turbine-type Petronas Jentram type neutral mineral oil for centrifugal pumps. Normally it should not contain any free acid, chlorine, sulfur, or more than a trace of free alkali. Based on tests by ASTM standard methods, the oil should at least have the physical characteristics given in Table 2-1.

Oil characteristics

Flash point, C............ Saybolt Viscosity at 37.7C Pour Point, C........... Steam Emulsion value

Napthene-base oil

Paraffin-base oil

165.5 min 65.5 SSU min -15 max 75 sec max

182.2 min 60 SSU min 1.6 max 75 sec max

Table 2.1 Recommended lube-oil characteristics for centrifugal pumps

Ambient Temp. C

Greasing Interval

8 hour day: 7-day week 7-day week

Pump Service

Low High

6 months 3 months

24 hour day: 7-day week 7-day week

Low High

6 weeks 3 weeks

Table 2.2 Greasing Interval for Vertical ball-bearing Pumps PP(T)SB PROPRIETARY DOCUMENT

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Grease-lubricated ball bearings are generally packed with lubricant before shipment. Over-greasing must be avoided because it is probably the most common cause of overheating. Many pump manufacturers recommend a sodium-soap grease carrying not less than 82 per cent of a high-grade filtered mineral oil having a viscosity of not less than 150 SSU at 37.7 C. The melting point must not be less than 148.8 C and The penetration at 25 C should be 275 to 330. In addition, the grease should not separate on standing, nor should it form gum, become sticky, harden, decompose, or corrode. Check to see that the grease is free of abrasive particles (sand, dirt, lime, etc.), resins, and mineral salts. Table 2-2 gives the greasing-interval recommendations of one pump manufacturer. Note that this is for a vertical pump. Time intervals for greasing bearings in horizontal pumps will vary somewhat, depending on the type of bearing, pump service conditions, and other factors.

Fig.6. (a) Double type outboard bearing. (b) Water lubricated bearing for deepwell turbine pump (c) ball bearing with cooling jacket (d) Oil lubricated bearing for deepwell turbine pump.

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Regreasing. To replace the grease in a new or used antifriction bearing, remove the bearing from the shaft. Using a brush, wash the race, balls, and other parts with warm kerosene or carbon tetrachloride. Soak the bearing in one of these solvents until the grease on it starts to dissolve. Use an innert-air hose to blow the grease off the bearing parts. Wash the bearing housing out with the same solvent. When both it and the bearing are clean, flusih both with clean mineral oil. Allow the oil to drain to waste. Check the cleanliness of the bearing by rotating it slowly. If it turns smoothly, you can assume it is clean. Do not use kerosene or carbon tetrachloride at a temperature higher fhan 50 C when cleaning the housing and bearing. Pump Exterior. Clean all external surfaces of The pump and its driver. Use rags or waste to remove dirt, dust, oil drippings, globs of grease, and similar matter. It is important to have the pump spotless if accidents and other operating troubles are to be avoided. Check the suction and discharge piping to see that flanges and screwed joints are tightly made up and will not leak. Where automatic valve operators are used, check their operation by opening and closing the valves several times. To be sure that the valve operating mechanism is working satisfactorily, use the manual controls to open and close the valves. Auxiliary Piping. Check the auxiliary piping and liquid supply by opening the cooling-water supply and discharge valves (Fig. 9) and observing the liquid flow. Where an independent oil pump is used, start it and check the pressure and flow in the various lube-oil lines. Check the cooling-water and oil flow at the pump bearings. See that the stuffing-box jackets and smothering glands, if used, have a sufficient supply of clear cool water. Pump Drive. Check the motor, turbine, engine, or other drive to see that it is lubricated and ready to operate. Whenever possible, solo-run the drive independently of the pump to see that it is in good operating order. Follow the manufacturer’s instructions for drive operation. Remember, the drive generally is not supplied by the pump manufacturer. So directions for drive operation may, or may not, be included with the pump instruction manual.

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Priming. Prime the pump by one of the methods given by the manufacturer’s recomendation. Take care to see that the suction pipe is full of liquid and that there is enough reserve in the supply system to keep the line full while the pump operates. Reduced flow into the suction pipe of a centrifugal pump can lead to overheating and extensive damage to the pump.

Position the Valves. Open the suction valve wide. Never use it as a throttling device for pump flow. With a medium- or high-head centrifugal pump it is best to start with the discharge gate valve closed. This is because the pump requires less power input when primed and operated at full speed with the discharge valve closed. Mixed-flow-type centrifugal pumps often require greater power input when started with the discharge valve closed than when it is open. Axial-flow-type centrifugal pumps almost always take more power when started with the discharge valve closed. So it is common practice to start these two types with the discharge gate valve open. But to be certain, check with the pump manufacturer. The Hydraulic Institute recommends that, except in the case of axial or mixed-flow pumps, units driven by squirrel-cage induction motors having reduced-voltage starting control should always be started with the discharge gate valve closed. With this type of motor using across-the-line starting, the discharge gate valve can be opened before the pump is started. But the length of time the electrical disturbance caused by the starting cycle lasts may be reduced by keeping the gate valve closed.

Standard Motors. General-purpose synchronous motors of ratings up to 500 hp (372.8 Kw) at 80 per cent power factor, and those having speeds of 500 rpm or higher at unity power factor, have sufficient pull-in torque to start centrifugal pumps with the discharge gate valve closed. At higher ratings and speeds below 500 rpm, standard synchronous motors do not have enough pull-in torque to start a pump when its discharge gate valve is open. Specially built motors can, however, be obtained to start under these conditions. The remarks above on the starting methods for induction motors also apply to synchronous motors. Where a centrifugal pump must be started with the discharge gate valve open and the starting current must be held to a minimum, use a wound-rotor induction motor where alternating current is available. This type of motor develops full-load torque without taking an excessive line current. A d-c motor can also be PP(T)SB PROPRIETARY DOCUMENT

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used, if a suitable power supply is available. It will develop full-load torque without the line current’s exceeding 125 to 200 per cent of the normal full-load current. Pump Warmup. With pumps handling hot water or other hot liquids, the casing, rotor, and other parts must be brought to within a temperature of 10 to 38 C of the liquid before the unit is started. This prevents unequal expansion, with the possibility of contact between the moving and stationary parts. Some long pumps actually bow or arch when the top half is hot while the bottom half is cold. Open the vent valve on top of the casing and admit warm liquid to the pump. Use one or more casing drains to increase the liquid flow from the pump, thereby reducing the time required for warm-up. Boiler-feed pumps, and units handling valuable or toxic liquids, cannot economically be warmed up in this manner. Instead, a jumper line around the discharge check valve is used. Hot liquid flows through this line, into the pump, and out the suction pipe. In a 4.542 Cubic meter/minute feed pump, a warmup flow of 37.8 liter/minute is enough to bring the pump temperature to within 38 C of the liquid handled. Pumps having labyrinth leakoffs in the stuffing box generally do not need a jumper line. Drainage through the leakoff is usually of sufficient quantity to keep the pump warm, if the leakage is made up with hot liquid from the suction line. At liquid temperatures of 175 C or higher, a period of 3 to 4 hrs. should be allowed for pump to warmup, when starting from a cold condition. Shorter intervals may be used when the pump is warm before heating is begun. Starting the Pump. The following steps are usually suitable for starting a centrifugal pump in good operating condition: (1) Turn on the cooling-water system for the pump bearings, stuffing boxes, and mechanical seals, if these parts are liquid-cooled. (2) Start the auxiliary lube-oil pump, if one is fitted, and check the oil flow to the bear- ings and other parts of the pump. (3) Open the suction gate valve, and close or open the discharge gate valve, depending on the starting procedure to be followed. (4) Close all the drains in the casing and suction and discharge piping. (5) Prime the pump. (6) Open the warmup valve if the pump is not at the right temperature. (7) Open the recirculating valve. (8) Start the driver and bring the pump up to speed. PP(T)SB PROPRIETARY DOCUMENT

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(9) As soon as the pump is up to rated speed, open the discharge gate valve slowly. (10) Check the leakage from the stuffing boxes. (11) Adjust sealing-liquid flow to ensure packing lubrication. A flow of 3 to 7 liter/min to each stuffing box is usually sufficient. (12) Check the pump bearings for lube-oil flow. (13) When there is sufficient flow through the pump, close the recirculating valve. (14) Check the pump suction, discharge, lube-oil, cooling-water, and sealing-water pressures and temperatures. Bearing temperatures generally should not exceed 65 C during pump operation. The above steps are suitable for almost all centrifugal pumps. Some steps may be omitted with smaller units not having separate cooling and oil systems. If the pump shows any signs of trouble while being started such as overheated bearings or packing, excessive vibration or noise stop the unit immediately. Inspect the pump for the cause of the trouble and take corrective action before starting the pump again.

Fig. 10. (a) Internal seal used with clean liquid

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Fig 10 (b) External seal for dirty liquids

Stuffing Boxes. On pumps using discharge liquid for the stuffing boxes (Fig. 10a) close the sealing-liquid valves while the pump is being started on a suction lift. When liquid is being discharged by the pump, open the sealing valves. Adjust until there is a slight but constant leakage of liquid from the stuffing-box glands. With an external seal-liquid supply (Fig. 10b) turn on the control valve before starting the pump. Figures 10c and 10d show two other sealing arrangements. Bypass Use. The recirculation or bypass connections should be cut in whenever the pump must run at shutoff or at 20 per cent or less of its rated capacity. Be sure the recirculated or bypassed liquid flows to a lower-pressure area where it can release some of its heat before returning to the pump. Then, there will be no danger of the pump’s overheating. The bypassed liquid is often returned to the source of suction supply, but not directly to the pump suction.

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Fig 10 (c) External grease seal for non-lubricating liquid, or where water cannot be used.

In boiler-feed pumps the maximum allowable temperature rise usually recommended is 10 C. But when the pump handles cold water a higher rise 15 to 38 C may be permitted. The exact rise allowable limits must be obtained from the pump manufacturer. A rule of thumb often used is: To limit the temperature rise of the water in a boiler-feed pump, do not reduce the capacity below 113 liters/min per 100 hp or 74.5 Kw input to the pump at shutoff. Many centrifugal boiler-feed pumps have an orifice in the bypass line. The orifice is sized to pass the minimum safe-flow quantity for the pump. Excessive throttling of the discharge gate valve can also lead to overheating of a pump. When only a small percentage of the rated flow is allowed to pass through the pump, the casing may be unable to radiate enough heat to keep the temperature constant. This is because the excess horsepower put into the pump, over that delivered by the liquid, appears as heat in the liquid. Noisy operation, over-heating, and shaft breakage are some of the operating troubles that may result.

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Fig. 10 (d) Seal cage next to impeller keeps abrasive liquid away from packing or mechanical seal.

Standby Pumps. When keeping a standby pump hot by circulating hot liquid, use the following operating methods: (1) Circulate cooling water through the bearings, packing boxes, and lantern rings. (2)Circulate oil to the bearings. (3) Start the pump once each 8-hr shift, bringing it up to full speed and operating it for 3 or 4 minutes if they are not in the RCM convention. When the pump is on hot-oil or other service where the liquid handled may coke or plug the passages, start the pump twice each 8-hr shift. Standby boiler-feed pumps are generally held in readiness with the suction and discharge gate valves open at all times. Be sure to have sufficient liquid in the suction well, tank, or pond when keeping a pump on standby.

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Fig. 11. Typical boiler-feed pump log

Fig. 12 Cooling jacket for stuffing box operating at 120 C, or higher

Operating Checks. While the pump runs, make the hourly checks listed below: bearing temperature, suction and discharge pressure, lube-oil temperature and pressure, leakoff flow, discharge flowmeter, stuffing-box leakage, cooling-water suction and discharge temperature and pressure, input to the pump driver and the oil level in the pump and driver bearings. Keep an hourly record of all these readings, using a log sheet (Fig. 11) developed for the particular installation. When the pump is fitted with water-quenched glands (Fig. 12) shut off the quenching-water supply before PP(T)SB PROPRIETARY DOCUMENT

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frying to determine the leakage from the stuffing box. Otherwise, the quenching-gland flow may be mistaken for leakage from the stuffing box (Fig. 10). The condition of the packing box can often be checked by feeling the gland while the pump is running (Fig 13). When the pump handles cool liquids, the gland handles cool liquids, the gland should not be too hot to touch. With hot liquids this test cannot be used because heat is transmitted from the liquid pumped to the gland.

Fig. 13. Checking stuffing-box temperature by hand.

Check ring-oiled or ball sleeve bearings every hour to see that the oil rings (Figs. 3 and 4) are taming freely and supplying enough oil to the bearings and shaft. With any pump having oil-lubricated bearings it is wise to rotate the shaft a few times by hand before starting for the first time when the oil is cold and the bearing surfaces dry. This starts a flow of lube oil to the bearing surfaces, reducing the possibility of overheating of the bearings during starting. When making the hourly round of a pump, listen to the sound it makes while running. Any change in the sound should be carefully checked because it may be the first sign of impending trouble. Increased vibration and sudden changes in the bearing temperatures are other indicators of possible operating troubles. Steam-turbine-driven Pumps. Warm the turbine before starting by opening the steam exhaust valve and all drains on the steam inlet and exhaust, and the turbine casing. Rotate the shaft at least once by hand to see that it is free. Heat the turbine casing by cracking the steam-inlet throttle valve and allowing a small amount of steam to pass through to the exhaust line. Allow steam to enter the casing until the unit reaches its operating temperature. Open all drain valves wide to remove all condensate from the casing and steam lines. When the lines and casing are dry, close all drains, open the throttle valve, and quickly bring the turbine up to rated speed. The governor will then take over control of the turbine speed. Prepare the pump driven by a turbine in the same manner as described earlier for other centrifugal pumps.

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Balancing-chamber Leakoff. Be sure that the leakoff from the balancing chamber of a multistage pump is always open while the unit runs. Shutting off flow from the chamber may prevent the drum from functioning, severely damaging the pump. When the pump is used in boiler-feed service, connect the balancing-chamber leakoff to the nearest convenient liquid tank on the suction side of the pump. Since this is at a lower pressure than in the chamber, leakoff liquid will automatically flow while the pump operates. In general, do not connect the leakoff line to the pump suction nozzle or pipe. Mechanical Seals. These are now used instead of packing in every pumps. When applied to high-pressure high-temperature service, some seals have a quench connection. Others, known as double seals, are used where the liquid is corrosive or abrasive. This type generally has an auxiliary pump and reservoir. The small auxiliary pump circulates lube oil through the seals to lubricate and cool them. Start the auxiliary pump and establish oil circulation before the main pump is started. Pumps handling clean cool liquids sometimes bypass a small portion of the discharge to the mechanical seals to cool and lubricate them. Be certain to open any valves in these and the quench connections before starting the main pump.

Casing Gasket. Before starting split-casing pumps for the first time, tighten the casing bolts. If the manufacturer recommends using a torque wrench, do so. This will ensure that sufficient pressure is obtained on the paper or asbestos flange gaskets. After tightening flange bolts, tighten the shaft-sleeve nuts on packed pumps. Then the sleeve will rotate properly with the shaft. Couplings. Some gear-type shaft couplings must be filled with oil or grease before the pump is started for the first time. Do not over-fill oil-type couplings because the pack-oil may leak out while the pump operates, damaging the pump room walls, ceilings, lights, or other fixtures. Figure 15 shows the three steps to be followed in filling grease-packed gear-type couplings.

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Fig. 15. Steps in packing a gear-type coupling. (a) Pack grease over grid and teeth. (b) Draw up cover and fasten with bolts. (c) Lubricate at least once every 6 months, using a grease gun.

Starting Marine Pumps. Centrifugal pumps in shipboard uses are usually motor- or turbine-driven. On some small vessels the pump may be belt-driven from an electric motor or an engine shaft. Use the same starting procedure given earlier, except that more than one valve may have to be opened in the suction line when the pump is connected to a manifold. In general, the discharge gate valve can be left open, if desired, if the pump is driven by a d-c motor. For most other drives the discharge gate valve should be left in the position described above. Belt-driven pumps need not be aligned quite so accurately with the drive shaft as coupled pumps. This is because the belts can operate with some misalignment of the shafts without excessive wear. But to obtain best life from the drive belts, align the shafts and sheaves as accurately as possible. Then the belts will last longer and will operate more quietly.

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Fig.16. Chart for best operating scheme.

When to Start Second Pump. In some services, like boiler-feed, condenser-circulating water, and others, there is a definite point at which it is economical to start a second pump to aid the first in delivering to the system they serve. With circulating-water pumps this condition usually occurs in the spring and fall of the year. Below a certain water temperature, running the second circulating pump reduces the overall plant efficiency because the input to the pump motor exceeds the additional generator output produced by the better vacuum. But above this temperature, the reverse holds true the extra output from the generator is greater than the input to the pump motor and the plant efficiency is higher. Figure 16, when used with the heat-rate correction chart furnished by the turbine manufacturer, helps show when to start or stop the second circulating pump, if the plant is fitted with two. Example:

A turbine-generator unit operating at 100,000 kw shows a net change of 0.3 per cent in the gross heat rate when its 600-hp second circulating pump is started. Should the pump be left running or be stopped?

Solution:

Enter Fig. 16 at the left at a gross generation of 100,000 kw and draw a straight line through the change in gross heat rate, 0.3 per cent in this case. Prolong this line until it intersects reference scale K. From this point draw a straight line through the circulating-pump horsepower. Extend it to the right. Scale PP(T)SB PROPRIETARY DOCUMENT

CENTRIFUGAL-PUMP OPERATION MYGAZEE ‘99

Q shows the circulating pump should be stopped to secure the best net heat rate for this particular unit. In actual operation, start or stop the second circulating pump, depending on whether it is idle or running. Then note the change in the condenser back pressure (vacuum). Find the corresponding change in the unit heat rate from the turbine manufacturer’s curve. Solve as described above. Pump Shutdown. When the liquid supply is above the pump centerline, close the discharge gate valve, then the suction gate valve. Shutoff the power immediately. When the liquid supply is below the pump (suction lift), close the suction gate valve, then the discharge gate valve. Immediately after, shut off the power or steam to the driver. This procedure keeps liquid in the pump, preventing damage should the unit be started before being primed. If the pump has a recirculating line, close its valve before the pump stops. Open the warmup valve if the pump is to be kept warm. When a pump operates against a high discharge pressure it should be stopped in three steps: (1) Partially close the discharge valve. (2) Shut off the power. (3) As soon as the power is off, close the discharge valve rapidly. The discharge valve should be tightly closed by the time the pump stops rotating. Then there is no possibility of backflow causing the pump to turn in a reverse direction This procedure also prevents water hammer in high-pressure lines. If the pump is not to be started, hang a suitable tag on it. Do not shut off the cooling and sealing liquid supply until the pump shaft has stopped turning. When the pump cannot be run against a closed gate valve, stop the motor or turbine before closing the discharge gate valve. Remember, an idle pump will partially drain through the glands while not operating. So be sure to prime it before starting again. Never run any centrifugal pump dry. Serious damage will almost always result. When the pump is on standby service, its suction and discharge gate valves should be left open. Do not tighten the gland nuts to prevent leakage from the pump unless they can be loosened quickly when the unit is started.

PP(T)SB PROPRIETARY DOCUMENT

CENTRIFUGAL-PUMP OPERATION MYGAZEE ‘99

Flushing Plant Piping. At PP(T)SB, it is mandatory to flush the piping with liquid discharged by one or more pumps (if used) . The end of the pipe is connected to the sewer or some other disposal point and water or other liquid is pumped through. This process serves two purposes it is a convenient way of cleaning any refuse or other matter from the new piping and it permits a good check of the performance of the pump. When flushing the piping a good opportunity is afforded for checking the stuffing-box packing (if used). Check the packing box by feel as soon as the pump starts and regularly thereafter while the pump runs. If the box overheats or begins to smoke, shut the pump down immediately. Allow the box to cool and start the pump again. Check its temperature as before (Fig.13). If it overheats, stop the pump again. It may be necessary to start and stop the pump several times before the packing is ‘worn-in’ and some of the liquid pumped can leak through to lubricate it. If the box continues to overheat, inspect and repack it. Pneumatic Pumps. Air-driven sump pumps are popular in our industrial and construction jobs. Before putting a unit of this type into service: (1) Blow out the air-supply hose to remove any water or dirt before it is attached to the pump. (2) Pour a small amount of clean ISO #10 oil into the live-air inlet. (3) Remove the oil plug and fill the lube-oil reservoir with the same grade as in (2), above. (4) Check the air-exhaust pipe to see that its outlet is above the water level and free of dirt and other obstructions. (5) Connect the air-supply hose. Maintain an air pressure of at least 6.5 Kg/Cm2 at the pump. Some pumps use other pressures be sure that the pressure at the pump air inlet is not more than 5 per cent below the rated inlet pressure. Mount the pump on a flat stone or board so that it is above any muck or settlings in the sump. If the liquid being pumped is extremely dirty, place the pump in a wire basket or screened box. When the pump is taken out of service overnight, drain all water from it and disconnect the air-supply and exhaust hoses. Pour a small amount of oil into the live-air inlet, connect the air hose, and allow the pump to idle for about a minute. Disconnect the hose.

PP(T)SB PROPRIETARY DOCUMENT

CENTRIFUGAL-PUMP OPERATION MYGAZEE ‘99

Spare Parts. The number of spare parts which should be kept on hand varies with the pump application. Thus, aboard ship more parts are generally carried than in a stationary plant. Most manufacturers recommend that the minimum number of spares should be one set of shaft bearings, one set of shaft sleeves, one set of wearing rings, and a supply of suitable set of packing or mechanical seals for the stuffing boxes. On vital jobs where a standby pump is not installed, stock a complete spare rotating element. When ordering spare parts, always remember to give the manufacturer the following information: serial number of pump, size and type of pump as given on the pump name-plate, the exact number of the part as listed in the pump instruction and parts manual, and the name of the part, as listed. Where possible, give the complete symbols stamped on the old part. This is important information hence our pumps are not purchased off-the shelve.

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References: (1) Hydraulic Institute. (2) Ingersoll-Rand Co. (3) Peerless Pump Div. (4) Layne & Bowler, Inc. (5) Westinghouse Electric Corp. (6) Allis-Chalmers Mfg. Co. (7) Power. (8) Worthington Corp. (9) Falk Coupling Div, Borg-Warner Corp. (10) Ebara Corporation. (11) Pacific-Hiro pump Div, Shin Nippon Machinery (12) EKK Eagle Industry Co. Ltd. (13) Lubrication Systems Company. (14) American Petroleum Institute. (15) Institution of Electrical and Electronic Engineers. (16) Institution of Diagnostic Engineers.

PP(T)SB PROPRIETARY DOCUMENT

PETRONAS PENAPISAN (T) SDN. BHD.

MAINTENANCE WORKBOOK FOR CARE OF CENTRUFUGAL PUMPS (APPLIES TO ALL MAKES)

The how and why of pump constructions how they affect pump maintenance. * How a change in liquid can blitz a pump. * Which way suction piping should slope - and why. Tips on easy ways to find leaks. * Common mistakes in packing stuffing boxes. * Quick diagnosis of pump ills...symptoms, causes, proven cures. * Maintenance time table * How tight is “too tight” for a gland? * How to figure head; with tables of losses in piping. valves, fittings, Valuable reference. * little known facts about cavitation; how to protect pumps against it * Vital role of water as a lubricant in pumps. © MYGAZEE ‘99

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CENTRIFUGAL FORCE IN ACTION..... ALL MOVING BODIES TEND TO TRAVEL IN A STRAIGHT LINE WHEN FORCED TO TRAVEL IN A CURVE, THEY CONSTANTLY TRY TO TRAVEL ON A TANGENT IN AN “AIRPLANE RIDE”

CENTRIFUGAL FORCE PUSHES DUMMY PLANES SWUNG IN A CIRCLE AWAY FROM CENTER OF ROTATION.

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CENTRUFUGAL PUMPS - BASIC © MYGAZEE ‘99

CENTRIFUGAL FORCE TENDS TO PUSH SWIRLING WATER OUTWARD........ FORMING VORTEX IN CENTER.

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CENTRUFUGAL PUMPS - BASIC © MYGAZEE ‘99

Let’s Build a Pump! A student of medicine spends long years learning exactly how the human body is built before he ever attempts to prescribe for its care. Knowledge of pump anatomy is equally basic in caring for centrifugal pumps! But whereas the medical students must take a body apart to learn its secrets, it will be far more instructive to us if we put a pump together (on paper, of course). Then we can start at the beginning--adding each new part as we need in its logical sequence. As we see what each part does, how it does it.... we’ll see how it must be cared for! Another analogy between medicine and maintenance: there are various types of human bodies, but if you know basic anatomy, you understand them all. The same is true of centrifugal pumps. In building one basic type, we’ll learn about all types. Part of this will be elementary to some maintenance men but they will find it a valuable “refresher” course, and, after all, maintenance just can’t be too good. __________________________________________________________ So with a side glance at the centrifugal principle on the first and second page, let’s get on with building our pump...

First we require a device to spin liquid at high speed... THAT paddle-wheel device (page 4) is called the “impeller” . . . .and it’s the heart of the pump. Note that the blades curves out from its hub. As the impeller spins, liquid between the blades is impelled outward liy centrifugal force. Note, too, that our impeller is open at the center--the “eye”. As liquid in the impeller moves outward, it will suck more liquid m behind it through this eye... provided it’s not clogged! That brings up Maintenance Rule No. 1: if there’s any danger that foreign matter (sticks, refuse, etc.) may be sucked into the pump---clogging or wearing the impeller unduly---provide the intake end of the suction piping with a suitable screen.

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A year from now, what will you wish you had done for your pump

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Now we need a shaft to support and turn the impeller.... Our shaft looks heavy... and it is. It must maintain the impeller in precisely the right place. But that raggedness does not protect the shaft from the corrosive or abrasive effects of the liquid pumped.... so we must protect it with sleeves slid on from either end. What these sleeves--and the impeller, too--are made of depends on the nature of the liquid we’re to pump. Generally they’re bronze, but various other alloys, ceramics, glass, or even rubber-coating are sometimes required. Maintenance Rule No 2: never pump a liquid for which the pump was not designed. Whenever a change in pump application is contemplated and there’s any doubt as to the pump’s ability to resist the different liquid, check with your pump manufacturer!

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We mount the shaft on either sleeve, ball or roller bearings... As we’ll see later, clearances between moving parts of our pump are quite small. If bearings supporting the turning shaft and impeller are allowed to wear excessively and lower the turning units within a pump’s closely-fitted mechanism, the life and efficiency of the pump will be seriously threatened. Maintenance Rule No.3: keep the right amount of the right lubricant in the bearings at all times. Follow your pump manufacturer’s lubrication instruction to the letter. Main points to keep in mind are... 1. Although too much oil won’t harm sleeve bearings, too much grease in anti-friction type bearings (ball or roller) will promote friction and heat. Main job of grease in anti-friction bearings is to protect steel elements against corrosion, not friction. 2. Operating conditions vary so widely that no one rule as to frequency of changing lubricant will fit all pumps. so play safe: if anything, change lubricant before it’s too worn or too dirty.

Lubricants are cheaper than new pumps—and easier to get.

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To connect with the motor, we add a coupling flange... Some pumps are built with pump and motor on one shaft, of course, and offer no alignment problem

But our pump is to be driven by a separate motor... and we attach a flange to one end of the shaft through which bolts will connect with the motor flange. Maintenance Rule No. 4: see that pump and motor flanges are parallel vertically and axially... And they’re kept that way! If shafts eccentricity are off-center or meet at an angle, every revolution throws tremendous extra load on bearings of both pump and motor. Flexible coupling will not correct this condition if excessive. Checking alignment should be regular procedure in pump maintenance. Foundation can settle unevenly, piping can change pump position, bolts can loosen. Misalignment is a major cause of pump and coupling wear. Always propose for the Reverse Periphery Alignment method to the conventional rim and face. ____________________________________________________

Now we need a “straw” through which liquid can be sucked... Notice two things (fig. 9) about the suction piping: 1) the horizontal piping slopes upward toward the pump; 2) any reducer which connects between the pipe and pump intake nozzle should be horizontal at the top---(eccentric reducer, not concentric). This up-sloping prevents air pocketing in the top of the pipe... which air might be drawn into the pump and cause loss of suction. Maintenance Rule No. 5: any down-sloping towards the pump in suction piping (as exaggerated in the diagrams below) should be corrected.

A flexible coupling will not compensate for misalignment.

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We contain and direct the spinning liquid with a casing.... We got a little ahead of our story on the page before... because we didn’t yet have a casing to which the suction piping bolts. And the manner in which it is attached is of great importance. Maintenance Rule No. 6: see that piping puts absolutely no strain on the pump casing. When the original installation is made, all piping should be in place and self-supporting before connection. Openings should meet with no force. Otherwise the casing is apt to be cracked... Or sprung enough to allow closely-fitted pump parts to rub. It’s good practice to check the piping supports regularly to see that loosening, or settling of the building, or plant structure, hasn’t put strains on the casing.

_____________________________________________________

Now our pump is almost complete, but it would leak like a sieve... We’re far enough along now to traèe the flow of water through our pump. It’s not easy to show suction piping in the cross-section view below, so imagine it stretching from your eye to the lower centre of the pump. Our pump happens to be a “double suction” pump, which means that water flow is divided inside the pump casing... reaching the eye of the impeller from either side. As water is sucked into the spinning impeller, centrifugal force causes it to flow outward... building up high pressure at the oUtside of the pump (which will force water out) and creating low pressure at the centre of the pUmp (which will suck water in.) This situation is diagrammed in the upper half of the pump illustrated on page 12. So far so good... except that water tends to be sucked back from pressure to suction through the space between impeller and casing---as diagrammed in the lower half of the pump illustrated on page 12,---and our next step must be to plug this leak, if our pump is to be very efficient! Never allow a pump to run dry. Page 11 of 31 PP(T)SB PROPRIETARY DOCUMENT

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So we add wearing rings to plug internal liquid leakage... You might ask why we didn’t build our parts closer fitting in the first place---instead of narrowing the gap between them by inserting wearing rings, The answer is that those rings are removable and replaceable... when wear enlarges the tiny gap between them and the impeller, (Sometimes rings are attached to impeller rather then casing, or rings are attached to both so they face each other.) Maintenance Rule No.7: Never allow a pump to run dry (either through lack of proper priming when starting or through loss of suction when operating). Water is a lubricant between rings and impeller. Maintenance Rule No. 8: examine wearing rings at regular intervals. When seriously worn, their replacement will greatly improve pump efficiency.

A pump that often loses its prime is soon past its prime.

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CENTRUFUGAL PUMPS - BASIC © MYGAZEE ‘99

To keep air from being sucked in, we use stuffing boxes... We have two good reasons for wanting to keep air out of our pump: 1) we want to pump water, not air; 2) air leakage is apt to cause our pump to lose suction. Each stuffing box (fig. 13) we use consist of a casing, rings of packing and a gland at the outside end. Maintenance Rule No. 9: Packing should be replaced periodically---depending on conditions---using the packing recommended by your pump manufacturer. Forcing in a ring or two of new packing instead of replacing worn packing is bad practice. It’s apt to displace the seal cage (see next page). Put each ring of packing in separately, seating it firmly before adding the next. Stagger adjacent rings so the points where their ends meet do not coincide. Maintenance Rule No. 10: never tighten a gland more than necessary.... as excessive pressure will wear shaft sleeves unduly. Maintenance Rule No. 11: if shaft sleeves are badly scored, replace them immediately... Or packing life will be entirely too short.

A lot depends on packing—- watch it!

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To make packing more air-tight, we add water seal piping... In the centre of each stuffing box is a “seal cage”. By connecting it with piping to a point near the impeller rim, we bring liquid under pressure to the stuffing box. This liquid acts both to block out air intake and at the same time to lubricate the packing. It makes both packing and shaft sleeves wear longer.., providing it’s clean liquid! Maintenance Rule No. 12: if the liquid being pumped contains grit. a separate source of sealing liquid should be obtained (e.g.. it may be possible to direct some of the pumped liquid into a container and settle the grit out). To control liquid flow, draw up the gland just tight enough so a thin stream flows from the stuffing box during pump operation.

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CENTRUFUGAL PUMPS - BASIC © MYGAZEE ‘99

Discharge piping completes the pump installation---and now we can analyse the various forces we’re dealing with.... SUCTION: At least 75% of centrifugal pump troubles trace to the suction side. To minimise them... 1. Total suction lift (distance between the centre line of pump and liquid level when pumping, plus friction losses) generally should not exceed 15 feet. 2. Piping should be at least a size larger than pump suction nozzle.

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3. Friction in piping should be minimised.... use as few and as easy bends as possible... Avoid scaled or corroded pipe. (The attached table I and II extracted from the standard edition of the Hydraulic Institute show friction losses). DISCHARGE: Lift, plus suction lift, Plus friction in the piping from the point where liquid enters the suction piping to the end of the discharge piping equals total head. Pumps should be operated near their rated heads. Otherwise, pump is liable to operate under unsatisfactory and unstable conditions which reduce efficiency and operating life of the unit. Note the description of “cavitation” next page--and directions for figuring the head of your pumps are working against on the following pages. PUMP CAPACITY: Generally (per API standard) they are measured in gallons per minute. A new pump is guaranteed to deliver its rating in capacity and head. But whether a pump retains its actual capacity depends to a great extent on its maintenance. Wearing ring must be replaced when necessary-- to keep internal leakage losses down. Friction must be minimised in bearings and stuffing boxes by proper lubrication.., and misalignment must not be allowed to force scraping between closely-fitted pump parts. POWER: Power of the driving motor, like capacity of the pump, will not remain at a constant level without proper maintenance. (If you use electric motor by all means please see your electrical specialist for advise and guide to care of Electric Motors!) Starting load on motors can be reduced by throttling or closing the pump discharge valve (never the suction valve!).., but the pump must not be operated for long with the discharge valve closed. Power then in converted into friction--- overheating the water with serious consequences.

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CENTRUFUGAL PUMPS - BASIC © MYGAZEE ‘99

MAINTENANCE RULES ON PUMPS... 1. IF THERE’S ANY DANGER THAT FOREIGN MATTER (STICKS, REFUSE, ETC.,) MAY BE SUCKED INTO THE PUMP--CLOGGING OR WEARING THE IMPELLER UNDULY—PROVIDE THE INTAKE END OF THE SUCTION PIPING WITH A SUITABLE SCREEN. 2. NEVER PUMP A LIQUID FOR WHICH THE PUMP WAS NOT DESIGNED. 3. KEEP THE RIGHT AMOUNT OF THE RIGHT LUBRICANT IN BEARINGS AT ALL TIMES, FOLLOW YOUR PUMP MANUFAGTURERS LUBRICATION INSTRUCTIONS TO THE LETTER. 4. SEE THAT PUMP AND MOTOR FLANGES / COUPLINGS ARE PARALLEL VERTICALLY AND AXIALLY AND THAT THEY ARE KEPT THAT WAY! 5. ANY DOWN-SLOPING TOWARD THE PUMP IN SUCTION PIPING SHOULD BE CORRECTED. 6. SEE THAT PIPING PUTS ABSOLUTELY NO STRAIN ON THE PUMP CASING. 7. NEVER ALLOW A PUMP TO RUN DRY (EITHER THROUGH LACK OF PROPER PRIMING WHEN STARTING OR THROUGH LOSS OF SUCTION WHEN OPERATING). WATER IS A LUBRICANT BETWEEN RINGS AND IMPELLER. 8. EXAMINE WEARING RINGS AT REGULAR INTERVALS. WHEN SERIOUSLY WORN, THEIR REPLACEMENT WILL GREATLY IMPROVE PUMP EFFiCIENCY. 9. PACKING SHOULD BE REPLACED PERIODICALLY —DEPENDING ON CONDiTIONS--USING THE PACKING RECOMMENDED BY YOUR PUMP MANUFACTURER. 1O. NEVER TIGHTEN A GLAND MORE THAN NECESSARY...AS EXCESSIVE PRESSURE WILL WEAR SHAFT SLEEVES UNDULY. 11. IF SHAFT SLEEVES ARE BADLY SCORED. REPLACE THEM IMMEDIATELY.. OR PACKING LIFE WILL BE ENTIRELY TOO SHORT. 12. IF THE LIQUID BEING PUMPED CONTAINS GRIT, A SEPARATE SOURCE OF SEALING LIQUID SHOULD BE OBTAINED. (eg: IT MAY BE POSSIBLE TO DIRECT SOME OF THE PUMPED LIQUID INTO A CONTAINER AND SETTLE THE GRIT OUT).

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How to figure Head 1. Determine Total Static Lift: You can measure the vertical height liquid is raised with tape measure or pressure gauge. Most pumps - have 1/4” or 1/2” tap holes at the pump nozzle flanges to permit use of gauges. 2. Determine Total Friction Loss: Losses depend on size, condition and length of piping in system in relation to gallons pumped per minute (see Table below) - plus valves and fittings used. 3. Add Totals and Compare with Pump Rating: If total head (total static lift plus total friction loss) fails to correspond to pump rating, either a pump of different capacity should be substitute or friction losses should be reduced. Reduction can be obtained by use of larger pipe, hewer pipe, or elimination of high-loss fittings. 4. An Exception: Hot Water: Water turns into steam at 212 F (100 C) at atmospheric pressure, which factor sharply limits the suction lift possible with hot water. The danger is that low pressure at the pump impeller eye will cause water to flash into steam at considerably lower temperature. Thus hot water can’t be lifted very high- and extremely hot water can’t be lifted at all. In many cases, the pump must be considerably lower than the source of hot water... so water will flow in under pressure. Pressure required depends on temperature of water, pump capacity, and type of impeller used. Safest bet with a hot water problem is to consult your pump manufacturer. Figuring hp and Input (the traditional way) 1. Actual brake horsepower required to drive a centrifugal pump delivering liquid with a specific gravity of 1.0 can be obtained with the following formula.... US GALLONS PER MIN. x HEAD IN FEET _______________________________

= Brake hp

3960 X PUMP EFFICIENCY PERCENT 2. To figure kW input for a motor driven pump delivering liquid with a specific gravity of 1.0... US GALLONS PER MIN. x HEAD IN FEET ______________________________________

= kW lnput

5308 x PUMP EFFICIENCY x MOTOR EFFICIENCY

Where the liquid pumped is lighter or heavier than water, multiply by specific gravity for correction. The above is a guide only, refer to library’s latest published data for information. Page 28 of 31 PP(T)SB PROPRIETARY DOCUMENT

CENTRUFUGAL PUMPS - BASIC © MYGAZEE ‘99

A Maintenance Timetable “A” TIMETABLE, not “the” timetable! Operating conditions vary so widely that to recommend one schedule of preventive maintenance for all pumps especially in PP(T)SB would be absurd. So consider this timetable as relative. The recommended inspection and service operations may seem too frequent, however, they are based on the severest applications. At any rate, a routine maintenance schedule should be set up and followed to really get the best service out of any pumping equipment. EVERY YEAR Remove rotating element. Inspect thoroughly for wear-- order replacement parts where necessary. Check wearing ring clearances. Generally, they should be approximately no more than .003” per inch on diameter of wearing rings. Remove any deposit or scaling. Clean out water seal piping. Measure total suction and discharge lifts as test of pipe conditions. Record figures and compare them with figures of next test. Inspect foot valves and check valves--especially check calves, which must safeguard against water hammer when pump stops. EVERY 6 MONTHS Replace packing- use grade recommended by pump manufacturer. Be sure seal cages are centred in stuffing boxes at the entrance of water seal piping. Check shaft sleeves for scoring that would accelerate packing wear--order new sleeves if needed. Check alignment of pump and motor--shim up units as needed. If misalignment recurs frequently, inspect entire piping system. Unbolt piping at suction and discharge nozzles to see if it springs away-- indicating strain on casing. Check all piping supports for soundness and effective support of load. EVERY 3 MONTHS Drain lubricant, wash out oil wells and bearings with kerosene. In the case of sleeve bearings, check to see that oil rings are free to turn with the shaft. Repair or replace if defective. Refill with new lubricant recommended by your pump manufacturer. Measure bearings for wear, replace if excessive. Generally, allow 2 thousandth of an inch clearance--plus 1 thousandth of an inch for each inch or fraction of an inch of shaft journal diameter. EVERY MONTH Check bearing temperature with thermometer--not by hand. If anti-friction bearings are running hot, probably they have too much lubricant. Relieve. If sleeve bearings are running too hot, probably they need more lubricant. If change in lubrication doesn’t correct condition, disassemble and inspect bearing. Or possibly the alignment of pump and motor should be checked.

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BASIC PUMP CARE MANUAL

-Goulds Pumps-

t

Pump Care Manual

MYG

A collection of centrifugal pump maintenance articles

TABLE OF CONTENTS Title

Page

Installation . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 5 MYG Says You Need A Good Foundation . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 5 Field Alignment . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 5 Baseplate Grouting . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 6 Initial Alignment. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 7 Align Pipe Flanges to Pump Flanges . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 9 Raised Face and Flat Face Flanges (Mating Combinations) . . . . . . . . . . . . . . . . . . . . . . . 10 Pump Start-up . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 12 Turbine Driven Units . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 12 Keep Air Out of Your Pump. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 13 Air Pockets in Suction Pipe . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 13 Testing for Air in Centrifugal Pumps . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 14 Importance of Proper Suction Pipe Submergence . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 15 Waterfall Effect . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 16 Excessive Throttling Can Shorten the Life of Your Pump . . . . . . . . . . . . . . . . . . . . . . . . . . 16 Throttling Accelerates Erosion When Pumping Liquids Containing Abrasives . . . . . . . . . . 18 Elbows at Pump Suction Opening Can Cause Trouble Unless Precautions Are Taken. . . 18 Use a Thermometer to Measure Bearing Temperatures . . . . . . . . . . . . . . . . . . . . . . . . . . 20 NPSH . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 21 Packed Stuffing Box Packing . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 22 Are You Increasing Your Stuffing Box Packing Problems?. . . . . . . . . . . . . . . . . . . . . . . . . 23 Stuffing Box Packing . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 24 Stuffing Box Cooling for High Temperature Applications . . . . . . . . . . . . . . . . . . . . . . . . . . 25 Mechanical Seals Part I - The Basic Seal. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 26 Part II - Types . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 28 Part III - Selection . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 30 Part IV - Trouble Shooting . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 31 Quench Glands . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 32 Constant Level Oilers - Operation and Adjustment. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 33 Ball Bearing - Handling, Replacement and Maintenance Suggestions. . . . . . . . . . . . . . . . 34 Grease Lubrication . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 36 Flexibly Mounted Baseplates . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 37 Impeller Clearance . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 38 Pump Vibration Anaylysis . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 39 Pump Maintenance Records Pay Dividends. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 42 Goulds Offers A Complete Line of Pumps . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 44

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INSTALLATION Very often problems with pump and motor bearings, shaft breakage and difficulty in aligning units stem from the improper installation of the pumping equipment. Inadequate foundations, bent or twisted baseplates, and excessive pipe strain can all work together to make an otherwise simple installation an operational and maintenance headache for the life of the unit.

Here are some tips on how you can prevent these things from causing increased downtime and higher than expected maintenance costs.

YOU NEED A GOOD FOUNDATION The importance of a good foundation cannot be overemphasized. The foundation must be heavy enough to provide rigid support for the pump, motor and base combination and to resist the normal forces encountered when the unit is in service.

A concrete foundation firmly attached to a well designed floor system would be the most satisfactory arrangement. When site conditions make it necessary to support the pump on a metal structure, the structural members should be selected and designed to provide adequate stiffness and mass to support the unit and dampen any vibration that may be present. Unsupported wood floors and thin concrete pads should be avoided.

BASEPLATE INSTALLATION Re-alignment of the pump and driver is required prior to final grouting of the baseplate. Even though the unit may have been aligned at the factory prior to shipment, components tend to shift in transit, during handling, or because of uneven foundation surfaces. Factory alignments cannot be depended upon during or after installation. With the baseplate installed on the foundation you need to adjust the baseplate as necessary to level it and provide a flat surface for the pump and driver to rest. In some installations there are strict baseplate flatness guidelines which need to be adhered to. A good rule

of thumb is to level the base to within .005” per foot in all directions. Other types of installations may require the base to be level to within .002” per foot. This is usually done to avoid problems later with a condition known as “soft foot” and to assure that oil lubricated bearings receive equal and adequate lubrication. It is usually necessary to remove the pump and motor from the base to attain this level of flatness. The baseplate is leveled using shims or wedges at various locations around the base close to where the anchor bolts are located. Some bases may be equipped with leveling screws which make this (cont'd) 5

process somewhat easier. The wedges, shims or leveling screws are alternately adjusted as necessary to bring the base to within the specified limit. Once done the components are re-installed on the base.

At this point, before grouting the unit, it is important to check the pump and motor for “Soft Foot” and to check the “Rough Alignment” of the components. If these checks are not made now and the unit is grouted in place, any changes that need to be made later become much more difficult and expensive. Taking a few minutes at this point to check and correct any problems that exist may save many hours of work later.

Figure 1

BASEPLATE GROUTING Once the baseplate is properly installed, it must be grouted. Grout is a concrete material or multi-part epoxy compound that is used to fill the cast iron or fabricated steel baseplate so that when hardened, the baseplate and foundation

become a unit. We recommend the use of a non-shrink grout for this purpose. The grout is mixed to a water-like consistency, and when poured should be puddled to insure that it flows evenly throughout the under portion of the (cont'd)

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baseplate, completely filling the base. It is recommended that the cementatious grout be allowed to cure for at least 48 hours. Refer to the epoxy grout manufacturer's instructions for the proper mixing and application of that product. Refer to your equipment instruction manual for further details.

Figure 2

INITIAL ALIGNMENT The proper alignment of your pump and driver is important to assure trouble-free mechanical operation. Noisy operation, vibration, reduced bearing life, shaft and coupling failures and wasted energy may result from faulty alignment. Before proceeding, you must know the thermal rise estimates of your equipment. Refer to the pump and driver manuals for these estimates. The most common methods of measuring misalignment are: Straight Edge and Feeler Gauges – This is the easiest and least expensive method of doing alignment but is also the least accurate. Used primarily for very small pump/motor combinations where there is not enough room to use more accurate, but larger alignment systems. The straight edge is laid across the flanges of the coupling hub and the feeler gauges are used between the faces of the coupling hubs. Shim changes are estimated and the alignment is attained through a process of “Trial & Error”.

through the use of a Straight Edge and Feeler Gauges. Dial Indicators – Used either singly or in pairs, as with the Reverse Indicator Method, the actual misalignment is measured by the indicator and through a few arithmetical calculations shim changes are determined. Lasers – Somewhat more complicated to set up but can be more accurate if properly used. The laser especially lends itself to aligning shafts that are separated by more than a few inches. Since the machines have the capability of calculating the shim changes required, once the operator becomes familiar with the set-up he or she can do the alignment of a pump / motor combination fairly quickly and accurately. The primary drawback of the Laser is its cost and, in some cases, its size, although continuous advances in technology are making the machines both smaller and less expensive.

Usually you cannot attain the equipment manufacturer's alignment specifications (cont'd) 7

Angular – With a non-spacer coupling, measure the space or gap between the hubs from top-to-bottom and from side-to-side using a feeler gauge. The measurements obtained should be very nearly the same. If they are not, shim or move the motor as necessary to satisfy the tolerance stated in your instruction manual for angular alignment. With spacer type couplings, a dial indicator should be used to obtain these readings. Mount the indicator base on the driver coupling hub so that the indicator reads on the face of the pump hub. Rotate both shafts and read the indicator at 4 points, 90° apart. The readings obtained will tell you how far out of alignment your equipment is from side-to-side as well as from top-to-bottom. Add or remove shims from under the driver to correct the alignment. The amount of shim adjustment needed depends on the amount of misalignment present and distance from the point of measurement to location of the shims.

Parallel — Using a straight edge rest the edge on top of the coupling hub and at 90° to the top. The amount of misalignment is shown by the gap present between the straight edge and the hub. On a spacer type coupling mount a dial indicator on the pump coupling hub so that the indicator reads on the motor coupling hub. Rotate both shafts and read the indicator at 4 points, 90° apart. The readings obtained will tell you how far out of alignment your equipment is from side-to-side as well as top-to-bottom (elevation). Correct the top-to-bottom alignment first by shimming the driver as needed one half the amount shown on the indicator. Correct the side-to-side alignment by moving the driver as needed, again, by one half the amount shown on the indicator. Recheck the top-to-bottom (elevation) to make sure that it did not change during the process. Be sure to allow for the thermal rise estimates of both pump and driver.

Figure 3

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ALIGN PIPE FLANGES TO PUMP FLANGES Never run pipe to the pump - always run pipe from the pump to a point several feet away where the final pipe connection can be made. This will help to minimize excessive pipe strain on the pump nozzles. The piping must be independently supported with an adequately designed pipe hanger system that will not allow the pump to carry the weight of the pipes or the liquid in them.

The system should also be designed to accommodate whatever thermal expansion or contraction is anticipated. After the piping has been made up, the alignment must be rechecked. By comparing this check with the initial alignment figures, the installer can determine the amount of pipe strain which has been put on the pump nozzles. The piping should be adjusted if there is any significant change in the alignment readings.

Figure 4

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RAISED FACE AND FLAT FACE FLANGES (MATING COMBINATIONS) Pumps of cast iron construction are furnished with 125 or 250 lb. flat face (F.F.) flanges. Since industry normally uses fabricated steel piping, the pumps are often connected to 150 or 300 lb. 1/16" raised face (R.F.) steel flanges. Difficulty can occur with this flange mating combination. The pump flange tends to pivot around the edge of the raised face as the flange bolts are tightened. This can cause the pump flange to break allowing leakage at the joint (Fig. 5). A similar problem can be encountered when a bronze pump with F.F. flanges is connected to R.F. steel flange (Fig. 6). Since the materials are not of equal strength, the bronze flange may distort, resulting in leakage. To avoid problems when attaching bronze or cast iron F.F. pump flanges to R.F. steel pipe flanges, the following steps should be taken (refer to Fig. 7): 1. Machine off the raised face on the steel pipe flange. 2. Use a full face gasket. If the pump is steel or stainless steel with F.F. flanges, no problem arises since materials of equal strength are being connected. Many customers, however, specify R.F. flanges on steel pumps for mating to R.F. companion flanges. This arrangement is technically and practically not required. The purpose of an R.F. flange is to concentrate more pressure on a smaller gasket area and thereby increase the pressure containment capability of the

joint. To create this higher gasket load, it is only necessary to have one-half of the flanged joint supplied with a raised face not both. The following illustrations show 4" steel R.F. and F.F. mating flange combinations and the gasket loading incurred in each instance. Assuming the force (F) from the flange bolts to be 10,000 lbs. and constant in each combination, the gasket stress is: P (Stress) = P (Fig. 8) =

Bolt Force (F) Gasket Area 10,000 lbs. = 49.4 sq. in.

203 psi

P (Fig. 9) = 10,000 lbs. = 630 psi P! (Fig. 10) = 15.9 sq. in. It can be readily seen that the smaller gasket, used with a raised face flange, increases the pressure containment capability of a flanged joint. However, it can also be noted that there is no difference in pressure capability between R.F. to R.F. and R.F. to F.F. flange combinations. In addition to being technically unnecessary to have a R.F. to R.F. mating combination, the advantages are: 1. The elimination of the extra cost for R.F. flanges. 2. The elimination of the extra delivery time required for a non-standard casing.

(cont'd) 10

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PUMP START-UP All of the factors that contribute to successful pump operation must be checked carefully before starting a new or rebuilt pump. The majority of failures or serious problems at start up can be avoided by just taking a few minutes to verify that all of the components and associated systems are in operational condition. Prior to coupling the unit, two checks which should always be made are: first, the unit turns freely by hand; and, second, the motor has been “bumped” and checked for the correct hand of rotation. Operating some pumps in the wrong direction of rotation can severely damage the unit and can lead to personal injury for those nearby. A pump that does not turn freely, with consideration given to the drag in the stuffing box or mechanical seal area, usually requires some attention prior to starting. If the pump shaft turned freely during the initial phases of the alignment process then something since that procedure has changed. Going back through each of the steps that followed the alignment in reverse will usually turn up the cause.

Causes can include pipe strain imposed due to the weight of the liquid in the piping, faults in the foundation or grouting, foreign material getting into the unit before the piping flanges were made up, etc. Whatever the cause, it must be identified and corrected before proceeding further. Just prior to start up always make sure the correct type, quantity and quality of lubrication and / or cooling is applied to the components as recommended by the manufacturer. This includes motor bearings, couplings, mechanical seals, packing and of course, the pump bearings. Once these checks have been made the Suction Valve on the pump may be fully opened while the air inside the pump is vented. This fills the unit with liquid, covering the impeller eye and makes it ready for start up. If the unit is in a “Suction Lift” type installation where the level of the liquid on the suction side of the pump is below the centerline of the pump casing, a vacuum priming system may be needed to evacuate the air from the pump casing or volute. In the case of “Self Priming” pumps the casing needs to have some liquid added to it before hand for the “Self Priming” feature to work.

TURBINE DRIVEN UNITS When new or rebuilt turbine driven units are being started, it is important to protect the pump from damage related to an improperly adjusted turbine. All turbine controls should be checked and verified prior to coupling it to the pump. Consult the turbine manufacturer's instruction for details. When started, the unit should be brought up to speed and flow as quickly as possible in order to flush the new or

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recently opened piping system and minimize the chance of pump failure caused by small amounts of dirt or foreign material in the sytem. Suction strainers are used to provide some protection against this, but normally the openings are much larger than the running clearances of the rotating parts in the pump.

KEEP AIR OUT OF YOUR PUMP Conventional Centrifugal Pumps are not designed to handle a mixture of liquid and gases. Pumping liquids containing a significant amount of entrained gas can lead to serious mechanical and hydraulic problems. A mixture of only 2% gas by volume will cause a 10% reduction in capacity and 4% will cause a reduction of over 43%. In addition to the loss of efficiency and wasted power the pump will probably be noisy and may vibrate excessively. Entrained gases can cause shaft breakage, seal failures and in some cases accelerate corrosion.

Air may be present in the liquid due to leaky suction lines on suction lift applications or a variety of other reasons. A free falling discharge into a tank or pit will also cause excessive gas entrainment and may cause problems for a pump drawing from that tank. There are Centrifugal Pumps designed specifically for applications where entrained gases are encountered.

AIR POCKETS IN SUCTION PIPE Air pockets can be a source of trouble on pump installations involving suction lift. They can cause a loss of prime on start-up as well as restrict flow in the suction pipe to the extent that a reduction in capacity is experienced. When installing piping, suction lines should always pass under interfering piping and where reducers are used, eccentric rather than straight reducers must be used.

Eccentric reducers are always installed with the horizontal side on top. Suction valves should be installed with the stems horizontal so that no air pockets are formed at the top of the valve near the bonnet. Figure 11 shows both the correct and incorrect method of installing piping.

Figure 11

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TESTING FOR AIR IN CENTRIFUGAL PUMPS The amount of air which can be handled with reasonable pump life varies from pump to pump. However, in no case is it expected that a pump will give better life with air present than it would if the liquid were entirely air-free. The elimination of air has greatly improved the operation and life of many troublesome pumps. When trouble occurs, it is common to suspect everything but air, and to consider air last, if at all.

can be fed into the "bubble bottle". The presence of air or vapor will show itself in the "bubble bottle". Obviously, the next step is to eliminate the source of air since quantities present in sufficient amount to be audible are almost certain to cause premature mechanical failure.

If air is present, the pump is likely to operate with a certain amount of internal noise. This noise can be described as a "gravel noise" - sounds very much must as though the pump were handling water full of gravel. This is the same type of noise generally associated with cavitation. In many cases a great deal of time, inconvenience, and expense can be saved by making a simple test for the presence of air. We will assume that calculations have already been made to assure that the NPSH available is greater than that required by the pump (the noise is not a result of cavitation). The next step should be to check for the presence of entrained air in the suction. When the source of suction supply is below the centerline of the pump, check for the conditions covered previously in this manual. When the source of suction supply is above the centerline of the pump, a check for air leaks can be made by collecting a sample in a "bubble bottle" as illustrated in Figure 12. Since the pressure at the suction chamber of the pump is above atmospheric pressure, a valve can be installed in one of the tapped openings at the high point in the chamber and liquid

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Figure 12

IMPORTANCE OF PROPER SUCTION PIPE SUBMERGENCE If velocity of the water at the suction pipe entrance is too high, a vortex (commonly referred to as a whirlpool) is created, through which air flows downward to the end of the suction pipe and into the pump casing. Air in the pump casing causes rough and noisy operation and severe vibration that in the course of a short time can result in a broken shaft as well as other damage. If sufficient air reaches the pump to completely air bind it, dry operation will cause metal seizure at the impeller hub and wearing ring. Unless corrective action is taken, the rotating parts will be damaged beyond repair. These difficulties can be avoided by selecting and installing a suction pipe through which the fluid handled enters at a velocity not exceeding that shown on the graph. A lower velocity and more submergence are good insurance against this type of air trouble.

In many cases, centrifugal pumps function improperly because air gets into the pump due to: 1. The installation of a suction pipe that is too small in diameter. 2. The end of the suction pipe not being submerged deeply enough. 3. Both of the above undesirable conditions.

Sump design for large double suction pumps can become more complex, and appropriate sources such as the "Hydraulic Institute Standards" should be consulted.

Figure 13

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WATERFALL EFFECT Liquid falling directly on the foot valve in Figure 14 would carry air into the suction line and could result in the pump losing its prime and seizing while running dry. This can also happen when a pump has a flooded suction as in Figure 15. In either case, the pump would be noisy. The free fall shown in Figures 14 and 15 is serious, but a line discharging vertically downward under pressure, such as an orifice return line to a feed water storage tank, has even more effect, as shown in Figure 16.

If the returned liquid is heavily laden with air, or the sump is small, baffles may also be required as shown in Figure 14 to allow the air to separate before entering the pump suction. The continued presence of entrained air and vapor in any liquid being pumped is extremely serious and should be corrected immediately. Any return line should be extended far enough away from the pump suction to prevent trouble, and below the surface of the water to minimize air entrainment. Check your installation today and make sure YOUR pump is not suffering from entrained air.

EXCESSIVE THROTTLING CAN SHORTEN THE LIFE OF YOUR PUMP A centrifugal pump should never be operated continuously near shutoff or zero capacity. To do so may shorten the life of the pump and greatly increase down time and maintenance.

allowed through the pump, the casing may be unable to radiate the heat generated, and the liquid and the pump may rise to a dangerously high temperature.

The difference between input horsepower and water horsepower is transferred to the liquid in the pump as heat. When only a small percentage of rated flow is

Hydraulic radial thrust is unbalanced when the pump is operating near shut-off, and this subjects the shaft to abnormal deflection. The pump will be noisy, will (cont'd)

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vibrate excessively, and may break shafts frequently. Pumps are frequently rated for capacities sufficient to handle maximum requirements. They are also occasionally selected to handle capacities required under emergency operation or at some future predicted flow far in excess of present demands. This procedure will minimize ultimate plant investment, duplication, and replacement with larger equipment at a future date. However, it may require excessive throttling of the discharge valve, resulting in severe punishment to the pump. Fortunately, there is a simple method to relieve the pump of undue strain. Simply extend a by-pass line from the pump discharge back to the source of supply. A throttle valve or an orifice plate should be placed in the by-pass line, and sufficient flow returned to allow the pump to operate at a capacity reasonably near its rating.

The bypassed liquid should never be returned to the suction line immediately upstream from the pump, but should always be returned back to the source of supply and discharged below the liquid level to avoid air entrainment. If excessive throttling at the discharge valve is required with the by-pass line open, actual system head and capacity requirements should be reviewed and a new rating selected. Excessive maintenance and shortened life can always be expected when a pump is operated continuously at or near shut-off. A return line by-passing part of the flow back to the source of supply will allow the pump to operate near rated capacity and extend the life of your pumping equipment.

Figure 17

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THROTTLING ACCELERATES ERROSION WHEN PUMPING LIQUIDS CONTAINING ABRASIVES When pumping liquids containing abrasive solids some errosion of the pump impeller and other parts is to be expected. The useful life of the unit is reduced not only by the quantity of abrasives present, but also the physical characteristics of the solids and the velocity of the mixture through the pump. Throttling a unit accelerates the wear even more by causing localized higher velocities and re-circulation within the impeller and casing. With a pump operating at its best efficiency point, the majority of solids make only one trip through the unit. Throttling the unit increases the internal re-circulation and allows the abrasive particles to scrape against the inside of the pump a number of times before finally being discharged.

Often a plant design calls for pumps to be oversized to accommodate future requirements or the occasional upset condition. Again, a discharge by-pass line should be considered to help allow the pump to operate at a better point on the curve and at better efficiency and reduced internal re-circulation. Often the re-circulation line is directed back into the tank or sump to provide additional agitation and mixing.

ELBOWS AT PUMP SUCTION CAN CAUSE TROUBLE UNLESS PRECAUTIONS ARE TAKEN Properly installed suction piping is of extreme importance in obtaining trouble-free operation in any double-suction type centrifugal pump. If an elbow is required at the pump suction flange, it should always be in a vertical position. Where, because of space limitations or other peculiarities of the installation, an elbow must be used near the pump suction flange in any position except vertical, it should never be installed unless there is a minimum of two diameters of straight pipe between the elbow and the pump flange (see Figures 18 and 19).

Without this straight run of pipe, the incoming liquid, because of centrifugal force, tends to pile up on the outside of the elbow. If the liquid enters the pump suction directly in this condition, it will not be evenly distributed between the two inlets of the double suction impeller. This unbalance will cause noise and generally unsatisfactory operation of the pump (see Figure 20). A device for relieving this condition is illustrated in Figure 21. It is a simple baffle that can be made in any shop. Inserted in the straight length of pipe adjacent to the pump flange, it will

(cont'd) 18

equalize the flow and assist materially in improving pump performance.

Figure 18

Figure 19

Figure 20

Figure 21

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USE A THERMOMETER TO MEASURE BEARING TEMPERATURES Pump bearing temperatures are frequently estimated by placing a hand lightly against the bearing housing or shell. If the bearing is “cool” or “warm,” we feel assured. If it feels “hot,” we may become concerned and spend considerable time and effort to reduce the temperature until the housing feels only “warm” without any clear idea of the actual operating temperature. Unfortunately, the human hand is not an accurate thermometer and can give a danger signal falsely. A temperature which feels “hot” varies from 120° to 130°F depending on the individual. Above this temperature the human hand is worthless in estimating temperature. Grease lubricated ball and roller bearings can be operated safely at temperatures up to at least 200°F. In fact, the upper operating limit on anti-friction bearings is determined solely by the temperature at which the lubricant fails and begins to carburize. Bearing temperatures up to 160°F are extremely safe. Operation of bearings at this temperature is not at all undesirable, as a better flow of lubricant can be expected. This gives a clear indication of why a thermometer is necessary to determine bearing temperature. All bearings operate at some temperature above that of the surrounding atmosphere, unless cooled. Heat is generated within the bearing due to rolling friction of the balls or rollers, and by the drag of the ball or roller separator cage. Some heat may be added by conduction along the pump shaft from the liquid within the pump and from external sources in close proximity. The amount of heat

which can be dissipated is dependent on the cooling area of the bearing housing and the temperature and motion of the surrounding air. The net result is a stable operating temperature. Once this temperature has been established, it will remain constant until one or more of the variables changes. A stable temperature, no matter how hot it may feel to the human hand, is not necessarily an indication of danger so long as it does not exceed the upper limit of the lubricant. The temperature should be established accurately by thermometer and recorded in a convenient location. A pronounced increase in temperature is an indication of danger and a signal to investigate. One shot of grease should be added to the bearing, but if this does not reduce the temperature immediately, no additional grease should be added. The unit should be checked for unnecessary loads, such as coupling misalignment, or improper packing adjustment. A temperature increase may not be an indication of impending bearing failure or excessive load, but could be due entirely to an increase in temperature of the liquid being pumped, or to an increase in the surrounding temperature during the warm summer months. Heat transfer due to increased liquid temperature can be minimized by connection of a quenching gland with cold water from the plant supply line. Bearing temperatures are frequently increased by lubrication practices. Excessive or over-greasing which results in complete packing of the bearing shell with grease greatly increases the resistance to rotation and the amount of (cont'd)

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heat generated. Removal of excessive grease from the bearing housing will lower the bearing temperature considerably. Most pumps handling hot liquids are equipped with quenching glands and water cooled bearings so that the actual operating temperature within the bearing may be lower than the ambient temperature of the surrounding air. High temperature greases should not be used with water cooled bearings as poor grease fluidity and inadequate lubrication may likely result.

Occasionally when pumps are first started the bearings seem to run extremely hot. This high temperature is frequently caused by the grease seals not the bearings. As soon as the seals are seated, the temperature will drop to a normal level.

NPSH The most maligned of all pump criteria is actually a simple physical phenomenon. NPSH, net positive suction head, is defined by the Hydraulic Institute as “the total suction head in feet of liquid absolute, determined at the suction nozzle and corrected to datum, less the vapor pressure of the liquid in feet absolute.” The key to the problem is vapor pressure. Every liquid exerts a pressure on its surroundings which is dependent on temperature. When this pressure is equal to the pressure of its environment, the liquid is said to boil or turn to vapor. For example, at sea level a pan of water on a stove will boil at 212°F. This means that the vapor pressure of water at 212°F is equal to the atmospheric pressure 14.7 psia. The amount of space a liquid occupies is very small compared to the amount of space occupied by the equivalent amount of liquid converted into vapor. For example, 1 pound of water at 50° F occupies a space equal to about a pint (16 oz.) of soda. One pound of water vapor at 50° F would occupy a space of over 1,700 cubic feet, nearly the size of a 15’ x 15’ x 8’ room. One pound of water

vapor occupies about 10,000 times the volume of an equal mass of water. Figure 22 is a profile and end view of a typical open impeller. An impeller is essentially a spinning wheel and as such centrifugal force creates a low pressure area at the center point “A”. In addition, pressure is lost due to liquid friction between the suction nozzle and the entrance to the vane. One other source of pressure loss is due to the shock of the vane hitting into the liquid stream.

Figure 22 (cont'd) 21

If the pressure at point “A” is lower than the vapor pressure of the liquid being pumped, the liquid boils or vaporizes. When this occurs, the vapor occupies many times the volume of the liquid, consequently the performance in terms of mass flow is greatly decreased. When the vapor enters the vane area, condensation or collapse of the vapor bubble occurs as soon as the pressure is higher than the vapor pressure. Since the pressure is generated by the vane working on the liquid or vapor, the collapse always happens at the surface of the vane. The crackling noise associated with cavitation is this collapse taking place.

When the vapor bubbles collapse on the vane surface, energy is released in the form of an implosion which will take particles of metal out of the surface. This causes the familiar pitting of the vane. Thus, NPSH required is simply the amount of pressure necessary to keep the liquid from vaporizing at the “eye” or center of the impeller. It is made up of the losses due to friction and shock plus the natural pressure reduction due to centrifugal force.

PACKED STUFFING BOX LUBRICATION Shaft sealing is the necessary evil to centrifugal pump operation. Mechanical seals and packing are the two methods used to accomplish a seal. It must be understood that while these sealing devices act in different ways to do the same job, there is work being done which generates friction and heat. Because of this, they require both cooling and lubrication. Packing does not and should not completely eliminate leakage, but rather throttles it to a tolerable limit. There should always be some drippage from the gland for good packing lubrication. The choice of stuffing box lubricant depends on the pumpage, temperature and pressure. Obviously, if the liquid pumped is a poor lubricant or contains abrasive solids, the packing will be severely damaged. Should the liquid be at or near its vapor pressure, the heat

generated in the bearing surface area may cause flashing with the resultant loss of lubricating properties. Figure 23 illustrates the most common method of supplying lubrication to the packing. The pumpage is simply by-passed from the discharge volute into the stuffing box. This is used when the liquid and operating conditions are such that the pumpage will provide adequate lubrication. Part A in Figure 23 is the cage ring. Its function is to provide an annular channel for even distribution of the lubricant. Figure 24 illustrates a method of lubrication used when the pumpage is heavy or thick, such as paper stock. In this case, the cooling fluid is supplied to a cage or lantern ring which is placed in the back of the stuffing box. This method prevents the pumpage from entering the throat bushing and will minimize wear and clogging. The cooling fluid is forced into (cont'd)

22

the stuffing box at a higher pressure than is in the area between the back of the impeller and the stuffing box with a portion entering the pumpage through the throat bushing. The remainder of the cooling fluid is forced along the shaft to cool and lubricate the packing. Care must be exercised in the choice of external lubricant. The primary concern will be compatibility with the pumpage. In most cases water will suffice for lubrication, but serious product damage may result if a chemical reaction occurs.

Figure 23

If water is undesirable, grease oil or a liquid with good lubricity and chemical compatibility should be selected. Figure 24

ARE YOU INCREASING YOUR STUFFING BOX PACKING PROBLEMS? Normal stuffing box leakage usually ranges from 40 to 60 drops per minute which is adequate to cool and lubricate the packing. Stuffing boxes for pumps with very large shaft diameters may need to leak as much as a small stream to provide sufficient cooling and lubrication. On units handling clear liquids, the pumpage itself becomes the coolant and lubricant. A seal flush line is piped up from a point on the pump casing to the seal water inlet. As the unit is operated, pumpage flows naturally to the stuffing box. An external sealing liquid is needed if the pressure at the stuffing box is negative, as in a Suction Lift situation, or if the liquid being pumped has suspended solids, or is highly volatile with poor lubricating qualities. If the pressure at the stuffing box is negative, air in the atmosphere will be

drawn into the packing. If enough air enters the pump through the packing it is possible for the unit to lose its prime or seize the pump altogether. A liquid barrier such as seal water, helps to prevent this. The sealing liquid is piped externally to the stuffing box and distributed between the rows of packing through a seal cage or lantern ring. Some of the sealing liquid will be drawn into the pump, and the remainder will trickle out of the stuffing box. In this case, the sealing liquid acts as a coolant, a lubricant and a barrier against the entrance of air as well. If the liquid pumped has solids in suspension, the stuffing box should be sealed with clean water from the plant water supply or some other source. The use of plant water assures a positive seal on the stuffing box during the priming period. (cont'd) 23

If the pressure at the stuffing box is positive and the liquid is clean, the natural flow of liquid trying to escape to the atmosphere will cool and lubricate the packing. In this case it is not necessary to connect either a flush line from the pump casing or an external supply of seal water. The additional pressure of the supply will increase the pressure at the stuffing box unnecessarily and require the packing to be tightened further to control the leakage. If the pumpage contains suspended solids special consideration needs to be given to both the type of packing being used and the supply of seal water. If allowed to leak, the solids in the pumped liquid will collect on the rings of packing and score the shaft sleeve. External sealing with clean plant water is recommended.

The pressure of the seal water supply, at the stuffing box connection, should be between 5 and 10 PSIG above the pressure acting on the stuffing box. This will insure a flow of liquid along the shaft sleeve and into the pump helping to prevent the entrance of solids. If the pumpage is toxic, volatile or cannot be contaminated, a non-soluble lubricant may be introduced at the seal cage or lantern ring. A quench type gland is often used to dilute and carry away any leakage of the pumped liquid which might escape past the packing. Recommendations for stuffing box sealing liquids are tabulated below. They should be used as a guide for new pump installations and when reviewing existing installations.

STUFFING BOX PACKING Packing is generally a rope-like material containing some type of lubricant, usually grease and graphite, Teflon™, mica or some other lubricating substance. The initial impregnation is intended to serve as the primary lubricant for the start-up and break-in period. After start-up other lubrication must be provided to prolong packing life. This lubrication is most often provided by the pumpage.

grease and graphite flakes. This packing is best suited for low speed, low temperature and low pressure applications on cold water and general services. It has a pH range of 4 to 10 which enables it to withstand mildly acidic or alkaline solutions. It’s the least expensive of the packing types available.

Packing supplied by ITT Industries Fluid Technologies can be broken down into two primary classifications: general and special.

A. Chemical and Solvent Applications

General Packing supplied for general service applications is braided, long filament vegetable fiber material impregnated with

Special

A wide variety of packing is available for severe chemical or solvent applications. Typical types include Teflon™ impregnated vegetable fiber packing, pure Teflon™, Grafoil™ and pure graphite, Kevlar™, lead packing, various combinations of these and other materials. Your choice of packing should (cont'd)

24

be based on the nature and concentration of the chemicals or solvents being pumped. B. High Pressure or Temperature Applications Packings of varied materials and arrangements can be supplied for most applications involving high pressures and/or temperatures. Combination sets of various types of foil, plastic or braided natural and man-made fibers are common for use with these applications. Your choice should be based on shaft speeds and the severity of the pressures and temperatures involved.

Often consultation with your ITT Industries Sales Representative or local packing supplier can provide alternative solutions to what may be considered an impossible situation.

Goulds Model 3196 with non-cooled stuffing box w/quench type gland. Standard configuration with 5 rings of packing and lantern ring illustrated.

Figure 25

STUFFING BOX COOLING FOR HIGH TEMPERATURE APPLICATIONS High temperature applications present certain inherent problems which do not arise when pumping lower temperature fluids. Special attention must be given to the stuffing box and packing arrangement to assure a long and useful life. In centrifugal pumps fitted with packing, there are high temperature options available, such as jacketed stuffing boxes, to help extend the life of the packing. As the temperature and pressures increase or as the pumpage becomes more volatile, corrosive or toxic, packing becomes more of a problem. Since stuffing boxes need to leak to provide the cooling and lubrication for the packing, this leakage can cause serious safety and health issues in the plant. In some applications the few drops of leakage can vaporize harmlessly into the atmosphere. In others, there can be no leakage to the environment at all.

Because they eliminate this leakage, mechanical seals have largely replaced packing for these sorts of applications. However, when using mechanical seals, care must be taken to be sure that there be liquid film between the seal faces at all times. If the liquid separating the seal face surfaces were to flash to vapor and allow the faces to touch, seal failure is nearly guaranteed. There are several ways to help keep the seal faces and sealing fluid cool: 1. Cooling liquid can be circulated through a Jacketed stuffing box which is an option for many of our models, including the 3196. The chart on the next page shows the cooling water required for various application temperatures. 2. A cool compatible liquid from an outside source can be injected directly into the seal chamber. (cont'd) 25

3. The pumpage can be cooled in a heat exchanger and returned to the seal chamber. Either (2) or (3) can be used by themselves or in combination with (1).

Figure 26

TYPE 5 Jacketed BigBore™ Seal Chamber

TYPE 4 TaperBore™ PLUS Seal Chamber

Maintains proper temperature control (heating or cooling) of seal environment.

Figure 27

26

MECHANICAL SEALS . . . PART I - THE BASIC SEAL A mechanical seal is a sealing device that provides a seal between rotating and stationary parts, thus keeping the liquid being pumped inside the pump. Today, the design of liquid handling equipment with rotating parts usually include the features necessary for the use of a mechanical seal. Some of the advantages mechanical seals offer over conventional packing include: 1. Reduced friction and power losses because of lower drag in the stuffing box. 2. Reduced leakage from the stuffing box.

O-ring to effect sealing and prevent rotation. 2. Rotation seal part-positioned on the shaft by the O-ring. The O-ring seals between it and the shaft and provides resiliency. 3. The mating faces - The faces are precisioned lapped for a flatness of 3 light bands and a surface finish of 5 microinches. 4. Spring Assembly - Rotates with the shaft and provides pressure to keep the mating faces together during periods of shut down or lack of hydraulic pressure.

3. Reduced shaft or sleeve wear. 4. Reduced maintenance. The wide variety of styles and designs together with extensive experience allows the use of seals on practically any sealing problem. A mechanical seal must seal at three points: 1. Static seal between the stationary part and the housing. 2. Static seal between the rotary part and the shaft. 3. Dynamic seal between the rotating seal face and the stationary seal face. Figure 28 shows a basic seal with these components: 1. Stationary seal part-positioned in the housing with preload on the

5. Driving member positions the spring assembly and the rotating face. It also provides the positive drive between shaft and the other rotating parts. 6. End movement - As wear takes place between the mating faces, the rotating face must move along the shaft to maintain contact with stationary face. The "O" ring must be free to move. The principle of operation of a mechanical seal is fairly straight forward: Liquid pressure in the seal chamber along with some type of spring arrangement forces the faces together. Due to capillary action, a thin film of lubricant separates them and keeps them from touching under normal circumstances. The pressure of the liquid, the spring pressure and this very thin layer of liquid between the faces reduces the leakage of the pumped fluid past the seal to a point near zero. These basic components are a part (cont'd) 27

of every seal, secondary sealing systems consisting of “O” rings, “V” rings, gaskets, etc. complete the assembly.

The form, shape, style and design will vary greatly depending upon the service and manufacturer. The basic theory, however, remains the same.

Figure 28

MECHANICAL SEALS . . . PART II - TYPES Mechanical seals can be classified into the general types and arrangements shown below. Understanding these classes provides the first step in proper seal selection.

Double seals Unbalanced or balanced

The pumpage is in direct contact with all parts of the seal and provides the lubrication for the faces. The full force of pressure in the box acts on the faces providing good sealing to approximately 200 PSIG. This is the most widely used type for services handling clear liquids. A circulation or by-pass line connected from the volute to the stuffing box provides continual flushing of the seal chamber.

Single Seals, Inside Unbalanced (Figure 29)

Single Seals, Outside Unbalanced (Figure 30)

The single inside seal mounts on the shaft or sleeve within the stuffing box housing.

This type mounts with the rotary part outside of the stuffing box. The springs and drive element are not in contact with

Single Seals Inside, outside, unbalanced, balanced

(cont'd) 28

the pumpage, thus reducing corrosion problems and preventing product accumulation in the springs. Pressures are limited to the spring rating, usually 35 PSIG. Usually the same style seal can be mounted inside or outside. The outside seal is easier to install, adjust and maintain. Single Seals, Balanced (Figure 31) Balancing a seal varies the face loading exerted by the box pressure, thus extending the pressure limits of the seal. A balanced rotating part utilizes a stepped face and a sleeve. Balanced seals are used to pressures of 200 PSIG. Their use is also extensive on light hydrocarbons which tend to vaporize easily. Balanced outside seals allow box pressure to be exerted toward the seal faces, thus allowing pressure ranges to above 150 PSI as compared to the PSIG limit for the unbalanced outside seal.

Double Seals Double seals use two seals mounted back to back in the stuffing box. The box is pressurized with a clear liquid from an outside source. This arrangement is often used on slurry service or where there can be no release of the pumpage to the environment because the condition of the inner seal can be monitored by checking the characteristics of the seal flush liquid. Balanced and unbalanced designs are available to meet specific pressures. Cartridge Seals Recent advancements in sealing technology have made Cartridge Type Seals the preferred type for many applications. The advantages include that they are available in all of the previously mentioned configurations, they are self contained, completely assembled on their own shaft sleeve, can accommodate pump impeller adjustments and are significantly easier to install. The problems of having to assemble the pump, mark the seal chamber location on the shaft or sleeve, disassemble the unit, install the seal and re-assemble the pump are eliminated. A further advantage is that you do not need to handle the delicate sealing faces or other components.

29

Manufacturers offer different configurations for different applications. A seal chamber suitable for a clear liquid application maybe totally unacceptable for a viscous application or one containing abrasive solids. Usually as the effectiveness of the seal chamber in cooling and lubricating the seal goes up, so does the cost of that design. That added cost can easily be more than offset by a reduction in the number of seal failure in the pump’s life time. Figure 32 Split Seals Sealing technology has advanced even further in the area of Split Seals. Split seals are usually a cartridge type seal with the added advantage of being split horizontally at the centerline. Since the seal can be separated into halves, placed around the shaft and reassembled, taking the pump apart to install it is unnecessary. The primary disadvantages are that the seal components need to be very accurately aligned and, since you are handling many of the individual parts of the seal, extra care needs to be taken to prevent damaging any of the components. Seal Chambers Seal life can be significantly extended by the proper selection of the seal environment. By providing better cooling and lubrication by increasing the circulation near the seal faces, the life of the seal is greatly increased. The design of the chamber can be as simple as just increasing the bore diameter in the area of the seal to increase the liquid capacity or as complicated as adding a taper to the sides and including ribs, vanes or fins to further induce circulation around the faces.

30

Tape Bore VPE

Figure 33

MECHANICAL SEALS . . . PART III - SELECTION The proper selection of a mechanical seal can be made only if the full operating conditions are known. These conditions are as follows: 1. Liquid 2. Pressure 3. Temperature 4. Characteristics of liquid 5. Pump model Liquid – Identification of the exact liquid to be handled provides the first step in seal selection. The metal parts must be resistant to the effects of the pumpage. These parts are available in steel, bronze, stainless steel, Hastelloy and a wide variety of other alloys to meet specific needs. The seal faces must be able to resist both corrosion and wear. Carbon, various types of ceramics, Stellite, silicon carbide and tungsten carbide are a few of the variety of materials which offer both excellent wear properties and corrosion resistance.

Temperature – The temperature will, in part, determine the material of the other sealing members. Synthetic rubbers are used to approximately 250°F, Teflon™ to 500°F. and other, more exotic elastomers to 750°F and above. Cooling the liquid in the seal chamber by use of cooling jackets or heat exchangers to cool flushing liquid , often extends seal life and allows for a wider selection of materials. Characteristics of Liquid – Liquids containing abrasive solids can create excessive wear and shorten seal life. Double seals, very hard faces and clear liquid flushing from an external source allows the use of mechanical seals on these difficult applications. On light hydrocarbons, balanced seals are often used even though pressures are low to promote longer seal life. Pump Model – Mechanical seal selection, installation and application varies with the pump model. Not all seals fit all pumps. Seal chamber configuration, pump discharge pressures and temperatures may not present ideal design conditions for the seal. Consult your ITT/FTC Pump Sales Engineer for selection of the proper pump and seal for the service.

The elastomers in the seal can be made of any number of different compounds including Teflon™, Viton, Buna N which completes the proper materials selection. Pressure – the proper type of seal configuration, unbalanced or balanced, is based on the pressure that the seal will be subjected to. Unbalanced seals are used in applications with pressures up to approximately 200 PSI. Balanced seals are usually required for pressures above 200 PSI. (cont'd) 31

MECHANICAL SEALS … PART IV - TROUBLESHOOTING When discussing mechanical seal problems, the following analogy may be made: Mechanical seals behave very much like bearings. Conditions that lead to bearing failure also lead to mechanical seal breakdown. Let’s consider these conditions. Misalignment - One of the main causes of mechanical seal failure is misalignment. This commonly occurs when seals are installed by personnel not trained in proper mechanical seal installation practices. If the seal is installed so that the mating faces are not parallel to each other and perpendicular to the shaft, the rotating face will try to square itself with the seal causing uneven wear on the mating faces. To prevent this problem, check to be sure that the seal is properly installed. The rotary unit must run parallel with the shaft and the stationary seat must be perpendicular to the shaft. Be sure the gland is drawn up evenly. This will eliminate the possibility of the improperly aligned stationary seat. Do not overtighten the gland - little more than finger tight is all that is generally needed in low pressure applications. Make sure all gaskets and/or O-Rings are properly installed. Lack of Lubrication - To operate properly, the mechanical seal mating faces must run on a liquid film. If there is no film, the seal will overheat and fail. Because this failure can occur in a matter of seconds, it is extremely important that the liquid film be present whenever the pump is in operation or about to be put into operation. Lack of lubrication at the seal faces is commonly experienced when handling hot water or light hydrocarbons

near their vapor pressure. To insure lubrication and dissipate heat, flushing must be provided at the seal faces. Be sure seal flush piping is used when employing balanced seals - the type usually recommended for liquids having low specific gravities. The slight additional cost of bypass piping is a good investment when handling liquids near their vapor pressure. Overheating - In some types of mechanical seal failures, overheating and lack of lubrication go hand in hand. This is especially true during the hard face heat checks. On examining a hard face that has heat checked you can see that the stress cracks are confined to the surface in direct contact with the carbon mating face. These surface cracks are caused when there is an intense heat buildup on the surface of the hard metal, while the subsurface of the part remains at ambient temperature. Since the surface metal is not allowed to expand in a lineal direction because of the subsurface temperature, the surface buckles upward and cracks. The edges of the cracks, which are slightly raised above the face, begin to shave material from the carbon mating face. Other symptoms of overheating are damaged elastomers (O-rings). This condition is easily recognized because the elastomers will show cracking on the surface and harden. To eliminate this type of failure, be sure the seal is operated in an environment having a lower temperature than the limits of any of its component parts.

(cont'd) 32

Abrasive Damage - Abrasive damage is a common cause of seal failure. When minute abrasive particles work their way between the seal faces, they can groove the mirror finish of the faces or they can leave deposits from evaporation of the liquid. When these deposits build up on the shaft they will freeze the sealing elements to the shaft and eliminate all seal flexibility. This condition can be particularly acute when teflon sealing wedges are incorporated in the seal. Abrasive particles may become imbedded in the teflon, and can cause shaft wear directly under the sealing member. Abrasive liquids must be kept out of the stuffing box if at all possible.

Corrosion - Corrosion failure is easily recognized in most cases. Metal parts show signs of pitting, or the sections are totally corroded away. We will not attempt to go into a discussion of corrosion, but it should be remembered that the metal sections of mechanical seals are usually much smaller than the metal section of a pump. Therefore, they can withstand much less corrosion attack. It is a good rule not to downgrade the metallurgy of the seal from that of the pump. When in doubt, select the more noble metal for the seals. (This is a general rule and does not hold true for all applications.)

QUENCH GLANDS A gland is a device used to compress the pump packing or provide support for a mechanical seal face in the stuffing box of rotating or reciprocating machinery. In particular, on centrifugal pumps, the primary function of a gland is to exert the necessary pressure on soft packing to hold it in place and to control the leakage of fluid from the stuffing box. It’s also used to prevent air from entering the pump casing through the stuffing box. In the case of a mechanical seal, the gland acts as a clamp holding the stationary seal face in concentric alignment with the shaft axis and in perpendicular alignment with the rotating seal face.

as it passes, it collects heat from the shaft and packing. Quenching in this manner permits pumping liquids at temperatures much higher than those normally allowed by the limits of bearing and packing lubrication. The quench lowers the temperature of the shaft at the packing and where it enters the bearing to permissible limits. The same quench gland may be used to prevent the escape of toxic or volatile liquid into the environment. Again water, oil or some other suitable fluid is fed through the gland, flushing away the undesirable leakage to a waste receiver.

The quench gland has a hollow portion adjacent to the shaft outside the stuffing box. This hollowed area can serve several purposes.

In some cases the gland is just used as a collector for the stuffing box leakage. There is no flushing connection - only a drain line to carry away any leakage from the stuffing box.

The gland can be used to remove heat from the shaft. Water, oil or other cooling is passed through the hollow portion and

Because of the quenching, flushing or collecting function, a quench gland must be made to fit the shaft closely on the (cont'd) 33

outboard end. If not close fitting, liquid would be allowed to spray or splash out along the shaft. Three separate methods may be used to achieve the close fit. First, the clearance between shaft and gland may be established by holding the I.D. of the outboard end of the gland to a close tolerance in comparison to the O.D. of the shaft in that area. This type of clearance cannot be restored without using a complete new gland and possibly a new shaft or new sleeve.

Second, the clearance may be established by using a replaceable bushing machined as above and pressed into a bored fit in the gland. This will allow restoring the clearance by changing only the bushing but would not compensate for any wear on the sleeve or the shaft. Third, auxiliary packing may be used. This may consist of a single ring of packing, a series of “V” rings and/or “O” rings - these are normally held in place by fitting them into a machined groove - or it may consist of several rings held in place by an auxiliary gland.

CONSTANT LEVEL OILERS - OPERATION AND ADJUSTMENT ITT Industries oil lubricated pumps come with a variety of methods of distributing the oil to the bearings. Some, such as the “X series” power frames use a flood oil arrangement with a sight glass mounted in the side of the bearing frame for checking the level. Others use the sight glass and oil rings to pick the oil up from the sump and distribute it to the bearings. Still some current models and many of the older models use a bottle - type constant level oiler.

oiler bottle mouth and therefore the oil level in the bearing housing. The oil setting must be checked in the field. To check setting remove oiler bottle – dust cap assembly, and lift leveling bar assembly from oiler body (see Figure 35). Setting should be the same as that shown in the pump instruction book. The oiler is usually set so that the oil either:

A cutaway of these oilers is shown in Figure 34. The oil stays in the bottle as long as the oil level in the bearing housing is level with the mouth of the bottle. When the oil level drops, air enters the bottle allowing oil to flow out until the oil level in the bearing housing is again up to the mouth of the bottle. The oil level in the bearing housing therefore stays constant.

2. is ¼” above the bottom of the oil ring in a ring oiled bearing.

The oiler bottle rests on a leveling bar that can be screwed up or down and locked in place. Raising the leveling bar raises the

1. covers half of the lowest ball in a flood oil lubricated ball bearing, or

The level must be set carefully. Too much oil is almost as bad as no oil at all. Be sure the oiler is clean when it is installed in the pump. Piping from the oiler to the pump must be level. If the oiler sags, the oil level in the housing will drop which could possibly ruin the bearing. Fill the oiler bottle with the proper oil, and place in the oiler body. The bearing housing is filled when the (cont'd)

34

bubbles of air entering the oiler stop and the level in the bottle remains constant. Never pour oil directly into the oiler body since you could easily overfill it.

Figure 35

Figure 34

BALL BEARINGS – HANDLING, REPLACEMENT AND MAINTENANCE SUGGESTIONS Ball and roller bearings are carefully designed and made to watch-like tolerances. They give long, trouble-free service when properly handled and maintained. They will not stand up to abuse. Keep Clean Dirt getting into the bearing during installation causes probably 90% of early bearing failures. Cleanliness is a must when working with bearings. Some things that will help are:

1. Do not open bearing housings unless absolutely necessary. 2. Spread clean plastic, newspapers or rags on the work benches and at the pump. Set your tools and the bearings on these covered surfaces only. 3. Wash your hands. Wipe dirt, chips and grease off tools.

(cont'd) 35

4. Keep the bearings, housings, and shaft covered with clean plastic or cloths whenever they are not being worked on. 5. Do not open the boxes or unwrap new bearings until you’re ready to install them. 6. Flush the shaft and housings with clean solvent before reassembly. Pull Bearings Carefully 1. Use a sleeve or puller which contacts just the inner race of the bearing. (The only exception to this is some double suction pumps which use the housing to pull the bearing.) 2. Never press against the balls or ball cages, only against the races. 3. Take care not to cock the bearing. Use a sleeve which is cut squarely, or puller which is adjusted to draw the bearing off the shaft squarely. Inspect Bearings and Shaft 1. Examine the bearing carefully. Scrap it if there are any flat spots, nicks or pits on the balls or races. Bearings should be in perfect condition to be reused. 2. Turn the bearing over slowly by hand. It should turn smoothly and quietly. If not, scrap the bearing. 3. Whenever in doubt about the condition of a bearing, scrap it. The relatively few dollars invested in new bearings may prevent serious loss from downtime and pump damage. In severe or critical

36

services, replace the bearings at each overhaul. 4. Check the condition of the shaft. Bearing surfaces should be smooth and free from burrs. Smooth burrs with emery cloth and polish with crocus cloth. Shaft shoulders should be square, free from nicks and burrs and smooth at the radii where smaller sections join larger sections. Check New Bearings Be sure the replacement bearing is of correct size and type. An angular contact bearing may be dimensionally interchangeable with a deep groove bearing and may fit perfectly in the pump. However, the angular contact bearing is not suitable for end thrust in both directions, and will probably fail quickly. Also check to see that the shields (if any) are the same as in the original unit. Refer to the pump bulletins and instruction book for the proper bearing to use and any notes regarding installation. Install Carefully 1. Oil the bearing surfaces on the shaft lightly. 2. Shielding, if any, must face in the proper direction. Angular contact bearings, on pumps where they are used, must also be oriented in the proper direction. Duplex bearing arrangements must be mounted with the proper faces together. Mounting arrangements vary from model to model. Consult the bulletin and instruction book for specific pump model.

3. Press the bearing on the shaft squarely. Be sure that the sleeve used to press the bearing on is clean, free from burrs, is square cut and contacts the inner race only. 4. Press bearing firmly against shaft shoulder. The shoulder helps locate the bearing on the shaft, support the inside race and square the bearing.

5. Be sure the snap rings, if used, are properly installed, flat side against bearing, and that the gap at the ends of the ring align with the oil return channel in the bearing housing to provide an open path for the oil to travel back into the housing. Make sure the lock nuts are tight. 6. Lubricate properly. Consult the instruction manual for specific information regarding lubrication requirements.

GREASE LUBRICATION ITT Industries grease lubricated pumps are designed for simple foolproof maintenance. Only one lubricant is required for a large variety of services. The correct lubricant is a premium ball, roller and plain bearing grease, which is oxidation resistant, has a wide temperature range and has a lithium soap base. The oil viscosity should be 200 – 250 SSU at 100°F. This would correspond to NLGI Grade 2 Grease. “Filled” greases such as those with molybdeum, graphite, Teflon™ and other fillers should be avoided. This one type of grease will suffice for nearly all applications; hot and cold; wet and dry; clean and dirty.

air flow around the pump are only a few. It is expected that the bearing temperature will stabilize at some level between 130° and 180°F which is perfectly acceptable for a grease or oil lubricated assembly. For most people anything above 120°F is uncomfortable to touch, so it is not unusual to have a properly lubricated, aligned, and operating bearing assembly which is too hot to hold your hand on. On the other hand, a sudden temperature rise without a corresponding rise in ambient or pumpage temperature can be a warning sign that there is something going wrong.

Grease should be added to the bearing after each 2000 hours of operation, or about every three months. More frequent lubrication can lead to shorten bearing life due overheating. The normal bearing temperature depends on many variables, the ambient temperature in the area of the unit, the temperature of the liquid being pumped, the speed of the shaft and the amount of Figure 36

37

FLEXIBLY MOUNTED BASEPLATES The flexibly mounted baseplate was first used in chemical plants in the mid 1950s. At that time, special extra heavy bases were used to provide adequate rigidity. Historically, pump and other rotating equipment suppliers have recommended a grouted base to provide a suitable platform to insure a permanent means of establishing and maintaining alignment between the rotating machine and its driver. The permanent grouted base is still recommended for large units where flange loads can be adequately limited, and where the base can be suitably maintained. In many chemical plants atmospheric corrodents make baseplates and foundations expensive and difficult to maintain. As a result, after a short time, the baseplate can deteriorate to a point where it no longer is able to maintain suitable alignment. Under adverse environmental conditions a well engineered flexibly mounted base can provide an economical solution to difficult problems. The flexible mounting system offers the following advantages: Reduced installation costs – no tiled or

acid proof brick foundation required, no grouting required, the pump can easily be adjusted to line up with piping. Lower maintenance costs – uniform flange loads will move the pump, base and motor as a unit reducing both the possibility of coupling misalignment and the seal problems normally associated with excess flange loading. The raised base can be easily hosed to remove corrosives and debris for simplified housekeeping and longer baseplate life. A successful installation requires that the base have sufficient rigidity to permit initial alignment and to maintain alignment over the operational life of the equipment. Goulds offers the flexibly mounted base arrangement as an option on our ANSI Family units and many other pump types. The standard cast iron or structural steel base normally furnished on the ANSI Family X-Series units have proved satisfactory under actual plant conditions up to 100 HP at 3550 RPM. Users report that the elimination of the expensive foundation enables them to use a standard horizontal ANSI type pump at the same installation cost as a vertical inline unit.

Figure 37

38

IMPELLER CLEARANCE The open impeller centrifugal pump offers several advantages over units equipped with other types of impellers. It is particularly suited to applications where the liquid contains abrasive solids. Abrasive wear on an open impeller is distributed over the diametrical area swept by the vanes. The resulting total wear has less effect on performance than the same total wear concentrated on the radial ring clearance of a closed impeller. Because of the impeller adjustment feature, the open impeller permits restoration of nearly “new pump” running clearance after wear has occurred without expensive parts replacement. A well designed open impeller pump will feature a simple positive means for axial adjustment without necessity of disassembling the unit to add shims or gaskets. Figure 38, with these typical instructions, shows one method employed on ITT Goulds open impeller centrifugal pumps for making this adjustment. Other methods using Dial Indicators, etc, may be used. Since the actual impeller clearance varies with pump type and temperature, consult the specific Operating Instruction for your pump for the proper settings. On units equipped with mechanical seals, follow the seal manufacturers instructions for the care of the seal during the impeller adjustment process.

1. Loosen the jam nuts and bolts (370D) located at the bearing housing (111). 2. Tighten the bolts (370C) evenly while slowly rotating shaft, until the impeller (101) just contacts the casing (100). 3. Snug all the bolts (370C) against the bearing housing (111). Loosen each bolt (370C) until a .015” feeler gauge can be placed between the bearing housing and the underside of the head of the bolt (370C). 4. Be sure the jam nuts on the bolts (370D) are loose. Tighten each bolt (370D) one flat at a time until the bearing housing is tight against the bolts (370C). Be sure all the bolts (370C and 370D) are tight. Tighten the jam nuts on bolts (370D). 5. The rotating element and consequently the impeller has been moved .015” away from the casing thus giving the required clearance between these two parts.

Figure 38 (cont'd) 39

PUMP VIBRATION ANALYSIS

Figure 39 Pump users have become increasingly aware of vibration and the use of vibration analysis in detecting problems, predicting failures and scheduling equipment outages. The vibration analyzer shown in Figure 39 is used to measure and indicate the amplitude, frequency and phase values of vibration. Furthermore, when vibration occurs at several frequencies, it separates one frequency from another so that each individual vibration characteristics can be identified and evaluated. The vibration pickup, or transducer, (A) senses the motion of the machine and converts it into an electrical signal. This signal may represent the amount of movement, which would be displacement, the velocity at which the machine is moving as it vibrates or the acceleration required to produce the particular velocity. The analyzer receives this signal and converts it to the corresponding amplitude and frequency.

Depending on the instrument and the settings used, the amplitude may be displayed in terms of peak-to-peak vibration, which would be the maximum values present or various methods of averaging such as RMS. Displacement is indicated in mils where 1 mil equals 1 one-thousandth of an inch, Velocity in Inches per Second and Acceleration in “G’s”. Many machines today have graphical displays for an immediate indication of both the amplitude and frequency of the vibration present and some have printers built into them to print out the information right on the spot. Often the units include a down loading feature so that the information collected may be saved on a PC or disk for later analysis. Most instruments are equipped with a device to indicate frequency which gives a direct readout of the predominant frequency of the vibration. Other instruments have tunable filters (C) which allow scanning the frequency scale and (cont'd)

40

reading amplitudes at any particular frequency. The strobe light (D) is used to determine the phase of the vibration which can be of great value when dealing with complicated vibration signatures and in depth analysis. The strobe can be made to flash at the predominant frequency of the vibration or at any frequency setting using an internal oscillator to fire it.

The first step in vibration analysis is to determine the severity of the vibration, then, if the vibration is serious, a complete set of vibration readings should be taken before attempting to analyze the cause. Figure 40 is a severity chart based on amplitude, frequency and velocity. Figure 41 is a general guide for centrifugal pumps as published by the Hydraulic Institute.

A reference mark on a rotating part viewed under the strobe flashing at the vibration frequency may appear as a single mark, multiple marks and either stationary or rotating. The number of marks viewed and whether they are rotating or stationary is useful in determining the source of the vibration. The relationship between the location of the mark on the part and where it is illuminated by the strobe can be determined and that information can be used when balancing rotating parts.

Figure 40

Figure 41 (cont'd) 41

The severity of vibration is a function of amplitude and frequency; however, it should be noted that a change in severity over a period of time is usually an indication that the condition of the machine is deteriorating. This change is often more important than the actual vibration values themselves. Vibration levels in the “slightly rough” or “rough” ranges that do not change with time maybe be perfectly acceptable. The opposite is also true; a machine that exhibits vibration in the “Very Smooth” area but moves to the “Very Good” area in a short period of time may be very close to failure. Complete pump vibration analysis requires taking vibration readings at each bearing in three planes (horizontal, vertical and axial). Readings at the pump suction and discharge flanges may also be useful in some cases. After all data has been tabulated, it can be analyzed to determine the most likely cause or causes of the vibration. Figure 42 lists the most common causes of vibration and the identifying characteristics of each. By analyzing the tabulated vibration data with the reference to the identification chart, one or more causes may be found. Each possibility must be checked, usually starting with the most likely cause or easiest or least expensive to correct.

For example, assume that axial vibration is 50% or more of the horizontal or vertical vibration and the predominant frequency is the same as the RPM of the pump. The chart indicates the most probable cause is misalignment or a bent shaft. Coupling misalignment is probably the most common single cause of pump vibration and is one of the easiest to check. If, after checking, the alignment is found to be within acceptable limits, then an inspection for excessive flange loading is warranted. Flange loading beyond the limits allowed can cause distortion and misaligment inside the pump leading to higher than expected vibration levels. If that inspection does not reveal the source then the next step is to check for a bent pump shaft. Cavitation in a pump can cause serious vibration. Vibration caused by it usually appears at random frequencies but it is most often identified by a loud crackling noise in the suction area of the pump rather than by any characteristic vibration. Vibration at random frequencies can also be caused by hydraulic disturbances in poorly designed suction or discharge systems. Vibration analysis, while a relatively new field, has rapidly gained acceptance in industry as a valuable tool in preventive and predictive maintenance and as a general troubleshooting tool.

(cont'd) 42

Vibration Identification Chart Cause

Amplitude

Frequency

Phase

Remarks

Unbalance

Largest in radial direction. Proportional to unbalance.

1 X RPM

Single reference mark

Misalignment of coupling or bearings and bent shaft

Axial direction vibration 50% or more of radial

1 X RPM normally

Single, double, or triple

Easily recognized by large axial vibration. Excessive flange loading can contribute to misalignment.

Bad anti-friction bearings

Unsteady

Very high. Several times RPM.

Erratic

Largest high-frequency vibration near the bad bearing.

2 X RPM

Two reference marks. Slightly erratic

Check grouting and baseplate bolting.

1, 2, 3, & 4 X RPM of belts

Unsteady

Use strobe light to freeze faulty belt.

Mechanical looseness Bad drive belts

Erratic or pulsing

Figure 42

PUMP MAINTENANCE RECORDS PAY DIVIDENDS Too often plant records of pumps are little more than an inventory record for accounting and insurance purposes listing not much more than the pump size and manufacturer. In other plants, maintenance records may be kept as part of a pump lubrication schedule. In either case complete pump records could be kept with very little additional effort. With the wide spread use of PC’s in industry and the development of inexpensive Data Base programs more and more organizations are going to Maintenance Tracking Systems in an effort to reduce downtime, lost production and maintenance costs. Complete maintenance records, filed in an

accessible location, are invaluable in diagnosing pump failure, in ordering repair parts and in establishing lubrication and maintenance schedules. In addition these maintenance record cards are helpful in determining a pump’s suitability for new requirements due to process changes. Notations of pump failures and the repairs required may be used to define the optimum period of any given pump before complete inspection and overhaul is required. If a unit has been on an annual overhaul program and the pump Maintenance record indicates that no failures have occurred over a period of years, perhaps a longer period between inspections would be warranted.

43

On other pumps, the records may indicate frequent failures. This may suggest that semi-annual inspections and repairs are required. It may also suggest that there is a need to take a closer look at this unit. Something may be wrong with the unit itself, its assembly, the application, the operation or the installation. The chronological notation of pump repairs can be used to develop a proper spare parts inventory and may, in some cases reduce inventory and the costs associated with it. Frequent replacement of a part subject to abrasive wear may indicate a different material selection would be more economical. Simple maintenance records for each pump, in both primary and secondary service, are much less expensive than trial and error solutions.

44

ITT INDUSTRIES / FTC OFFERS A COMPLETE LINE OF PUMPS TO MEET EVERY INDUSTRIAL NEED ANSI Process

Magnetic Drive

Model 3196

Model 3298

This is the original ANSI pump that has become the standard of the industry. Over 600,000 installations attest to the remarkable performance of the 3196.

The 3298 is designed specifically to handle moderate to severe corrosives with or without solids. As a sealless design, it’s an effective alternative to pumps with mechanical seal problems. Meets strictest EPA regulations.

Abrasive Slurry

In-Line ANSI Process

Model SRL

Model 3996

SRL rubber lined pumps are designed specifically for handling abrasive and corrosive slurries found in the mining and mineral process industries, and in pulp and paper, aggregate, pollution control and chemical applications.

For corrosives, abrasives and high temperatures. Fully open impeller, back pull-out design, heavy duty construction. Field alignment not required.

45

Splitcase Double Suction Model 3420

Vertical Industrial Model VIT (Turbine)

For a wide range of industrial, municipal, marine services .. cooling tower, raw water supply, booster, fan pump, high/low lift, fire pump, pipelines.

A wide range of hydraulic conditions allows meeting requirements of virtually every pump service. These pumps are designed to meet custom user specifications. Model VIC can-type turbine meets API-610 specifications. Model VIC (Can-type)

Paper Stock Process Model 3180 All customer requirements were considered in the design of this paper stock/process pump. The result is a true world class pump line. . .a product without compromise.

46

Vertical Sump and Process Model 3171 Vertical industrial pumps for wide range liquid handling capabilities. For drainage and general sump, tank unloading and transfer, process circulation, outside tank mounting, high temperature service.

ISO Lined Sealless

General Industry

Model Richter PCK Model 3355 The ICK is designed to handle moderate to sever corrosives, providing excellent service in a variety of process plants worldwide.

Multi-Stage Model 3600 Advanced design with proven operating history. Axially split, with many enhanced features that make it an extremely reliable, high performance, pump well suited to a wide range of services.

Goulds Model 3355 is a multi-stage ring section pump designed for high-pressure services including: boiler feed, reverse osmosis, shower service, pressure boosting, plus much more. Its modular design and multiple configurations make it ideal for your system.

API-610 (8th Edition) Process Model 3700 High temperature and high pressure process pumps designed to meet the requirements of API-610 (8th Edition). Centerline support for high temperature stability, maximum rigidity. Tangential discharge for maximum hydraulic efficiency.

47

Nature of Pumpage

Abrasive Slurry/Solids Handling

Multi-Stage & Double Suction

Sump/ Abrasives/ Solids Handling

General Service/Vertical Turbine

® Tefzel® and TEFLON are registered trademarks for fluorpolymer resins, films and fibers made by DuPont.

48

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q q q

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Non-Abrasive

q q q q q

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m

High Temperature 500°F (260°C) or Greater

q q q q q q q q q q q q q q q q q q

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z z z z z z z z z z z z z z z z z z z z z z z

Primary Metals

m

Oil Refining & Gas Processing

q

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API Process

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Food & Beverage

Paper Stock/ Process

Process Pumping System Repair & Service of all Rotating Equipment Types of Manufacture 3196 ANSI Chemical Process LF 3196 Low Flow ANSI Process IC ISO Chemical Process NM 3196 FRP Process 3796 Self-Priming Process 3996 In-Line Process 3296 Magnetic Drive Process 3298/SP3298 Tefzel® Lined Sealless V 3298 Tefzel® Lined Sealless 3299 PFA Lined Sealless MPB/F Peripheral with Magnetic Drive MNK Lined Sealless MDK Lined Sealless PCK/SCK Lined/Heavy Duty Lined Process 3198 PFA Teflon® Process 3200 Light Duty Sealless CV 3196 Non-Clog Process AF Axial Flow 3171 Vertical Sump & Process NM 3171 FRP Vert. Sump/Process 3175 Paper Stock/Process 3180/3185 Paper Stock/Process 3181/3186 High Temperature 3500 Heavy Duty Paper Stock 3700/3710 API-610 Process 3910 API 610-In-Line 3620 High Temp. Double Suction 3640 High Temp. Two-Stage CPR API Process JC Medium Duty Abrasive Slurry SRL Rubber-Lined Abrasive Slurry 5000/RX Side Suction Abrasive Slurry Slurry Pro Heavy Duty Abrasive Slurry 5500 Severe Duty Abrasive Slurry HS Non-Clog Solids Handling 3310H High Pressure Multi-Stage 3600 Heavy Duty Multi-Stage 3335 Diffuser Type Multi-Stage 3935 3316 Two-Stage 3408 3409 3410 Single-Stage, Double Suction 3415 3420 3498 MP Multi-Stage P Multi-Stage Whirl-Flo Vortex Prime Line Industrial Self-Priming Trash Hog Solids Handling, Self-Priming VHS 5100 Vertical Cantilever 5150 VJC HSU HSUA Submersible HSUL JCU 3100 General Service VIT & VIC Vertical Turbine/Can Type VMF/VAF Mixed Flow/Axial Flow VMP Vertical Marine VIS Vertical Submersible

Mining & Minerals

Chemical Process

Pump Type Pulp & Paper

PumpSmart® PRO Services™

Goulds Model

Chemical

Pump Category

Power Generation

Solids

x

x

x

x x x x x

x x x

x

x x

x x x

x x x x

x

x x x x x x

x x x x x x x x x x x x

x x x x x x x x x x x x x x x x

x x x x x x x x x x x x x x

Ideally Suited for Service Indicated

x

x x x

x

x

x x x x x x x x x x

NOTES

49

NOTES

50

Visit our website at www.gouldspumps.com

© 2002 Goulds Pumps Form # BRCARE

PRINTED IN U.S.A. 9/2002

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TEXTRON

FLUID HANDLING PRODUCTS

Service Centers for the New Millennium

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Top Quality Parts, Personnel, Equipment and Experience Serving you around the clock and around the world with quality service and repair parts for your Centrifugal or Reciprocating Pumps and related equipment needs. Experienced, factorytrained, authorized TEXTRON Fluid Handling Products technicians and

TEXTRON Fluid Handling Products can be your source for replacement parts and fast turn-around repairs or rebuilding of pumps and related equipment, no matter who manufactured them or how old they are. TEXTRON Fluid Handling Products works with you to overcome obstacles and minimize costly downtime. Here’s what we offer: Complete Service and Repair Centers, fully stocked and equipped ...strategically located. Extensive inventory of parts and stocking programs for other non-traditional equipment for TEXTRON Fluid Handling Products models ...fast fabrication of parts not in stock, ours or “theirs”. Large inventory of heat treated materials in stock — for quick production of replacement parts, with no delays. tchibouela.

Fast parts order processing — our computerized inventory control provides on-line answers, immediately.

engineers are on call 24

Unsurpassed pump engineering expertise — our engineers as well as our servicemen have the experience and know-how to get the job done right.

hours a day. For a prompt

Revamp, Upgrade & Retrofit packages to improve the performance, efficiency & reliability of your equipment.

response to your pump

Direct access to a service engineer who will assist you with your problem and be responsible for making sure your expectations are fully met.

repair needs, call our toll

Instruction & training of your operations & maintenance team. Pick-up and delivery for equipment repair.

free number today!

1•800•877•PUMP

Types of Equipment Repaired or Rebuilt (partial listing) Centrifugal Process API and ANSI Pumps

Compressors

Power Pumps

Boiler Feed Pumps

Horizontal and Vertical Pumps Single and Double-Acting Pumps Steam Pumps Homogenizers Blowers Steam Turbines Gas Turbines and Expanders Pumping Equipment on Off-shore Oil Platform Vertical Turbine Pumps

Condensate Pumps Descaler Pumps Large Circulating Pumps Reverse Osmosis Service Pumps Controlled Volume Pumps Piston and Plunger Pumps Various Rotating Equipment Service of other manufacturers’ equipment.

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Capabilities—Types of Repairing & Rebuilding Operations Following is a partial listing of the most common services we provide. If your requirements are not listed, please contact us — chances are, we’ve got you covered. New pump start-up Pump repair/overhaul Teardown, cleaning, inspection of units Precision machining of all parts Assembly and disassembly of multistage rotating elements Hardcoating, metalizing of shafts, pistons, and shaft sleeves, etc... Certified welding, piping and plate fabrication Dynamic balancing of rotating elements On-site balancing Pump destaging Rebuilding, rechroming and grinding of crankshafts Rebuilding plungers, fluid end components, pistons, etc... Trim and balance impellers Hydro testing equipment Reline, rebore and hone cylinders Compressor repair Casting repair Overhaul fuel and air starting valves Chrome Steel Multi-stage casing rewelding and reboring Non-destructive testing Field machining

usmnfg12

Field repair Horizontal milling Retrofitting/design upgrading Preventive maintenance programs On-site technical services & performance testing Modal and structural analysis

Upgrading Your Present Pumps: A Dollar-wise Productivity Enhancer Many older facilities are operating steadily, but without the benefit of the latest advances in pump technology. By rebuilding your pumps with advanced designs and improved metallurgy, performance levels and efficiencies can be greatly increased. This can result in enhanced productivity of many processes. TEXTRON Fluid Handling Products has the capability to completely upgrade most of your older pumps. Applying the latest technology can greatly improve the efficiency and productivity of the unit. If they represent the latest technology in their field, complete rebuilding can often restore original performance parameters, resulting in increased output.

impwork

Examples of such upgrading — to name just a few — include: retrofitting packed pumps to sealed pumps and upgrading pumps to meet new API or MIL standards. Call our toll-free hotline (1-800-877-7867) for a free estimate of the cost and potential payback of upgrading your older pumps.

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A Leader in Satisfying Your Replacement Part Requirements As one of the few pump companies in the world specializing in all three major types of pumps — centrifugal, steam and power, TEXTRON Fluid Handling Products has a special advantage by supplying your replacement parts needs.

Largest Inventory of Quality Parts As the number and diversity of customers we service has grown, so has our inventory of stocked parts. We’re very proud of our quick response to meet your needs. Maintaining a large inventory of parts helps keep our service the best available. Nothing is more irritating than waiting for a spare part to arrive. We understand this and work hard to keep if from happening to you.

Preventive Maintenance Agreements TEXTRON Fluid Handling Products Company can provide on-site inspection and maintenance programs to help prevent unscheduled shutdowns. In such an approach, we work with your management personnel to develop a detailed inspection and maintenance schedule tailored to the present conditions, age and demands place on a given facility. Under many circumstances, there’s a good chance that we can handle such a program at lower total cost than an in-house program. As with all our services, a single call tour toll-free hotline can start the ball rolling toward a feasibility study of this approach for one or more of your facilities.

Fastest Fabrication of Non-Stock Parts When the needed part is a non-stock TEXTRON Fluid Handling Products part, our team of shop professionals can manufacture it from original drawings. There’s no guess work as to part specifications or materials. We also apply this same expertise to producing other-make parts. This means that hard-to-get, even non-existent parts become available to you on a rush basis through TEXTRON Fluid Handling Products. Using the latest in diagnostic, measurement and production technology, we can manufacture to exact specifications, in the shortest possible time, parts for centrifugal pumps, power pumps, steam pumps and compressors.

DVMXexploded

Maintenance Training Seminars We can also help if you have the personnel available for preventative maintenance, but turnover and continuous demands on their time have prevented them from acquiring the necessary knowledge and understanding to perform these critical functions. Our training seminars can help you cut repair costs and minimize costly downtime by instructing your maintenance personnel in proper operating, inspection and maintenance procedures. These seminars are custom-designed to fit your specific needs. They are conducted by highly qualified and experienced TEXTRON technicians and engineers. They can be presented at your facilities or ours. If you suspect that one or more of your facilities could benefit from such training, pick up the phone and dial our toll-free hotline today at (1-800-877-PUMP).

Large supply of stock and custom parts for quick repair or replacement needs.

The most cost-effective approach: stopping trouble before it starts No one disputes that a carefully conceived maintenance program to prevent or minimize breakdowns is preferable to reacting after trouble strikes. But lack of qualified personnel or lack of adequate training, among others often interfere with such a program. To help you overcome these obstacles, we offer to programs: maintenance training seminars for your people and on-site, turnkey maintenance agreements using TEXTRON technicians.

dzQx200 or Z1.bmp

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Se 1.

2.

TEXTRON Fluid Handling Products is people... committed to winning as a team. TEXTRON Fluid Handling Products Company supports its product lines with a worldwide network of professional field sales engineers. They are committed to prompt, on-site analysis or your fluid-handling problems and dedicated to supplying long-term, cost-effective solutions. We pride ourselves on having the ability to work with our customers “hands on” to satisfy their needs whenever and wherever they need help. Our service technicians have access to all areas of expertise from the factory by personal contact with any member of the TEXTRON Fluid Handling Products Team. TEXTRON Fluid Handling Products Company has a long and distinguished history in the process industries. We feel proud that since 1885 we have been satisfying your increasingly sophisticated requirements. We have proven our ability to keep pace with, and surpass our competition in the design and manufacture of outstanding products and in providing unexcelled service, whenever and wherever it’s needed around the globe.

3.

4.

5.

6. Our past experience will continue to fuel our future achievements as we further modernize our manufacturing facilities and add additional qualified people throughout our organization. From first to last, we are dedicated to the conviction that our primary goal is to serve our customers and their pumping needs.

7.

8.

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Service Center Locations 1.

Southeast Service Center 12742 Ronaldson Road Baton Rouge, Louisiana 70807 Telephone: 504•775•2660 Toll Free: 800•842•7867

2.

10723 Rockley Road Houston, Texas 77099 Telephone: 281•933•2900 Toll Free: 800•637•6656 3.

Canadian Eastern Service Center 4211 Mainway Drive Burlington, Ontario L7L 5N9 Canada Telephone: 905•335•2580

4.

24-Hour Service

Southwest Service Center

Canadian Western Service Center

At TEXTRON Fluid Handling Products Company, we match our schedule to yours. Because breakdowns don’t always occur between 8 and 5 in industries where 24-hour production is the norm, you shouldn’t have to wait for repairs until the following morning or the following Monday. Our service technicians are on call 24 hours a day, seven days a week, to give you help when you need it. Contact us at anytime. We’ll be there for you. Prompt, dependable service is why our customers count on us to help maintain their equipment. Next time you have an equipment repair or parts problem, call TEXTRON’s toll-free, 24-hour hotline (1•800•887•PUMP). You’ll be glad you did!

3525 62nd Avenue S.E. Calgary, Alberta T2C 1P5 Canada Telephone: 403•236•8725 5.

Midwest Service Center 4602 West Dickman Road Battle Creek, Michigan 49015 Telephone: 616•966•4644 Toll Free: 800•957•8957

6.

Western Service Center

9 3 5

4 6

2

8

1 7

9838 Firestone Boulevard Downey, CA 90241 Telephone: 562•622•2380 Fax: 562•622•2375 Toll Free: 800•258•7867 7.

Shreveport Service Center 5030 Floumoy-Lucas Road Shreveport, Louisiana 71129 Telephone: 800•256•4944

DAVID BROWN PUMPS

TEXTRON

FLUID HANDLING PRODUCTS

GUINARD PUMPS

8.

Penistone Service Center (UK) Green Road Penistone, Sheffield S36 6BJ UK Telephone: +44 (0) 1226 763311

TEXTRON

FLUID HANDLING PRODUCTS

UNION PUMPS

TEXTRON

FLUID HANDLING PRODUCTS

9.

Annecy Service Center (FRANCE) UNION PUMPS

39, Avenue du Pont de Tasset B.P. 435-74020 Annecy cedex FRANCE Telephone: +33 (0) 4 50 05 56 68

TEXTRON

FLUID HANDLING PRODUCTS

TEXTRON

FLUID HANDLING PRODUCTS

CENTRIFUGAL PUMP TROUBLE SHOOTING MYG’99

Where Trouble Occurs..... Failure of a centrifugal pump in service may be sudden, as when a shaft breaks, or gradual, as when brackish water causes blistering of the impeller or casing. Fortunately, neither condition is common in well-operated and maintained pumps. Nor is slow corrosion of the impeller vanes too often met. Failures of the materials used in construction of a pump are usually traceable to incorrect application, instead of poor operating and maintenance procedures. But where pumps are wrongly applied for example, using a bronze-fitted pump for an alkaline solution having a high pH thorough maintenance should quickly show up the error. As a result, wrongly applied pumps are generally replaced before too much trouble occurs in the plant. Hydraulic troubles, like failure to deliver any liquid, low discharge pressure, and others, are more common and more difficult to solve. One manufacturer estimates that, except for mechanical defects, about 85% of the troubles met with centrifugal pumps occur on the suction side of the unit. So this directive has been specially prepared as a handy troubleshooting guide for all types and sizes of centrifugal pumps handling any liquid at any flow rate, and at any temperature and pressure for which pumps of this type are built today.Each of the major troubles is listed separately; below it the possible causes are given in italics, followed by the recommended remedy or cure. Horizontal pumps are considered first, vertical pumps second. Some troubles, however, are common to both types of pumps. NO LIQUID DELIVERED Lack of Prime: Fill the pump and its suction pipe completely with the liquid being handled. To rid the casing and piping of air, open all Centrifugal Pumps the vent valves while filling the pump and pipe. Leave the vents open until clear bubble-free liquid flows from them. Close the vents and start the pump. Speed of Pump Driver Too Low: With a motor drive, check to see that it is connected directly across the line and receives full rated voltage. With alternating-current (AC) motors, check the frequency it may be too low. Or the motor may have an open phase, causing it to run at a speed lower than its rated value. With a turbine-driven pump check the governor setting and throttle valve. A loose or damaged throttle-valve disk may plug the steam passage. causing wire drawing and a loss of pressure. With engine-driven pumps, check the fuel supply, governor setting. intake-air supply, scavenging-air pressure, spark plugs, magneto, carburetor, or fuel-injection pumps. Air-driven centrifugal pumps run at too low a speed when the air pressure is lower than that required at the air-motor inlet. A fall in air pressure may be caused by reduced output of the compressor, excessive use of air in other parts of the system, use of undersize hose or pipe connections, or open drain valves in the system. Discharge Head Too High: Check all valves in the discharge line to see that they are wide open. Be sure that gage valves are not stuck closed by some obstruction in the pipe. If the discharge head is still too high and no new devices have been installed in the system, check the piping for obstructions, from either solids contained in the liquid, or scale buildup.

PP(T)SB PROPRIETARY DOCUMENT

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CENTRIFUGAL PUMP TROUBLE SHOOTING MYG’99

Suction Lift Too High: Check the pump inlet for clogging by mud or some other obstruction. If a foot valve is used, cheek it for broken disks or a clogged strainer. When the pump is being started for the first time and it will not draw water from the suction pit, check the maximum suction lift recommended by the manufacturer. With centrifugal pumps this is usually 15 ft. (4.572 m) As Fig. 1-A shows, atmospheric pressure forces liquid into the pump suction when an open suction pit is used. If a perfect vacuum could be produced in the suction eye of a pump handling water at ordinary atmospheric temperatures, the atmospheric pressure could support a column of water 34 ft (10.3632 m) high (Fig. 1-B). But a standard suction lift of only 15 ft (4.572 m) is usually recommended for centrifugal pumps, giving a 19-ft reserve (Fig. 1-C). Remember, suction lift is made up of the static lift, friction head, and velocity head. So check the vertical distance between the liquid surface and pump inlet, as well as any possible clogging of the suction pipe by dirt or scale.

FIG. 1. (a) Action of atmospheric pressure on liquid in an open suction tank.(b)Maximum suction lift possible at sea level is 34 ft (10.3632 m) of water. (c) Standard suction lift.

Impeller Plugged: Solids in the liquid may accumulate on the impeller, preventing it from discharging liquid. Open the casing and clean all parts of the impeller. Wrong Direction of Rotation: Despite all the warnings of pump manufacturers, wrong rotation still occurs when some pumps are first started. See that the pump turns in the direction of the arrow on its casing. Other, less common causes of no liquid being delivered are an air or vapor pocket in the pump suction line, suction pipe not sufficiently submerged, available net positive suction head (npsh) not high enough, and the total head against which the pump works higher than that for which the pump is designed. PP(T)SB PROPRIETARY DOCUMENT

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CENTRIFUGAL PUMP TROUBLE SHOOTING MYG’99

NOT ENOUGH LIQUID DELIVERED With this trouble the pump delivers some liquid but the amount is less than the rated capacity of the unit at the head at which it is operating. This can be almost as serious as no delivery at all because valuable units served by the pump may be endangered. Possible causes, covered above, include wrong direction of rotation, speed too low, discharge head too high, impeller clogged, suction lift too high, air or vapor pocket in suction line, suction pipe not sufficiently submerged, available npsh insufficient, and clogged foot valve. The remedies to be used are the same as those recommended above. Other causes are given below. Air Leaks: These may occur in two places the suction line or the pump stuffing boxes. Check the flanges and screwed joints with a flame or match only if the liquid handled is nonexclusive and no explosive dust or gas is present in the atmosphere. The flame will be drawn toward any leaks, if it is held close to the pipe and flanges. Where an explosive liquid is handled, or an open flame cannot be used around the pump, check for leaks by shutting the suction valve or plugging the inlet of the suction pipe and putting the line under pressure from another source like a small hand pump. Attach a pressure gage to the suction line and observe it to see if the pressure falls over a period of time say 30 minute. Plug all leaks found in the suction piping. If the suction piping has no leaks, check the pump stuffing boxes. The box serving the pressure side of the main shaft should leak a small amount of liquid during pump operation. Adjust the gland to give a suitable flow from the box. If this adjustment does not give the desired results, stop the pump and check the packing. New packing is probably needed. Install as directed in the maintenance manual . Recheck leakage from the box. If it is still unsatisfactory, disconnect the water-seal piping at the box and check liquid flow. Clean out plugged piping. Next, check the position of the seal cage. It should be directly under the seal-liquid inlet. Inspect shaft sleeve for excessive wear. Deep grooves and pits allow air to leak into the suction side of the casing. Replace the shaft sleeve. Another way of testing for stuffing-box or mechanical-seal air leaks requires somewhat more work but is effective. Disconnect the suction line at the pump and screw a plug into the suction coupling, or fit a blind flange to the suction flange. Dope the threads or gasket. Drill and tap a ¼-in (6.35 mm), hole into the plug or flange and connect a vacuum gage to it. Fill the pump completely with liquid, start its drive, and run it at the normal operating speed. The pump should develop a vacuum of 20 in (508 mm) or more within 1 minute after starting, if there is no leak in the stuffing box or mechanical seal. Low NPSH: Connect a compound pressure gage to the suction pipe. If the needle fluctuates rapidly, the liquid in the suction pipe is flashing into vapor. Safest remedy is to check with the pump manufacturer. Worn Wearing Rings: Inspect the rings visually. If they are badly worn, permitting leakage in the pump, replace all rings. Damaged impeller: Remove casing and inspect the impeller. Replace with a new one if vanes or other parts are damaged or worn. PP(T)SB PROPRIETARY DOCUMENT

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CENTRIFUGAL PUMP TROUBLE SHOOTING MYG’99

Undersize Foot Valve: Area of foot-valve openings should be at least equal to that of the suction pipe, and preferably 1.5 to 2 times as large. The net clear area of the strainer should be 3 to 4 times that of the suction pipe. Wrong Direction of Rotation: A pump turning in the wrong direction may deliver some liquid, but its head will be low and the driver will be overloaded. The pump usually will not deliver more than about one-third of its rated capacity. Check the directional arrow on the casing. Suction-bay Disturbances: Vertical centrifugal pumps must often be located a specified distance from the walls of the suction bay. If the pump does not deliver its rated capacity, check the location against the manufacturer’s recommendations. Viscosity Too High: A pump designed to handle water will deliver less head and capacity when pumping a thick oil or other viscous liquid. So if the pump is moved from one job to another, or the liquid it handles has been changed, check the viscosity to see that it is not too high for the unit. Worn Gaskets: Replace all gaskets during a pump overhaul. Tighten the hold-down bolts to the desired or recommended value, using a torque wrench, if necessary. Impeller Eye Too Small: The larger the diameter of the impeller eye, the greater its capacity (Fig. 2). Incorrect choice of a pump, or moving it from one job to another may cause this trouble. The only solution is installation of a pump having a suitable capacity for the job.

Fig. 2 : How Suction-eye size affects capacity.

PUMP DISCHARGE PRESSURE LOW Typical causes of this trouble include too low a speed, worn wearing rings, damaged impeller, worn packing, gas or vapor in the liquid. too viscous a liquid, wrong direction of rotation, and worn gaskets. Other possible causes are given below. Gas or Air in Liquid: Bubbles will form imi the liquid when it enters the suction pipe. Check for this condition by reducing the pressure on the surface of a small amount of liquid and

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CENTRIFUGAL PUMP TROUBLE SHOOTING MYG’99

observe if any bubbles form. A gas-separation chamber on the suction line near the pump may be needed to remove the gas or air from the liquid. Pump Water Passages Obstructed: Open the casing and check the water passages for freeness. Remove any obstructions and replace the casing. Impeller Diameter Too Small: This traces back to the assumptions made in the selection of the pump. Check with the pump manufacturer. It may be possible to use a larger-diameter impeller in the same casing. Or by reducing the pipe friction losses or static suction or discharge head, or increasing the speed of the pump, a higher pressure can be obtained. But be extremely careful when changing the system head or pump speed not to overload the driver beyond its safe rated limit. PUMP LOSES PRIME AFTER STARTING There are a number of common causes of this trouble. These are incomplete priming, too high a suction lift, air leaks in the suction pipe or packing glands, gas or air in the liquid, suction line not filled with liquid, air or vapor pockets in the suction line, inlet not sufficiently submerged, low available npsh, plugged seal-liquid piping, or a misplaced lantern ring in the stuffing box. Various remedies are discussed above under these or similar headings. PUMP OVERLOADS DRIVER Discharge Head Low: With too low a discharge head the pump delivers too much liquid, overloading the driver. It may be possible to turn down the outside diameter of the impeller, thereby reducing its capacity. But never do this without complete advice from the pump manufacturer. Serious damage can occur otherwise. Wrong Liquid: If either the specific gravity or viscosity of the liquid is different from that for which the pump is rated, there is a chance of the driver’s being overloaded. Use a larger driver, after consulting the manufacturer for the recommended size. Be sure to check the viscosity and specific gravity of the liquid before ordering a larger driver. Another difficulty may be the cause. Speed Too High, or Wrong Direction of Rotation: Correct as described above. Packing Too Tight: Release the gland pressure; then retighten reasonably. Check the leakage of the seal liquid from the packing. If there is no leakage while the pump runs, replace the packing as directed in the maintenance manual. Check for a scored shaft sleeve if the packing wears rapidly. Replace worn sleeves with new or refinished ones. Distorted Casing: Poorly aligned suction and discharge piping can distort the pump casing, causing excessive friction between the impeller and casing. Check the PP(T)SB PROPRIETARY DOCUMENT

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CENTRIFUGAL PUMP TROUBLE SHOOTING MYG’99

piping and drive alignment. Inspect for worn wearing rings; replace damaged or worn parts. Bent Shaft: Thermal distortion damage during pump overhaul, or wrong assembly of the rotating element can cause bending of the shaft. Check the shaft deflection by means of a dial gage by turning the shaft between lathe centers. (refer Manual) Recommended average runout varies from one manufacturer and pump design to another, but 0.003 in. (0.9144 mm) is usually the maximum runout allowed on the shaft of a high speed pump. On a slow-speed pump it should not exceed 0.006 in (1.8288 mm). Mechanical Failures: These can increase the drag on the shaft, raising the power input. Check all rotating and stationary parts for failure, including the pump bearings, wearing rings, packing gland, bushing, and impeller. Replace, as necessary. Misalignment: Realign the pump and its driver. Pump Speed Too High: Power input to a pump increases as the cube of the speed. So a slight increase in pump speed can mean a measurable rise in power input. Check the line voltage on motor-driven pumps and the governor on turbine and engine-driven units. Also check the speed of the driver to see that it is correct. STUFFING BOXES OVERHEAT Common causes of this trouble include packing that is too tight, not enough packing lubricant, wrong grade of packing, not enough seal liquid flowing to the packing, and incorrect installation of the packing. Remember, there is a correct way of installing each type of packing. Check with the manufacturer manual. EXCESSIVE VIBRATION Gas or air in the liquid leads to a starved suction, as do insufficient npsh, not enough submergence of the end of the suction pipe, and gas or vapor pockets in the suction line. Other causes of vibration include pump misalignment, worn or loose bearings, rotor unbalanced because of a plugged or damaged Impeller, bent shaft, Improper positioning of a control valve in the discharge, and a non rigid foundation. BEARINGS OVERHEAT Many lubricating troubles cause overheated bearings. These include too low an oil level, a poor or wrong grade of oil, dirt in the bearings or the oil, moisture in the oil, a clogged or scaled oil cooler, failure of the oiling system, not enough bearing cooling water, bearings too tight, misalignment, or oil seals fitted too closely on the shaft.

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CENTRIFUGAL PUMP TROUBLE SHOOTING MYG’99

BEARINGS WEAR RAPIDLY Excessive wear of sleeve, ball, or roller bearings may be caused by misalignment, a bent shaft, vibration, excessive thrust from a mechanical failure inside the pump, lack of lubrication, wrong bearing installation procedures, dirt in the bearings, moisture in the oil, and excessive cooling of the bearings. Vertical Centrifugal Pumps. Though subject to many of the same troubles as horizontal pumps, vertical units are likely to have different causes of their troubles. This is because of differences in construction, installation, and operating conditions. Typical troubles, causes, and remedies are given below. PUMP WILL NOT START

impeller Locked: Sand causes’ many locks. Try to raise and lower the impeller adjusting nut This may free them. If it does not, backwash the pump, using clear water or whatever other liquid the pump normally handles. Try turning the shaft at its top, using a small pipe wrench. Be careful the shaft is easily damaged by a wrench. If the impellers cannot be freed, pull the pump and tear down the bowl assembly to get at the rotating parts. Trash in Casing: Rags, wood, Or metal jammed, in the pump may prevent it from turning. Tear down the pump and remove the obstruction. Fit the suction with a strainer to keep trash out of the pump Corrosion or Growths: Pumps that are out of service for long periods may be locked tight. Use acid or other recommended chemicals to remove corroded matter or growths from the pump. Packing Too Tight: Adjust so there is enough leakage for cooling and lubrication of the pump shaft. Too Much Bearing Friction: Use the right oil; consult the pump builder for the correct viscosity range. Check the tube tension nut for tightness. See if the pump shaft is bent; replace, if necessary. Check the anchoring of the pump head to see that it has not caused bending and distortion of the pump. Return bent shafts and columns to the builder’s factory for new ones. See that water-lubricated rubber bearings are wetted and free of sand. The wrong tension on the shaft enclosing tube of oil-lubricated pumps may throw the bearings out of line. If the well in which the pump is installed is so crooked that it causes misalignment of the pump, have it reamed to a larger diameter or install a smaller pump. Motor or Wiring Faulty: Check the circuit breaker or fuses for an open line. If the starter overload relays have tripped, reset them. Disconnect the motor from the pump and see if it starts. If it does not, have an electrician look it over or check it as directed in the manual PP(T)SB PROPRIETARY DOCUMENT

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CENTRIFUGAL PUMP TROUBLE SHOOTING MYG’99

Impellers Not Adjusted Properly: Set the impellers high enough so there is room for the shaft stretch caused by hydraulic thrust. This adjustment should allow the shaft to turn freely; then the stretch caused by the rotor and shaft weight will not bind the pump. Well Cave-in: Outside help is needed to correct this condition. PUMP DOES NOT DELIVER LIQUID Wrong Rotation: Change the rotation of the motor. With a 3-phase motor, just switch any two power leads. Speed Too Low: Check voltage and frequency of the power supply. See if excessive bearing friction, corrosion, or obstruction of the impellers slows the pump. Check the gear ratio and motor speed if the pump is being operated for the first time. Look over belt-driven units for slippage or wrong pulley size. Pump Not Primed: Vent the well to the atmosphere so there is not a vacuum at the pump suction. All impellers of vertical turbine pumps must be under water or the liquid handled because these units will not ordinarily start discharging against a suction lift. A 4 to 10-ft (1.2192 to 3.048 m) npsh is needed for good operation. Have enough head on the pump to allow it to discharge at rated capacity. Well Over pumped: With excessive draw down, the pump may break suction and fail to deliver water. Reduce the pump capacity by throttling the discharge. Two other common causes of trouble of this type are failure of pump parts and too high a pumping head. Check for a broken shaft, broken bowl assembly, and loose column-pipe joints. Tighten loose impellers. Check discharge valves to see that they are open and that the check valves do not stick. If the water table has fallen, the suction lift may be too high for the unit. Clear a clogged suction pipe or impeller by back washing. If the well screen is plugged, help from an experienced well driller is probably needed. PUMP USES TOO MUCH POWER Causes of this trouble include over speeding, wrong lubricant, tight packing, impeller rub, wrong rotation, misalignment, tight bearings, excessive vibration of the pump or piping, incorrectly chosen pump, and high discharge pressure. Correct as directed. On vertical pumps fitted with water-lubricated shaft bearings, an air vent or air-relief valve may be needed on the column to allow water to enter the bearings. This is true of all vertical pumps where the liquid handled acts as the bearing lubricant.

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CENTRIFUGAL PUMP TROUBLE SHOOTING MYG’99

PUMP CAPACITY LOW Low Liquid Level: Vent the well. Check the pump inlet for excessive turbulence, vortexing, or eddies. The velocity of the liquid entering the pump must be that recommended by the manufacturer and the suction submergence sufficient. Check the bowls and well screen for sand, rust, or bacterial blocking. Impeller Wear: Metal loss from the outer tips of the impeller vanes reduces pump capacity. Loss at the inner or suction end has not much effect. If fully enclosed impellers have the usual wearing rings, the trouble may be in them. Look for excessive clearance. With semi-open impellers not having bottom shroud or wearing rings, a close running clearance is needed at the bottom of the vanes. Faulty Instruments: Make sure that the water-level reading is correct. See that flow meters for measuring pump capacity are adjusted and read properly. Check the pressure gages. Other causes include too great a head on the pump, piping and pump leaks, and too low a speed. Check and correct as given above. PUMP VIBRATES EXCESSIVELY Rough Operation: Check to see that the impeller and bowl passages are free of wood, rags, sand, and other material that might throw the pump out of balance. Check the driver by disconnecting it and operating it alone. Look for excessive wear in the rotating parts. Pump Taking Air: Check the water velocity at the pump inlet. See if there are any leaks in the well vent. Check if the suction head on the pump is sufficient. Over pumping a well so the water level is intermittently drawn down may cause the pump to “grab air” and is a common source of severe vibration. Bearing Troubles: Check the . lube oil or grease for grade and quantity. Look for too much sand in the water or liquid handled. Check pump alignment.

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Set-up: Fixtures and Analyzer

To use this quick reference guide, make sure that your fixtures are setup as shown. Always place Head A on the left and Head B on the right. STEP 1: Job Definition Enter machine dimensions and RPM as shown in the diagram. Use ⇑ and ⇓ to move between values. Use ⇐ and ⇒ to change machine descriptions. Press ENTER when complete.

Note: Use the slash (/) to enter fractions. Ex. 14 7/8 = 1 4. 7/ 8

STEP 2: Foot Pre-Check

STEP 3: Alignment Data

CSI recommends that you perform a foot check before attempting alignment. Use any key to toggle answer to “YES”.

To collect alignment data, rotate the heads to one of the four positions. Averaging automatically begins when the heads are within range of one of the positions and stops when the heads are moved to another position.

Press DELETE to start. Loosen hold-down bolt on indicated foot and wait for dial (on right) to stop. Press any number key to accept the value. Retighten bolt and proceed to the next foot. Press ENTER when finished. Use feeler gauges to determine shim correction. X = SOFT, XX = VERY SOFT, XXX = EXTREME.

After data has been collected from at least 3 of the 4 positions, press ENTER. CSI recommends using all four data points whenever possible. STEP 4: View Data This provides a graphic display of how your results compare to the specified tolerance levels. (Example shown is for a direct coupled machine.)

STEP 5: Machine Moves

Other Useful Information:

The Machine Moves screens indicate how much to move each foot.

Extensive on-line help is available for each step of the alignment process. Simply press the HELP key for complete, context sensitive user guidance. Consult manual for complete instructions.

Press INSERT to perform a live move.

The Bull’s eye shows job status: Center = Excellent 1st ring = Acceptable 2nd ring = Close

www.CSImeansReliability.com

QuickAlign Quick Start Guide

A little bit more about PUMP Cavitation Cavitation means different things to different people. It has been described as: • • • • •

A reduction in pump capacity. A reduction in the head of the pump. The formation of bubbles in a low pressure area of the pump volute. A noise that can be heard when the pump is running. Damaged that can be seen on the pump impeller and volute.

Just what then is this thing called cavitation? Actually it is all of the above. In another section of this series I described the several types of cavitation, so in this paper I want to talk about another side of cavitation and try to explain why the above happens. Cavitation implies cavities or holes in the fluid we are pumping. These holes can also be described as bubbles, so cavitation is really about the formation of bubbles and their collapse. Bubbles form when ever liquid boils. Be careful not to associate boiling with hot to the touch. Liquid oxygen will boil and no one would ever call that hot. Fluids boil when the temperature of the fluid gets too hot or the pressure on the fluid gets too low. At an ambient sea level pressure of 14.7 psia (one bar) water will boil at 212°F. (100°C) If you lower the pressure on the water it will boil at a much lower temperature and conversely if you raise the pressure the water will not boil until it gets to a higher temperature. There are charts available to give you the exact vapor pressure for any temperature of water. If you fall below this vapor pressure the water will boil. As an example: Fahrenheit

Centigrade

Vapor pressure lb/in2 A

Vapor pressure (Bar) A

40

4.4

0.1217

0.00839

100

37.8

0.9492

0.06546

180

82.2

7.510

0.5179

212

100

14.696

1.0135

300

148.9

67.01

4.62

Please note that I am using absolute not gauge pressure. It is common when discussing the suction side of a pump to keep everything in absolute numbers to avoid the use of minus signs. So instead of calling atmospheric pressure zero, we say one atmosphere is 14.7 psia at seal level and in the metric system the term commonly used is one bar, or 100 kPa if you are more comfortable with those units. Now we will go back to the first paragraph and see if we can clear up some of the confusion:

The capacity of the pump is reduced • •

This happens because bubbles take up space and you cannot have bubbles and liquid in the same place at the same time. If the bubble gets big enough at the eye of the impeller, the pump will lose its suction and will require priming.

The head is often reduced •

Bubbles unlike liquid are compressible. It is this compression that can change the head.

The bubbles form in a lower pressure area because they cannot form in a high pressure area. •

You should keep in mind that as the velocity of a fluid increase, the pressure of the fluid decreases. This means that high velocity liquid is by definition a lower pressure area. This can be a problem any time a liquid flows through a restriction in the piping, volute, or changes direction suddenly. The fluid will accelerate as it changes direction. The same acceleration takes place as the fluid flows in the small area between the tip of the impeller and the volute cut water.

A noise is heard •

Any time a fluid moves faster than the speed of sound, in the medium you are pumping, a sonic boom will be heard. The speed of sound in water is 4800 feet per second (1480 meters/sec) or 3,273 miles per hour (5,267 kilometers per hour).

Pump parts show damage •

•

The bubble tries to collapse on its self. This is called imploding, the opposite of exploding. The bubble is trying to collapse from all sides, but if the bubble is laying against a piece of metal such as the impeller or volute it cannot collapse from that side, so the fluid comes in from the opposite side at this high velocity proceeded by a shock wave that can cause all kinds of damage. There is a very characteristic round shape to the liquid as it bangs against the metal creating the impression that the metal was hit with a "ball peen hammer". This damage would normally occur at right angles to the metal, but experience shows that the high velocity liquid seems to come at the metal from a variety of angles. This can be explained by the fact that dirt particles get stuck on the surface of the bubble and are held there by the surface tension of the fluid. Since the dirt particle has weakened the surface tension of the bubble it becomes the weakest part and the section where the collapse will probably take place.

The higher the capacity of the pump the more likely cavitation will occur. Some plants inject air into the suction of the pump to reduce this capacity and lower the possibility of

cavitation. They choose this solution because they fear that throttling the discharge of a high temperature application will heat the fluid in the pump and almost guarantee the cavitation. In this case air injection is the better choice of two evils. High specific speed pumps have a different impeller shape that allows them to run at high capacity with less power and less chance of cavitating. You normally find this impeller in a pipe shaped casing rather than the volute type of casing that you commonly see.

A little bit more about troubleshooting centrifugal pumps and mechanical seals. One of the U. S. based Japanese automobile manufacturers has a unique method of troubleshooting any type of mechanical failure. The system is called the "Five Whys". It is a simple but powerful concept, nothing has been solved until the question "why ?" has been asked at least five times and a sensible answer has been given for each of the "why" questions. As an example: 1. Why did the seal fail? •

The lapped faces opened and solids penetrated between them. (solids can't get in until the faces open)

2. Why did the faces open? •

The set screws holding the rotary unit slipped due to a combination of vibration and system pressure.

3. Set screws are not supposed to slip. Why did the set screws slip? •

The seal was installed on a hardened sleeve.

4. Why was the seal installed on a hardened sleeve? •

This was a packing conversion and a stock sleeve was used.

5. Why couldn't the mechanic tell the difference between a hardened sleeve and a soft one? •

They were both stored in the same bin.

6 Why were they stored in the same bin? •

Because they had the same part number.

7. Why did they have the same part number? •

They should have had different part numbers. Once that problem is corrected, the failures will stop.

Now you get the idea! Needless to say you may have to go further than just five "whys". Let's try another example: 1. Why did the seal fail?

•

The pump was cavitating and the vibration caused the carbon face to crack.

2. Why was the pump cavitating? •

It did not have enough suction head.

3. Why didn't it have enough suction head? •

The level in the tank got too low.

4. Why did the level in the tank get too low? •

I don't know.

You have not finished five "whys" so you better go find out why the level in the tank go too low or the problem is going to repeat its self. In the above example the float got stuck on a corroded rod, giving an incorrect level indication. One more example should do it: 1. Why did the seal start to leak? •

The elastomer got hard and cracked.

2. Why did the elastomer get hard and crack? •

It got too hot.

3. Why did it get too hot? •

The pump stuffing box ran dry.

4. Why did the stuffing box run dry? •

It was running under a vacuum and it was not supposed to.

5. Why was it running under a vacuum? •

A Goulds pump impeller was adjusted backwards to the back plate and the impeller pump-out rings emptied the stuffing box.

6. Why was it adjusted backwards? •

Most of the pumps in the facility are of the Duriron brand and they normally adjust to the back plate. The mechanic confused the impeller adjustment method. He has since been retrained

This is a powerful trouble shooting technique. I hope you make good use of it.

A quick reference guide for mechanical seal failure Of all the seal related activities, analyzing mechanical seal failure continues to be the single greatest problem for both the consumer and the seal company representative. I have addressed this problem in several of my other technical papers. If you will take a little bit of time to familiarize yourself with the following outline you should feel a lot more comfortable the next time you are called upon to do some seal troubleshooting. I should mention here in the beginning that as you look over the failed seal components keep in mind that a rebuilt seal may have some marks that occurred during a previous failure, making them especially difficult to analyze, but regardless of the design mechanical seals fail for only two reasons: • •

Damage to one of the components The seal faces open prematurely.

We will start with damage. This damage is almost always visible. Look for : Corrosion - The elastomer swells or the other seal parts become "sponge like" or pitted. • • • •

•

The product you are sealing is attacking one of the seal components. The attack is coming from the cleaner or solvent used to clean the lines between batches or at the end of a "run". The attack is coming from lubricants put on the elastomers or seal faces. Petroleum grease on Ethylene Propylene O-rings will cause them to "swell up". Galvanic corrosion - Happens with dissimilar materials in physical contact and connected by an electrolyte. As an example: stainless steel can attack the nickel binder in a tungsten carbide face. Oxidizers and Halogens attack all forms of carbon including black O-rings.

KEEP IN MIND THE CORROSION INCREASES WITH TEMPERATURE Physical damage. • • • • • • • •

Wear or rubbing of a flexible component. Thermal shock of some seal face materials. Especially those that are hard coated or plated. Thermal expansion of the shaft or sleeve can break a stationary seal face or interfere with the free movement of a dynamic elastomer. The rotating seal hits something because of shaft deflection. Temperature extremes (both high and cryogenic) will destroy elastomers and some seal face materials. Erosion from solids in the product you are pumping. Fretting caused by the dynamic elastomer removing the passivated layer from the corrosion resistant shaft or sleeve. Fluid abrasion that can weaken materials and destroy critical tolerances.

•

• • •

A discharge recirculation line circulates high velocity liquid with entrained solids that can break a metal bellows and injure lapped seal faces, as well as interfere with the free movement of the seal. The elastomer or rubber part can swell and breaks the face. Problems at installation. This includes mishandling, setting at the wrong compression, putting the wrong lubricant on the elastomer etc. Fatigue of the springs caused by misalignment.

The seal faces opening prematurely is the second cause Scoring or wear of the hard face is the most common symptom of this failure. The scoring occurs because the solids imbed into the softer carbon face after they open. The seal faces must stay in contact, but there are all kinds of conditions that are trying to force or pull them open. Physical causes • • • • • • • • • • • • • • • • • • • • •

Axial shaft movement (end play or thrust). This is normal at start up. Radial shaft movement (run out or misalignment) Operating off of the pump's best efficiency point. Hysteresis caused by a viscous (thick) product. Centrifugal force tries to separate the faces in a rotating seal application. Hydrodynamic forces generated between the lapped faces. Pressure distortion caused during pressure peaks such as water hammer and cavitation. Thermal distortion that can cause the seal face to separate from its holder or "go out of flat". A failure to provide equal and opposite clamping across the stationary seal face will cause distortion. A hardened sleeve can cause the seal set screws to slip. A wrong initial setting of the face load. Springs can clog if they are located in the product. Loose set screws. If the sleeve is too soft they can vibrate out. Shaft tolerance and finish is out of specifications. The rotating shaft or seal hits something. The discharge recirculation line can force open the faces. Outside springs painted by maintenance people. A cartridge seal installation method can compress one set of faces and open the other. Vibration. Fretting hang up. Cartridge mounted stationary seals move excessively unless they have some type of "built in" self aligning feature.

Product problems . With a loss of an environmental control the fluid can:

• • • • • • • • •

Vaporize between the lapped faces forcing them open and causing a "chipping" of the carbon outside diameter as well as leaving solids between the lapped faces. Become viscous preventing the faces from following normal "run out". Solidify between the lapped faces or around the faces. Crystallize between the faces or around the dynamic portions of the seal. Build a film on the sliding components or between the faces causing them to separate. Be a slurry and/ or abrasive Operate in a vacuum causing the ingestion of air between the faces of some unbalanced seal designs. Swell up the dynamic elastomer, locking up the seal . Cause slipstick between the faces if the sealed fluid is a non, or poor lubricant

The common causes of shaft displacement. • • • • • • • • • • • • • • • • • • • • • •

Operating off the pump's best efficiency point (B.E.P.). Misalignment between the pump and its driver. The rotating assembly is out of balance. A bent shaft. A non concentric sleeve or seal. Vibration Slip stick Harmonic Induced Passing through, or operating at a critical speed. Water hammer in the lines. The stuffing box is not square to the shaft, causing misalignment problems. Pipe strain. An impeller adjustment is made to compensate for normal impeller wear. Thermal growth of the shaft in both a radial and axial direction. Bad bearings or a poor bearing fit. Two direction axial thrust at start up is normal. The motor is finding its magnetic center. Cavitation - there are five separate types of damage that can be observed. The sleeve moved when the impeller was tightened. The unit is pulley driven causing excessive side thrust The impeller is positioned too far from the bearings. This is a severe problem in mixer or agitator applications.

How to preventing product problems that cause premature seal failure. Control the environment in the stuffing box. • • •

Control the temperature in the seal area Use the correct spring or bellows compression. Use only hydraullically balanced seals.

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Select a low friction face combination. Avoid "dead ending" the stuffing box. Jacket the stuffing box Quench behind the seal with the correct temperture steam or fluid Use a gland jacket Utilize two seals with a barrier fluid between them Use heat tape around the stuffing box Use a heat pipe to remove heat from the stuffing box. Vent the stuffing box, especially in a vertical application Flush in a cool compatible liquid. Control the pressure in the seal area o Be sure to use only hydraulically balanced seals. o Discharge recirculation will raise the pressure if you put a restrictive bushing into the bottom of the stuffing box. o Suction recirculation will lower the pressure in the stuffing box. o Use two seals and let the barrier fluid control the pressure between the seals. o Cross connect the stuffing boxes to equalize the stuffing box pressures in a multi stage pump. o Stage the stuffing box pressure with tandem seals. o Impeller pump out vanes will lower stuffing box pressure. Give the seal more radial space o Bore out the existing stuffing box if it is possible. o Make or buy a new back plate with the large stuffing box cast into it. o Make or buy a large bore stuffing box and attach it to the back plate after you have machined the old one off. Flush the product away if you are unable to control it. o Suction recirculation will bring fluid into the stuffing box from behind the impeller, where it is usually cleaner. This works on most closed impeller pump applications and those open impeller pump applications where the impeller adjusts to the volute rather than the back plate. o Flush with a clean liquid from an outside source. o A pressurized barrier fluid between two seals can keep solids from penetrating between the faces if the faces should open. This application will also work if the solid particles are less than one micron in diameter (Kaoline is such a product).

Build the seal to compensate for operating extremes. Slurry features that can be part of seal design. • • • •

Springs out of the fluid Teflon coating the metal parts so particles will not stick to sliding components. The elastomer moves to a clean surface as the face wears. Keep the sealing fluid on the outside diameter of the seal to take advantage of centrifugal force that will throw solids away from the lapped faces.

• • • •

Rotate the fluid with the seal to prevent erosion of the seal components. A simple vane arrangement can accomplish this. Use two hard faces if you find it impossible to keep the lapped seal faces together. Use a pumping ring to keep solids away from the faces. Mount the seal closer to the bearings to diminish the affect of shaft deflection.

Design for higher temperature capability • • • • • • • •

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Eliminate elastomers when ever possible. If you cannot eliminate elastomers, the O-ring location becomes important. Try to move the elastomer away from the faces. Hydraulically balanced seals generate less heat. Select low friction faces. Fool proof, correct installation dimensions are necessary. A cartridge design is your best choice. Keep a good product circulation around the components. A good lapping technique will keep the faces flat at high and cryogenic temperatures. Pumping rings will keep fluid circulating between two seals. If you are using balanced seals a simple convection tank is usually more than adequate. An air operated diaphragm pump can be used in the line to increase the circulation. Try to avoid the use of petroleum based fluids as the barrier or buffer fluid between the seals. Petroleum based fluids have a very low specific heat that will increase the temperature between the seals, Gland features such as quenching, recirculation, venting and flushing help. Choose well designed faces that will resist thermal distortion. The closer you get to a "square block" design, the better off you are going to be. Do not insulate the faces with an elastomer.

Design for pressure resistance • • • • • •

Limit the number of diameters in any single seal component Laminated bellows will allow you to keep a low spring rate while maintaining pressure capability, if you are using a welded metal bellows design. Finite element analysis of the seal components will prevent pressure and temperature distortion. Use more mass to resist hoop stresses. Higher modulus materials will resist bending and deformation. Use a tandem seal design for pressure break down between two seals.

Design for corrosion resistance • • •

Choose good materials, clearly identified by type and grade. Eliminate elastomers when possible. Elastomers are the most corrosion sensitive part of the seal. Design non stressed parts when ever possible

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Try not to weld any of the metal components. If it is necessary, monitor the temperature to prevent inter granular corrosion Control the temperature. Corrosion increases with temperature. Use non metallic materials for non metallic equipment. Watch out for galvanic corrosion when using dissimilar materials. Do not use stainless steel springs. Stick with Hastelloy "C" if the metal parts of the seal are manufactured from iron, steel, stainless steel, or bronze. If the seal is manufactured from a different metal, use springs manufactured from that material. Do not depend upon flushing to provide corrosion resistance. Use the correct materials, keeping in mind that solvents and steam are sometimes used to flush the lines. Any materials that you select must be compatible with these flushing or cleaning fluids also.

If you need cryogenic capability • • • • •

Go to a welded metal bellows configuration to eliminate all elastomers. You will need a special carbon/ graphite face that has an organic material impregnated to assist in the release of the graphite. Avoid plated or coated hard faces. Differential expansion will cause them to crack. Always lap the faces at a cryogenic temperature. Do not coat the faces with grease or oil. It will freeze at cryogenic temperatures.

An overview of seal troubleshooting Seal problems are almost always associated with face leakage, but as we will soon learn there are other leak paths in addition to the obvious one between the lapped seal faces. In the following paragraphs, we will be looking at all of these leak paths. Keep in mind that seals are classified into many categories : stationary, rotary, balanced, unbalanced, inside, outside, metallic, non-Metallic, single, dual, elastomer, metal bellows, rubber bellows, cartridge, split, solid, etc.. These classifications were described in another paper in this series Try to keep these classifications in mind as we investigate the cause of seal failure. As with former papers, I will be presenting the troubleshooting hints in an outline form. You should not find these terms confusing because I have assumed you have a pretty good knowledge of mechanical seals or otherwise you would not be attempting to trouble shoot them. In the event you do have trouble with some of the terms or techniques any representative of a reputable seal company should be able to explain them to you. LEAKAGE AT THE SEAL FACES. The seal face is not flat. (Flatness should be measured within three helium light bands, (0,000033" or 1 micron) • •

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The face was damaged by mishandling. Poor packaging. The seal should be able to survive a 39" (1 meter.) drop. To insure this, the seal must be shipped in a reusable box insulated with plenty of foam or any other adequate insulation. The face was distorted by high pressure or surges in pressure. "Water hammer" would be an example. It was distorted when you tightened the stationary face against an uneven surface. The clamping is not "equal and opposite" across the stationary face. This is a common problem with "L" shaped and "T" shaped stationary faces. The "hard" seal face has been installed backwards. You are running on a non lapped seal face. It is common practice to lap only one side of a hard face. The face is being distorted by a change in temperature. This happens when you forget to vent a vertical pump. The face never was flat. You have a bad part. The carbon metal composite was not stress relieved after the carbon was "pressed in".

The face has been chemically attacked. • •

Oxidizing agents attack all forms and grades of carbon graphite. Some de ionized water will attack any form of carbon.

• • •

Corrosion increases with any temperature increase. A 10 ° Centigrade (18°F.) rise in temperature will double the corrosion rate of most corrosives. A cleaner or solvent is being flushed through the lines and it is attacking the carbon. You are using a poor grade of Carbon. Go to an unfilled grade such as Pure Carbon Company grade 658 RC. This is a common occurrence if the seal is being repaired by some one other than the original manufacturer.

The plating or hard coating is coming off of the hard face. •

• •

All coatings are porous. The chemical is penetrating this porous coating and attacking the bond between the coating and the base material, or the base material its self. An inferior plating was originally put on the base material. Differential expansion of the dissimilar materials is causing them to separate.

The seal face is cracked, pitted or damaged. •

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High temperature is heat checking (cracking) the plated face. This is a common problem with cobalt based tungsten carbide. The nickel base version is less likely to crack. The product is solidifying between the faces and they are breaking at start up. Most face materials have high compressive strength, but tend to be weak in tension. Excessive vibration is causing the drive pins to crack the face. Low cost seals experience this problem quite often. There is a high temperature differential across the ceramic. 7 to 10 cycles can break even good ceramics in hot water or hot petroleum products. Air is trapped in the Carbon. Heat is causing it to expand and blow out pieces of the carbon face. The carbon usually blisters prior to blowing out. The solution is to go to a more dense carbon. The product is vaporizing and allowing solid material to blow across the lapped face. This is a common occurrence in boiler feed water applications. The seal faces have opened, solids penetrated and imbedded into the soft carbon causing rapid wear in the hard face. The same problem occurs if the carbon was relapped using lapping powder. Lubricant, on the faces is freezing in cryogenic (cold) applications. The elastomer is being chemically attacked and swelling up. This can break the face in those seal applications where the elastomer is positioned in the seal inside diameter. In some instances the swelling elastomer will open up the two faces, allowing the solids to penetrate. This can be a problem with boot mounted faces The rotating shaft, or sleeve, is hitting the stationary face. This can happen if the pump is running off of its B.E.P., which almost always occurs at start up. The seal is being mishandled during installation. Good packaging and proper training can solve many of these problems.

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The crack may have occurred during disassembly. Check to see if there is discoloration deep in the crack. Discoloration means that it occurred during, or before, operation. Petroleum products can "coke" at the face causing pieces of carbon to be pulled out as the face rotates. You will have to select two hard faces for this application. The rotating face is not centered in the stationary face and is running off the edge of the stationary face. Look for rubbing marks around the O.D. of the rotary unit. A bent shaft or out of balance rotating assembly is the most common cause You will notice a much wider wear track if you are experiencing this problem. The seal will appear to "spit" as lubricant is dragged across the face and off the seal outside diameter. Dirt can be dragged across the faces as they separate.

The movable face is not free to follow whip, wobble or run out. • • • • • • • • • • • • • • • • • • • • •

The rotating face is hitting the I.D. of the stuffing box. The recirculation line from the pump discharge is aimed at the seal faces and interfering with their free movement. Dirt or solids are clogging the movable components. Magnetite is a very big problem in most hot water applications. The product is interfering with the free movement of the components. It is: Crystallizing ( like sugar) Solidifying (like glue) Viscous (molasses) Building a film on the sliding components ( hard water or paint) Coking (oil or any other petroleum product) The elastomer has been chemically attacked causing it to swell up and interfere with free movement of the face. Temperature growth of the shaft is interfering with the free movement of the movable face. The shaft or sleeve is the problem. It is over size - + 0.00" - 0.002" ( 0,00-0,05 mm.) is ideal. It is too rough; it should be at least 32 R.M.S. (0,8 microns) It is fretted, corroded or damaged in some way. Solids have attached themselves to that portion of the shaft where the dynamic elastomer is located. A gasket or fitting is protruding into the stuffing box. Solids from outside the stuffing box are getting under the faces. This is a common problem with vertical pumps. The elastomer is spring loaded and the interference on the shaft is restricting the face movement. The elastomer has extruded because of high pressure or excessive clearance. A foreign object has passed into the seal chamber and is interfering with the free movement of the seal.

The product has plated, or formed on the face and a piece of it has broken off.

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This problem occurs with products that are sensitive to temperature and/ or pressure changes.

The set screws have come loose. • • • •

The shaft has been hardened. They have worked loose in a sleeve that is too soft. The hardened set screws have corroded. They were not replaced when the seal was rebuilt and as a result are not "digging" into the shaft.

The face has lost its spring load. • • • • •

The initial setting was wrong. Temperature growth of the shaft has altered the original setting. The impeller has been adjusted towards the wet end of the pump. The sleeve moved when the impeller was tightened to the shaft. The cartridge seal was pushed on the shaft by pushing on the gland and the seal is now over compressed. In a dual seal application this will over compress the inner seal and open up, or unload the outer seal.

The product is vaporizing and blowing the faces open. This happens in hot applications if there is water in the product. •

It can also occur if the pump/seal was hydrostatically tested with a water base fluid.

The inner seal, of a dual seal application was not balanced in both directions and is opening up with reversing pressure. This is a common problem in unbalanced seals that are subject to both vacuum and pressure or if the barrier fluid pressure varies. The single spring, found in some seal designs, was wound in the wrong direction for the shaft rotation. The Bellows seal has lost cooling and the anti vibration lugs are engaging the shaft. Shaft movement will cause the faces to open. LEAKAGE AT THE ELASTOMER LOCATION. Compression set ( the elastomer has changed shape). • •

Either the product is too hot or there is too much heat being generated at the seal faces. You must vent vertical pumps to prevent this problem. This is a common problem with most grades of Dupont's Kalrez® material, preventing it from being free to flex and roll.

The elastomer is cracked. • • • • • •

The shelf life has been exceeded. Buna N (Nitrile) has a shelf life of only twelve months because of its sensitivity to ozone attack. High heat is the main cause. Chemical attack. In most cases the elastomer swells but cracking and shrinking does occur in isolated cases. Cryogenic (cold) temperatures freeze the elastomer and it will crack when hit.. The rubber bellows did not stick to the shaft because the wrong lubricant was used. The shaft turned inside the bellows causing high heat. The seal faces stuck together. The shaft was turning inside the rubber bellows causing excessive heat.

The elastomer is cut or damaged. • • • • •

Mishandling. The elastomer was slid over a rough spot on the shaft or sleeve. Be careful of old set screw marks, splined shafts, key ways, etc. It was extruded by high pressure. You may need a backup ring. The product is penetrating into the elastomer and blowing out the other side. This problem is a common occurrence when you are trying to seal ethylene oxide. Teflon jacketed O-Rings can split in the presence of halogenated fluids. The halogen will cause the elastomer to swell up, inside of the Teflon jacket. Halogens can be recognized because most of them end in the letters "ine", such as bromine, astintine, chlorine, fluorine, iodine, etc..

The elastomer is not seated properly. • • • • • • •

It was twisted during installation. High pressure can cause the elastomer to extrude or twist in the O-Ring groove. Solids have "built up" or penetrated between the elastomer and the shaft. The shaft is corroded, damaged, or fretted. The shaft is oversized. Excessive travel can cause the elastomer to "snake". Most o-rings can roll up to one half of their diameter. The O-ring groove is damaged or coated with a solid material.

The elastomer has swollen or changed color. • • • •

Product attack. This is the most common cause and usually occurs within five to ten days The wrong lubricant was used at installation. As an example, you should never put petroleum grease on EPR O-rings. Solvents or chemicals used to clean the lines are not compatible with the elastomer. Steam can harm many elastomers including most grades of Viton®.

•

Oxidizers can attack the carbon black in O-Rings and other elastomers.

The elastomer leaks when pressurized in the opposite direction. • •

A common problem with unbalanced, dual seal applications. Two way balanced seals are recommended for these applications. Remember that O-Rings are the only common elastomers that seal in both directions. Wedges, U cups, and chevrons do not have this ability.

OTHER LEAK PATHS TO CONSIDER Between the carbon and its metal holder. • •

Some seal companies, and most seal repair facilities, glue the carbon in place. The glue may not be compatible with the product you are sealing. "Pressed in" carbons can leak in a high temperature application because of the differential expansion between carbon and metal. Low expansion metal is available for these applications

Between the shaft and the sleeve. • •

Damaged gasket or gasket surface. Distorted sleeve or shaft.Many packed, double ended pumps have this problem because there is no gasket between the impeller and the sleeve that is holding it in place.

Stationary face gasket or elastomer leaking. •

This leak path is not always visible. It often looks like face leakage.

Gland gasket or gasket surface leakage. •

This leak path should always be visible.

Pipe flange leaking above the seal and dripping into the seal area. •

I found this one after every other avenue was exhausted.

At the weld location if a seal face holder is welded to the cartridge sleeve. At the pipe connections, ancillary hardware, A.P.I. Gland fittings, and recirculation lines. A scratch or nick in the o-ring groove. Remember that up to 100 p.s.i. (6 bar) o-rings seal on the O.D. and the I.D. not the sides.

Seal faces will not leak visibly if they are lapped flat and we keep them in total contact. Shaft movement is the main contributor to the opening of the seal faces and allowing solids to penetrate. Shaft movement is caused by many factors. In the following paragraphs we will be looking at most of them. CAUSES OF EXCESSIVE SHAFT MOVEMENT, INCLUDING VIBRATION. Cavitation • • • •

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Vaporization caused by too high a product temperature or too low a suction head. Air is entering the stuffing box. A common problem with pumps that run in a vacuum or taking a suction from an evaporator or condenser. Internal recirculation. Occurs when the Suction Specific Speed is too high, or when either the impeller or wear ring clearance becomes excessive. The vane passing syndrome occurs if the O.D. of the impeller is too close to the pump cutwater. This clearance should be at least 4% of the impeller diameter in the smaller size impellers and at least 6% in the larger diameter impellers (greater than 14 inch or 355 mm.) Turbulence. Occurs if there is not laminar flow in the lines.

The bearings are worn excessively. • • •

Contamination of the lubricant is the biggest cause. Grease or lip seals have a useful life of only 2000 hours (84 days). Poor fit or installation. Serious misalignment. The misalignment can be the result of pipe strain or misalignment between the pump and its driver.

The shaft is bent. • • •

Usually occurs during sleeve removal or if the bearing was installed with an arbor press. Improper storage with the long shaft supported only on the ends. Heating the shat to remove the sleeve is another common cause

The impeller is out of balance. • • • •

The impeller was damaged by either wear, corrosion or cavitation. Product has built up on the vanes or in the balance holes. The impeller diameter was reduced and the impeller was not re balanced The impeller never was balanced.

An unbalanced rotating assembly. Pressure surges or water hammer.

Worn coupling. The pump is operating off of its best efficiency point. Rubbing of a rotating component. • • • • •

The shaft is hitting the wear ring, or a stationary wear ring is contacting a rotating wear ring. The shaft is hitting the seal gland or stationary face. A seal rotating component is hitting the stuffing box I.D.. A recirculation line aimed at the seal faces is causing a pulse each time the impeller vane passes the fitting. A gasket or fitting is protruding into the stuffing box.

The stationary seal face is not perpendicular to the rotating shaft. This causes the spring loaded, rotating face to move back and forth twice per revolution. • • • • • • • • •

The stuffing box face is not square to the shaft. The stuffing box face is often a rough casting. Tightening the gland bolts through a gasket is cocking the stationary face. Pipe strain. Temperature growth. A convection tank, or some other heavy device is hanging off of the gland. Bearing fit or wear. Coupling alignment. Shaft deflection. The deflection can be caused by operating the pump off of its best efficiency point, the rotating assembly is out of balance, or the shaft is bent. Poor installation technique.

VIBRATION AT THE SEAL FACES. Harmonic vibration. •

The seal is vibrating in harmony with some rotating component. The same thing that causes a rear view mirror to vibrate in an automobile. Most harmonic vibration can be stopped by changing the speed of the equipment or "damping" the vibrating component.

Slipstick--caused by: • • • • • •

Poor lubricating fluids. Hot water. Solvents. Some detergents. Gases Dry running applications.

• • • • •

Too high a face load. Using unbalanced seals. Poor installation technique. Face load has changed because of temperature growth or impeller adjustment. You are using a high friction face combination. Often occurs if you use two hard faces.

A discharge recirculation line aimed at the seal faces. •

Each time the impeller passes the recirculation connection it causes a pulse of fluid at the seal face.

Vaporization of the product at the seal face. • •

Happens frequently with products that contain water, and are operated at elevated temperature. Can occur at the seal face because of high face load or if you use unbalanced seals.

EXCESSIVE AXIAL MOVEMENT OF THE SEAL • • • • • •

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Temperature growth. The impeller was adjusted, after the seal was installed, to compensate for wear. The rotor motor, moved to its magnetic center at start up. The equipment is equipped with sleeve or babbitted bearings and has excessive end play. Shaft thrust. There is a thrust towards the bearings caused by the combination of the fluid changing direction in the impeller and acting on the shaft and/or impeller surfaces. This thrust is offset by a thrust towards the wet end caused by the impeller shape. In centrifugal pumps the resulting force can be in either direction, depending upon how close the pump is operating to its best efficiency point. Above 65% of its best efficiency, the thrust is towards the wet end. Below 65% of the best efficiency the thrust is towards the power or bearing end. There is little to no movement at 65% of the pumps best efficiency. This means that at start up the shaft moves in both directions accounting for a higher percentage of seal failure at start up. Vertical mixer shafts often lift vertically when solids are mixed with liquid.

SHAFT NOT CONCENTRIC WITH THE STUFFING BOX, this will cause a wiping action in stationary seals. • • • •

The shaft is bending as you move away from the pump B.E.P. It bends at 240 degrees, from the cutwater, at low flow and high head. It bends at 60 degrees, from the cutwater, at high flow and low head. Coupling misalignment.

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Poor bearing fit. Pipe strain. Temperature growth causes the stuffing box to move relative to the shaft. The sleeve is not concentric with the shaft. The seal is not concentric with the sleeve/ shaft. A bolted on stuffing box has slipped. The back plate is not machined concentric to the stuffing box.

Heat is always an indication of wasted energy, but it can also have a disastrous affect on seal life and performance. Let's take a look at what is causing this heat CAUSES OF HIGH HEAT AT THE SEAL FACES. Too much spring compression. • • • • • • •

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Installation error. No print was used or the mechanic cannot read the print he was given. The shaft was marked in the wrong location. The mechanic used the wrong marking tool. The mark is too wide. The sleeve moved when the impeller was tightened. The impeller was adjusted after the seal was installed. A cartridge seal was installed on the shaft, by pushing on the gland. Interference from the sleeve elastomer has caused an over compression of the seal. In some dual seal applications the outer seal will become under compressed. The shaft moved because of thrust. Thermal growth of the shat.

Problems with some seal designs. • • • • • • •

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Unbalanced seals are supplied by original equipment companies. They generate more heat than balanced seals. The elastomer is located too close to the seal faces. The heat generated at the faces is affecting both the elastomer and the seal face. The carbon face is insulated by an elastomer. The face is too wide causing the hydraulic force to generate excessive heat. The Carbon seal face is too narrow causing excessive heat from the spring pressure. A vertical seal installation is not being vented. The faces are running dry in a bubble. Speeds above 5000 F.P.M. (25 m/sec) require a special balance and less spring load. A 60/40 balance and a face load of 8 psi. to 15 psi. ( 0,07 to 0,2 n/mm2) would be normal. An outside metal or elastomer bellows seal is almost impossible to vent. Spring loaded elastomers cause varying seal face loads. The actual load depends upon shaft tolerance and installation dimension.

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Some seal faces are glued in. The glue acts as an insulator preventing the face heat from conducting to the metal holder. Many single spring designs are uni-directional requiring both right handed and left handed seals on a double ended pump. Many metal bellows designs lack effective vibration damping. Stationary seal designs require clean flushing if solids are present. centrifugal force does not throw the solids away from the moveable (spring loaded) components.

Problems with face materials. • • •

Heat conductivity is low in some materials. (ceramic, carbon, Teflon) The coefficient of friction varies with face combinations and various sealing products. Carbon/ metal composite faces conduct heat better than plain carbon/ graphite, as long as there is a true interference fit and they are not glued together to hold them in place.

Problems with the pump operation that causes high heat at the faces. • • • • • • • • • • • • •

Operating off of the B.E.P The degree of the problem is determined by the L3/D4 ratio. Operating at vapor point, causing cavitation. Running dry. Gases. Dry solids. Pumping a tank dry. Losing barrier fluid in a dual seal application. Shutting off the flushing water. Vacuum applications. Vertical pumps not vented in the stuffing box. The liquid is not a lubricant. Pump out rings on the back of the impeller running too close to the pump back plate.

Other causes of high heat. • • • • • • • • •

The shaft, or sleeve is rubbing a stationary component. The gland. The bushing in the bottom of the stuffing box. The bushing in the A.P.I. gland. A pump wear ring. A protruding gasket. A fitting. The stationary portion of a mechanical seal. The shaft, or sleeve, is not straight.

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It is bending, because the pump is operating off of its best efficiency point. It is bent. This often happens when the sleeve is removed. The rotating assembly is not balanced. The shaft never was straight. There is not enough circulation around the seal. Install a large diameter stuffing box. You should be able to get at least 1" (25 mm.) all around the rotating unit. Connect a recirculation line from the bottom of the stuffing box to the suction side of the pump. You can do this in almost every case except when you are pumping a product at its vapor point or if the solids have a specific gravity lower than the fluid. The cooling jacket is clogged. There is no carbon restriction bushing in the bottom of the stuffing box and you are using the cooling jacket. The restriction bushing slows down the heat transfer. Loss of an environmental control. The flush is not constant. The pressure is changing. Quenching steam or water has been shut off during pump shut down. The double seal barrier fluid is not circulating. The cooling jacket has become clogged by the calcium in the hard water. Try condensate instead. The filter, or separator, is clogged. Either the suction or discharge recirculation line is clogged. If you are using double seals, remember that two seals generate twice as much heat and conventional cooling may not be sufficient. Contact the manufacturer for the rules when using convection tanks and dual seals. You may need a "built in" pumping ring. Solids in the stuffing box are interfering with a rotating component.

® DuPont Dow elastomer

Causes of overheating in cartridge mechanical seals Too much heat can cause multiple problems with mechanical seals: • • • • • • •

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The elastomer (rubber part) can be damaged. Some seal faces can be damaged. Carbon-graphite faces can pit as trapped air expands within the carbon, or the product carbonizes and pulls out pieces of the seal face. Plated faces can heat check and crack causing rapid carbon face wear. The filler in some carbon /graphite compounds can melt or oxidize at elevated temperatures. Critical dimensions can change causing the lapped seal faces to go out of flat and leak prematurely (especially fugitive emissions). The sealed product can change state and : o Vaporize between the faces opening them. o Crystallize on the moving components, restricting their movement. o Change fluid viscosity restricting the ability of the seal to follow run out. o Solidify, making the seal inoperable. o Build a film on sliding components and the lapped seal faces. o Carbonize or coke restricting the seal movement and opening the lapped faces. Corrosion always increases with increasing temperature.

Some heat problems are not seal design or seal installation related: •

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•

• • •

An inefficient heating or cooling jacket on the pump. o A layer of calcium or some other similar product has built up on the jacket walls, interfering with the heat transfer. o The coolant is flowing too rapidly through the cooling jacket. o A thermal bushing was not located in the bottom or end of the stuffing box. o If steam is being used as the coolant, the pressure is too high. o The fluid is not "dead ended" in the stuffing box. There is either suction or discharge recirculation of the pumping fluid. o Clearance between the seal outside diameter and the stuffing box bore is not sufficient. The shaft material is conducting the product heat to the cartridge static elastomer and other components. As an example: carbon steel conducts heat much better than a stainless steel shaft. The dual seal convection tank is not convecting. o The convection tank is running backwards. o The dual seal barrier or buffer fluid has been shut off. The quench has failed. The product has a low specific heat and poor conductivity. Oil is a good example of such a product. The seal faces were over-compressed during the installation process.

• • • •

A wrong installation measurement was used. The mechanic did not read, or understand the print dimension. The pump sleeve moved as the impeller was tightened on the shaft. The measurement was taken at the wrong place. The stuffing box face is the only safe reference point.

The cartridge seal design has a major affect on heat generation and heat sensitivity: • • • • •

Unbalanced seals generate more heat than hydraulically balanced mechanical seals. Two hard faces generate more heat than carbon/graphite vs. a hard face. Silicone carbide and tungsten carbide dissipate heat faster than 99.5 ceramic or carbon&endash;graphite. The location as well as the grade of the elastomer can be critical in temperature sensitive applications. In dual seal applications, convection systems are not as efficient as pumping rings or forced circulation of the barrier fluid system. When oil is used as a barrier fluid forced circulation or the use of a pumping ring is mandatory.

The above problems are not unique to cartridge seals, there are however some problems that are unique: •

•

• • •

•

Pushing the seal gland along the shaft and against the stuffing box face can over compress the seal because of the friction between the shaft and the cartridge sleeve static elastomer. In dual seal applications the inner seal can over compress as the outside seal looses some of its compression. Be sure to reset the spacing device (usually "clips" of some type) prior to locking the seal to the shaft. Some open impeller pump designs (Duriron as an example) adjust to the back plate rather than the volute. Be sure to reset the cartridge seal after the impeller adjustment. Cartridge set screws can slip on a hardened sleeve. The system pressure can then over compress the seal. Higher pressure applications, or water hammer can move the set screws and over compress the seal faces. Be sure to re-tighten the adjusting nuts after making the impeller micrometer adjustment on those pump that uses that type of adjustment method. The Chesterton System #1 pump is a good example of this design. Make sure the centering-positioning clips are in place when installing or resetting the seal for proper face loading.

Cavitation Cavitation means that cavities are forming in the liquid that we are pumping. When these cavities form at the suction of the pump several things happen all at once. • • • •

We experience a loss in capacity. We can no longer build the same head (pressure) The efficiency drops. The cavities or bubbles will collapse when they pass into the higher regions of pressure causing noise, vibration, and damage to many of the components.

The cavities form for five basic reasons and it is common practice to lump all of them into the general classification of cavitation. This is an error because we will learn that to correct each of these conditions we must understand why they occur and how to fix them. Here they are in no particular order : • • • • •

Vaporization Air ingestion Internal recirculation Flow turbulence The Vane Passing Syndrome

Vaporization . A fluid vaporizes when its pressure gets too low, or its temperature too high. All centrifugal pumps have a required head (pressure) at the suction side of the pump to prevent this vaporization. This head requirement is supplied to us by the pump manufacturer and is calculated with the assumption that fresh water at 68 degrees Fahrenheit (20 degrees Centigrade) is the fluid being pumped. Since there are losses in the piping leading from the source to the suction of the pump we must determine the head after these losses are calculated. Another way to say this is that a Net Positive Suction Head is Required (N.P.S.H.R.) to prevent the fluid from vaporizing. We take the Net Positive Suction Head Available (N.P.S.H.A.) subtract the Vapor Pressure of the product we are pumping, and this number must be equal to or greater than the Net Positive Suction Head Required. To cure vaporization problems you must either increase the suction head, lower the fluid temperature, or decrease the N.P.S.H. Required. We shall look at each possibility: Increase the suction head • • •

Raise the liquid level in the tank Raise the tank Put the pump in a pit

•

• •

Reduce the piping losses. These losses occur for a variety of reasons that include : o The system was designed incorrectly. There are too many fittings and/or the piping is too small in diameter. o A pipe liner has collapsed. o Solids have built up on the inside of the pipe. o The suction pipe collapsed when it was run over by a heavy vehicle. o A suction strainer is clogged. o Be sure the tank vent is open and not obstructed. Vents can freeze in cold weather o Something is stuck in the pipe, It either grew there or was left during the last time the system was opened . Maybe a check valve is broken and the seat is stuck in the pipe. o The inside of the pipe, or a fitting has corroded. o A bigger pump has been installed and the existing system has too much loss for the increased capacity. o A globe valve was used to replace a gate valve. o A heating jacket has frozen and collapsed the pipe. o A gasket is protruding into the piping. o The pump speed has increased. Install a booster pump Pressurize the tank

Lower the fluid temperature • • •

Injecting a small amount of cooler fluid at the suction is often practical. Insulate the piping from the sun's rays. Be careful of discharge recirculation lines, they can heat up the suction fluid.

Reduce the N.P.S.H. Required • • • • •

Use a double suction pump. This can reduce the N.P.S.H.R. by as much as 27% or in some cases it will allow you to raise the pump speed by 41% Use a lower speed pump Use a pump with a larger impeller eye opening. If possible install an Inducer. These inducers can cut N.P.S.H.R. by almost 50%. Use several smaller pumps. Three half capacity pumps can be cheaper than one large pump plus a spare. This will also conserve energy at lighter loads.

It is a general rule of thumb that hot water and gas free hydrocarbons can use up to 50% of normal cold water N.P.S.H. requirements, or 10 feet (3 meters), whichever is smaller. I would suggest you use this as a safety margin rather than design for it. Air ingestion A centrifugal pump can handle 0.5% air by volume. At 6% air the results can be disastrous. Air gets into as system in several ways that include :

• • • • • •

Through the stuffing box. This occurs in any packed pump that lifts liquid, pumps from a condenser, evaporator or any piece of equipment that runs in vacuum. Valves above the water line. Through leaking flanges Vortexing fluid. A bypass line has been installed too close to the suction. The suction inlet pipe is out of fluid. This can occur when the level gets too low or there is a false reading on the gauge because the float is stuck on a corroded rod.

Both vaporization and air ingestion have an affect on the pump. The bubbles collapse as they pass from the eye of the pump to the higher pressure side of the impeller. Air ingestion seldom causes damage to the impeller or casing. The main effect of air ingestion is loss of capacity. Although air ingestion and vaporization both occur they have separate solutions. Air ingestion is not as severe as vaporization and seldom causes damage, but it does lower the capacity of the pump. Internal Recirculation This condition is visible on the leading edge of the impeller, and will usually be found at the discharge tip working its way back to the suction. It can also be found at the suction eye of the pump. As the name implies the fluid recirculates increasing its velocity until it vaporizes and then collapses in the surrounding higher pressure. This has always been a problem with low NPSH pumps and the term SPECIFIC SUCTION SPEED was coined to give you a guide in determining how close you have to operate to the B.E.P. of a pump to prevent the problem. The higher the number the smaller the window in which you have to operate. The numbers range between 3,000 and 20,000. Water pumps should stay between 3,000 and 12,000. Here is the formula to determine the suction specific speed number of your pump:

rpm = Pump speed gpm = Gallons per minute or liters per second of the largest impeller at its BEP Head= Net positive suction head required at that rpm • •

For a double suction pump the flow is divided by 2 since there are 2 impeller eyes Try to buy pumps lower than 8500.(5200 metric ) forget those over 12000 (8000 metric) except for extreme circumstances.

• • • •

Mixed hydrocarbons and hot water at 9000 to 12000 (5500 to 7300 metric) or higher, can probably operate satisfactorily. High specific speed indicates the impeller eye is larger than normal, and efficiency may be compromised to obtain a low NPSH required. Higher values of specific speed may require special designs, and operate with some cavitation. Normally a pump operating 50% below its best efficiency point (B.E.P.) is less reliable.

With an open impeller pump you can usually correct the problem by adjusting the impeller clearance to the manufacturers specifications. Closed impeller pumps present a bigger problem and the most practical solution seems to be to contact the manufacturer for an evaluation of the impeller design and a possible change in the design of the impeller or the wear ring clearances. Turbulence We would prefer to have liquid flowing through the piping at a constant velocity. Corrosion or obstructions can change the velocity of the liquid and any time you change the velocity of a liquid you change its pressure. Good piping layouts would include : • •

• • • • • • • •

Ten diameters of pipe between the pump suction and the first elbow. In multiple pump arrangements we would prefer to have the suction bells in separate bays so that one pump suction will not interfere with another. If this is not practical a number of units can be installed in a single large sump provided that : The pumps are located in a line perpendicular to the approaching flow. There must be a minimum spacing of at least two suction diameters between pump center lines. All pumps are running. The upstream conditions should have a minimum straight run of ten pipe diameters to provide uniform flow to the suction bells. Each pump capacity must be less than 15,000 gpm.. Back wall clearance distance to the centerline of the pump must be at least 0.75 of the suction diameter. Bottom clearance should be approximately 0.30 (30%) of the suction diameter The minimum submergence should be as follows: FLOW

MINIMUM SUBMERGENCE

20,000 GPM

4 FEET

100,000 GPM

8 FEET

180,000 GPM

10 FEET

200,000 GPM

11 FEET

250,000 GPM

12 FEET

FLOW

The metric numbers are : MINIMUM SUBMERGENCE

4,500 M3/HR

1.2 METERS

22,500 M3/HR

2.5 METERS

40,000 M3/HR

3.0 METERS

45,000 M3/HR

3.4 METERS

55,000 M3/HR

3.7 METERS

The Vane Passing Syndrome You will notice damage to the tip of the impeller caused by its passing too close to the pump cutwater. The velocity of the liquid increases if the clearance is too small lowering the pressure and causing local vaporization. The bubbles collapse just beyond the cutwater and there is where you should look for volute damage. You will need a flashlight and mirror to see the damage unless it has penetrated to the outside of the volute. The damage is limited to the center of the impeller and does not extend into the shrouds. You can prevent this problem if you keep a minimum impeller tip to cutwater clearance of 4 % of the impeller diameter in the smaller impeller sizes (less than 14' or 355 mm.) and 6% in the larger impeller sizes (greater than 14" or 355 mm.). To prevent excessive shaft movement bulkhead rings can be installed in the suction eye. At the discharge rings can be manufactured to extend from the walls to the impeller shrouds.

A.P.I. (American Petroleum Institute) and C.P.I. (Chemical Process Industry) merger

Any prediction about the future of the pump and seal business would have to include the high probability that the CPI will adopt the API seal standard. The adoption of this standard will be enthusiastically supported by the CPI insurance companies and will dramatically increase the price of mechanical seals to the consumer as well as bring seals into a commodity status which has been the goal of some of the largest pump and seal manufacturers all along. Recent pump/seal mergers, buy outs, and alliances hint that the adoption of these new standards will also dramatically increase the profits of these highly competitive manufacturers. The API (American Petroleum Institute) standard is the one universal standard being used by oil refineries throughout the world. There is on going talk about combining this standard with the chemical industry ANSI (American National Standards Institute) standard for a single unified pump standard. The problem with all standards of this type is that they have produced a failure rate in mechanical seals that exceeds 85%. The only part of a mechanical seal that is sacrificial is the carbon face and in better than 85% of the cases there is plenty of carbon face left when the seal begins to leak. The A.P.I. specification addresses just about everything about mechanical seals. The subjects include: • • • • • •

Seal design Materials Accessories Instrumentation Inspection, testing and preparation for shipment. Manufacturing.

In this section we will be looking at just a few of those parts of the A.P.I. standard 682 that when combined with the C.P.I. standard, will be affecting your seal purchases in the near future. Most of this information was taken from A.P.I. Standard 682, First Edition, dated October 1994. I recommend you get hold of a copy of this and any future updates to learn the full particulars. 2.1.1 •

2.1.2

All standard mechanical seals, regardless of type or arrangement, shall be of the cartridge design.

•

The standard single arrangement pusher seal shall be an inside-mounted balanced cartridge seal.

2.1.5 •

• •

The standard, un-pressurized dual mechanical seal shall be an inside, balanced, cartridge mounted mechanical seal (with two rotating flexible elements and two mating rings in series). Outer seals shall be designed to the same operating pressure as the inner seal, but do not have to be balanced. Cooling for the inboard seal is achieved by a seal flush. Cooling for the outside seal is accomplished by a circulating device moving a buffer fluid through an external seal flush system.

2.1.6 •

The standard pressurized dual mechanical seal shall be an inside, balanced, cartridge mounted mechanical seal (with two rotating flexible elements and two mating rings in series). The inner seal shall have an internal (reverse) balance feature designed and constructed to withstand reverse pressure differentials without opening.

2.1.7 •

The standard configuration for API single pusher and all dual mechanical seals is for the flexible elements to rotate. For seals having a seal face surface speed greater than 25 meters per second (5000 feet per minute), the standard alternative of stationary flexible elements shall be provided.

2.2.6 •

O-ring grooves shall be sized to accommodate perfluoroelastomer O-rings.

2.27 •

For vacuum services, all seal components shall be designed with a positive means of retaining the sealing components to prevent them from being dislodged.

2.3.3.1 •

2.3.5.1

Seal chambers shall conform to the minimum dimensions shown in Table 1 or Table 2. With these dimensions the minimum radial clearance between the rotating member of the seal and the stationary surfaces of the seal chamber and gland shall be 3 mm (1/8 inch).

•

For horizontally split pumps, slotted glands shall be provided to make disassembly easier.

2.3.5.2 •

Provisions shall be made for centering the seal gland and/or chamber with either an inside-or outside diameter register fit. The register fit surface shall be concentric to the shaft and shall have a total indicated run out of not more than 125 micrometers (0.005 inch). Shaft centering of mechanical seal components or the use of seal gland bolts is not acceptable.

2.3.10 •

Seal chamber pressure for single seals, and for the inner un-pressurized dual seal, shall be a minimum of 3.5 bar (50 psi.) or 10 percent above the maximum fluid vapor pressure at seal chamber fluid temperature. This margin shall be achieved by raising the seal chamber pressure and/or lowering the seal chamber temperature. Lowering the temperature is always preferable. Pumps which develop less than 3.5 bar (50 psi) differential pressure may not meet this requirement and alternate requirements shall be agreed upon by the purchaser and the seal manufacturer

2.3.18.1 •

On vertical pumps the seal chamber or gland plates shall have a port no less than 3 mm, (1/8") above the seal faces to allow the removal of trapped gas. The port must be orificed and valved.

2.3.20 •

For single seals and when specified for dual seals, a non-sparking, floatingthrottle bushing shall be installed in the seal gland or chamber and positively retained against blowout to minimize leakage if the seal fails.

•

Shaft sleeves shall be supplied by the seal manufacturer.

2.4

2.4.1 •

2.4.3

Unless otherwise specified a shaft sleeve of wear, corrosion, and erosion resistant material shall be provided to protect the shaft. The sleeve shall be sealed at one end. The shaft sleeve assembly shall extend beyond the outer face of the seal gland plate.

•

Shaft sleeves shall have a shoulder or shoulders for positively locating the rotating element or elements.

2.4.4.4 •

Shaft to sleeve sealing devices shall be elastomeric O-rings or flexible graphite rings.

2.4.5 •

Standard seal sizes shall be in even increments of ten millimeters. It is preferred that alternate seals be sized in increments of 0.635 mm (0,25 inches) starting with 38.0 mm (1.5 inches).

2.4.6 •

Sleeves shall have a minimum radial thickness of 2.5 mm (0.100 inches).

2.4.8 •

Sleeves shall be relieved along their bore leaving a locating fit at or near each end.

2.4.9 •

Shaft to sleeve diametral clearance shall be 25 micrometers to 75 micrometers (0.001 inch to 0.003 inch

2.4.10.2 •

Drive collar set screws shall be of sufficient hardness to securely embed in the shaft.

2.4.9 •

Shaft to sleeve diametrical clearance shall be 25 micrometers to 75 micrometers (0.001 inch to 0.003 inches)

2.5.1 •

Seal and mating rings shall be of one homogeneous material. Overlays and coatings shall not be used as the sole source of wear resistant material. Materials such as silicone or tungsten carbide may be enhanced by applying additional coating.

2.6.1 •

The type A standard pusher seal shall incorporate multiple springs with O-rings as the secondary sealing elements. When specified on the date sheet option, a single spring shall be furnished.

3.2.2 •

One of the seal face rings shall be premium grade, blister resistant carbon graphite with suitable binders and impregnates to reduce wear and provide chemical resistance. Several grades are available; therefore, the manufacturer shall state the type of carbon offered for each service.

3.2.3 •

The mating ring should be reaction bonded silicone carbide (RBSiC). When specified, self sintered silicone carbide (SSSiC) shall be furnished.

3.2.4 •

Abrasive service may require two hard materials. Unless otherwise specified for this service, the seal ring shall be reaction bonded silicone carbide and tungsten carbide (WC) with nickel binder

•

Unless otherwise specified, metal bellows for the type B seal shall be Hastelloy C. For the type C seal, Inconel 718.

3.6

3.7.2 •

Unless otherwise specified, gland plate to seal chamber seal shall be fluoroelastomer O-ring for services below 150°C (300°F). For temperatures over 150°C (300°F) or when specified, graphite-filled type 304 stainless steel spiral wound gaskets shall be used.

4.2.1 •

4.2.3

If you are using dual mechanical seals, only mechanically forced seal flush and barrier/buffer fluid systems shall be provided. Systems that rely upon a thermosyphon to maintain circulation during normal operation are not allowed.

•

Seal systems that utilize internal circulating devices, such as a pumping ring, that rely upon the rotation of the mechanical seal to maintain circulation shall be designed to thermo-syphon when the seal is not running.

4.5.4.1.1 •

If a dual seal buffer/barrier fluid reservoir is specified, a separate barrier/buffer fluid reservoir shall be furnished for each mechanical seal

Section 4.4.4 contains numerous references to dual seal system reservoirs. 4.5.5.1 •

The purchaser will specify on the date sheets the characteristics of the buffer/barrier fluid.

Section 4.6 addresses the circulation of the buffer/barrier fluid. There will be some benefits to the user when the API specification is adopted in to the CPI industry • • • • •

The decision to standardize on balanced seals is a wise one. It will reduce the seal inventory of most consumers and prevent a lot of premature seal failures. Allowing slotted glands for horizontally split pumps is a good idea. It should also extend to end suction centrifugal pumps. Requiring seal chamber vents on vertical pump installations makes sense. Banning coated or plated seal faces makes sense. Requiring the manufacturer to specify the carbon he is supplying is an excellent idea.

What is the problem with this API specification as a standard for the Chemical Process Industry? There are a lot of things I do not like about it in its present form. If combining with the CPI means a complete re-writing of the API specification that will be fine depending upon the final result. •

•

2.1.1 Some seal designs do not lend themselves to a cartridge design. Split seals as an example. You could mount a split seal on a split cartridge, but that would be "over kill" in most cases. 2.1.2 I do not like the definition of pusher seal in this standard. The term "pusher seal" is emotionally charged and misleading. It is used to describe a reliable Oring seal in the same category as spring loaded Teflon® wedges, or chevrons, and non-elastomer "U" cup designs. The implication is that the "non-pusher" metal bellows seal is a better choice. The fact is that O-ring seals are usually a better choice because of their ability to flex and roll and the O-ring provides a built in vibration damper that eliminates the need for letting a bellows metal face holder bounce off the shaft or sleeve.

•

• • • • •

•

•

•

• •

• •

•

2.1.5 The dual seal specification recognizes only tandem or series mounted rotating seals. It ignores concentric and "face to face" designs that make sense in some applications where space is not available for tandem configurations. Over the years the API has failed to recognize that there are four ways to install dual seals in a pump. They have played with the terminology over the years but have never got it simplified. It should be: Face to face Tandem or series Back to back Concentric, or one inside of the other. 2.1.6 The specification calls for the inner seal of a dual seal to be either balanced or reverse balanced depending upon whether high pressure barrier fluid or lower pressure buffer fluid is circulated between the dual seals. It totally ignores two way balance of the inner seal that would allow the consumer his choice between barrier or buffer fluid. 2.1.6 The specification call for the dual seals to be mounted in series (tandem), but almost all gas dual seals supplied to refineries to date have been supplied in the "back to back" configuration which is the worst possible installation method for slurry and abrasive service. 2.1.7 The specification approves rotating seals only and recommends stationary seals for speeds above 5000 fpm (25 m/sec). The fact is that stationary seals are almost always a better choice for leak free and the more severe fugitive emission sealing. 2.1.7 Stationary seals (the spring or springs do not rotate with the shaft) can be cartridge mounted if you take precautions to insure that the rotating face stays square to the shaft when the cartridge sleeve is set screwed or tightened to the shaft. It is not an easy problem to solve, but there are several solutions to the problem. Please see "stationary cartridge seals". 2.2.6 The specification calls for O-ring grooves with a larger groove dimension than normally used to accommodate perfluoroelastomer O-rings. 2.3.5.2 The specification assumes all pump manufacturers have provided a machined diameter concentric to the pump shaft so that the seal gland can be machined to register on an inside or outside diameter. The fact is that most pumps were manufactured for packing and do not have these concentric machined surfaces available to the seal manufacturer. In the CPI industry, shaft centering makes the most sense. 2.3.10 Maintaining a seal chamber 50 psi (3.5) bar above vapor pressure does not make any sense in the majority of balanced seal applications. 2.4.1 The specification calls for a shaft sleeve and allows the manufacturer to reduce the diameter of the solid shaft to accommodate the sleeve. This increasing of the pump shaft L3/D4 adversely affects the pump and seal performance. 2.4.1 The specification calls for sealing the sleeve on one end, but fails to specify the impeller end except in the case of O-ring seals. If the seal is on the outboard end, the space between the sleeve and shaft can fill with solids and hamper the removal of the sleeve. This can be a major concern in hot oil type applications where "coking" is always a problem.

•

•

•

•

• • • •

2.4.9 A shaft to sleeve diametral clearance of 0.001 inch to 0.003 inch is not practical. You will never be able to remove the sleeve once some solids get between the sleeve and shaft, and they will get there! 2.6.1 The standard seal is equipped with multiple springs, but the standard does not specify the springs must be located outside the fluid. If located in the fluid they can easily clog with solids. 3.2.3 Reaction bonded silicone carbide is specified as the standard hard face even though it is sensitive to caustic and other high pH chemicals frequently used to clean lines and systems. In most cases alpha sintered would be a much better choice. 4.2.1 The term "flush" is misleading. Over the years the API has failed to recognize the differences in bringing liquid to the pump stuffing box area and lumped them all under the common term "Flush". There is better terminology: o Discharge recirculation connects the discharge of the pump to the stuffing box to raise stuffing box pressure. o Suction recirculation connects the bottom of the stuffing box to the suction side of the pump usually allowing clean fluid to circulate from behind the impeller into the stuffing box. o Barrier fluid describes a higher-pressure fluid that is circulated between dual seals. o Buffer fluid describes a low-pressure fluid circulating between dual seals. o Quenching fluid is introduced into the seal gland outboard the seal to wash away leakage and control the environment outboard the seal. o Jacketing fluid circulates around the outside the stuffing box to control stuffing box temperature. o Flushing fluid is fluid from an outside source introduced into the stuffing box that dilutes the pumpage. It is seldom desirable, but sometimes necessary. The specification allows spring-loaded elastomers (O-rings) that do not have the ability to flex and roll. The specification allows a single spring seal design even if it is sensitive to the direction of rotation. The specification does not prohibit the use of mechanical seals that frett (damage) shafts and sleeves. The specification should call for the seal's dynamic O-ring to move towards a clean surface to prevent "hang up".

®

Pump Division

Types:

CPX, CPXR and CPXN

FRAME MOUNTED CHEMICAL PROCESS PUMPS

USER INSTRUCTIONS: INSTALLATION, OPERATION, MAINTENANCE C937KH013 - 09/03 (E) (incorporating C937KH054) These instructions should be read prior to installing, operating, using and maintaining this equipment.

USER INSTRUCTIONS CPX, CPXR and CPXN ENGLISH C937KH013 - 09/03 ®

CONTENTS

Page

Page

1 INTRODUCTION AND SAFETY ............................4 1.1 General............................................................4 1.2 CE marking and approvals ..............................4 1.3 Disclaimer........................................................4 1.4 Copyright .........................................................4 1.5 Duty conditions ................................................4 1.6 Safety ..............................................................5 1.7 Nameplate and warning labels ........................8 1.8 Specific machine performance ........................9 1.9 Noise level .......................................................9

6 MAINTENANCE................................................... 19 6.1 General ......................................................... 19 6.2 Maintenance schedule .................................. 20 6.3 Spare parts .................................................... 21 6.4 Recommended spares .................................. 21 6.5 Tools required................................................ 21 6.6 Fastener torques ........................................... 21 6.7 Setting impeller clearance............................. 22 6.8 Disassembly.................................................. 23 6.9 Examination of parts ..................................... 23 6.10 Assembly..................................................... 24 6.11 Sealing arrangements ................................. 26

2 TRANSPORT AND STORAGE ..............................9 2.1 Consignment receipt and unpacking ...............9 2.2 Handling ........................................................10 2.3 Lifting .............................................................10 2.4 Storage ..........................................................10 2.5 Recycling and end of product life ..................10 3 DESCRIPTION.....................................................10 3.1 Configurations ...............................................10 3.2 Name nomenclature ......................................10 3.3 Design of major parts .................................... 11 3.4 Performance and operating limits.................. 11 4 INSTALLATION .................................................... 11 4.1 Location ......................................................... 11 4.2 Part assemblies ............................................. 11 4.3 Foundation..................................................... 11 4.4 Grouting.........................................................12 4.5 Initial alignment..............................................12 4.6 Piping.............................................................13 4.7 Final shaft alignment check ...........................14 4.8 Electrical connections....................................14 4.9 Protection systems ........................................15

7 FAULTS; CAUSES AND REMEDIES .................. 28 8 PARTS LISTS AND DRAWINGS......................... 30 8.1 CPX and CPXN............................................. 30 8.2 CPXR ............................................................ 32 8.3 General arrangement drawing ...................... 33 9 Certification.......................................................... 33 10 OTHER RELEVANT DOCUMENTATION AND MANUALS.................................................. 33 10.1 Supplementary User Instruction manuals ... 33 10.2 Change notes.............................................. 33 10.3 Additional sources of information ................ 33

5 COMMISSIONING, START-UP, OPERATION AND SHUTDOWN ..............................................15 5.1 Pre-commissioning procedure.......................15 5.2 Pump lubricants.............................................16 5.3 Open impeller clearance ...............................17 5.4 Direction of rotation ........................................17 5.5 Guarding........................................................17 5.6 Priming and auxiliary supplies.......................17 5.7 Starting the pump ..........................................17 5.8 Running the pump .........................................18 5.9 Stopping and shutdown .................................19 5.10 Hydraulic, mechanical and electrical duty ...19

Page 2 of 36

USER INSTRUCTIONS CPX, CPXR and CPXN ENGLISH C937KH013 - 09/03 ®

INDEX PAGE

PAGE

Additional sources (10.3) .........................................33 Alignment of shafting (4.3, 4.5 and 4.7) Assembly (6.10) .......................................................24 ATEX marking (1.6.4.2)..............................................6 Bearing sizes and capacities (5.2.2) ........................16 CE marking and approvals (1.2) ................................4 Certification (9) .........................................................33 Change notes (10.2) ................................................33 Clearances, impeller (6.7) ........................................22 Commissioning and operation (5) ............................15 Compliance, ATEX (1.6.4.1) ......................................6 Configurations (3.1)..................................................10 Copyright (1.4) ...........................................................4 Design of major parts (3.3) ......................................11 Direction of rotation (5.4)..........................................17 Disassembly (6.8) ....................................................23 Disclaimer (1.3) ..........................................................4 Dismantling (6.8, Disassembly) ...............................23 Drawings (8) .............................................................30 Duty conditions (1.5) ..................................................4 Electrical connections (4.8) ......................................14 End of product life (2.5)............................................10 Examination of parts (6.9) ........................................23 Fastener torques (6.6)..............................................21 Faults; causes and remedies (7)..............................28 Foundation (4.3).......................................................11 General arrangement drawing (8.3) .........................33 General assembly drawings (8) ...............................30 Grouting (4.4) ...........................................................12 Guarding (5.5) ..........................................................17 Handling (2.2)...........................................................10 Hydraulic, mechanical and electrical duty (5.10) .....19 Impeller clearance (5.3 and 6.7) Inspection (6.2.1 and 6.2.2) .....................................20 Installation (4)...........................................................11 Lifting (2.3) ...............................................................10 Location (4.1) ...........................................................11 Lubrication (5.1.1, 5.2 and 6.2.3) Lubrication schedule (5.2.5).....................................17 Maintenance (6) .......................................................19 Maintenance schedule (6.2).....................................20 Name nomenclature (3.2) ........................................10 Nameplate (1.7.1) ......................................................8 Operating limits (3.4.1).............................................11 Ordering spare parts (6.3.1).....................................21 Part assemblies (4.2) ...............................................11 Parts lists (8) ............................................................30 Performance (3.4) ....................................................11 Piping (4.6) ...............................................................13 Pre-commissioning (5.1) ..........................................15 Priming and auxiliary supplies (5.6) .........................17 Protection systems (4.9) ..........................................15

Reassembly (6.10, Assembly) ................................. 24 Receipt and unpacking (2.1).................................... 10 Recommended fill quantities (see 5.2.2) ................. 16 Recommended grease lubricants (5.2.3) ................ 16 Recommended oil lubricants (5.2.1)........................ 16 Recommended spares (6.4) .................................... 21 Recycling (2.5)......................................................... 10 Replacement parts (6.3 and 6.4) ............................. 21 Running the pump (5.8) ........................................... 18 Safety action (1.6.3)................................................... 5 Safety markings (1.6.1) ............................................. 5 Safety, protection systems (1.6 and 4.9) Sealing arrangements (6.11) ................................... 26 Sectional drawings (8) ............................................. 30 Setting impeller clearance (6.7)............................... 22 Sound pressure level (1.9, Noise level)..................... 9 Sources, additional information (10.3)..................... 33 Spare parts (6.3)...................................................... 21 Specific machine performance (1.8).......................... 9 Starting the pump (5.7) ............................................ 17 Stop/start frequency (5.8.5) ..................................... 18 Stopping and shutdown (5.9)................................... 19 Storage, pump (2.4)................................................. 10 Storage, spare parts (6.3.2)..................................... 21 Supplementary manuals or information sources ..... 33 Supplementary User Instructions (10.1) .................. 33 Thermal expansion (4.5.1)....................................... 12 Tools required (6.5) ................................................. 21 Torques for fasteners (6.6) ...................................... 21 Trouble-shooting (see 7) ......................................... 28 Vibration (5.8.4) ....................................................... 18 Warning labels (1.7.2)................................................ 8

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USER INSTRUCTIONS CPX, CPXR and CPXN ENGLISH C937KH013 - 09/03 ®

1 INTRODUCTION AND SAFETY

1.3 Disclaimer

1.1 General

Information in these User Instructions is believed to be reliable. In spite of all the efforts of Flowserve Corporation to provide sound and all necessary information the content of this manual may appear insufficient and is not guaranteed by Flowserve as to its completeness or accuracy.

These instructions must always be kept close to the product’s operating location or directly with the product. Flowserve’s products are designed, developed and manufactured with state-of-the-art technologies in modern facilities. The unit is produced with great care and commitment to continuous quality control, utilising sophisticated quality techniques, and safety requirements. Flowserve is committed to continuous quality improvement and being at service for any further information about the product in its installation and operation or about its support products, repair and diagnostic services. These instructions are intended to facilitate familiarization with the product and its permitted use. Operating the product in compliance with these instructions is important to help ensure reliability in service and avoid risks. The instructions may not take into account local regulations; ensure such regulations are observed by all, including those installing the product. Always co-ordinate repair activity with operations personnel, and follow all plant safety requirements and applicable safety and health laws/regulations. These instructions should be read prior to installing, operating, using and maintaining the equipment in any region worldwide. The equipment must not be put into service until all the conditions relating to safety noted in the instructions, have been met.

1.2 CE marking and approvals

It is a legal requirement that machinery and equipment put into service within certain regions of the world shall conform with the applicable CE Marking Directives covering Machinery and, where applicable, Low Voltage Equipment, Electromagnetic Compatibility (EMC), Pressure Equipment Directive (PED) and Equipment for Potentially Explosive Atmospheres (ATEX). Where applicable, the Directives and any additional Approvals, cover important safety aspects relating to machinery and equipment and the satisfactory provision of technical documents and safety instructions. Where applicable this document incorporates information relevant to these Directives and Approvals. To confirm the Approvals applying and if the product is CE marked, check the serial number plate markings and the Certification. (See section 9, Certification.)

Flowserve manufactures products to exacting International Quality Management System Standards as certified and audited by external Quality Assurance organisations. Genuine parts and accessories have been designed, tested and incorporated into the products to help ensure their continued product quality and performance in use. As Flowserve cannot test parts and accessories sourced from other vendors the incorrect incorporation of such parts and accessories may adversely affect the performance and safety features of the products. The failure to properly select, install or use authorised Flowserve parts and accessories is considered to be misuse. Damage or failure caused by misuse is not covered by Flowserve’s warranty. In addition, any modification of Flowserve products or removal of original components may impair the safety of these products in their use.

1.4 Copyright

All rights reserved. No part of these instructions may be reproduced, stored in a retrieval system or transmitted in any form or by any means without prior permission of Flowserve Pump Division.

1.5 Duty conditions

This product has been selected to meet the specifications of your purchaser order. The acknowledgement of these conditions has been sent separately to the Purchaser. A copy should be kept with these instructions. The product must not be operated beyond the parameters specified for the application. If there is any doubt as to the suitability of the product for the application intended, contact Flowserve for advice, quoting the serial number. If the conditions of service on your purchase order are going to be changed (for example liquid pumped, temperature or duty) it is requested that the user seeks Flowserve’s written agreement before start up.

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USER INSTRUCTIONS CPX, CPXR and CPXN ENGLISH C937KH013 - 09/03 ®

1.6 Safety 1.6.1 Summary of safety markings These User Instructions contain specific safety markings where non-observance of an instruction would cause hazards. The specific safety markings are: This symbol indicates electrical safety instructions where non-compliance will involve a high risk to personal safety or the loss of life. This symbol indicates safety instructions where non-compliance would affect personal safety and could result in loss of life. This symbol indicates “hazardous and toxic fluid” safety instructions where non-compliance would affect personal safety and could result in loss of life. This symbol indicates safety instructions where non-compliance will involve some risk to safe operation and personal safety and would damage the equipment or property. This symbol indicates explosive atmosphere zone marking according to ATEX. It is used in safety instructions where non-compliance in the hazardous area would cause the risk of an explosion. This sign is not a safety symbol but indicates an important instruction in the assembly process. 1.6.2 Personnel qualification and training All personnel involved in the operation, installation, inspection and maintenance of the unit must be qualified to carry out the work involved. If the personnel in question do not already possess the necessary knowledge and skill, appropriate training and instruction must be provided. If required the operator may commission the manufacturer/supplier to provide applicable training. Always coordinate repair activity with operations and health and safety personnel, and follow all plant safety requirements and applicable safety and health laws and regulations. 1.6.3 Safety action This is a summary of conditions and actions to help prevent injury to personnel and damage to the environment and to equipment. For products used in potentially explosive atmospheres section 1.6.4 also applies. NEVER DO MAINTENANCE WORK WHEN THE UNIT IS CONNECTED TO POWER GUARDS MUST NOT BE REMOVED WHILE THE PUMP IS OPERATIONAL

DRAIN THE PUMP AND ISOLATE PIPEWORK BEFORE DISMANTLING THE PUMP The appropriate safety precautions should be taken where the pumped liquids are hazardous. FLUORO-ELASTOMERS (When fitted.) When a pump has experienced temperatures over 250 ºC (482 ºF), partial decomposition of fluoro-elastomers (example: Viton) will occur. In this condition these are extremely dangerous and skin contact must be avoided. HANDLING COMPONENTS Many precision parts have sharp corners and the wearing of appropriate safety gloves and equipment is required when handling these components. To lift heavy pieces above 25 kg (55 lb) use a crane appropriate for the mass and in accordance with current local regulations. THERMAL SHOCK Rapid changes in the temperature of the liquid within the pump can cause thermal shock, which can result in damage or breakage of components and should be avoided. NEVER APPLY HEAT TO REMOVE IMPELLER Trapped lubricant or vapour could cause an explosion. HOT (and cold) PARTS If hot or freezing components or auxiliary heating supplies can present a danger to operators and persons entering the immediate area action must be taken to avoid accidental contact. If complete protection is not possible, the machine access must be limited to maintenance staff only, with clear visual warnings and indicators to those entering the immediate area. Note: bearing housings must not be insulated and drive motors and bearings may be hot. If the temperature is greater than 68 ºC (175 ºF) or below 5 ºC (20 ºF) in a restricted zone, or exceeds local regulations, action as above shall be taken. HAZARDOUS LIQUIDS When the pump is handling hazardous liquids care must be taken to avoid exposure to the liquid by appropriate siting of the pump, limiting personnel access and by operator training. If the liquid is flammable and or explosive, strict safety procedures must be applied. Gland packing must not be used when pumping hazardous liquids. PREVENT EXCESSIVE EXTERNAL PIPE LOAD Do not use pump as a support for piping. Do not mount expansion joints, unless allowed by Flowserve in writing, so that their force, due to internal pressure, acts on the pump flange.

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USER INSTRUCTIONS CPX, CPXR and CPXN ENGLISH C937KH013 - 09/03 ®

ENSURE CORRECT LUBRICATION (See section 5, Commissioning, startup, operation and shutdown.) START THE PUMP WITH OUTLET VALVE PART OPENED (Unless otherwise instructed at a specific point in the User Instructions.) This is recommended to minimize the risk of overloading and damaging the pump or motor at full or zero flow. Pumps may be started with the valve further open only on installations where this situation cannot occur. The pump outlet control valve may need to be adjusted to comply with the duty following the run-up process. (See section 5, Commissioning start-up, operation and shutdown.) NEVER RUN THE PUMP DRY INLET VALVES TO BE FULLY OPEN WHEN PUMP IS RUNNING Running the pump at zero flow or below the recommended minimum flow continuously will cause damage to the pump and mechanical seal. DO NOT RUN THE PUMP AT ABNORMALLY HIGH OR LOW FLOW RATES Operating at a flow rate higher than normal or at a flow rate with no back pressure on the pump may overload the motor and cause cavitation. Low flow rates may cause a reduction in pump/bearing life, overheating of the pump, instability and cavitation/ vibration. 1.6.4 Products used in potentially explosive atmospheres • • • • •

Where Flowserve has supplied only the bare shaft pump, the Ex rating applies only to the pump. The party responsible for assembling the pump set shall select the coupling, driver and any additional equipment, with the necessary CE Certificate/ Declaration of Conformity establishing it is suitable for the area in which it is to be installed. The output from a variable frequency drive (VFD) can cause additional heating effects in the motor and so, for pumps sets with a VFD, the ATEX Certification for the motor must state that it is covers the situation where electrical supply is from the VFD. This particular requirement still applies even if the VFD is in a safe area. 1.6.4.2 Marking An example of ATEX equipment marking is shown below. The actual classification of the pump will be engraved on the nameplate. II 2 GD c 135 ºC (T4) Equipment Group I = Mining II = Non-mining Category 2 or M2 = High level protection 3 = normal level of protection Gas and/or Dust G = Gas; D= Dust c = Constructional safety (in accordance with prEn13463-5) Maximum surface temperature (Temperature Class) (see section 1.6.4.3.)

Measures are required to: Avoid excess temperature Prevent build up of explosive mixtures Prevent the generation of sparks Prevent leakages Maintain the pump to avoid hazard

1.6.4.3 Avoiding excessive surface temperatures ENSURE THE EQUIPMENT TEMPERATURE CLASS IS SUITABLE FOR THE HAZARD ZONE

The following instructions for pumps and pump units when installed in potentially explosive atmospheres must be followed to help ensure explosion protection. Both electrical and non-electrical equipment must meet the requirements of European Directive 94/9/EC. 1.6.4.1 Scope of compliance Use equipment only in the zone for which it is appropriate. Always check that the driver, drive coupling assembly, seal and pump equipment are suitably rated and/or certified for the classification of the specific atmosphere in which they are to be installed.

Pumps have a temperature class as stated in the ATEX Ex rating on the nameplate. These are based on a maximum ambient of 40 ºC (104 ºF); refer to Flowserve for higher ambient temperatures. The surface temperature on the pump is influenced by the temperature of the liquid handled. The maximum permissible liquid temperature depends on the temperature class and must not exceed the values in the table that follows. The temperature rise at the seals and bearings and due to the minimum permitted flow rate is taken into account in the temperatures stated.

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USER INSTRUCTIONS CPX, CPXR and CPXN ENGLISH C937KH013 - 09/03 ®

Temperature class to prEN 13463-1 T6 T5 T4 T3 T2 T1

Maximum surface temperature permitted 85 °C (185 °F) 100 °C (212 °F) 135 °C (275 °F) 200 °C (392 °F) 300 °C (572 °F) 450 °C (842 °F)

Temperature limit of liquid handled (* depending on material and construction variant - check which is lower) Consult Flowserve Consult Flowserve 115 °C (239 °F) * 180 °C (356 °F) * 275 °C (527 °F) * 400 °C (752 °F) *

The responsibility for compliance with the specified maximum liquid temperature is with the plant operator. Do not attempt to check the direction of rotation with the coupling element/pins fitted due to the risk of severe contact between rotating and stationary components. Where there is any risk of the pump being run against a closed valve generating high liquid and casing external surface temperatures it is recommended that users fit an external surface temperature protection device. Avoid mechanical, hydraulic or electrical overload by using motor overload trips, temperature monitor or a power monitor and make routine vibration monitoring checks. In dirty or dusty environments, regular checks must be made and dirt removed from areas around close clearances, bearing housings and motors. 1.6.4.4 Preventing the build up of explosive mixtures ENSURE THE PUMP IS PROPERLY FILLED AND VENTED AND DOES NOT RUN DRY Ensure the pump and relevant suction and discharge pipeline system is totally filled with liquid at all times during the pump operation, so that an explosive atmosphere is prevented. In addition it is essential to make sure that seal chambers, auxiliary shaft seal systems and any heating and cooling systems are properly filled. If the operation of the system cannot avoid this condition the fitting of an appropriate dry run protection device is recommended (for example liquid detection or a power monitor). To avoid potential hazards from fugitive emissions of vapour or gas to atmosphere the surrounding area must be well ventilated.

1.6.4.5 Preventing sparks To prevent a potential hazard from mechanical contact, the coupling guard must be non-sparking and anti-static for Category 2. To avoid the potential hazard from random induced current generating a spark, the earth contact on the baseplate must be used. Avoid electrostatic charge: do not rub non-metallic surfaces with a dry cloth; ensure cloth is damp. The coupling must be selected to comply with 94/9/EC and correct alignment must be maintained. 1.6.4.6 Preventing leakage The pump must only be used to handle liquids for which it has been approved to have the correct corrosion resistance. Avoid entrapment of liquid in the pump and associated piping due to closing of suction and discharge valves, which could cause dangerous excessive pressures to occur if there is heat input to the liquid. This can occur if the pump is stationary or running. Bursting of liquid containing parts due to freezing must be avoided by draining or protecting the pump and ancillary systems. Where there is the potential hazard of a loss of a seal barrier fluid or external flush, the fluid must be monitored. If leakage of liquid to atmosphere can result in a hazard, the installation of a liquid detection device is recommended. 1.6.4.7 Maintenance to avoid the hazard CORRECT MAINTENANCE IS REQUIRED TO AVOID POTENTIAL HAZARDS WHICH GIVE A RISK OF EXPLOSION The responsibility for compliance with maintenance instructions is with the plant operator. To avoid potential explosion hazards during maintenance, the tools, cleaning and painting materials used must not give rise to sparking or adversely affect the ambient conditions. Where there is a risk from such tools or materials, maintenance must be conducted in a safe area. It is recommended that a maintenance plan and schedule is adopted. (See section 6, Maintenance.)

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USER INSTRUCTIONS CPX, CPXR and CPXN ENGLISH C937KH013 - 09/03 ®

1.7 Nameplate and warning labels 1.7.1 Nameplate For details of nameplate, see the Declaration of Conformity. 1.7.2 Warning labels

Oil lubricated units only:

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USER INSTRUCTIONS CPX, CPXR and CPXN ENGLISH C937KH013 - 09/03 ®

1.8 Specific machine performance

For performance parameters see section 1.5, Duty conditions. Where performance data has been supplied separately to the purchaser these should be obtained and retained with these User Instructions if required.

1.9 Noise level

When pump noise level exceeds 85 dB(A) attention must be given to prevailing Health and Safety Legislation, to limit the exposure of plant operating personnel to the noise. The usual approach is to control exposure time to the noise or to enclose the machine to reduce emitted sound. You may have already specified a limiting noise level when the equipment was ordered, however if no noise requirements were defined then machines above a certain power level will exceed 85 dB(A). In such situations consideration must be given to the fitting of an acoustic enclosure to meet local regulations. Pump noise level is dependent on a number of factors - the type of motor fitted, the operating capacity, pipework design and acoustic characteristics of the building. Typical sound pressure levels measured in dB, and A-weighted are shown in the table below (LpfA). The figures are indicative only, they are subject to a +3dB tolerance, and cannot be guaranteed. The values are based on the noisiest ungeared electric motors that are likely to be encountered. They represent sound pressure levels at 1 m (3.3 ft) from the directly driven pump, for "free field over a reflecting plane".

If a pump unit only has been purchased, for fitting with your own driver, then the "pump only" noise levels from the table should be combined with the level for the driver obtained from the supplier. If the motor is driven by an inverter it may show an increase in noise level at some speeds. Consult a Noise Specialist for the combined calculation.



For units driven by equipment other than electric motors or units contained within enclosures, see the accompanying information sheets and manuals.

2 TRANSPORT AND STORAGE 2.1 Consignment receipt and unpacking

Immediately after receipt of the equipment it must be checked against the delivery/shipping documents for its completeness and that there has been no damage in transportation. Any shortage and/or damage must be reported immediately to Flowserve Pump Division and must be received in writing within one month of receipt of the equipment. Later claims cannot be accepted. Check any crate, boxes or wrappings for any accessories or spare parts that may be packed separately with the equipment or attached to side walls of the box or equipment. Each product has a unique serial number. Check that this number corresponds with that advised and always quote this number in correspondence as well as when ordering spare parts or further accessories.

Typical sound pressure level, LpfA – (dB, A-weighted) Motor size and speed kW <0.55 0.75 1.1 1.5 2.2 3 4 5.5 7.5 11 15 18.5 22 30 37 45 55 75 90 110 150

(hp) (<0.75) (1) (1.5) (2) (3) (4) (5) (7.5) (10) (15) (20) (25) (30) (40) (50) (60) (75) (100) (120) (150) (200)

3550 r/min Pump and Pump motor only LpfA LpfA 71 66 74 66 74 68 77 70 78 72 81 74 82 75 90 77 90 78 91 80 92 83 92 83 92 83 100 85 100 86 100 87 102 88 100 90 97 90 100 91 101 92

2900 r/min Pump and Pump motor only LpfA LpfA 64 62 67 62 67 64 70 66 71 68 74 70 75 71 83 73 83 74 84 76 85 79 85 79 85 79 93 81 93 82 93 83 95 84 95 86 92 86 95 87 96 88

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1750 r/min Pump and Pump motor only LpfA LpfA 64 62 67 62 67 64 70 66 71 68 74 74 75 75 76 75 77 76 78 77 80 79 80 79 81 79 84 80 84 80 84 80 86 81 88 81 90 81 91 83 91 83

1450 r/min Pump and Pump motor only LpfA LpfA 63 62 63 62 65 64 66 66 68 68 70 70 71 71 72 71 73 72 74 73 76 75 76 75 77 75 80 76 80 76 80 76 82 77 83 78 85 78 86 79 86 79

USER INSTRUCTIONS CPX, CPXR and CPXN ENGLISH C937KH013 - 09/03 ®

2.2 Handling

Boxes, crates, pallets or cartons may be unloaded using fork lift vehicles or slings dependent on their size and construction.

2.3 Lifting The pump and cast iron baseplate set should be lifted as shown.

2.4 Storage Store the pump in a clean, dry location away from vibration. Leave piping connection covers in place to keep dirt and other foreign material out of pump casing. Turn pump at intervals to prevent brinelling of the bearings and the seal faces, if fitted, from sticking. The pump may be stored as above for up to 6 months. Consult Flowserve for preservative actions when a longer storage period is needed.

2.5 Recycling and end of product life

At the end of the service life of the product or its parts, the relevant materials and parts should be recycled or disposed of using an environmentally acceptable method and in accordance with local requirements. If the product contains substances that are harmful to the environment, these should be removed and disposed of in accordance with current regulations. This also includes the liquids and/or gases that may be used in the "seal system" or other utilities. Make sure that hazardous substances are disposed of safely and that the correct personal protective equipment is used. The safety specifications must be in accordance with the current regulations at all times.

3 DESCRIPTION 3.1 Configurations

The pump is a modular designed centrifugal pump that can be built to achieve almost all chemical liquid pumping requirements. (See 3.2 and 3.3 below.)

3.2 Name nomenclature

The pump size will be engraved on the nameplate typically as below:

80-50CPX200

Nominal suction size in mm Where the baseplate is folded steel there are no specific lifting points provided for this complete machine set (unless so identified). Any lifting points that can be seen are provided only for dismantling parts for servicing. Slings, ropes and other lifting gear should be positioned where they cannot slip and where a balanced lift is obtained. A crane must be used for all pump sets in excess of 25 kg (55 lb). Fully trained personnel must carry out lifting, in accordance with local regulations. The driver and pump weights are recorded on their respective nameplates or massplates.

Nominal discharge size in mm Configuration – see 3.3.1 and 3.3.2 below Nominal ISO maximum impeller diameter The typical nomenclature above is the general guide to the CPX configuration description. Identify the actual pump size and serial number from the pump nameplate. Check that this agrees with the applicable certification provided.

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USER INSTRUCTIONS CPX, CPXR and CPXN ENGLISH C937KH013 - 09/03 ®

3.3 Design of major parts 3.3.1 Pump casing The pump casing is designed with a horizontal centreline end inlet and a vertical centreline top outlet which makes it self venting. For ease of maintenance, the pump is constructed so that pipe connectors do not have to be disturbed when internal maintenance is required. On the CPX and CPXR the casing feet pads are underneath the casing. On the CPXN they are on the shaft centreline.

3.3.9 Accessories Accessories may be fitted when specified by the customer. Fan cooling is available for high temperature operation. (This is a fan fitted within the coupling guard to blow cooling air over the bearing housing and shaft.)

3.4 Performance and operating limits

This product has been selected to meet the specifications of the purchase order. See section 1.5.

3.3.2 Impeller An open impeller is fitted. (On the CPXR the impeller is recessed into the back of the casing.)

The following data is included as additional information to help with your installation. It is typical, and factors such as temperature, materials, and seal type may influence this data. If required, a definitive statement for your particular application can be obtained from Flowserve.

3.3.3 Shaft The large diameter stiff shaft, mounted on bearings, has a keyed drive end.

3.4.1 Operating limits Maximum ambient temperature: +40 ºC (104°F) Maximum pump speed: refer to the nameplate.

3.3.4 Bearing housing The bearing housing enables adjustment of impeller face clearance via the bearing carrier jacking screws.

4 INSTALLATION

3.3.5 Pump bearings and lubrication The pump is fitted with ball and/or roller type bearings which may be configured differently dependent on use. The bearings may be oil or grease lubricated.

Equipment operated in hazardous locations must comply with the relevant explosion protection regulations. See section 1.6.4, Products used in potentially explosive atmospheres.

4.1 Location

3.3.6 Seal housing The seal housing has spigots between the pump casing and bearing housing for optimum concentricity.

The pump should be located to allow room for access, ventilation, maintenance and inspection with ample headroom for lifting and should be as close as practicable to the supply of liquid to be pumped. Refer to the general arrangement drawing for the pump set.

A fully confined gasket forms the seal between the pump casing and the seal housing.

4.2 Part assemblies

The seal housings designs provide improved performance of mechanical seals.

On baseplated pump sets the coupling elements are supplied loose. It is the responsibility of the installer to ensure that the pump set is finally alined up as detailed in section 4.5.2, Alignment methods.

The design enables one of a number of sealing options to be fitted.

4.3 Foundation

3.3.7 Shaft seal The mechanical seal(s) attached to the drive shaft seals the pumped liquid from the environment. Gland packing may be fitted as an option. 3.3.8 Driver The driver is normally an electric motor. Different drive configurations may be fitted such as internal combustion engines, turbines, hydraulic motors etc driving via couplings, belts, gearboxes, drive shafts etc.

There are many methods of installing pump units to their foundations. The correct method depends on the size of the pump unit, its location and noise and vibration limitations. Non-compliance with the provision of correct foundation and installation may lead to failure of the pump and, as such, would be outside the terms of the warranty. Ensure the following are met: a) The baseplate should be mounted onto a firm foundation, either an appropriate thickness of quality concrete or sturdy steel framework. (It should NOT be distorted or pulled down onto the surface of the foundation, but should be supported to maintain the original alignment.)

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USER INSTRUCTIONS CPX, CPXR and CPXN ENGLISH C937KH013 - 09/03 ®

b) Install the baseplate onto packing pieces evenly spaced and adjacent to foundation bolts.

c) Level with shims between baseplate and packing pieces. d) The pump and driver have been aligned before dispatch however the alignment of pump and motor half coupling must be checked. If this is incorrect, it indicates that the baseplate has become twisted and should be corrected by re-shimming. e) If not supplied, guarding shall be fitted as necessary to meet the requirements of EN292 and EN953.

Motor and pump centre line height adjustment:

Graph based on the assumptions that: 1 Operating temperature rise of the motor frame is 50 °C (122 °F). 2 Packing piece/motor stool is not affected. Operation 1 Enter graph at base to shaft centre line height. 2 Read line for frame material. 3 Set motor shaft and coupling LOW by figure on left-hand side.

4.4 Grouting

Where applicable, grout in the foundation bolts. After adding pipework connections and rechecking the coupling alignment, the baseplate should then be grouted in accordance with good engineering practice. Fabricated steel, cast iron and epoxy baseplates can be filled with grout. Folded steel baseplates should be grouted to locate their packing pieces. If in any doubt, please contact your nearest service centre for advice. Grouting provides solid contact between the pump unit and foundation, prevents lateral movement of vibrating equipment and dampens resonant vibrations. Foundation bolts should only be fully tightened when the grout has cured.

4.5 Initial alignment 4.5.1 Thermal expansion The pump and motor will normally have to be aligned at ambient temperature with an allowance for thermal expansion at operating temperature. (See chart.) In pump installations involving high liquid temperatures, the unit should be run at the actual operating temperature, shut down and the alignment checked immediately.

4.5.2 Alignment methods Pump and driver must be isolated electrically and the half couplings disconnected. The alignment MUST be checked. Although the pump will have been aligned at the factory it is most likely that this alignment will have been disturbed during transportation or handling. If necessary, align the motor to the pump, not the pump to the motor. Alignment is achieved by adding or removing shims under the motor feet and also moving the motor horizontally as required. In some cases where the alignment cannot be achieved it will be necessary to move the pump before recommencing the above procedure. For couplings with narrow flanges use a dial indicator as shown. The alignment values are maximums for continuous service. Parallel

Angular

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USER INSTRUCTIONS CPX, CPXR and CPXN ENGLISH C937KH013 - 09/03 ®

Permissible misalignment limits at working temperature: • Parallel alignment - 0.25 mm (0.010 in.) TIR maximum • Angular alignment - 0.3 mm (0.012 in.) TIR maximum for couplings not exceeding 100 mm (4 in.) flange diameter - 0.5 mm (0.020 in.) TIR maximum for couplings over 100 mm (4 in.) diameter When checking parallel alignment, the total indicator read-out (TIR) shown is twice the value of the actual shaft displacement. Align in the vertical plane first, then horizontally by moving motor. Maximum pump reliability is obtained by near perfect alignment of 0.05 - 0.075 mm (0.002 0.003 in.) parallel and 0.05 mm (0.002 in.) per 100 mm (4 in.) of coupling flange diameter as angular misalignment. When performing final alignment, check for soft-foot under the driver. An indicator placed on the coupling, reading in the vertical direction, should not indicate more than 0.05 mm (0.002 in.) movement when any driver foot fastener is loosened. Complete piping as below and see sections 4.7, Final shaft alignment check up to and including section 5, Commissioning, startup, operation and shutdown, before connecting driver and checking actual rotation.

4.6 Piping Protective covers are fitted to the pipe connections to prevent foreign bodies entering during transportation and installation. Ensure that these covers are removed from the pump before connecting any pipes. 4.6.1 Suction and discharge pipework In order to minimize friction losses and hydraulic noise in the pipework it is good practice to choose pipework that is one or two sizes larger than the pump suction and discharge. Typically main pipework velocities should not exceed 2 m/s (6 ft/sec) suction and 3 m/s (9 ft/sec) on the discharge. Take into account the available NPSH which must be higher than the required NPSH of the pump.

piping.

Never use the pump as a support for

Maximum forces and moments allowed on the pump flanges vary with the pump size and type. To minimize these forces and moments that may, if excessive, cause misalignment, hot bearings, worn couplings, vibration and the possible failure of the pump casing, the following points should be strictly followed:

• • •

Prevent excessive external pipe load Never draw piping into place by applying force to pump flange connections Do not mount expansion joints so that their force, due to internal pressure, acts on the pump flange

before use.

Ensure piping and fittings are flushed

Ensure piping for hazardous liquids is arranged to allow pump flushing before removal of the pump. 4.6.2 Suction piping a) The inlet pipe should be one or two sizes larger than the pump inlet bore and pipe bends should be as large a radius as possible. b) On suction lift the piping should be inclined up towards the pump inlet with eccentric reducers incorporated to prevent air locks. c) On positive suction, the inlet piping must have a constant fall towards the pump. d) The pipe next to the pump should be the same diameter as the pump suction and have a minimum of two pipe diameters of straight section between the elbow and the pump inlet flange. Where the NPSH margin is not large, it is recommended that the pipe straight is 5 to 10 pipe diameter. (See section 10.3, Reference 1.) Inlet strainers when used, should have a net ’free area’ of at least three times the inlet pipe area. e) Fitting isolation and non-return valves will allow easier maintenance. f) Never throttle pump on suction side and never place a valve directly on the pump inlet nozzle. 4.6.3 Discharge piping A non-return valve should be located in the discharge pipework to protect the pump from excessive back pressure and hence reverse rotation when the unit is stopped. Fitting an isolation valve will allow easier maintenance. 4.6.4 Auxiliary piping The connections that are to be piped up will have been fitted with protective metal or plastic plugs which will need to be removed. 4.6.4.1 Pumps fitted with packed glands When suction pressure is below ambient pressure and differential head is less than 10 m (32.8 ft), it may be necessary to feed gland packing with liquid to provide lubrication and prevent the ingress of air. 4.6.4.2 Pumps fitted with mechanical seals The conical design of the single internal seal housing provides excellent liquid circulation around the seal and will not normally require a separate flush.

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USER INSTRUCTIONS CPX, CPXR and CPXN ENGLISH C937KH013 - 09/03 ®

Single seals requiring re-circulation will normally be provided with the auxiliary piping from pump casing already fitted.

For pumping hot liquids, to avoid seal damage, it is recommended that any external flush/cooling supply be continued after stopping the pump.

Flowserve seal connections are designated as follows: Q - quench F - flush D - drain outlet BI - barrier fluid in (double seals) BO - barrier fluid out (double seals) H - heating jacket C - cooling jacket

Tandem seals require a barrier liquid between the seals compatible with the pumped liquid. 4.6.4.3 Pumps fitted with heating/cooling jackets Connect the heating/cooling pipes from the site supply. The top connection should be used as the outlet to ensure complete filling/venting of the annulus with heating/cooling liquids; steam is usually in at the top, out at the bottom.

Seal housings/covers having an auxiliary quench connection, require connection to a suitable source of liquid flow, low pressure steam or static pressure from a header tank. Recommended pressure is 0.35 bar (5 psi) or less. Check General arrangement drawing.

4.6.5 Final checks Check the tightness of all bolts in the suction and discharge pipework. Check also the tightness of all foundation bolts.

Double seals require a barrier liquid between the seals, compatible with the pumped liquid.

After connecting piping to the pump, rotate the shaft several times by hand to ensure there is no binding and all parts are free.

With back-to-back double seals, the barrier liquid should be at a minimum pressure of 1 bar above the maximum pressure on the pump side of the inner seal. (See chart.) The barrier liquid pressure must not exceed limitations of the seal on the atmospheric side. For toxic service the barrier liquid supply and discharge must be in a safe area.

4.7 Final shaft alignment check

Recheck the coupling alignment, as previously described, to ensure no pipe strain. If pipe strain exists, correct piping.

4.8 Electrical connections Electrical connections must be made by a qualified Electrician in accordance with relevant local national and international regulations. It is important to be aware of the EUROPEAN DIRECTIVE on potentially explosive areas where compliance with IEC60079-14 is an additional requirement for making electrical connections.

Notes: a) Total seal pressure is equal to pressure at seal plus suction pressure. b) For pumped liquid viscosities greater than 440 Centistokes multiply the generated pressure by 1.25 for 125, 160 and 200 size pumps and by 2.0 for larger sizes. c) Differential pressure in bar equals head in metres multiplied by specific gravity all divided by 10.19. d) Ensure to check the seal minimum and maximum seal pressure limits are not exceeded and the pressure is agreed with Flowserve Pump Division.

Special seals may require modification to auxiliary piping described above. Consult Flowserve if unsure of correct method or arrangement.

It is important to be aware of the EUROPEAN DIRECTIVE on electromagnetic compatibility when wiring up and installing equipment on site. Attention must be paid to ensure that the techniques used during wiring/installation do not increase electromagnetic emissions or decrease the electromagnetic immunity of the equipment, wiring or any connected devices. If in any doubt contact Flowserve for advice. The motor must be wired up in accordance with the motor manufacturer’s instructions (normally supplied within the terminal box) including any temperature, earth leakage, current and other protective devices as appropriate. The identification nameplate should be checked to ensure the power supply is appropriate. A device to provide emergency stopping must be fitted.

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USER INSTRUCTIONS CPX, CPXR and CPXN ENGLISH C937KH013 - 09/03 ®

If not supplied pre-wired to the pump unit, the controller/starter electrical details will also be supplied within the controller/starter. For electrical details on pump sets with controllers see the separate wiring diagram. See section 5.4, Direction of rotation before connecting the motor to the electrical supply.

4.9 Protection systems The following protection systems are recommended particularly if the pump is installed in a potentially explosive area or is handling a hazardous liquid. If in any doubt consult Flowserve.

When fitted with a constant level oiler, the bearing housing should be filled by unscrewing or hinging back the transparent bottle and filling the bottle with oil. Where an adjustable body Denco oiler is fitted this should be set to the height shown in the following diagram:

If there is any possibility of the system allowing the pump to run against a closed valve or below minimum continuous safe flow a protection device should be installed to ensure the temperature of the liquid does not rise to an unsafe level. If there are any circumstances in which the system can allow the pump to run dry, or start up empty, a power monitor should be fitted to stop the pump or prevent it from being started. This is particularly relevant if the pump is handling a flammable liquid. If leakage of product from the pump or its associated sealing system can cause a hazard it is recommended that an appropriate leakage detection system is installed. To prevent excessive surface temperatures at bearings it is recommended that temperature or vibration monitoring are carried out.

The oil filled bottle should then be refitted so as to return it to the upright position. Filling should be repeated until oil remains visible within the bottle. Approximate oil volumes are shown in section 5.2.2, Bearing sizes and capacities. Grease lubricated pumps and electric motors are supplied pre-greased. Other drivers and gearboxes, if appropriate, should be lubricated in accordance with their manuals.

5 COMMISSIONING, START-UP, OPERATION AND SHUTDOWN These operations must be carried out by fully qualified personnel.

5.1 Pre-commissioning procedure 5.1.1 Lubrication Determine the mode of lubrication of the pump set, eg grease, oil, product lubrication etc. For oil lubricated pumps, fill the bearing housing with correct grade of oil to the correct level, ie sight glass or constant level oiler bottle.

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USER INSTRUCTIONS CPX, CPXR and CPXN ENGLISH C937KH013 - 09/03 ®

5.2 Pump lubricants

Centifugal pump lubrication

5.2.1 Recommended oil lubricants Oil

Splash lubrication

Viscosity mm²/s 40ºC Temp. max. ºC (ºF)

32

68

46

65 (149)

80 (176)

-

HL/HLP 32

HL/HLP 68

HL/HLP 46

BP Energol HL32 BP Energol HLP32 Anstron HL32 Anstron HLP32 OLNA 32 HYDRELEF 32 TURBELF 32 ELFOLNA DS32 TERESSO 32 NUTO H32 Mobil DTE oil light Mobil DTE13 MobilDTE24 Q8 Verdi 32 Q8 Haydn 32 Shell Tellus 32 Shell Tellus 37 Rando Oil HD 32 Rando Oil HD-AZ-32 Wiolan HN32 Wiolan HS32

BP Energol HL68 BP Energol HLP68 Anstron HL68 Anstron HLP68

BP Energol HL46 BP Energol HLP46 Anstron HL46 Anstron HLP46

TURBELF SA68

TURBELF SA46

ELFOLNA DS68 TERESSO 68 NUTO H68 Mobil DTE oil heavy medium Mobil DTE26

ELFOLNA DS46 TERESSO 46 NUTO H46 Mobil DTE oil medium Mobil DTE15M Mobil DTE25 Q8 Verdi 46 Q8 Haydn 46 Shell Tellus 01 C 46 Shell Tellus 01 46 Rando Oil 46 Rando Oil HD B-46 Wiolan HN46 Wiolan HS46

Designation according to DIN51502 ISO VG BP

Oil companies and lubricants

DEA Elf Esso Mobil Q8 Shell Texaco Wintershall (BASF Group)

Force feed lubrication

Q8 Verdi 68 Q8 Haydn 68 Shell Tellus 01 C 68 Shell Tellus 01 68 Rando Oil 68 Rando Oil HD C-68 Wiolan HN68 Wiolan HS68

5.2.2 Bearing sizes and capacities Grease lubricated Grease lubricated medium duty bearings heavy duty bearings Frame size Pump end Drive end Pump end Drive end* 3306 Z C3 6207 Z C3 7306 pair back-to-back 1 6207 Z C3 6309 Z C3 3309 Z C3 6309 Z C3 7309 pair back-to-back 2 6311 Z C3 3311 Z C3 6311 Z C3 7311 pair back-to-back 3 6313 Z C3 3313 Z C3 6313 Z C3 7313 pair back-to-back 4 * Nilos ring fitted into bearing nut outer (3712/2) Oil lubricated Oil lubricated medium duty bearings heavy duty bearings Frame size Pump end Drive end Pump end Drive end 1 6207 C3 3306 C3 6207 C3 7306 pair back-to-back 6309 C3 7309 pair back-to-back 2 6309 C3 3309 C3 6311 C3 3311 C3 6311 C3 7311 pair back-to-back 3 4 6313 C3 3313 C3 6313 C3 7313 pair back-to-back Note: The bearing sizes do not constitute a purchasing specification.

Grease lubricated bearing capacities Pump end Drive end 45 cm3 75 cm3 150 cm3 105 cm3 150 cm3 300 cm3 240 cm3 450 cm3

Oil lubricated optional heavy duty bearings Pump end Drive end NUP 207 C3 7306 pair back-to-back NUP 309 C3 7309 pair back-to-back NUP 311 C3 7311 pair back-to-back NUP 313 C3 7313 pair back-to-back

Frame oil capacity (approx.) 0.7 litre 1.8 litre 1.4 litre 2.8 litre

5.2.3 Recommended grease lubricants Grease Temp. range ºC (ºF) Designation according to DIN BP DEA Elf Esso Mobil Q8

Grease nipples NLGI 2 * NLGI 3 ** -20 to +100 -20 to +100 (-4 to +212) (-4 to +212) K2K-20

K2K 30

Energrease LS2 Glissando 20 Elfmulti 2 Beacon 2 Mobilux 2 Rembrandt 2

Energrease LS3 Glissando 30 Elfmulti 3 Beacon 3 Mobilux 3 Rembrandt 3

Alvania Fett G2 Alvania Fett R2 Multilak 20 Multilak EP2

Multilak 30 Multilak EP3

Wintershall (BASF Group)

Wiolub LFK 2

-

SKF

LGMT 2

LGMT 3

Shell Texaco

Alvania R3

Silkolene G55/T G56/T * NLGI 2 is an alternative grease and is not to be mixed with other grades ** Factory packed bearings for the temperature range with grease nipples

5.2.4 Recommended fill quantities Refer to section 5.2.2, Bearing sizes and capacities.

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USER INSTRUCTIONS CPX, CPXR and CPXN ENGLISH C937KH013 - 09/03 ®

5.2.5

Lubrication schedule

5.2.5.1 Oil lubricated bearings Normal oil change intervals are 4 000 operating hours or at least every 6 months. For pumps on hot service or in severely damp or corrosive atmosphere, the oil will require changing more frequently. Lubricant and bearing temperature analysis can be useful in optimizing lubricant change intervals. The lubricating oil should be a high quality mineral oil having foam inhibitors. Synthetic oils may also be used if checks show that the rubber oil seals will not be adversely affected.

The pump is shipped with the coupling element removed. Ensure the direction of rotation of the motor is correct before fitting the coupling element. Direction of rotation must correspond to the direction arrow. If maintenance work has been carried out to the site' s electricity supply, the direction of rotation should be re-checked as above in case the supply phasing has been altered.

5.5 Guarding Guarding is supplied fitted to the pump set. If this has been removed or disturbed ensure that all the protective guards are securely refitted.

The bearing temperature may be allowed to rise to 50 ºC (122 ºF) above ambient, but should not exceed 82 ºC (180 ºF) (API 610 limit). A continuously rising temperature, or an abrupt rise, indicates a fault.

5.6.1 Filling and priming

Pumps which handle high temperature liquids may require their bearings to be cooled to prevent bearing temperatures exceeding their limits.

Ensure inlet pipe and pump casing is completely full of liquid before starting continuous duty operation.

5.2.5.2 Grease lubricated bearings When grease nipples are fitted, one charge between grease changes is advisable for most operating conditions; ie 2 000 hours interval.

Priming may be carried out with an ejector, vacuum pump interceptor or other equipment, or by flooding from the inlet source.

Normal intervals between grease changes are 4 000 hours or at least every 6 months.

When in service, pumps using inlet pipes with foot valves may be primed by passing liquid back from the outlet pipe through the pump.

The characteristics of the installation and severity of service will determine the frequency of lubrication. Lubricant and bearing temperature analysis can be useful in optimising lubricant change intervals. The bearing temperature may be allowed to rise to 55 ºC (131 ºF) above ambient, but should not exceed 95 ºC (204 ºF). For most operating conditions, a quality grease having a lithium soap base and NLGI consistency of No 2 or No 3 is recommended. The drop point should exceed 175 ºC (350 ºF). Never mix greases containing different bases, thickeners or additives.

5.3 Open impeller clearance

The impeller clearance is set in the factory. This may require adjustment because of piping attachment or increase in temperatures. For setting instructions see section 6.7, Setting impeller clearance.

5.4 Direction of rotation

5.6 Priming and auxiliary supplies

5.6.2 Auxiliary supplies Ensure all electrical, hydraulic, pneumatic, sealant and lubrication systems (as applicable) are connected and operational.

5.7 Starting the pump a) b) c) d) e) f)

Ensure flushing and/or cooling/ heating liquid supplies are turned ON, before starting pump. CLOSE the outlet valve. OPEN all inlet valves. Prime the pump. Start motor and check the outlet pressure. If the pressure is satisfactory, slowly OPEN the outlet valve.

g)

Do not run the pump with the outlet valve closed for a period longer than 30 seconds. h) If NO pressure, or LOW pressure, STOP the pump. Refer to section 7, Faults; causes and remedies for fault diagnosis.

Serious damage can result if the pump is started or run in the wrong direction of rotation.

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USER INSTRUCTIONS CPX, CPXR and CPXN ENGLISH C937KH013 - 09/03 ®

5.8 Running the pump

•

5.8.1 Pumps fitted with packed gland If the pump has a packed gland there must be some leakage from the gland. Gland nuts should initially be finger-tight only. Leakage should take place soon after the stuffing box is pressurised.

•

The gland must be adjusted evenly to give visible leakage and concentric alignment of the gland ring to avoid excess temperature. If no leakage takes place the packing will begin to overheat. If overheating takes place the pump should be stopped and allowed to cool before being re-started. When the pump is re-started, check to ensure leakage is taking place at the packed gland. If hot liquids are being pumped it may be necessary to slacken the gland nuts to achieve leakage. The pump should be run for 30 minutes with steady leakage and the gland nuts tightened by 10 degrees at a time until leakage is reduced to an acceptable level, normally 30 to 120 drops per minute. Bedding in of the packing may take another 30 minutes. Care must be taken when adjusting the gland on an operating pump. Safety gloves are essential. Loose clothing must not be worn to avoid being caught up by the pump shaft. Shaft guards must be replaced after the gland adjustment is complete. a short time.

Never run gland packing dry, even for

5.8.2 Pumps fitted with mechanical seal Mechanical seals require no adjustment. Any slight initial leakage will stop when the seal is run in. Before pumping dirty liquids it is advisable, if possible, to run the pump in using clean liquid to safeguard the seal face. For external flush or quench, this should be started before the pump is run and allowed to flow for a period after the pump has stopped.

•

Record the bearing temperature (t) and the ambient temperature (ta) Estimate the likely maximum ambient temperature (tb) Set the alarm at (t+tb-ta+5) ºC (t+tb-ta+10) ºF and the trip at 100 ºC (212 ºF) for oil lubrication and 105 ºC (220 ºF) for grease lubrication.

It is important, particularly with grease lubrication, to keep a check on bearing temperatures. After start up the temperature rise should be gradual, reaching a maximum after approximately 1.5 to 2 hours. This temperature rise should then remain constant or marginally reduce with time. Refer to section 6.2.3.1 for further information. 5.8.4 Normal vibration levels, alarm and trip For guidance, pumps generally fall under a classification for rigid support machines within the International rotating machinery standards and the recommended maximum levels below are based on those standards. Alarm and trip values for installed pumps should be based on the actual measurements taken on the pump in the fully commissioned as new condition. Measuring vibration at regular intervals will then show any deterioration in pump or system operating conditions. Vibration velocity – unfiltered Normal

N

Alarm

N x 1.25

Horizontal pumps 15 kW mm/s (in./s) r.m.s. ≤ 3.0 (0.12)

> 15 kW mm/s (in./s) r.m.s.

≤ 3.8 (0.15)

≤ 5.6 (0.22)

≤ 6.0 (0.24)

≤ 9.0 (0.35)

Shutdown trip N x 2.0

≤ 4.5 (0.18)

Where a grease lubricated unit is utilised in a vertical shaft configuration with a duck-foot bend onto the pump suction, the following apply: Vibration velocity – unfiltered Normal N Alarm

N x 1.25

Shutdown trip N x 2.0

Vertical configurations mm/s (in./s) r.m.s. ≤ 7.1 (0.28) ≤ 9.0 (0.35) ≤ 14.2 (0.56)

5.8.5 Stop/start frequency Pump sets are normally suitable for the number of equally spaced stop/starts per hour shown in the table below. Check capability of the driver and control/starting system before commissioning.

Never run a mechanical seal dry, even for a short time. 5.8.3 Bearings If the pumps are working in a potentially explosive atmosphere temperature or vibration monitoring at the bearings is recommended

Motor rating kW (hp)

If bearing temperatures are to be monitored it is essential that a benchmark temperature is recorded at the commissioning stage and after the bearing temperature has stabilized.

Up to 15 (20) Between 15 (20) and 90 (120) Above 90 (120)

Maximum stop/starts per hour 15 10 6

Where duty and standby pumps are installed it is recommended that they are run alternately every week.

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USER INSTRUCTIONS CPX, CPXR and CPXN ENGLISH C937KH013 - 09/03 ®

5.9 Stopping and shutdown a)

Close the outlet valve, but ensure that the pump runs in this condition for no more than a few seconds. b) Stop the pump. c) Switch off flushing and/or cooling/heating liquid supplies at a time appropriate to the process. d)

For prolonged shut-downs and especially when ambient temperatures are likely to drop below freezing point, the pump and any cooling and flushing arrangements must be drained or otherwise protected.

5.10 Hydraulic, mechanical and electrical duty

This product has been supplied to meet the performance specifications of your purchase order, however it is understood that during the life of the product these may change. The following notes may help the user decide how to evaluate the implications of any change. If in doubt contact your nearest Flowserve office.

5.10.1 Specific gravity (SG) Pump capacity and total head in metres (feet) do not change with SG, however pressure displayed on a pressure gauge is directly proportional to SG. Power absorbed is also directly proportional to SG. It is therefore important to check that any change in SG will not overload the pump driver or over-pressurize the pump. 5.10.2 Viscosity For a given flow rate the total head reduces with increased viscosity and increases with reduced viscosity. Also for a given flow rate the power absorbed increases with increased viscosity, and reduces with reduced viscosity. It is important that checks are made with your nearest Flowserve office if changes in viscosity are planned. 5.10.3 Pump speed Changing pump speed effects flow, total head, power absorbed, NPSHR, noise and vibration. Flow varies in direct proportion to pump speed, head varies as speed ratio squared and power varies as speed ratio cubed. The new duty, however, will also be dependent on the system curve. If increasing the speed, it is important therefore to ensure the maximum pump working pressure is not exceeded, the driver is not overloaded, NPSHA > NPSHR, and that noise and vibration are within local requirements and regulations. 5.10.4 Net positive suction head (NPSHA) NPSH available (NPSHA) is a measure of the head available in the pumped liquid, above its vapour pressure, at the pump suction branch.

NPSH required (NPSHR) is a measure of the head required in the pumped liquid, above its vapour pressure, to prevent the pump from cavitating. It is important that NPSHA > NPSHR. The margin between NPSHA > NPSHR should be as large as possible. If any change in NPSHA is proposed, ensure these margins are not significantly eroded. Refer to the pump performance curve to determine exact requirements particularly if flow has changed. If in doubt please consult your nearest Flowserve office for advice and details of the minimum allowable margin for your application. 5.10.5 Pumped flow Flow must not fall outside the minimum and maximum continuous safe flow shown on the pump performance curve and or data sheet.

6 MAINTENANCE 6.1 General It is the plant operator’s responsibility to ensure that all maintenance, inspection and assembly work is carried out by authorized and qualified personnel who have adequately familiarized themselves with the subject matter by studying this manual in detail. (See also section 1.6.2.) Any work on the machine must be performed when it is at a standstill. It is imperative that the procedure for shutting down the machine is followed, as described in section 5.9. On completion of work all guards and safety devices must be re-installed and made operative again. Before restarting the machine, the relevant instructions listed in section 5, Commissioning, start up, operation and shut down must be observed. Oil and grease leaks may make the ground slippery. Machine maintenance must always begin and finish by cleaning the ground and the exterior of the machine. If platforms, stairs and guard rails are required for maintenance, they must be placed for easy access to areas where maintenance and inspection are to be carried out. The positioning of these accessories must not limit access or hinder the lifting of the part to be serviced. When air or compressed inert gas is used in the maintenance process, the operator and anyone in the vicinity must be careful and have the appropriate protection.

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USER INSTRUCTIONS CPX, CPXR and CPXN ENGLISH C937KH013 - 09/03 ®

Do not spray air or compressed inert gas on skin.

6.2.1 Routine inspection (daily/weekly)

Do not direct an air or gas jet towards other people.

The following checks should be made and the appropriate action taken to remedy any deviations: a) Check operating behaviour. Ensure noise, vibration and bearing temperatures are normal. b) Check that there are no abnormal fluid or lubricant leaks (static and dynamic seals) and that any sealant systems (if fitted) are full and operating normally. c) Check that shaft seal leaks are within acceptable limits. d) Check the level and condition of oil lubricant. On grease lubricated pumps, check running hours since last recharge of grease or complete grease change. e) Check any auxiliary supplies eg heating/cooling (if fitted) are functioning correctly.

Never use air or compressed inert gas to clean clothes. Before working on the pump, take measures to prevent an uncontrolled start. Put a warning board on the starting device with the words: "Machine under repair: do not start". With electric drive equipment, lock the main switch open and withdraw any fuses. Put a warning board on the fuse box or main switch with the words: "Machine under repair: do not connect". Never clean equipment with inflammable solvents or carbon tetrachloride. Protect yourself against toxic fumes when using cleaning agents.

Refer to the manuals of any associated

6.2 Maintenance schedule

equipment for routine checks needed.

It is recommended that a maintenance plan and schedule is adopted, in line with these User Instructions, to include the following: a) Any auxiliary systems installed must be monitored, if necessary, to ensure they function correctly. b) Gland packings must be adjusted correctly to give visible leakage and concentric alignment of the gland follower to prevent excessive temperature of the packing or follower. c) Check for any leaks from gaskets and seals. The correct functioning of the shaft seal must be checked regularly. d) Check bearing lubricant level, and if the hours run show a lubricant change is required. e) Check that the duty condition is in the safe operating range for the pump. f) Check vibration, noise level and surface temperature at the bearings to confirm satisfactory operation. g) Check dirt and dust is removed from areas around close clearances, bearing housings and motors. h) Check coupling alignment and re-align if necessary. Our specialist service personnel can help with preventative maintenance records and provide condition monitoring for temperature and vibration to identify the onset of potential problems. If any problems are found the following sequence of actions should take place: a) Refer to section 7, Faults; causes and remedies, for fault diagnosis. b) Ensure equipment complies with the recommendations in this manual. c) Contact Flowserve if the problem persists.

6.2.2 Periodic inspection (six monthly) a)

Check foundation bolts for security of attachment and corrosion. b) Check pump running records for hourly usage to determine if bearing lubricant requires changing. c) The coupling should be checked for correct alignment and worn driving elements.

Refer to the manuals of any associated

equipment for periodic checks needed. 6.2 3 Re-lubrication Lubricant and bearing temperature analysis can be useful in optimizing lubricant change intervals. In general however, the following is recommended. 6.2.3.1 Oil lubricated bearings • Normal oil change intervals are 4 000 operating hours or at least every six months. For pumps on hot service or in severely damp or corrosive atmosphere, the oil will require changing more frequently. Lubricant and bearing temperature analysis can be useful in optimising lubricant change intervals. • The lubricating oil should be a high quality oil having oxidisation and foam inhibitors, or synthetic oil. • The bearing temperature may be allowed to rise to 50 °C (122 °F) above ambient, but should not exceed 82 °C (180 °F) (API 610 limit). A continuously rising temperature, or an abrupt rise, indicate a fault. • Pumps which handle high temperature liquids may require their bearings to be cooled to prevent bearing temperatures exceeding their limits.

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USER INSTRUCTIONS CPX, CPXR and CPXN ENGLISH C937KH013 - 09/03 ®

6.2.3.2 Grease lubricated bearings • When grease nipples are fitted, one charge between grease changes is advisable for most operating conditions; ie 2 000 hours interval. • Normal intervals between grease changes are 4 000 hours. The characteristics of the installation and severity of service will determine the frequency of lubrication. Lubricant and bearing temperature analysis can be useful in optimising lubricant change intervals. • For most operating conditions, a quality grease having a lithium soap base and NLGI consistency of No.2 or No.3 is recommended. The drop point should exceed 175 °C (347 °F). Never mix greases containing different bases, thickeners or additives. 6.2.4 Mechanical seals When leakage becomes unacceptable the seal will need replacement. 6.2.5 Gland packing The stuffing box split gland can be completely removed for re-packing or to enable the addition of extra rings of packing. The stuffing box is normally supplied with a lantern ring to enable a clean or pressurised flush to the centre of the packing. If not required, this can be replaced by an extra 2 rings of packing.

6.3 Spare parts 6.3.1 Ordering of spares Flowserve keep records of all pumps that have been supplied. When ordering spares the following information should be quoted. 1) Pump serial number 2) Pump size 3) Part name – taken from section 8 4) Part number – taken from section 8 5) Number of parts required The pump size and serial number are shown on the pump nameplate. To ensure continued satisfactory operation, replacement parts to the original design specification should be obtained from Flowserve. Any change to the original design specification (modification or use of a non-standard part) will invalidate the pump’s safety certification. 6.3.2 Storage of spares Spares should be stored in a clean dry area away from vibration. Inspection and re-treatment of metallic surfaces (if necessary) with preservative is recommended at 6 monthly intervals.

6.4 Recommended spares for two years operation (as per VDMA 24296) Number of pumps (including stand-by) 3 4 5 6/7 8/9 10(+)

Part no. Designation 2 2200

Impeller

2100

Shaft

1

1

2

3

30%

3

3712

Bearing nut

1

30%

4

50%

2450

Shaft sleeve

3042

Pump side bearing

1

2

3

4

50%

4

3041

Drive side bearing

1

2

50%

3

4

50%

2 2

3

2

3

4590/1* Pump casing gasket

4

6

8

9

12

150%

4610/1

4

6

8

9

12

150%

4

6

8

9

10

100%

O-ring - impeller

4610/10* O-ring - carrier 2540/2

Pump side liquid flinger

1

2

4130

Gland packing - set

4120

Gland halves

1

2

4200

Mechanical seals

1

2

3

2

3

4610/2

4 3

40% 30%

3

Power end - * NB for CPXR replace with the following parts: 4590

30%

30%

-

1

2

Pump casing gasket

8

12

16

18

24

300%

O-ring - carrier

4

6

8

9

10

100%

6.5 Tools required

A typical range of tools that will be required to maintain these pumps is listed below. Readily available in standard tool kits, and dependent on pump size: • Open ended spanners (wrenches) to suit up to M 48 screws/nuts • Socket spanners (wrenches), up to M 48 screws • Allen keys, up to 10 mm (A/F) • Range of screwdrivers • Soft mallet More specialized equipment: • Bearing pullers • Bearing induction heater • Dial test indicator • C-spanner (wrench) - for removing shaft nut. (If difficulties in sourcing are encountered, consult Flowserve.) • Coupling grip/shaft spanner

6.6 Fastener torques Screw position Casing, seal cover and others

Page 21 of 36

Screw size M8 M 10 M 12 M 16 M 20

Torque Nm (lbf•ft) 16 25 35 80 130

(12) (18) (26) (59) (96)

USER INSTRUCTIONS CPX, CPXR and CPXN ENGLISH C937KH013 - 09/03 ®

6.7 Setting impeller clearance 6.7.1 Setting CPX and CPXN impeller front clearance This procedure may be required after the pump has been dismantled or a different clearance is required. Before carrying out this procedure ensure that the mechanical seal(s) fitted can tolerate a change in their axial setting, otherwise it will be necessary to dismantle the unit and reset the seal axial position after adjusting the impeller clearance.

h) Evenly tighten the bearing housing screws keeping the dial indicator or feeler gauges reading the correct setting. Then tighten the hex nuts to lock the jacking screws in position. i) Compare the original and final gaps between the bearing carrier and housing to check if the movement of the shaft has exceeded the seal capability (over/under compression of seal). Re-position the seal to correct this. j) Check that the shaft can turn freely without binding. k) If a cartridge seal is fitted it should be reset at this point. l) Ensure the coupling distance between shaft ends (DBSE) is correct. Reset/re-align if necessary. 6.7.2 Setting CPXR impeller clearance a) The impeller does not have a fine front clearance setting and adjustment of the impeller is not normally required. b) The shaft build position should be as described in section 6.10.1.

Temp ºC (ºF) 50 (122) 100 (212) 150 (302) 200 (392) 250 (482)

Clearance mm (in.) Impellers up to 210 mm 0.3 (0.012) 0.4 (0.016) 0.5 (0.020) 0.6 (0.024) 0.7 (0.028)

Impellers Impellers (*)150CPX400 211 mm to over 260 mm (*)200CPX400 260 mm (except *) (*)150CPX500 0.4 (0.016) 0.5 (0.020) 1.0 (0.040) 0.5 (0.020) 0.6 (0.024) 1.0 (0.040) 0.6 (0.024) 0.7 (0.028) 1.1 (0.044) 0.7 (0.028) 0.8 (0.032) 1.2 (0.048) 0.8 (0.032) 0.9 (0.036) 1.3 (0.052)

a) Disconnect the coupling if it has limited axial flexibility. b) Record the gap between the bearing carrier and bearing housing using feeler gauges. c) Loosen the bearing carrier nuts and screws and back off the bearing carrier jacking screws by 2 mm (0.08 in.). d) Tighten the bearing carrier screws evenly, drawing the bearing carrier towards the bearing housing, until the impeller contacts the pump casing. Turn the shaft, during this procedure, until a detectable rub is obtained. This is the zero clearance position. e) Set a dial indicator to zero on the shaft end or measure the bearing carrier to bearing housing gap and record the measurement. f) Slacken the bearing carrier screws. g) Tighten jacking screws evenly (about one flat at a time) until the dial indicator or feeler gauge shows the correct impeller clearance from the zero clearance position. This clearance should be between 0.3 and 2 mm (0.008 and 0.080 in.) depending on the nature of the pumped fluid. See table above.

c) If the back clearance is altered, ensure that the mechanical seal(s) fitted can tolerate a change in their axial setting, otherwise it will be necessary to dismantle the unit and reset the seal axial position after adjusting the impeller clearance. d) Disconnect the coupling if it has limited axial flexibility. e) Record the gap between the bearing carrier and bearing housing using feeler gauges. f) Loosen the bearing carrier nuts and screws and back off the bearing carrier jacking screws by 2 mm (0.08 in). g) Tighten jacking screws evenly (about one flat at a time) until the feeler gauge shows the correct impeller clearance. h) Evenly tighten the bearing housing screws keeping the feeler gauges reading the correct setting. Tighten the hex nuts to lock the jacking screws in position. i) Compare the original and final gaps between the bearing carrier and housing to check if the movement of the shaft has exceeded the seal capability (over/under compression of seal). Re-position the seal to correct this.

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USER INSTRUCTIONS CPX, CPXR and CPXN ENGLISH C937KH013 - 09/03 ®

j)

Check that the shaft can turn freely without binding. k) If a cartridge seal is fitted it should be reset at this point. l) Ensure the coupling distance between shaft ends (DBSE) is correct. Reset/re-align if necessary.

6.8 Disassembly Refer to Safety section before dismantling the pump. Before dismantling the pump for overhaul, ensure genuine Flowserve replacement parts are available. Refer to sectional drawings for part numbers and identification. See section 8, Parts lists and drawings. 6.8.1 Bearing housing assembly To remove, proceed as follows: a) Disconnect all auxiliary pipes and tubes where applicable. b) Remove coupling guard and disconnect coupling. c) If oil lubricated frame, drain oil by removing drain plug. d) Record the gap between the bearing carrier and bearing housing so that this setting can be used during workshop assembly. e) Place hoist sling through bearing housing window. f) Remove casing screws and support foot to baseplate screws. g) Remove bearing housing assembly from pump casing. h) The two threaded holes in the bearing housing flange can be used for jacking screws to assist with removal. i) Remove pump casing gasket and discard. A replacement gasket will be required for assembly. j) Clean gasket mating surfaces. 6.8.2 Impeller removal NEVER APPLY HEAT TO REMOVE THE IMPELLER. TRAPPED OIL OR LUBRICANT MAY CAUSE AN EXPLOSION. a) Fit a chain wrench or bolt a bar to the holes in the coupling half, or fit a keyed shaft wrench directly to the shaft. b) Using gloved hands, raise the wrench above the work bench by turning the impeller clockwise as viewed from the impeller end of the shaft. c) Give the impeller a quick turn counter-clockwise to strike the wrench handle against the work bench surface or a wooden block. This will free the impeller from the shaft. d) The loosened impeller has an O-ring that should be discarded. Use a new O-ring for assembly.

6.8.3 Seal housing and seal The seal manufacturer’s instructions should be followed for dismantling and assembly, but the following guidance should assist with most seal types: a) Remove shaft guard (if fitted). b) Remove the seal cover nuts, if a separate seal cover is fitted, and slide the seal cover away. c) Remove the seal housing screws. d) Loosen the grub screws (used in most mechanical seals). e) Carefully pull off the seal housing and mechanical seal rotating element(s). f) Remove the seal cover. g) Remove shaft sleeve (if fitted). h) On non-cartridge seals the stationary seat remains in the seal housing/cover with its sealing member. Remove only if damaged or worn out. i) On pumps fitted with gland packing, the packing and lantern ring should be removed only if the packing is to be replaced. 6.8.4 Bearing housing a) Pull off the pump half of the coupling and remove the coupling key. b) Remove support foot (if necessary). c) Remove the pump side liquid flinger and/or labyrinth seal rotary half (depending on option fitted). d) Slacken the nuts and remove bearing carrier screws. e) Tighten bearing carrier jacking screws evenly to initiate bearing carrier release. f) Remove bearing carrier and shaft assembly from the bearing housing by pulling it towards the coupling end. g) Remove bearing circlip (or bearing carrier locking ring if paired angular contact bearings are fitted). Bearing carrier locking rings are lefthand thread. h) Remove drive side liquid flinger and/or labyrinth seal rotary half (depending on option fitted). i) Remove bearing carrier. j) Remove pump side bearing. k) Release the self locking drive side bearing nut and remove drive side bearing. l) When pressing bearings off the shaft, use force on the inner race only.

6.9 Examination of parts Used parts must be inspected before assembly to ensure the pump will subsequently run properly. In particular, fault diagnosis is essential to enhance pump and plant reliability.

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USER INSTRUCTIONS CPX, CPXR and CPXN ENGLISH C937KH013 - 09/03 ®

6.9.1 Casing, seal housing and impeller Inspect for excessive wear, pitting, corrosion, erosion or damage and any sealing surface irregularities. Replace as necessary. 6.9.2 Shaft and sleeve (if fitted) Replace if grooved or pitted. With the bearing mounting diameters (or bearing outer) supported by V-blocks, check that the shaft runouts are within 0.025 mm (0.001 in.) at the coupling end and 0.050 mm (0.002 in.) at the sleeve end.

c) Press drive side bearing(s) on to shaft. d) The double row thrust bearing will not normally have a single filling slot, as such bearings are limited to taking thrust in only one direction. If such a bearing replacement is used, it must be positioned on the shaft so that the bearing filling slot faces the impeller end of the shaft. e) If the pair of angular contact bearings are to be fitted, these must be mounted back-to-back, as shown below:

6.9.3 Gaskets and O-rings After dismantling, discard and replace. 6.9.4 Bearings It is recommended that bearings are not re-used after any removal from the shaft. 6.9.5 Bearing labyrinths/isolators The lubricant, bearings and bearing housing seals are to be inspected for contamination and damage. If oil bath lubrication is utilised, these provide useful information on operating conditions within the bearing housing. If bearing damage is not due to normal wear and the lubricant contains adverse contaminants, the cause should be corrected before the pump is returned to service. Labyrinth seals and bearing isolators should be inspected for damage but are normally non-wearing parts and can be re-used. Bearing seals are not totally leak free devices. Oil from these may cause staining adjacent to the bearings. 6.9.6 Bearing housing and carrier Inspect the bearing carrier circlip groove. Ensure it is free from damage and that housing lubrication passages are clear. Replace grease nipples or the filter breather (where fitted) if damaged or clogged. On oil lubricated versions, the oil level sight glass should be replaced if oil stained.

Nilos ring (clearance type) is only fitted on grease lubricated option units

f)

g) h) i) j) k) l) m)

6.10 Assembly

To assemble the pump consult the sectional drawings. See section 8, Parts lists and drawings. Ensure threads, gasket and O-ring mating faces are clean. Apply thread sealant to non-face sealing pipe thread fittings. 6.10.1 Bearing housing and rotating element assembly a) Clean the inside of the bearing housing, bearing carrier and bores for bearings. b) Attach bearing housing support foot.

n)

o) p)

The following methods are recommended for fitting the bearings onto the shaft: Method 1: Use a hotplate, hot bath, oven or induction heater to heat the bearing race so it can easily be placed in position then allowed to shrink and grip the shaft. It is important that the temperature is not raised above 100 ºC (212 ºF). Method 2: Press the bearing onto the shaft using equipment that can provide a steady, even load to the inner race. Take care to avoid damaging the bearing and shaft. With bearings at ambient temperature, screw on the drive side self-locking bearing nut (with its polyamide insert facing away from the bearing) until tight. With double row thrust bearings place the inner bearing circlip over the shaft, with the tapered face facing the impeller end. With the heavy duty bearing option, the locking ring should be placed between the bearings with the larger diameter end facing the impeller end. Press pump side bearing onto the shaft using Method 1 or 2 above. With the NUP roller bearing option, the loose ring should be against the shaft shoulder. Fit O-ring on the bearing carrier. Lightly lubricate the bearing carrier bore and O-ring. If a separate labyrinth type bearing housing seal is used there may be a drain hole that should be at the 6 o' clock position facing the bearing. (See manufacturer’s drawing if in doubt.) Ensure the shaft keyway edges are free of burrs. During installation, use shimming or tape over the keyway to avoid damaging the drive side bearing seals. Slide the bearing carrier onto the shaft/bearing assembly and insert inner circlip into the carrier groove or screw up the bearing locking ring. On grease lubricated pumps, pump grease through the grease nipple in the bearing carrier until grease is visible in the bearing races.

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USER INSTRUCTIONS CPX, CPXR and CPXN ENGLISH C937KH013 - 09/03 ®

q) Check shaft for free rotation. r) Fit the pump side labyrinth into the bearing housing ensuring the drain hole faces the bearing and is at the 6 o’clock position. s) Install the shaft assembly into the bearing housing until the gap is approximately 5 mm (0.2 in.).

t) Fit the bearing carrier screws but do not tighten. u) Press drive side liquid flinger and pump side liquid flinger onto shaft where applicable. These should be set 0.5 to 2 mm (0.02 to 0.08 in.) (light contact for elastomer type) from the bearing carrier and bearing housing respectively. v) The pump side flinger (this feature is integral with some proprietary labyrinth seals) should only be set in its final position after setting the shaft axial position. w) Temporarily fit the seal housing (with any internal anti-vortex rib at the topmost position). The shaft may now be positioned in relation to the seal housing face, as shown below:

Bearing housing Frame 1 Frame 2 Frame 3 Frame 4

Dia. X mm (in.) 24 (0.945) 32 (1.260) 42 (1.654) 48 (1.890)

Z mm (in.) 9 (0.354) 17 (0.669) 9 (0.354) 22 (0.866)

x) The pump side flinger may then be moved towards the bearing housing and set with its clearance. 0.5 to 2 mm (0.02 to 0.08 in.)

6.10.2 Seal housing and seal assembly a) Extreme cleanliness is required. The sealing faces and shaft or sleeve surface must be free from scratches or other damage. b) Refer to the seal arrangement section for seal diagrams. c) Carefully press the stationary seat into the mechanical seal housing or cover, ensuring that the seating ring is not deformed. Where an antirotation pin is fitted ensure that correct engagement with the slot is achieved. d) Place any separate seal covers over the shaft. e) Refer to manufacturer’s instructions to position the mechanical seal rotating elements. Tighten any drive screws in the seal drive collar. For precise compression most cartridge seals should be set after complete pump assembly. f) Fit the seal housing into the bearing housing and tighten all fasteners. 6.10.3 Gland packed stuffing box assembly a) Assemble the gland packing into the stuffing box housing before fitting on to the shaft. b) Stagger the joints in the gland packing by 90 degrees to each other. c) The lantern ring halves (if required) should be positioned mid-way along the packing. d) Position the gland squarely against the last ring and tighten the gland nuts finger-tight only. Install into bearing housing assembly and fit the two screws to hold the seal housing in place. e) Check that the shaft rotates freely. 6.10.4 Impeller assembly and setting a) Fit a new O-ring into the impellers using a small amount of grease to hold it in place. Apply antigalling compound (which does not contain copper) to the impeller thread to help subsequent removal. b) Assemble impeller onto the shaft. c) Tighten the impeller. Use the same method as in disassembly but rotating in opposite direction. A few sharp strikes will tighten it to the correct level. 6.10.5 Assembly of bearing housing into casing a) Fit a new gasket into the casing. b) Install the bearing housing assembly into the pump casing. Coat the screws with anti-galling compound and tighten into casing. c) Check impeller front clearance against original setting or process requirement and adjust as necessary. (See section 6.7, Setting impeller clearance.) d) Ensure that all other items have been re-attached and all fasteners tightened, then follow the instructions in the sections on Installation and Commissioning.

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USER INSTRUCTIONS CPX, CPXR and CPXN ENGLISH C937KH013 - 09/03 ®

6.11 Sealing arrangements

6.11.1d Single seal with external lip seal

The following section shows details of the seal arrangements. The dimensions provided are for nonstep balanced mechanical seals conforming to DIN 24960. Contact your nearest Flowserve sales office or service centre if you require further information, such as a mechanical seal dimensional drawing, or are unsure of the specific arrangement supplied. Refer also to section 4.6.5, Auxiliary piping. 6.11.1 Single seal types 6.11.1a Single stepped balanced seal

Q - Rp ¼ in. quench D - Rp ¼ in. drain F - Rp ¼ in. flush Z - Position of lip seal hard sleeve NB: Lever flange away after fitting hard sleeve to shaft. Setting dimension (mm) X Y 23.5 11.0 34.0 19.0 33.5 11.0 51.5 24.0

Bearing housing

6.11.1b Single unbalanced (or inherently balanced) seal

Frame 1 Frame 2 Frame 3 Frame 4 Pump size 125 160 200 250 315 400 500

Bearing housing Frame 1 Frame 2 Frame 3 Frame 4

Setting dimension (mm) X Y 23.5 11.0 34.0 19.0 33.5 11.0 51.5 24.0

Frame 1 41.5 41.5 36.5 -

Setting dimension Z (mm) Frame 2 Frame 3 Frame 4 49.0 49.0 44.0 45.0 44.0 45.0 65.0 36.5 57.0 44.0 45.0 65.0

6.11.1e Single internal seal with internal and external neck bush Q - Rp ¼ in. quench D - Rp ¼ in. drain F - Rp ¼ in. flush

6.11.1c Single seal with external neck bush

Q - Rp ¼ in. quench D - Rp ¼ in. drain F - Rp ¼ in. flush Bearing housing Frame 1 Frame 2 Frame 3 Frame 4

Pump size Setting dimension (mm) X Y 23.5 11.0 34.0 19.0 33.5 11.0 51.5 24.0

125 160 200 250 315 400 500

Page 26 of 36

Setting dimensions (mm) Frame 1 X Y 12.5 0 12.5 0 17.5 5.0 -

Frame 2 X Y 5.5 -9.5 5.5 -9.5 10.6 -4.4 10.6 -4.4 10.6 -4.4

Frame 3 X Y 18.3 -4.3 18.3 -4.3 27.0 4.3 18.3 -4.3

Frame 4 X Y -4.7 -32.3 3.5 -24.0 -4.7 -32.3

USER INSTRUCTIONS CPX, CPXR and CPXN ENGLISH C937KH013 - 09/03 ®

6.11.2 Cartridge seal types

6.11.4 Double seal types

6.11.2a Cartridge seal in conical seal housing

6.11.4a Double back-to-back seal with Flowserve eccentric pumping annulus circulation BI - Rp ¼ in. barrier liquid inlet BO - Rp ¼ in. barrier liquid outlet

6.11.2b DIN 24960 ’C’ cartridge seal Pump size 125 160 200 250 315 400 500

Setting dimension X (mm) Frame 1 11.0 11.0 6.0 -

Frame 2 17.5 17.5 12.4 12.4 12.4

Frame 3 14.4 14.3 5.7 14.3

For S see seal supplier’s instructions.

6.11.5 External seal types

6.11.3 Tandem seal types 6.11.3a Tandem seal with Flowserve eccentric pumping annulus circulation

6.11.5a External seal

D - drain

6.11.6 Packed gland seal types 6.11.6a Packed gland with fibre packing

BI - Rp ¼ in. barrier liquid inlet BO - Rp ¼ in. barrier liquid outlet F - Rp ¼ in. flush Pump size 125 160 200 250 315 400 500

Setting dimensions (mm) Frame 1 X Y 20.0 31.5 20.0 31.5 20.0 26.5 -

Frame 2 X Y 28.0 41.5 28.0 41.5 28.0 36.4 28.0 36.4 28.0 36.4

Frame 3 Frame 4 X Y X Y 27.5 33.7 27.5 33.7 45.5 56.7 27.5 25.3 45.5 48.3 27.5 33.7 45.5 56.7

Page 27 of 36

F - Rp ¼ in. flush

Frame 4 32.3 24.0 32.3

USER INSTRUCTIONS CPX, CPXR and CPXN ENGLISH C937KH013 - 09/03 ®

7 FAULTS; CAUSES AND REMEDIES FAULT SYMPTOM P u m p o v e r h e a ts a n d s e i z e s Bearings have short life





Pump vibrates or is noisy





Mechanical seal has short life





Mechanical seal leaks excessivel y





Pump requires excessive power





P u m p l o s e s p r i m e a ft e r s ta r t i n g





Insufficient pressure developed





Insuffi cient capaci t y deli vered





Pump does not deliver liquid









PROBABLE CAUSES

POSSIBLE REMEDIES A. System troubles



Pump not primed.























Pump or suction pipe not completely filled with liquid.









Insufficient margin between suction pressure and vapour pressure.



























Suction lift too high or level too low.































Check suction line design for vapour pockets.

Air leaks into suction line.

Check suction pipe is airtight.

Air leaks into pump through mechanical seal, sleeve joints, casing joint or pipe plugs.

Check and replace faulty parts. CONSULT FLOWSERVE.

Foot valve too small.

Investigate replacing the foot valve.

Foot valve partially clogged.

Clean foot valve.

Inlet of suction pipe insufficiently submerged.

Check out system design.

Speed too low.

CONSULT FLOWSERVE.

Speed too high.

CONSULT FLOWSERVE.

Total head of system higher than differential head of pump.



Viscosity of liquid differs from that for which designed.







Check and purge pipes and system.

Air or vapour pocket in suction line.

Specific gravity of liquid different from design. 



Check NPSHA > NPSHR, proper submergence, losses at strainers/fittings.

Excessive amount of air or gas in liquid.

Total head of system lower than pump design head. 

Check complete filling. Vent and/or prime.

Check system losses. Remedy or CONSULT FLOWSERVE.

Check and CONSULT FLOWSERVE.

Operation at very low capacity.

Measure value and check minimum permitted. Remedy or CONSULT FLOWSERVE.

Operation at high capacity.

Measure value and check maximum permitted. Remedy or CONSULT FLOWSERVE. B. Mechanical troubles













































Misalignment due to pipe strain.

Check the flange connections and eliminate strains using elastic couplings or a method permitted.

Improperly designed foundation.

Check setting of baseplate: tighten, adjust, grout base as required.

Shaft bent.

Check shaft runouts are within acceptable values. CONSULT FLOWSERVE.

Rotating part rubbing on stationary part internally.

Check and CONSULT FLOWSERVE, if necessary.

Bearings worn

Replace bearings.

Wearing ring surfaces worn.

Replace worn wear ring/surfaces.

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USER INSTRUCTIONS CPX, CPXR and CPXN ENGLISH C937KH013 - 09/03 ®

FAULT SYMPTOM P u m p o v e r h e a ts a n d s e i z e s Bearings have short life





Pump vibrates or is noisy





Mechanical seal has short life





Mechanical seal leaks excessivel y





Pump requires excessive power





P u m p l o s e s p r i m e a ft e r s ta r t i n g





Insufficient pressure developed





Insuffi cient capaci t y deli vered





Pump does not deliver liquid



























































































PROBABLE CAUSES

POSSIBLE REMEDIES

Impeller damaged or eroded.

Replace or CONSULT FLOWSERVE for improved material selection.

Leakage under sleeve due to joint failure.

Replace joint and check for damage.

Shaft sleeve worn or scored or running off centre.

Check and renew defective parts.

Mechanical seal improperly installed.

Check alignment of faces or damaged parts and assembly method used.

Incorrect type of mechanical seal for operating conditions.

CONSULT FLOWSERVE.

Shaft running off centre because of worn bearings or misalignment.

Check misalignment and correct if necessary. If alignment satisfactory check bearings for excessive wear.

Impeller out of balance resulting in vibration. Abrasive solids in liquid pumped. 

Check and CONSULT FLOWSERVE.

Internal misalignment of parts preventing seal ring and seat from mating properly. Mechanical seal was run dry.

Check mechanical seal condition and source of dry running and repair.

Internal misalignment due to improper repairs causing impeller to rub.

Check method of assembly, possible damage or state of cleanliness during assembly. Remedy or CONSULT FLOWSERVE, if necessary.

Excessive thrust caused by a mechanical failure inside the pump.

Check wear condition of impeller, its clearances and liquid passages.

Excessive grease in ball bearings.

Check method of regreasing.

Lack of lubrication for bearings.

Check hours run since last change of lubricant, the schedule and its basis.

Improper installation of bearings (damage during assembly, incorrect assembly, wrong type of bearing etc).

Check method of assembly, possible damage or state of cleanliness during assembly and type of bearing used. Remedy or CONSULT FLOWSERVE, if necessary.

Damaged bearings due to contamination.

Check contamination source and replace damaged bearings.

C. MOTOR ELECTRICAL PROBLEMS 

















Wrong direction of rotation.

Reverse 2 phases at motor terminal box.

Motor running on 2 phases only.

Check supply and fuses.

Motor running too slow.

Check motor terminal box connections and voltage.

Page 29 of 36

USER INSTRUCTIONS CPX, CPXR and CPXN ENGLISH C937KH013 - 09/03 ®

8 PARTS LISTS AND DRAWINGS 8.1 CPX and CPXN 8.1.1 CPX and CPXN exploded view drawing

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USER INSTRUCTIONS CPX, CPXR and CPXN ENGLISH C937KH013 - 09/03 ®

8.1.2 CPX and CPXN parts list Item no 1 2 3* 4 5 6 7 8 9 10

Description

Drive side liquid flinger Hex screw Shaft sleeve Bearing carrier O-ring Bearing nut Drive side bearing Inner circlip Bearing housing Oil filler plug/breather (customer option) 11 Hex screw 12 Hex screw 13 Coupling key 14 Shaft 15 Pump side bearing 16 Pump side labyrinth seal 17 Pump side liquid flinger 18 Jacking screw (use item 11) 19 Nameplate 20 Support foot 21 Hex screw 22 Hex nut 23 Hex screw (jacking point) 24 Mechanical seal housing 25 Mechanical seal 26 O-ring 27 Impeller 28 Pump casing gasket 29 Pump casing (Note: CPXN has centre line mounting feet) 30 Casing drain plug and sealing washer (customer option) * Not illustrated: 2450 optional shaft sleeve

8.1.3 CPX bearing housing sealing details Europump part no 2540/1 9906/2 2450 3240 4610/2 3712/1 3041 6546 3130 3854 9906/1 9906/5 6742 2100 3042 4300/2 2540/2 6575 3134 9906/4 9923 9906/3 4210 4200 4610/1 2200 4590/1 1111 6515

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USER INSTRUCTIONS CPX, CPXR and CPXN ENGLISH C937KH013 - 09/03 ®

8.2 CPXR 8.2.1 CPXR sectional drawing

8.2.2 CPXR parts list Europump no.

Description

4210

Mechanical seal housing

1111

Pump casing

4330

Labyrinth seal (pump side)

2100

Shaft

4590

Pump casing gasket

2200

Impeller

4610/1

O-ring

2510/1

Distance ring

4610/2

O-ring

2510/2

Seal setting collar (L1K)

6515/1

Drain plug (optional)

2540/1

Flinger (liquid) drive side

6515/2

Drain plug (magnetic) (oil lubrication only)

2540/2

Flinger (liquid) pump side

6546

Inner circlip

3041

Drive side bearing

6742

Coupling key

3042

Pump side bearing

9906/01

Hex screw

3130

Bearing housing

9906/02

Hex screw

3134

Support foot

9906/03

Hex screw

3240

Bearing carrier

9906/04

Hex screw

3712/1

Bearing nut

9923/1

Hex nut

3712/2

Bearing nut outer

9923/2

Hex nut

3854

Oil filler plug (oil lubrication only)

9951

Stud

3855

Constant level oiler (oil lubrication only)

3858

Sight glass (oil lubrication only)

2450

Shaft sleeve

4200

Mechanical seal

3853

Grease nipples (grease lubrication only)

Items not illustrated

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USER INSTRUCTIONS CPX, CPXR and CPXN ENGLISH C937KH013 - 09/03 ®

8.3 General arrangement drawing

The typical general arrangement drawing and any specific drawings required by the contract will be sent to the Purchaser separately unless the contract specifically calls for these to be included into the User Instructions. If required, copies of other drawings sent separately to the Purchaser should be obtained from the Purchaser and retained with these User Instructions.

9 CERTIFICATION

Certificates determined from the Contract requirements are provided with these Instructions where applicable. Examples are certificates for CE marking, ATEX marking etc. If required, copies of other certificates sent separately to the Purchaser should be obtained from the Purchaser for retention with these User Instructions.

10 OTHER RELEVANT DOCUMENTATION AND MANUALS

10.3 Additional sources of information Reference 1: NPSH for Rotordynamic Pumps: a reference guide, Europump Guide No. 1, Europump & World Pumps, Elsevier Science, United Kingdom, 1999. Reference 2: th Pumping Manual, 9 edition, T.C. Dickenson, Elsevier Advanced Technology, United Kingdom, 1995. Reference 3: nd Pump Handbook, 2 edition, Igor J. Karassik et al, McGraw-Hill Inc., New York, 1993. Reference 4: ANSI/HI 1.1-1.5 Centrifugal Pumps - Nomenclature, Definitions, Application and Operation. Reference 5: ANSI B31.3 - Process Piping.

10.1 Supplementary User Instruction manuals

Supplementary instruction determined from the contract requirements for inclusion into User Instructions such as for a driver, instrumentation, controller, sub-driver, seals, sealant system, mounting component etc are included under this section. If further copies of these are required they should be obtained from the purchaser for retention with these User Instructions. Where any pre-printed set of User Instructions are used, and satisfactory quality can be maintained only by avoiding copying these, they are included at the end of these User Instructions such as within a standard clear polymer software protection envelope.

10.2 Change notes

If any changes, agreed with Flowserve Pump Division, are made to the product after its supply, a record of the details should be maintained with these User Instructions.

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USER INSTRUCTIONS CPX, CPXR and CPXN ENGLISH C937KH013 - 09/03 ®

Notes:

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USER INSTRUCTIONS CPX, CPXR and CPXN ENGLISH C937KH013 - 09/03 ®

Notes:

Page 35 of 36

USER INSTRUCTIONS CPX, CPXR and CPXN ENGLISH C937KH013 - 09/03 ®

FLOWSERVE REGIONAL SALES OFFICES: Europe, Middle East & Africa Flowserve Limited (Pump Division) Harley House, 94 Hare Lane Claygate, Esher, Surrey KT10 0RB United Kingdom

Latin America Flowserve S.A. de C.V. Avenida Paseo de la Reforma 30 2nd Floor, Colonia Juarez Centro Mexico, D.F.Z.C. 06040

Tel +44 (0)1372 463 700 Fax +44 (0)1372 460 190

Tel +52 5705 5526 Fax +52 5705 1125

USA and Canada Flowserve Corporation (Pump Division) Millennium Center, 222 Las Colinas Blvd. 15th Floor, Irving, TX 75039-5421, USA

Asia Pacific Flowserve Pte Ltd (Pump Division) 200 Pandan Loop, 06-03/04 Pantech 21, Singapore 128388

Tel +1 972 443 6500 Toll free 800 728 PUMP (7867) Fax +1 972 443 6800

Tel +65 775 3003 Fax +65 779 4607

Visit our web site at: www.flowserve.com Your Flowserve factory contact:

Your local Flowserve representative:

To find your local Flowserve representative, please use the Sales Support Locator System found at www.flowserve.com

Page 36 of 36

®

Flowserve Pump Division

I.O.M. Installation, Operation and Maintenance IDP® Pumps Type CPXP SELF PRIMING, FRAME MOUNTED CHEMICAL PROCESS PUMP

Instruction Manual C961KH001 - 01/03

INSTRUCTION MANUAL CPXP ENGLISH

C961KH001 - 01/03

INTRODUCTION

Flowserve Flowserve's products and brands are the leading names in their fields: the CPX range of process pumps specifically focus on demanding chemical process applications. The pumps are manufactured at modern facilities, utilising state of art equipment and sophisticated quality control techniques. Flowserve is proud of earning preferred supplier status to many of world's leading processing companies. Engineered, manufactured, sold and serviced to ISO 9001 quality certification, Flowserve pumps are truly world class products. With more than 120 years of experience in servicing the needs of world-wide process industries, Flowserve has become the unchallenged leader in hydraulic design engineering, materials expertise and application know-how. Committed to continuous quality improvement, Flowserve controls the complete product life cycle - from application engineering, design, melting and casting, to cellular manufacturing, to assembly and testing, to the supply of aftermarket products, repair and diagnostic services. Flowserve is on hand to provide technical support and special services specific to the needs of its customers. Copyright All rights reserved. No part of this manual may be reproduced, stored in a retrieval system or transmitted in any form or by any means without prior permission of Flowserve Pump Division. CE Mark System It is a legal requirement that machinery and equipment put into service within the European Union shall conform with the applicable European Union Directives covering Machinery, Low Voltage Equipment and (EMC), Pressure Equipment Directive (PED) and Equipment for Potentially Explosive Atmospheres (ATEX). Where applicable the European Union Directives cover important Safety aspects relating to machinery and equipment and the satisfactory provision of technical documents and safety instructions. This document incorporates information relevant to these Directives. The Manual should be read prior to installing, operating, using and maintaining the equipment. The equipment must not be put into service until all the conditions relating to safety noted in the Manual have been met. Disclaimer Flowserve manufactures products to exacting International Quality Management System Standards (ISO 9001). Genuine parts and accessories have been designed, tested and incorporated into the products to ensure their continued product quality and performance in use. As Flowserve cannot test parts and accessories sourced from other vendors the incorrect incorporation of such parts and accessories may adversely affect the performance and safety features of the products. The failure to properly select, install or use authorised Flowserve parts and accessories is considered to be misuse. Damage or failure caused by misuse is not covered by Flowserve's warranty. In addition, any modification of Flowserve products or removal of original components may impair the safety of these products in their use.

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1 NAMEPLATE & WARNING LABELS 1.1 Nameplate For details of nameplate, see the Declaration of conformity. 1.2 Warning labels

Oil lubricated units only

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INSTRUCTION MANUAL CPXP ENGLISH 2 SAFETY

the shaft, packing or mechanical seal to run hot and fail within a short time.

2.1 Duty conditions This pump has been selected to meet the duty and service conditions advised on your order. The acknowledgement of these conditions has been sent separately to the Purchaser. A copy should be kept with this manual.

2.2.7 DO NOT RUN THE PUMP AT ABNORMALLY HIGH OR LOW FLOW RATES Operating at a flow rate higher than normal or at a flow rate with no back pressure on the pump may overload the motor and cause cavitation. Low flow rates may cause a reduction in pump/bearing life, overheating of the pump, instability and cavitation/vibration.

If there is any doubt as to the suitability of the pump for the application intended, contact Flowserve for advice, quoting the pump serial number.

2.2.8 NEVER DO MAINTENANCE WORK WHILST THE UNIT IS CONNECTED TO POWER

2.2 Safety action Always coordinate repair activity with operations personnel, and follow all plant safety requirements and applicable safety and health laws/regulations.

2.2.9 NEVER APPLY HEAT TO REMOVE IMPELLER Trapped lubricant or vapour could cause an explosion.

THIS IS A SUMMARY OF CONDITIONS AND ACTIONS TO PREVENT INJURY TO PERSONNEL AND DAMAGE TO EQUIPMENT.

2.2.10 HANDLING COMPONENTS Many precision parts have sharp corners and the wearing of appropriate safety gloves and equipment is required when handling these components. To lift heavy pieces above 30kg (66lbs) use a crane corresponding to the mass and in accordance with current local regulations.

This sign indicates safety instructions where non-compliance would affect personal safety. This symbol indicates electrical safety instructions where non-compliance would affect personal safety.

2.2.11 DRAIN PUMP AND ISOLATE PIPEWORK BEFORE DISMANTLING THE PUMP The appropriate safety precautions should be taken where the pumped liquids are hazardous.

This symbol indicates safety instructions where non-compliance would affect the safe operation or protection of the pump or pump unit.

2.2.12 FLUORO-ELASTOMERS (When fitted to high temperature units). When a pump has experienced temperatures over 250°C (482ºF), partial decomposition of fluoroelastomers (eg viton) will occur. In this condition these are extremely dangerous and skin contact must be avoided.

2.2.1 PREVENT EXCESSIVE EXTERNAL PIPE LOAD Do not use pump as a support for piping. Do not mount expansion joints so that their force, due to internal pressure, acts on the pump flange.

2.2.13 THERMAL SHOCK Rapid changes in the temperature of the liquid within the pump can cause thermal shock, which can result in damage or breakage of components. Thermal shock should be avoided, particularly so where the material of the pump is not resistant to such loading.

2.2.2 ONLY CHECK DIRECTION OF MOTOR ROTATION WITH COUPLING ELEMENT/ PINS REMOVED Starting in reverse direction of rotation will damage the pump. 2.2.3 START THE PUMP WITH OUTLET VALVE CLOSED This is recommended to avoid the risk of overloading and damaging the pump motor at full flow. Pumps may be started with the valve open only on installations where this situation cannot occur.

2.2.14 HOT (and cold) PARTS If hot or freezing components or auxiliary heating supplies can present a danger to operators, they must be shielded to avoid accidental contact. If complete protection is not possible, machine access must be limited to maintenance staff only. Note: drive motors and bearings may be hot.

2.2.4 ENSURE CORRECT LUBRICATION (See: "Making ready for operation - Lubrication".) 2.2.5

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IF THE TEMPERATURE IS GREATER THAN 80°C (175°F) OR BELOW 5°C (20°F), A VISUAL WARNING INDICATOR SUCH AS A WARNING PLATE MUST BE PLACED CLEARLY ON THE EQUIPMENT.

NEVER RUN THE PUMP DRY

2.2.6 INLET VALVES TO BE FULLY OPEN WHEN PUMP IS RUNNING Running the pump at zero flow or below the recommended minimum flow continuously will cause 4

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INSTRUCTION MANUAL CPXP ENGLISH

C961KH001 - 01/03

2.2.15 HAZARDOUS LIQUIDS When the pump is handling hazardous liquids care must be taken to avoid liquid contact using the appropriate health and safety procedures. Pump location and personnel access/training should consider and address these site dangers. 2.3 Potentially explosive atmospheres Always check that the driver, drive coupling assembly and pump equipment are suitably rated and/or certified for the classification of the specific atmosphere in which they are to be installed. See section 17, Certification.

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3. CONTENTS PAGE

PAGE

INTRODUCTION

2

11 STOPPING AND SHUTDOWN

14

1 NAMEPLATE & WARNING LABELS

3

2 SAFETY Duty conditions Safety action Potentially explosive atmospheres

4 4 5

3 CONTENTS

6

12 PREVENTATIVE MAINTENANCE & SERVICING Maintenance schedule Routine inspection (daily/weekly) Periodic inspection (6 monthly) Lubrication data Gland packing Mechanical seals Setting impeller front clearance

14 14 14 14 15 15 15

4 PUMP TECHNICAL DATA Performance Noise level Pressure limits Flange loads Pump lubricants Recommended screw torques

7 7 7 7 8 8

13 DISMANTLING AND ASSEMBLY Dismantling Examination of parts Assembly

15 16 17

14 SEALING ARRANGEMENTS Single seal types Cartridge seal types Tandem seal types Double seal types External seal types Packed gland seal types

19 20 20 20 20 20

15 SPARE PARTS Ordering of spares Storage of spares Recommended spares

21 21 21

16 GENERAL ARRANGEMENT DRAWING

21

17 CERTIFICATION

21

5 PRODUCT DESCRIPTION General Pump casing Impeller Shaft Bearing housing Pump bearings and lubrication Seal housing Shaft seal Driver Accessories

9 9 9 9 9 9 9 9 9 9

6 STORAGE

9

7 INSTALLATION Unpacking and inspection Handling Location Foundation Grouting Alignment of couplings Electrical connections Pipework connections Final piping check Auxiliary piping

9 9 10 10 10 10 11 11 12 12

18 SUPPLEMENTARY INSTRUCTION MANUALS

21

19 CHANGE NOTES

21

20 OPERATING DIFFICULTIES

22

21 SECTIONAL ARRANGEMENT DRAWINGS AND PARTS LISTS CPXP pump CPXP pump (diffuser-casing sizes)

23 24

8 MAKING READY FOR OPERATION Lubrication Direction of rotation Guarding Open impeller clearance Primary and auxiliary supplies Filling and self priming

12 13 13 13 13 13

21 PARTS INTERCHANGEABILITY

25

9 STARTING THE PUMP

13

10 RUNNING Pumps fitted with packed glands Pumps fitted with mechanical seals Stop/start frequency

14 14 14

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INSTRUCTION MANUAL CPXP ENGLISH 4 PUMP TECHNICAL DATA

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For units driven by equipment other than electric motors or units contained within enclosures, see the accompanying information sheets and manuals.

4.1 Performance For performance parameters see the paragraph on Safety - Duty conditions. When specified by the contract, performance data has been supplied separately to the purchaser and should be obtained and retained with this manual if required.

4.3 Pressure limits 4.3.1 The operating pressure has been selected to meet your specified requirements. See the paragraph on Safety - Duty conditions for details.

4.2 Noise level When pump noise level exceeds 85dBA attention must be given to prevailing Health and Safety Legislation, to limit the exposure of plant operating personnel to the noise. The usual approach is to control exposure time to the noise or to enclose the machine to reduce emitted sound.

4.3.2 The pressure and temperature operating limits for the flanges are in accordance with the relevant National or International standards unless advised otherwise. 4.3.3 Heating/cooling jackets are designed for operation up to 6.0 bar (87psi).

You may have already specified a limiting noise level when the equipment was ordered, however if no noise requirements were defined then machines above a certain power level will exceed 85dBA.

4.4 Flange loads The permissible flange loading is dependent on a number of factors such as dimensions, flange rating, pressure, temperature, material, pump configuration etc. The recommendations contained in the section on pipework connections should be followed to eliminate these loads.

Pump noise level is dependent on a number of factors - the type of motor fitted, the operating capacity, pipework design and acoustic characteristics of the building. The levels specified in the table below are estimated and not guaranteed.

When requested the permissible flange loading will have been supplied separately to the purchaser and should be obtained and retained with this manual.

The dBA values are based on the noisiest ungeared electric motors which are likely to be encountered. They are sound pressure levels at 1m (3.3 ft) from the directly driven pump, for "free field over a reflecting plane".

If in doubt contact Flowserve for information.

If a pump unit only has been purchased, for fitting with your own driver, then the "pump only" noise levels from the table should be combined with the level for the driver obtained from the supplier. Consult a Noise Specialist for this calculation Motor size kW <0.55 0.75 1.1 1.5 2.2 3 4 5.5 7.5 11 15 18.5 22 30 37 45 55 75 90 110 150

(hp) (<0.75) (1) (1.5) (2) (3) (4) (5) (7.5) (10) (15) (20) (25) (30) (40) (50) (60) (75) (100) (120) (150) (200)

3550 rpm Pump & Pump motor only dBA dBA 71 66 74 66 74 68 77 70 78 72 81 74 82 75 90 77 90 78 91 80 92 83 92 83 92 83 100 85 100 86 100 87 102 88 100 90 97 90 100 91 101 92

2900 rpm Pump & Pump motor only dBA dBA 64 62 67 62 67 64 70 66 71 68 74 70 75 71 83 73 83 74 84 76 85 79 85 79 85 79 93 81 93 82 93 83 95 84 95 86 92 86 95 87 96 88

1750 rpm Pump & Pump motor only dBA dBA 64 62 67 62 67 64 70 66 71 68 74 74 75 75 76 75 77 76 78 77 80 79 80 79 81 79 84 80 84 80 84 80 86 81 88 81 90 81 91 83 91 83

1450 rpm Pump & Pump motor only dBA dBA 63 62 63 62 65 64 66 66 68 68 70 70 71 71 72 71 73 72 74 73 76 75 76 75 77 75 80 76 80 76 80 76 82 77 83 78 85 78 86 79 86 79

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4.5 Pump lubricants

Centifugal pump lubrication

4.5.1 Recommended oil lubricants Oil

Splash lubrication

Viscosity mm²/s 40ºC Temp. max. ºC (ºF) Designation according to DIN51502 ISO VG BP

Oil companies and lubricants

DEA Elf Esso Mobil Q8 Shell Texaco Wintershall (BASF Group)

Force feed lubrication

32

68

46

65 (149)

80 (176)

-

HL/HLP 32

HL/HLP 68

HL/HLP 46

BP Energol HL32 BP Energol HLP32 Anstron HL32 Anstron HLP32 OLNA 32 HYDRELEF 32 TURBELF 32 ELFOLNA DS32 TERESSO 32 NUTO H32 Mobil DTE oil light Mobil DTE13 MobilDTE24 Q8 Verdi 32 Q8 Haydn 32 Shell Tellus 32 Shell Tellus 37 Rando Oil HD 32 Rando Oil HD-AZ-32 Wiolan HN32 Wiolan HS32

BP Energol HL68 BP Energol HLP68 Anstron HL68 Anstron HLP68

BP Energol HL46 BP Energol HLP46 Anstron HL46 Anstron HLP46

TURBELF SA68

TURBELF SA46

ELFOLNA DS68 TERESSO 68 NUTO H68

ELFOLNA DS46 TERESSO 46 NUTO H46 Mobil DTE oil medium Mobil DTE15M Mobil DTE25 Q8 Verdi 46 Q8 Haydn 46 Shell Tellus 01 C 46 Shell Tellus 01 46 Rando Oil 46 Rando Oil HD B-46 Wiolan HN46 Wiolan HS46

Mobil DTE oil heavy medium Mobil DTE26 Q8 Verdi 68 Q8 Haydn 68 Shell Tellus 01 C 68 Shell Tellus 01 68 Rando Oil 68 Rando Oil HD C-68 Wiolan HN68 Wiolan HS68

4.5.2 Bearing sizes and capacities Grease lubricated Grease lubricated medium duty bearings heavy duty bearings Frame size Pump end Drive end Pump end Drive end* 1 6207 Z C3 3306 Z C3 6207 Z C3 7306 pair back-to-back 2 6309 Z C3 3309 Z C3 6309 Z C3 7309 pair back-to-back 3 6311 Z C3 3311 Z C3 6311 Z C3 7311 pair back-to-back * Nilos ring fitted into bearing nut outer (3712/2) Oil lubricated Oil lubricated medium duty bearings heavy duty bearings Frame size Pump end Drive end Pump end Drive end 1 6207 C3 3306 C3 6207 C3 7306 pair back-to-back 2 6309 C3 3309 C3 6309 C3 7309 pair back-to-back 3 6311 C3 3311 C3 6311 C3 7311 pair back-to-back NB: The bearing sizes do not constitute a purchasing specification.

4.5.3 Recommended grease lubricants Grease Temp. range ºC (ºF) Designation according to DIN BP DEA Elf Esso Mobil Q8 Shell Texaco

Grease lubricated bearing capacities Pump end Drive end 45 cm3 75 cm3 3 105 cm 150 cm3 150 cm3 300 cm3

Oil lubricated optional heavy duty bearings Pump end Drive end NUP 207 C3 7306 pair back-to-back NUP 309 C3 7309 pair back-to-back NUP 311 C3 7311 pair back-to-back

Frame oil capacity (approx.) 0.7 litre 1.8 litre 1.4 litre

4.6 Recommended screw torques Screw position

Grease nipples NLGI 2 * NLGI 3 ** -20 to +100 -20 to +100 (-4 to +212) (-4 to +212) K2K-20

K2K 30

Energrease LS2 Glissando 20 Elfmulti 2 Beacon 2 Mobilux 2 Rembrandt 2 Alvania Fett G2 Alvania Fett R2 Multilak 20 Multilak EP2

Energrease LS3 Glissando 30 Elfmulti 3 Beacon 3 Mobilux 3 Rembrandt 3

Casing and seal cover

Screw size M8 M10 M12 M16 M20

Torque Nm (lbf ft) 16 (12) 25 (18) 35 (26) 80 (59) 130 (96)

Alvania R3 Multilak 30 Multilak EP3

Wintershall Wiolub LFK 2 (BASF Group) SKF LGMT 2 LGMT 3 Silkolene G55/T G56/T * NLGI 2 is an alternative grease and is not to be mixed with other grades ** Factory packed bearings for the temperature range with grease nipples

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INSTRUCTION MANUAL CPXP ENGLISH 5 PRODUCT DESCRIPTION

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Fan cooling is available for high temperature operation. This is a fan fitted within the coupling guard to blow cooling air over the bearing housing and shaft.

5.1 General The pump is a modular designed centrifugal pump which can be built to achieve almost all chemical liquid pumping requirements.

6 STORAGE Store the pump in a clean, dry location away from vibration. Leave piping connection covers in place to keep dirt and other foreign material out of pump casing. Turn pump at intervals to prevent brinelling of the bearings and the seal faces, if fitted, from sticking.

5.2 Pump casing The pump casing is designed with a self priming action which works on the reflux principle for suction lifts up to 7m (23 ft). It has a horizontal centreline end inlet and a vertical centreline top outlet which makes it self venting.

The pump may be stored as above for up to 6 months. Consult Flowserve for preservative actions when a longer storage period is needed.

For ease of maintenance, the pump is constructed so that pipe connectors do not have to be disturbed when internal maintenance is required.

Warranty for the pumps will normally be for 12 months. Extension of this period can only be achieved with the prior agreement of Flowserve and would necessitate inspection, prior to putting the pump into service.

5.3 Impeller An open impeller is fitted. 5.4 Shaft The large diameter stiff shaft, mounted on bearings, has a keyed drive end.

7 INSTALLATION

5.5 Bearing housing The bearing housing enables adjustment of impeller face clearance (open impellers only) via the bearing carrier jacking screws.

7.1 Unpacking and inspection The pump should be checked against the delivery advice note and any damage or shortage reported immediately to Flowserve. Any crate, carton or wrappings should be checked for spare parts or accessories that may be packed with the pump.

5.6 Pump bearings and lubrication The pump is fitted with ball and/or roller type bearings which may be configured differently dependent on use. (The bearings may be oil or grease lubricated.)

7.2 Handling Boxes, crates, pallets or cartons may be unloaded using fork lift vehicles or slings dependent on their size and construction.

5.7 Seal housing The seal housing has spigots between the pump casing and bearing housing for optimum concentricity.

The pump and cast iron baseplate set should be handled as shown in the appropriate diagram. Where the baseplate is folded steel there are no specific lifting points provided for this complete machine set (unless so identified). Any lifting points that can be seen are provided only for dismantling parts for servicing. Slings, ropes and other lifting gear should be positioned where they cannot slip and where a balanced lift is obtained.

A fully confined gasket forms the seal between the pump casing and the seal housing. The seal housing designs provide improved performance of mechanical seals. The design enables one of a number of sealing options to be fitted. 5.8 Shaft seal The mechanical seal(s) attached to the drive shaft seals the pumped liquid from the environment. (Gland packing may be fitted as an option.) 5.9 Driver The DRIVER is normally an electric motor. Different drive configurations may be fitted such as an internal combustion engine, turbines, hydraulic motors etc driving via couplings, belts, gearboxes, drive shafts etc. 5.10 Accessories Accessories may be fitted when specified by the customer. 9

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INSTRUCTION MANUAL CPXP ENGLISH 7.3 Location The pump should be located to allow room for access, ventilation, maintenance and inspection with ample headroom for lifting and should be as close as practicable to the supply of liquid to be pumped.

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Motor and pump centre line height adjustment:

7.4 Foundation 7.4.1 There are many methods of installing pump units to their foundations. The correct method depends on the size of the pump unit, its location and noise vibration limitations. Non-compliance with the provision of correct foundation and installation may lead to failure of the pump and as such, would be outside the terms of the warranty. 7.4.2 The baseplate should be mounted onto a firm foundation, either an appropriate thickness of quality concrete or sturdy steel framework. It should NOT be distorted or pulled down onto the surface of the foundation, but should be supported to maintain the original alignment. 7.4.3 Install the baseplate onto packing pieces evenly spaced and adjacent to foundation bolts. Level with shims between baseplate and packing pieces. The pump and driver have been aligned before dispatch. Check alignment of pump and motor half coupling. If this is incorrect, it indicates that the baseplate has become twisted and should be corrected by reshimming.

Graph based on the assumptions that: a) Operating temperature rise of the motor frame is 50°C (122ºF). b) Packing piece/motor stool is not affected. Operation 1 Enter graph at base to shaft centre line height. 2 Read line for frame material. 3 Set motor shaft and coupling LOW by figure on left-hand side.

7.6.2 Alignment methods

7.5 Grouting

7.6.2.1 Ensure the pump and motor half couplings are disconnected.

7.5.1 Where applicable, grout in the foundation bolts.

7.6.2.2 The alignment MUST be checked. Although the pump will have been aligned at the factory, it is most likely that this alignment will have been disturbed during transportation or handling. Align the motor to the pump, not the pump to the motor. Alignment is achieved by adding or removing shims from under the motor feet and also moving the motor horizontally as required. In some cases, where the alignment cannot be achieved, it will be necessary to move the pump before recommencing the above procedure.

7.5.2 After adding pipework connections and rechecking the coupling alignment, the baseplate should then be grouted in accordance with good engineering practice. Fabricated steel, cast iron and epoxy baseplates can be filled with grout. Folded steel baseplates should be grouted to locate their packing pieces. If in any doubt, please contact your nearest service centre for advice. 7.5.3 Grouting provides solid contact between the pump unit and foundation, prevents lateral movement of vibrating equipment and dampens resonant vibrations. 7.6 Alignment of couplings 7.6.1 Thermal expansion The pump and motor will normally have to be aligned at ambient temperature and should be corrected to allow for thermal expansion at operating temperature. In pump installations involving high liquid temperatures, the unit should be run at the actual operating temperature, shut down and the alignment checked immediately.

7.6.2.3 For couplings with narrow flanges, use a dial indicator gauge as shown. The alignment values are maximums for continuous service.

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Permissible misalignment limits at working temperature: • Parallel alignment - 0.25mm (0.010in) TIR maximum • Angular alignment - 0.3mm (0.012in) TIR maximum for couplings not exceeding 100mm (4in) flange diameter - 0.5mm (0.020in) TIR maximum for couplings over 100mm (4in) diameter

misalignment, hot bearings, worn couplings, vibration and the possible failure of the pump casing, the following points should be strictly followed: • Prevent excessive external pipe load. • Never draw piping into place by applying force to pump flange connections. • Do not mount expansion joints so that their force, due to internal pressure, acts on the pump flange.

7.6.2.4 When checking parallel alignment, the total indicator read-out (TIR) shown is twice the value of the actual shaft displacement.

7.8.3 The inlet pipe should be as short as possible, airtight and the smallest volume as practical for the pump flow rate.

7.7 Electrical connections

It is recommended that the pump inlet pipe is no larger than the pump inlet bore or such that the suction velocity is in the range of 3 to 5m/sec (10 to 16ft/sec). The piping should be inclined up towards the pump inlet.

7.7.1 Electrical connections should be made by a qualified Electrician in accordance with the relevant local national and international regulations. 7.7.2 It is important to be aware of the EUROPEAN DIRECTIVE on electromagnetic compatibility when wiring up and installing equipment on site.

7.8.4 Allow a minimum of two pipe diameters of straight section between the elbow and inlet flange. Use of an inlet strainer is not recommended as this can prevent the priming process.

Attention must be paid to ensure that the techniques used during wiring/installation do not increase electromagnetic emissions or decrease the electromagnetic immunity of the equipment, wiring or any connected devices. If in any doubt contact Flowserve for advice.

7.8.5 Fitting an isolator and non-return valves can allow easier maintenance. Never throttle pump on suction side and never place a valve directly on the pump inlet nozzle.

7.7.3 The motor must be wired up in accordance with the motor manufacturer's instructions (normally supplied within the terminal box) including any temperature, earth leakage, current and other protective devices as appropriate. The identification nameplate should be checked to ensure the power supply is appropriate.

7.8.7 A regulating valve should be fitted in the discharge pipework, unless pump flow is controlled by the delivery system design.

7.8.6 If a non-return valve is located in the discharge pipework then a vent/bleed pipe should be fitted from the discharge pipe back to the sump or source tank.

7.8.8 The delivery pipework must permit priming air to escape unhindered from the pump during the priming cycle, without back pressure and prevent excessive run-back of liquid on shutdown to minimise syphoning.

7.7.4 A device to provide emergency stopping shall be fitted. 7.7.5 If not supplied pre-wired to the pump unit the controller/starter electrical details will also be supplied within the controller/starter.

Priming air may be vented in one of the following ways: 1) The discharge pipework regulating valve, if fitted, may be partly opened during the priming cycle to freely vent the air. 2) An automatic air release valve may be fitted to the discharge pipework, between the pump and any valves, providing that gases and vapours given off are environmentally safe and acceptable for release into the atmosphere. 3) An air bleed pipe may be run from the discharge pipework, between the pump and any valves, back to the suction tank or sump. This arrangement has a disadvantage in that normal manual/automatic control will be necessary during operation in order to prevent continuous recirculation of the pumped liquid.

7.7.6 For electrical details on pump sets with controllers see the wiring diagram. 7.7.7 See paragraphs on 'direction of rotation' before connecting the motor to the electrical supply. 7.8 Pipework connections 7.8.1 Protective covers are fitted to the pipe connections to prevent foreign bodies entering during transportation and installation. Ensure that these covers are removed from the pump before connecting any pipes.

7.8.9 Piping and fittings should be flushed before use.

7.8.2 Maximum forces and moments allowed on the pump flanges vary with the pump size and type. To minimise these forces and moments that may cause

7.8.10 Piping for corrosive liquids should be arranged to allow pump flushing before removal of a unit. 11 ®

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7.9 Final piping check After connecting piping to the pump, rotate the shaft several times by hand to ensure there is no binding and all parts are free. Recheck the coupling alignment, as previously described, to ensure no pipe strain. If pipe strain exists, correct piping. 7.10 Auxiliary piping 7.10.1 The connections which are to be piped up will have been fitted with protective metal or plastic plugs which will need to be removed. 7.10.2 Pumps fitted with packed glands When suction pressure is below ambient pressure and differential head is less than 10m (32.8ft), it may be necessary to feed gland packing with liquid to provide lubrication and prevent the ingress of air.

NOTES: a) Total seal pressure is equal to pressure at seal plus suction pressure. b) For pumped liquid viscosities greater than 440 Centistokes multiply the generated pressure by 1.25 for 125, 160 and 200 size pumps and by 2.0 for larger sizes. c) Differential pressure in bar equals head in metres multiplied by specific gravity all divided by 10.19. d) For psi multiply the pressure in bar by 14.504.

7.10.3 Pumps fitted with mechanical seals 7.10.3.1 The conical design of the single internal seal housing provides excellent liquid circulation around the seal and will not normally require a separate flush.

7.10.3.8 Special seals may require modification to auxiliary piping described above. Consult Flowserve if unsure of correct method or arrangement.

7.10.3.2 Single seals requiring re-circulation will normally be provided with the auxiliary piping from pump casing already fitted.

7.10.3.9 For pumping hot liquids, to avoid seal damage, it is recommended that any external flush/cooling supply be continued after stopping the pump.

7.10.3.3 Flowserve seal connections are designated as follows: Q - quench F - flush D - drain outlet BI - barrier fluid in (double seals) BO - barrier fluid out (double seals) H - heating jacket C - cooling jacket

7.10.3.10 Tandem seals require a barrier liquid between the seals that is compatible with the pumped liquid. 7.10.4 Pumps fitted with heating/cooling jackets Connect the heating/cooling pipes from the site supply. The top connection should be used as the outlet to ensure complete filling/venting of the annulus.

7.10.3.4 Seal housings/covers having an auxiliary quench connection, require connection to a suitable source of liquid flow, low pressure steam or static pressure from a header tank. Recommended pressure is 0.35 bar (5psi) or less.

8 MAKING READY FOR OPERATION 8.1 Lubrication 8.1.1 Determine the mode of lubrication of the pump set, eg grease, oil, product lubrication etc.

7.10.3.5 Double seals require a barrier liquid between the seals, compatible with the pumped liquid.

8.1.2 For oil lubricated pumps, fill the bearing housing with the correct grade of oil to the correct level, ie sight glass or constant level oiler bottle.

7.10.3.6 With back-to-back double seals, the barrier liquid should be at a minimum pressure of 1 bar (14.5psi) above the maximum pressure on the pump side of the inner seal. The barrier liquid pressure must not exceed limitations of the seal on the atmospheric side. For toxic service the barrier liquid supply and discharge must be in a safe area. 7.10.3.7 Seal chamber pressure v generated head: MECHANICAL SEAL GLAND PACKING

8.1.2.1 When fitted with a constant level oiler, the bearing housing should be filled by unscrewing or hinging back the transparent bottle and filling the bottle with oil. Where an adjustable body Denco oiler is fitted this should be set to the height shown in the following diagram.

Use seal manufacturer's limits or ask seal manufacturer to verify seal pressure Maximum stuffing box pressure = 5 bar (3500rpm), 7 bar (2900rpm) and 10 bar (1450 & 1750rpm)

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8.6.2 The pump has self priming action for which a separate air pump or foot valve or non-return valve is not normally required. 9 STARTING THE PUMP 9.1 Ensure flushing and/or cooling/ heating liquid supplies are turned ON, before starting pump. 9.2 CLOSE the outlet valve. The oil filled bottle should then be refitted so as to return it to the upright position. Filling of the bottle should be repeated until oil remains visible within the bottle.

9.3 OPEN all inlet valves. 9.4 Prime the pump. The pump casing must initially be filled with compatible liquid before starting the unit. Damage will occur if the pump is run dry or for prolonged periods with no incoming liquid.

8.1.3 To fill the bearing housing with oil, unscrew the oil filler/breather and fill through the orifice. 8.1.4 Grease lubricated pumps and electric motors are supplied pre-greased.

Subsequent filling should not be necessary unless the pump has been emptied or drained of fluid.

8.1.5 Other drivers and gearboxes, if appropriate, should be lubricated in accordance with their manuals. 8.2 Direction of rotation

Pump housing filling hole.

8.2.1 Serious damage can result if the pump is started or run in the wrong direction of rotation.

When the initial fill reaches the suction pipe, excess liquid will flow out of the casing.

8.2.2 The pump is shipped with the coupling element removed. Ensure the direction of rotation of the motor is correct before fitting the coupling element. Direction of rotation must correspond to the direction arrow. 8.2.3 If maintenance work has been carried out to the site's electricity supply, the direction of rotation should be re-checked as above in case the supply phasing has been altered.

Pump size 40-40CPXP125 80-80CPXP125 40-40CPXP160 80-80CPXP160 40-40CPXP200 65-65CPXP200 80-80CPXP250 100-100CPXP250 100-100CPXP315 150-150CPXP315

8.3 Guarding Guarding is supplied fitted to the pump set. If this has been removed or disturbed ensure that all the protective guards are securely refitted. 8.4 Open impeller clearance The impeller clearance is set in the factory. This may require adjustment because of piping attachment or increase in temperatures. For setting instructions refer to the Preventative Maintenance and Servicing section of this book.

Initial fill in Litres (US gal) 2.5 (0.65) 6.0 (1.50) 3.0 (0.80) 6.5 (1.75) 5.0 (1.35) 8.5 (2.25) 12.0 (3.20) 36.0 (9.5) 14.8 (3.95) 18.0 (4.8)

9.5 Start the motor and, if no specific provision has been made in the delivery pipework for evacuating the primed air, open the delivery valve by approximately 10% to allow priming air to escape. Check outlet pressure.

8.5 Primary and auxiliary supplies Ensure all electrical, hydraulic, pneumatic, sealant and lubrication systems (as applicable) are connected and operational.

9.6 If pressure is satisfactory, slowly OPEN outlet valve. 9.7 It is recommended that the priming time is noted. Priming times in excess of 5 minutes will indicate a pump or system fault. Any noticeable increases in priming time on subsequent starts will also indicate a fault. Irregular use could lead to the danger of 'evaporation' of the priming fluid.

8.6 Filling and self priming 8.6.1 Fill the pump with liquid to be pumped, or compatible liquid, via the filling plug, before starting continuous duty operation. 13

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INSTRUCTION MANUAL CPXP ENGLISH

9.8 Do not run the pump with the outlet valve closed for a period longer than 30 seconds. (If the pump has to self prime the system it may take a short time before the outlet is pressurised.)

C961KH001 - 01/03

11.4 For prolonged shut-downs and especially when ambient temperatures are likely to drop below freezing point, the pump and any cooling and flushing arrangements must be drained or otherwise protected.

9.9 If NO pressure, or LOW pressure, STOP the pump. Refer to fault finding chart for fault diagnosis.

12 PREVENTATIVE MAINTENANCE & SERVICING

10 RUNNING 10.1 Pumps fitted with packed glands 10.1.1 If the pump has a packed gland there must be some leakage from the gland. Gland nuts should initially be finger-tight only. Leakage should take place soon after the stuffing box is pressurised. If no leakage takes place the packing will begin to overheat. If overheating takes place the pump should be stopped and allowed to cool before being re-started. When the pump is restarted it should be checked to ensure leakage is taking place at the packed gland. 10.1.2 If hot liquids are being pumped it may be necessary to slacken the gland nuts to achieve leakage. 10.1.3 The pump should be run for ten minutes with steady leakage and the gland nuts tightened by 10 degrees at a time until leakage is reduced to an acceptable level, normally 30 to 120 drops per minute. Bedding in of the packing may take another 15 minutes. 10.2 Pumps fitted with mechanical seals 10.2.1 Mechanical seals require no adjustment. Any slight initial leakage will stop when the seal is run in. Seals will always leak in operation.

12.1 Maintenance schedule Our specialist service personnel can help with preventative maintenance records and provide condition monitoring for temperature and vibration to identify the onset of potential problems. 12.2 Routine inspection (daily/weekly) The following checks should be made and the appropriate action taken to remedy any deviations: • Check operating behaviour; ensure noise, vibration and bearing temperatures are normal. • Check that there are no abnormal fluid or lubricant leaks (static and dynamic seals) and that any sealant systems (if fitted) are full and operating normally. • Check that shaft seal leaks are within acceptable limits. • Check the level and condition of oil lubricant. On grease lubricated pumps, check running hours since last recharge of grease or complete grease change. • Check any auxiliary supplies eg heating/cooling (if fitted) are functioning correctly. • Refer to the manuals of any associated equipment for routine checks needed. 12.3 Periodic inspection (6 monthly) • Check foundation bolts for security of attachment and corrosion. • Check pump running records for hourly usage to determine if bearing lubricant requires changing. • The coupling should be checked for correct alignment and worn driving elements. • Refer to the manuals of any associated equipment for periodic checks needed.

10.2.2 Before pumping dirty liquids, it is advisable, if possible, to run the pump in using clean liquid to safeguard the seal face. 10.2.3 For external flush or quench, this should be started before the pump is run and allowed to flow for a period after the pump has stopped.

12.4 Lubrication data 12.4.1 Oil lubricated bearings • Normal oil change intervals are 4000 operating hours. For pumps on hot service or in severely damp or corrosive atmosphere, the oil will require changing more frequently. Lubricant and bearing temperature analysis can be useful in optimising lubricant change intervals. • The lubricating oil should be a high quality oil having oxidisation and foam inhibitors, or synthetic oil. • The bearing temperature may be allowed to rise to 50°C (122ºF) above ambient, but should not exceed 82°C (180ºF), API 610 limit. A continuously rising temperature, or an abrupt rise, indicate a fault. • Pumps which handle high temperature liquids may require their bearings to be cooled to prevent bearing temperatures exceeding their limits.

10.3 Stop/start frequency Generally 6 stop/starts per hour may be satisfactory. Refer frequent stop/starting to motor manufacturer. STANDBY PUMPS SHOULD BE RUN ALTERNATELY. 11 STOPPING AND SHUTDOWN 11.1 Close the outlet valve, but ensure that the pump runs in this condition for no more than a few seconds. 11.2 Stop the pump. 11.3 Switch off flushing and/or cooling/heating liquid supplies at a time appropriate to the process. 14

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INSTRUCTION MANUAL CPXP ENGLISH 12.4.2 Grease lubricated bearings • When grease nipples are fitted, one charge between grease changes is advisable for most operating conditions; ie 2000 hours interval. • Normal intervals between grease changes are 4000 hours. The characteristics of the installation and severity of service will determine the frequency of lubrication. • Lubricant and bearing temperature analysis can be useful in optimising lubricant change intervals. • For most operating conditions, a quality grease having a lithium soap base and NLGI consistency of No.2 or No.3 is recommended. The drop point should exceed 175°C (347°C).

•

Never mix greases containing different bases, thickeners or additives.

•

• • •

12.5 Gland packing The stuffing box split gland can be completely removed for re-packing or to enable the addition of extra rings of packing.

•

The stuffing box is normally supplied with a lantern ring to enable a clean or pressurised flush to the centre of the packing. If not required, this can be replaced by an extra 2 rings of packing.

• • •

12.6 Mechanical seals When leakage becomes unacceptable the seal will need replacement.

C961KH001 - 01/03

Tighten the bearing carrier screws evenly, drawing the bearing carrier towards the bearing housing, until the impeller contacts the pump casing. Turn the shaft, during this procedure, until a detectable rub is obtained. This is the zero clearance position. Set a dial indicator to zero on the shaft end or measure the bearing carrier to bearing housing gap and record the measurement. Slacken the bearing carrier screws. Tighten jacking screws evenly (about one flat at a time) until the dial indicator or feeler gauge shows the correct cold impeller clearance setting from the zero clearance position. The clearance setting is given in the table in this section. Evenly tighten the bearing housing screws keeping the dial indicator or feeler gauges reading the correct setting. Then tighten the hex nuts to lock the jacking screws in position. Compare the original and final gaps between the bearing carrier and housing to check if the movement of the shaft has exceeded the seal capability (over/under compression of seal). Re-position the seal to correct this. Check that the shaft can turn freely without binding. If a cartridge seal is fitted it should be reset at this point. Ensure the coupling distance between shaft ends (DBSE) is correct. Reset/re-align if necessary.

13 DISMANTLING AND ASSEMBLY

12.7 Setting impeller front clearance This procedure may be required after the pump has been dismantled or a different clearance is required.

13.1 Dismantling 13.1.1 Refer to Safety section before dismantling the pump. 13.1.2 Before dismantling the pump for overhaul, ensure genuine Flowserve replacement parts are available.

Temp. ºC (ºF) 50 (122) 100 (212) 150 (302) 200 (392) 250 (482)

13.1.3 Bearing housing assembly To remove bearing housing assembly, proceed as follows: • Disconnect all auxiliary pipes and tubes where applicable. • Remove coupling guard and disconnect coupling. • If oil lubricated frame, drain oil by removing drain plug. • Record the gap between the bearing carrier and bearing housing so that this setting can be used during workshop assembly. • Place hoist sling through bearing housing window. • Remove casing screws. • Remove bearing housing assembly from pump casing. • The two threaded holes in the bearing housing flange can be used for jacking screws to assist with removal. • Remove pump casing gasket 4590/1 and discard. A replacement gasket will be required for assembly. • Clean gasket mating surfaces. • On diffuser casing sizes it is not normally necessary to remove the diffuser 1410, 4590/2 and 9906/5.

Clearance mm (inches) Impellers Impellers Impellers up to 210mm 211 to 260mm over 260mm 0.3 (0.012) 0.4 (0.016) 0.5 (0.020) 0.4 (0.016) 0.5 (0.020) 0.6 (0.024) 0.5 (0.020) 0.6 (0.024) 0.7 (0.028) 0.6 (0.024) 0.7 (0.028) 0.8 (0.032) 0.7 (0.028) 0.8 (0.032) 0.9 (0.036)

Before carrying out this procedure ensure that the mechanical seal(s) fitted can tolerate a change in their axial setting, otherwise it will be necessary to dismantle the unit and reset the seal axial position after adjusting the impeller clearance. • Disconnect the coupling if it has limited axial flexibility. • Record the gap between the bearing carrier and bearing housing using feeler gauges. • Loosen the bearing carrier nuts and screws and back off the bearing carrier jacking screws by 2mm (0.080in). 15

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INSTRUCTION MANUAL CPXP ENGLISH

C961KH001 - 01/03

13.1.4 Impeller removal

13.2 Examination of parts

NEVER APPLY HEAT TO REMOVE THE IMPELLER. TRAPPED OIL OR LUBRICANT MAY CAUSE AN EXPLOSION. • Fit a chain wrench or bolt a bar to the holes in the coupling half, or fit a keyed shaft wrench directly to the shaft. • Using gloved hands, raise the wrench above the work bench by turning the impeller clockwise as viewed from the impeller end of the shaft. • Give the impeller a quick turn counter-clockwise to strike the wrench handle against the work bench surface or a wooden block. This will free the impeller from the shaft. • The loosened impeller has an O-ring that should be discarded. Use a new O-ring for assembly.

13.2.1 Used parts must be inspected before assembly to ensure the pump will subsequently run properly. In particular, fault diagnosis is essential to enhance pump and plant reliability. 13.2.2 Casing, seal housing and impeller • Inspect for excessive wear, pitting, corrosion, erosion or damage and any sealing surface irregularities. • Replace as necessary. 13.2.3 Shaft and sleeve (if fitted) • Replace if grooved or pitted. • With the bearing mounting diameters (or bearing outer) supported by V blocks, check that the shaft runouts are within 0.025mm (0.001in) at the coupling end and 0.050mm (0.002in) at the sleeve end.

13.1.5 Seal housing and seal The seal manufacturer's instructions should be followed for dismantling and assembly, but the following guidance should assist with most seal types: • Remove the seal housing screws. • Remove the seal cover nuts, if a separate seal cover is fitted, and slide the seal cover away. • Loosen the grub screws (used in most mechanical seals). • Carefully pull off the seal housing and mechanical seal rotating element(s). Remove the seal cover. • Remove shaft sleeve (if fitted). • On non-cartridge seals the stationary seat remains in the seal housing/cover with its sealing member. Remove only if damaged or worn out. • On pumps fitted with gland packing, the packing and lantern ring should be removed only if the packing is to be replaced.

13.2.4 Gaskets and O-rings • After dismantling, discard and replace. 13.2.5 Bearings • It is recommended that bearings are not re-used after any removal from the shaft. 13.2.6 Bearing isolators, labyrinths or lip seals (if fitted) • The lubricant, bearings and bearing housing seals are to be inspected for contamination and damage. If oil bath lubrication is utilised, these provide useful information on operating conditions within the bearing housing. • If bearing damage is not due to normal wear and the lubricant contains adverse contaminants, the cause should be corrected before the pump is returned to service. • Labyrinth seals and bearing isolators should be inspected for damage but are normally nonwearing parts and can be re-used. • Bearing seals are not totally leak free devices. Oil from these may cause staining adjacent to the bearings.

13.1.6 Bearing housing • Pull off the pump half of coupling and remove the coupling key. • Remove support foot. • Remove the pump side liquid flinger and/or labyrinth seal rotary half (depending on option fitted). • Slacken the nuts and remove bearing carrier screws. • Tighten bearing carrier jacking screws evenly to initiate bearing carrier release. • Remove bearing carrier and shaft assembly from the bearing housing by pulling it towards the coupling end. • Remove bearing circlip (or bearing carrier locking ring if paired angular contact bearings are fitted). NB: Bearing carrier locking rings are left-hand thread. • Remove drive side liquid flinger and/or labyrinth seal rotary half (depending on option fitted). • Remove bearing carrier. • Remove pump side bearing. • Release the self locking drive side bearing nut and remove drive side bearing. • When pressing bearings off the shaft, use force on the inner race only.

13.2.7 Bearing housing and carrier • Inspect the bearing carrier circlip groove, ensure it is free from damage and that housing lubrication passages are clear. • Replace grease nipples or the filter breather (where fitted) if damaged or clogged. • On oil lubricated versions, the oil level sight glass should be replaced if oil stained.

16 ®

INSTRUCTION MANUAL CPXP ENGLISH 13.3 Assembly

•

13.3.1 To assemble the pump consult the sectional drawings.

•

13.3.2 Ensure threads, gasket and O-ring mating faces are clean. Apply thread sealant to non-face sealing pipe thread fittings.

•

13.3.3 Bearing housing and rotating element assembly • Clean the inside of the bearing housing, bearing carrier and bores for bearings. • Attach bearing housing support foot. • Press drive side bearing(s) on to shaft. • The double row thrust bearing will not normally have a single filling slot, as such bearings are limited to taking thrust in only one direction. If such a bearing replacement is used, it must be positioned on the shaft so that the bearing filling slot faces the impeller end of the shaft. • If duplex bearings are to be fitted, these must be mounted back-to-back, as shown below.

• • • • •

C961KH001 - 01/03

If a separate labyrinth type bearing housing seal is used there may be a drain hole that should be at the 6 o'clock position facing the bearing. Fit drive side radial lip seal (if fitted) into the bearing carrier, having filled the position between the two lips with grease. Ensure the shaft keyway edges are free of burrs. During installation, use shimming or tape over the keyway to avoid damaging the drive side bearing seals Slide the bearing carrier onto the shaft/bearing assembly and insert inner circlip into the carrier groove or screw up the bearing locking ring. On grease lubricated pumps, pump grease through the grease nipple in the bearing carrier until grease is visible in the bearing races. Check shaft for free rotation. Fit the pump side labyrinth into the bearing housing ensuring the drain hole faces the bearing and is at the 6 o'clock position. Install the shaft assembly into the bearing housing until gap is approximately 5mm (0.2in).

Nilos ring (clearance type) is only fitted on grease lubricated option units

•

•

• •

• • •

The following methods are recommended for fitting the bearings onto the shaft: Method 1 - Use a hotplate, hot bath, oven or induction heater to heat the bearing race so it can easily be placed in position then allowed to shrink and grip the shaft. It is important that the temperature is not raised above 100ºC (212ºF). Method 2 - Press the bearing onto the shaft using equipment that can provide a steady, even load to the inner race. Care is to be taken to avoid damage to the bearing and shaft. With bearings at ambient temperature, screw on the drive side self-locking bearing nut (with its polyamide insert facing away from the bearing) until tight. With double row thrust bearings place the inner bearing circlip over the shaft with the tapered face facing the impeller end. With the heavy duty bearing option, the locking ring should be placed between the bearings with the face marked 'left-hand thread' facing the impeller end. Press pump side bearing onto the shaft using method 1 or 2 above. With the NUP roller bearing option, the loose ring should be against the shaft shoulder. Fit O-ring on the bearing carrier. Lightly lubricate the bearing carrier bore and O-ring.

• •

•

•

Fit the bearing carrier screws but do not tighten. Press drive side liquid flinger and pump side liquid flinger onto shaft where applicable. These should be set 0.5 to 2mm (0.02 to 0.08in) (light contact for elastomer type) from the bearing carrier and bearing housing respectively. The pump side flinger (this feature is integral with some proprietary labyrinth seals) should only be set in its final position after setting the shaft axial position. Temporarily fit the seal housing (with any internal anti-vortex rib at the topmost position). The shaft may now be positioned in relation to the seal housing face, as shown below.

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INSTRUCTION MANUAL CPXP ENGLISH

•

Bearing housing

Dia. X mm (in)

Frame 1

24 (0.945)

9 (0.354)

Frame 2

32 (1.260)

17 (0.669)

Frame 3

42 (1.654)

9 (0.354)

• •

Z mm (in)

C961KH001 - 01/03

Assemble impeller onto the shaft. Tighten the impeller. Use the same method as in disassembly but rotating in opposite direction. A few sharp strikes will tighten it to the correct level.

13.3.7 Assembly of bearing housing into casing • Fit a new gasket into the casing. • Install the bearing housing assembly into the pump casing. Coat the screws with anti-galling compound and tighten into casing. • Check impeller front clearance against original setting, or process requirement and adjust as necessary. See section on Preventative maintenance and servicing. • Ensure that all other items have been re-attached and all fasteners tightened, then follow instruction in the Installation sections of this manual.

The pump side flinger may then be moved towards the bearing housing and set with its clearance. 0.5 to 2mm (0.02 – 0.08”)

13.3.4 Seal housing and seal assembly • Extreme cleanliness is required. The sealing faces and shaft or sleeve surface must be free from scratches or other damage. • Refer to the Seal arrangement section for seal diagrams. • Carefully press the stationary seat into the mechanical seal housing or cover, ensuring that the seating ring is not deformed. Where an antirotation pin is fitted ensure that correct engagement with the slot is achieved. • Place any separate seal covers over the shaft. • Refer to manufacturer's instructions to position the mechanical seal rotating elements. Tighten any drive screws in the seal drive collar. For precise compression most cartridge seals should be set after complete pump assembly. • To set, or reset, a cartridge mechanical seal having a PTFE 'setting ring - throttle' and no separate setting clips, finger tighten the seal cover stud nuts, then fully torque up the sleeve screws. The seal cover stud nuts can then be fully torqued up. • Fit the seal housing into the bearing housing and tighten all fasteners. 13.3.5 Gland packed stuffing box assembly • Assemble the gland packing into the stuffing box housing before fitting on to the shaft. • Stagger the joints in the gland packing by 90 degrees to each other. • The lantern ring halves (if required) should be positioned mid-way along the packing. • Position the gland squarely against the last ring and tighten the gland nuts finger-tight only. Install into bearing housing assembly and fit the two screws to hold the seal housing in place. • Check that the shaft rotates freely. 13.3.6 Impeller assembly and setting • Fit a new O-ring into the impellers using a small amount of grease to hold it in place. Apply antigalling compound (which does not contain copper) to the impeller thread to help subsequent removal. 18 ®

INSTRUCTION MANUAL CPXP ENGLISH

C961KH001 - 01/03

14.1.4 Single seal with external lip seal

14 SEALING ARRANGEMENTS This Section shows details of the seal arrangements. Contact your nearest Flowserve sales office or service centre if you require further information or are unsure of the specific arrangement supplied. Refer also to section on auxiliary piping in this manual. 14.1 Single seal types 14.1.1 Single stepped balanced seal

Q - Rp 1/4" quench D - Rp 1/4" drain F - Rp 1/4" flush Z - Position of lip seal hard sleeve NB: Lever flange away after fitting hard sleeve to shaft. Setting dimension X Y 23.5 11.0 34.0 19.0 33.5 11.0

Bearing housing Note: “L1K” and “L1N” are seal lengths defined within seal standard DIN 24960.

Frame 1 Frame 2 Frame 3

14.1.2 Single unbalanced (or inherently balanced) seal

Pump size 125 160 200 250 315

Frame 1 41.5 41.5 36.5 -

Setting dimension Z Frame 2 Frame 3 49.0 49.0 44.0 45.0 44.0 45.0

14.1.5 Single internal seal with internal and external neck bush Q - Rp 1/4" quench D - Rp 1/4" drain F - Rp 1/4" flush

VARIANTS Self setting collar. Separate seal drive collar, set to Dimension ‘X’. Integral seal drive collar with screws, set to Dimension ‘X’. Bearing housing Frame 1 Frame 2 Frame 3

Setting dimension X Y 23.5 11.0 34.0 19.0 33.5 11.0

14.1.3 Single seal with external neck bush

Pump size

Q - Rp 1/4" quench D - Rp 1/4" drain F - Rp 1/4" flush Bearing housing Frame 1 Frame 2 Frame 3

Setting dimension X 23.5 34.0 33.5

Y 11.0 19.0 11.0

125

Frame 1 X Y 12.5 0

160 200 250 315

12.5 17.5 -

0 5.0 -

Setting dimensions Frame 2 Frame 3 X Y X Y 5.5 5.5 10.6 10.6

-9.5 -9.5 -4.4 -4.4

18.3 18.3

-4.3 -4.3

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INSTRUCTION MANUAL CPXP ENGLISH

C961KH001 - 01/03

14.2 Cartridge seal types

14.4 Double seal types

14.2.1 Cartridge seal in conical seal housing

14.4.1 Double back-to-back seal with EPA circulation BI - Rp 1/4" barrier liquid inlet BO - Rp 1/4" barrier liquid outlet

14.2.2 DIN 24960 "C" cartridge seal Setting dimension X

Pump size Frame 1

Frame 2

Frame 3

125

11.0

-

-

160

11.0

17.5

-

200

6.0

17.5

-

250

-

12.4

14.4

315

-

12.4

14.3

14.5 External seal types

For 'S' see seal supplier's instructions.

14.5.1 External seal

14.3 Tandem seal types 14.3.1 Tandem seal with EPA circulation

D - Drain

14.6 Packed gland seal types 14.6.1 Packed gland with modern fibre packing

BI - Rp 1/4" barrier liquid inlet. BO - Rp 1/4" barrier liquid outlet. F - Rp 1/4" flush. Pump size

125 160 200 250 315

Setting dimensions Frame 1 Frame 2 Frame 3 X Y X Y X Y 20.0 31.5 20.0 31.5 28.0 41.5 20.0 26.5 28.0 41.5 28.0 36.4 27.5 33.7 28.0 36.4 27.5 33.7

F - Rp 1/4" flush

20 ®

INSTRUCTION MANUAL CPXP ENGLISH 15 SPARE PARTS

C961KH001 - 01/03

16 GENERAL ARRANGEMENT DRAWING The typical general arrangement drawing and any specific drawings required by the Contract will be sent to the Purchaser separately. If required these should be obtained from the Purchaser and retained with this manual.

15.1 Ordering of spares Flowserve keeps records of all pumps that have been supplied. When ordering spares the following information should be quoted: 1. Pump serial number. 2. Pump size. 3. Part name. 4. Part number. 5. Number of parts required.

17 CERTIFICATION Any certificates eg materials, hydraulic tests, conformities, Ex protection for an explosive atmosphere, performance test curves etc as determined by the contract requirements, will be sent to the Purchaser separately. If required, copies of these should be obtained from the Purchaser for retention with this manual.

15.2 The pump size and serial number are shown on the pump nameplate. 15.3 To ensure continued satisfactory operation, replacement parts to the original design specification should be obtained from Flowserve. Any change to the original design specification (modification or use of a non-standard part) will invalidate the pump's safety certification.

18 SUPPLEMENTARY INSTRUCTION MANUALS See also the supplementary instruction manuals supplied with this manual eg for electric motors, controllers, engines, gearboxes, sealant systems etc. 19 CHANGE NOTES Change notes and errata (if any) will be included on a separate page(s) within the manual. If changes are made to the pump after supply, this manual will require updating.

15.4 Storage of spares Spares should be stored in a clean dry area away from vibration. Inspection and retreatment of metallic surfaces (if necessary) with preservative is recommended at 6 monthly intervals.

15.5 Recommended spares for two years operation (as per VDMA 24296) Part no.

Designation

2200 2100 3712 2450 3042 3041 4590/1 4610/1 4610/10 2540/2

Impeller Shaft Bearing nut Shaft sleeve Bearing - pump side Bearing - drive side Pump casing gasket O-ring - impeller O-ring - carrier Pump side liquid flinger

4130 4120 4200 -

Gland Packing - set Gland halves Mechanical seals Power end

Number of pumps (including stand-by) 2

3

4

5

1 1 1

2 2

3

2

3

1 1 4 4 4 1

2 2 8 8 8

6 6 6

3 3 9 9 9

2 2

1 1 -

6/7 2

-

10(+)

3 3 4 4 4 4 12 12 10

30% 30% 50% 50% 50% 50% 150% 150% 100% 30%

4

40% 30% 30% 2

3 3

2 2 -

8/9

3 3 -

-

1

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INSTRUCTION MANUAL CPXP ENGLISH

C961KH001 - 01/03

20 OPERATING DIFICULTIES SYMPTOMS PUMP OVERHEATS AND SEIZES BEARINGS HAVE SHORT LIFE ⇓ PUMP VIBRATES OR IS NOISY ⇓ MECHANICAL SEAL HAS SHORT LIFE ⇓ MECHANICAL SEAL LEAKS EXCESSIVELY ⇓ PUMP REQUIRES EXCESSIVE POWER ⇓ PUMP DOES NOT SELF PRIME ⇓ INSUFFICIENT PRESSURE DEVELOPED ⇓ INSUFFICIENT CAPACITY DELIVERED ⇓ PUMP DOES NOT DELIVER LIQUID ⇓ ⇓ SUCTION TROUBLES Pump not primed or filled with liquid. Suction lift too high. Insufficient margin between suction pressure and vapour pressure. Excessive amount of foaming or gas in liquid. Air or vapour pocket in suction line. Air leaks into suction line. Air leaks into pump through mechanical seal, sleeve joints, casing joint or pipe lugs. Suction pipe blocked. Inlet of suction pipe insufficiently submerged. SYSTEM TROUBLES Speed too low. Speed too high. Total head of system higher than head of pump. Total head of system lower than pump design head.

l l l l l

l l l

l l

l l l l l l l l

l

l

l

l

l

l

l l l

l

l l l l l

l l l

l

l

Specific gravity of liquid different from design. Viscosity of liquid differs from that for which designed. Operation at very low capacity. Operation at high capacity.

l

l l l

l

l l

l

MECHANICAL TROUBLES Misalignment due to pipe strain. Improperly designed foundation. Shaft bent. Rotating part rubbing on stationary part internally. Bearings worn. Wearing ring surfaces worn. Impeller damaged or eroded. Leakage under sleeve due to joint failure. Shaft sleeve worn or scored or running off centre. Mechanical seal improperly installed. Incorrect type of mechanical seal for operating conditions. Shaft running off centre because of worn bearings or misalignment. Impeller out of balance resulting in vibration. Abrasive solids in liquid pumped. Internal misalignment of parts preventing seal ring and seat from mating properly. Mechanical seal was run dry. Internal misalignment due to improper repairs causing impeller to rub. Excessive thrust caused by mechanical failure inside the pump. Excessive grease in ball bearings. Lack of lubrication for bearings.

l l

l l

l l

l

l l l l l

l

l l l

l

l l l l l l l

l l l l l l l l

l l l l

l l l l l l

l l l l l

Improper installation of bearings (damage during assembly, incorrect assembly, wrong type of bearing etc). Damaged bearings due to contamination. MOTOR ELECTRICAL PROBLEMS Wrong direction of rotation. Motor running on 2 phases only. Motor running too slow, check terminal box.

l

l l l l l

l l

l

l l

l

l l

l

l l

l l l l

l l l l l

l

l

l l

l

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INSTRUCTION MANUAL CPXP ENGLISH

C961KH001 - 01/03

21 SECTIONAL ARRANGEMENT DRAWINGS AND PARTS LISTS 21.1 CPXP pump

Europump No

Description

1111

Pump casing

2100

Shaft

2200

Impeller

2450

Shaft sleeve (if fitted)

2510

Seal setting collar (L1K)

2540/1

Flinger (liquid) drive side

2540/2

Flinger (liquid) pump side

3041

Bearing (drive side) ang. cont.

3042

Bearing (pump side) ball

3130 3134

Bearing housing Support foot

3240

Bearing carrier

3712 3854

Bearing nut Oil filler plug

3855

Constant level oiler (optional)

3858

Sight glass

4200

Mechanical seal

4210 4330

Mechanical seal housing Labyrinth seal (pump side)

4590

Pump casing gasket

4610/1

D-ring

4610/2

D-ring

6511

Filler plug

6515/1

Drain plug

6515/2

Drain plug (magnetic)

6546

Inner circlip

6742

Coupling key

9906/01

Hex screw

9906/02

Hex screw

9906/03

Hex screw

9906/04

Hex screw

9923/1

Hex nut

9923/2 9951

Hex nut Stud

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INSTRUCTION MANUAL CPXP ENGLISH

C961KH001 - 01/03

20.2 CPXP pump (diffuser-casing sizes)

Europump No

Description

1111

Pump casing

1410

Diffuser

4610/3

D-ring

4590/1

Pump casing gasket

4590/2 9906/1

Diffuser gasket Hex screw

9906/5

Hex screw

6515

Drain plug

24 ®

INSTRUCTION MANUAL CPXP ENGLISH

C961KH001 - 01/03

21 CPXP PARTS INTERCHANGEABILITY

25 ®

INSTRUCTION MANUAL CPXP ENGLISH

C961KH001 - 01/03

Notes:

26 ®

INSTRUCTION MANUAL CPXP ENGLISH

C961KH001 - 01/03

Notes:

27 ®

INSTRUCTION MANUAL CPXP ENGLISH

C961KH001 - 01/03

Europe, Middle East & Africa Flowserve Limited (Pump Division) Harley House, 94 Hare Lane Claygate, Esher, Surrey KT10 0RB United Kingdom

Latin America Flowserve S.A. de C.V. Avenida Paseo de la Reforma #30 2nd Floor, Colonia Juarez Centro Mexico, D.F.Z.C. 06040

Tel +44 (0)1372 463 700 Fax +44 (0)1372 460 190

Tel +52 5705 5526 Fax +52 5705 1125

USA and Canada Flowserve Corporation (Pump Division) Millennium Center, 222 Las Colinas Blvd. 15th Floor, Irving, TX 75039-5421, USA

Asia Pacific Flowserve Pte Ltd (Pump Division) 200 Pandan Loop, #06-03/04 Pantech 21, Singapore 128388

Tel +1 972 443 6500 Toll free 800 728 PUMP (7867) Fax +1 972 443 6800

Tel +65 775 3003 Fax +65 779 4607

Visit our web site at: www.flowserve.com Your Flowserve factory contact:

Your local Flowserve representative:

Flowserve Pumps Limited PO Box 17, Newark Notts NG24 3EN United Kingdom Telephone (24 hours) +44 (0)1636 494 600 Sales & Admin Fax +44 (0)1636 705 991 Repair & Service Fax +44 (0)1636 494 833 E.mail [email protected]

To find your local Flowserve representative, please use the Sales Support Locator System found at www.flowserve.com

28 ®

®

Pump Division

Types:

LNN, LNNV and LNNC

CENTRIFUGAL PUMPS

USER INSTRUCTIONS: INSTALLATION, OPERATION, MAINTENANCE User Instructions C953KH001 - 09/03 (E)

LNN, LNNV and LNNC USER INSTRUCTIONS ENGLISH C953KH001 – 09/03 ®

CONTENTS

PAGE

1 INTRODUCTION AND SAFETY............................3 1.1 General............................................................3 1.2 CE marking and approvals ..............................3 1.3 Disclaimer........................................................3 1.4 Copyright .........................................................3 1.5 Duty conditions ................................................3 1.6 Safety ..............................................................4 1.7 Nameplate and warning labels ........................7 1.8 Specific machine performance ........................8 1.9 Noise level .......................................................8 2 TRANSPORT AND STORAGE..............................9 2.1 Consignment receipt and unpacking ...............9 2.2 Handling ..........................................................9 2.3 Lifting ...............................................................9 2.4 Storage ............................................................9 2.5 Recycling and end of product life ....................9 3 PUMP DESCRIPTION .........................................10 3.1 Configurations ...............................................10 3.2 Name nomenclature ......................................10 3.3 Design of major parts ....................................10 3.4 Performance and operating limits..................11 4 INSTALLATION ....................................................11 4.1 Location .........................................................11 4.2 Part assemblies .............................................11 4.3 Foundation.....................................................11 4.4 Grouting.........................................................12 4.5 Initial alignment..............................................12 4.6 Piping.............................................................13 4.7 Final shaft alignment check ...........................15 4.8 Electrical connections....................................15 4.9 Protection systems ........................................16 5 COMMISSIONING, START-UP, OPERATION AND SHUTDOWN............................................16 5.1 Pre-commissioning procedure.......................16 5.2 Pump lubricants.............................................17 5.3 Direction of rotation .......................................18 5.4 Guarding........................................................18 5.5 Priming and auxiliary supplies.......................18 5.6 Starting the pump ..........................................18 5.7 Running the pump .........................................19 5.8 Stopping and shutdown .................................20 5.9 Hydraulic, mechanical and electrical duty .....20 6 MAINTENANCE ...................................................20 6.1 General..........................................................20 6.2 Maintenance schedule ..................................21 6.3 Spare parts ....................................................23 6.4 Recommended spares and consumable items...................................23 6.5 Tools required ................................................23 6.6 Fastener torques ...........................................24 6.7 Renewal clearances ......................................24 6.8 Disassembly ..................................................24 6.9 Examination of parts......................................25 6.10 Assembly .....................................................26

PAGE 7 FAULTS; CAUSES AND REMEDIES .................. 29 8 PARTS LIST AND DRAWINGS ........................... 31 8.1 Sectional drawings - LNN and LNNC ........... 31 8.2 Parts list - LNN and LNNC ............................ 32 8.3 Sectional drawings - LNNV sleeve bearing type................................................... 33 8.4 Parts list - LNNV............................................ 34 8.5 General arrangement drawing ...................... 35 9 CERTIFICATION ................................................. 35 10 OTHER RELEVANT DOCUMENTATION AND MANUALS........................................................ 35 10.1 Supplementary User Instructions................ 35 10.2 Change notes.............................................. 35 10.3 Additional sources of information ................ 35

INDEX

PAGE

Alignment of shafting (see 4.5, 4.7 and 4.3) CE marking and approvals (1.2)................................ 3 Clearances (see 6.7, Renewal clearances)............. 24 Commissioning and operation (see 5)..................... 16 Configurations (3.1) ................................................. 10 Direction of rotation (5.3) ......................................... 18 Dismantling (see 6.8, Disassembly) ........................ 24 Duty conditions (1.5).................................................. 3 Electrical connections (4.8) ..................................... 15 Examination of parts (6.9) ....................................... 25 Faults; causes and remedies................................... 29 General assembly drawings (see 8) ........................ 31 Grouting (4.4)........................................................... 12 Guarding (5.4).......................................................... 18 Handling (2.2) ............................................................ 9 Hydraulic, mechanical and electrical duty (5.9) ....... 20 Lifting (2.3) ................................................................. 9 Location (4.1)........................................................... 11 Lubrication schedule (see 5.2, Pump lubricants) .... 17 Maintenance schedule (6.2) .................................... 21 Piping (4.6) .............................................................. 13 Priming and auxiliary supplies (5.5)......................... 18 Reassembly (see 6.10, Assembly) .......................... 26 Replacement parts (see 6.3 and 6.4) ...................... 23 Safety, protection systems (see 1.6 and 4.9) Sound level (see 1.9, Noise level) ............................. 8 Specific machine performance (1.8).......................... 8 Starting the pump (5.6) ............................................ 18 Stopping and shutdown (5.8)................................... 20 Storage (2.4).............................................................. 9 Supplementary manuals or information sources ..... 35 Tools required (6.5) ................................................. 23 Torques for fasteners (6.6) ...................................... 24

Page 2 of 36

LNN, LNNV and LNNC USER INSTRUCTIONS ENGLISH C953KH001 – 09/03 ®

1 INTRODUCTION AND SAFETY

1.3 Disclaimer

1.1 General

Information in these User Instructions is believed to be reliable. In spite of all the efforts of Flowserve Pump Division to provide sound and all necessary information the content of this manual may appear insufficient and is not guaranteed by Flowserve as to its completeness or accuracy.

These instructions must always be kept close to the product’s operating location or directly with the product. Flowserve’s products are designed, developed and manufactured with state-of-the-art technologies in modern facilities. The unit is produced with great care and commitment to continuous quality control, utilising sophisticated quality techniques, and safety requirements. Flowserve is committed to continuous quality improvement and being at service for any further information about the product in its installation and operation or about its support products, repair and diagnostic services. These instructions are intended to facilitate familiarization with the product and its permitted use. Operating the product in compliance with these instructions is important to help ensure reliability in service and avoid risks. The instructions may not take into account local regulations; ensure such regulations are observed by all, including those installing the product. Always coordinate repair activity with operations personnel, and follow all plant safety requirements and applicable safety and health laws and regulations. These instructions should be read prior to installing, operating, using and maintaining the equipment in any region worldwide. The equipment must not be put into service until all the conditions relating to safety noted in the instructions, have been met.

1.2 CE marking and approvals

It is a legal requirement that machinery and equipment put into service within certain regions of the world shall conform with the applicable CE Marking Directives covering Machinery and, where applicable, Low Voltage Equipment, Electromagnetic Compatibility (EMC), Pressure Equipment Directive (PED) and Equipment for Potentially Explosive Atmospheres (ATEX). Where applicable the Directives and any additional Approvals cover important safety aspects relating to machinery and equipment and the satisfactory provision of technical documents and safety instructions. Where applicable this document incorporates information relevant to these Directives and Approvals. To confirm the Approvals applying and if the product is CE marked, check the serial number plate markings and the Certification. (See section 9, Certification.)

Flowserve manufactures products to exacting International Quality Management System Standards as certified and audited by external Quality Assurance organisations. Genuine parts and accessories have been designed, tested and incorporated into the products to help ensure their continued product quality and performance in use. As Flowserve cannot test parts and accessories sourced from other vendors the incorrect incorporation of such parts and accessories may adversely affect the performance and safety features of the products. The failure to properly select, install or use authorised Flowserve parts and accessories is considered to be misuse. Damage or failure caused by misuse is not covered by Flowserve’s warranty. In addition, any modification of Flowserve products or removal of original components may impair the safety of these products in their use.

1.4 Copyright

All rights reserved. No part of these instructions may be reproduced, stored in a retrieval system or transmitted in any form or by any means without prior permission of Flowserve Pump Division.

1.5 Duty conditions

This product has been selected to meet the specifications of your purchaser order. The acknowledgement of these conditions has been sent separately to the Purchaser. A copy should be kept with these instructions. The product must not be operated beyond the parameters specified for the application. If there is any doubt as to the suitability of the product for the application intended, contact Flowserve for advice, quoting the serial number. If the conditions of service on your purchase order are going to be changed (for example liquid pumped, temperature or duty) it is requested that the user seeks Flowserve’s written agreement before start up.

Page 3 of 36

LNN, LNNV and LNNC USER INSTRUCTIONS ENGLISH C953KH001 – 09/03 ®

1.6 Safety 1.6.1 Summary of safety markings These User Instructions contain specific safety markings where non-observance of an instruction would cause hazards. The specific safety markings are: This symbol indicates electrical safety instructions where non-compliance will involve a high risk to personal safety or the loss of life. This symbol indicates safety instructions where non-compliance would affect personal safety and could result in loss of life. This symbol indicates “hazardous and toxic fluid” safety instructions where non-compliance would affect personal safety and could result in loss of life. This symbol indicates safety instructions where non-compliance will involve some risk to safe operation and personal safety and would damage the equipment or property. This symbol indicates explosive atmosphere zone marking according to ATEX. It is used in safety instructions where non-compliance in the hazardous area would cause the risk of an explosion. This sign is not a safety symbol but indicates an important instruction in the assembly process. 1.6.2 Personnel qualification and training All personnel involved in the operation, installation, inspection and maintenance of the unit must be qualified to carry out the work involved. If the personnel in question do not already possess the necessary knowledge and skill, appropriate training and instruction must be provided. If required the operator may commission the manufacturer/supplier to provide applicable training. Always coordinate repair activity with operations and health and safety personnel, and follow all plant safety requirements and applicable safety and health laws and regulations. 1.6.3 Safety action This is a summary of conditions and actions to prevent injury to personnel and damage to the environment and to equipment. For products used in potentially explosive atmospheres section 1.6.4 also applies.

NEVER DO MAINTENANCE WORK WHEN THE UNIT IS CONNECTED TO POWER GUARDS MUST NOT BE REMOVED WHILE THE PUMP IS OPERATIONAL DRAIN THE PUMP AND ISOLATE PIPEWORK BEFORE DISMANTLING THE PUMP The appropriate safety precautions should be taken where the pumped liquids are hazardous. FLUORO-ELASTOMERS (When fitted.) When a pump has experienced temperatures over 250 ºC (482 ºF), partial decomposition of fluoroelastomers (example: Viton) will occur. In this condition these are extremely dangerous and skin contact must be avoided. HANDLING COMPONENTS Many precision parts have sharp corners and the wearing of appropriate safety gloves and equipment is required when handling these components. To lift heavy pieces above 25 kg (55 lb) use a crane appropriate for the mass and in accordance with current local regulations. THERMAL SHOCK Rapid changes in the temperature of the liquid within the pump can cause thermal shock, which can result in damage or breakage of components and should be avoided. NEVER APPLY HEAT TO REMOVE IMPELLER Trapped lubricant or vapour could cause an explosion. HOT (and cold) PARTS If hot or freezing components or auxiliary heating supplies can present a danger to operators and persons entering the immediate area action must be taken to avoid accidental contact. If complete protection is not possible, the machine access must be limited to maintenance staff only, with clear visual warnings and indicators to those entering the immediate area. Note: bearing housings must not be insulated and drive motors and bearings may be hot. If the temperature is greater than 68 °C (175 °F) or below 5 °C (20 °F) in a restricted zone, or exceeds local regulations, action as above shall be taken. HAZARDOUS LIQUIDS When the pump is handling hazardous liquids care must be taken to avoid exposure to the liquid by appropriate siting of the pump, limiting personnel access and by operator training. If the liquid is flammable and/or explosive, strict safety procedures must be applied. Gland packing must not be used when pumping hazardous liquids.

Page 4 of 36

LNN, LNNV and LNNC USER INSTRUCTIONS ENGLISH C953KH001 – 09/03 ®

PREVENT EXCESSIVE EXTERNAL

1.6.4.1 Scope of compliance

PIPE LOAD Do not use pump as a support for piping. Do not mount expansion joints, unless allowed by Flowserve in writing, so that their force, due to internal pressure, acts on the pump flange.

Use equipment only in the zone for which it is appropriate. Always check that the driver, drive coupling assembly, seal and pump equipment are suitably rated and/or certified for the classification of the specific atmosphere in which they are to be installed.

ENSURE CORRECT LUBRICATION (See section 5, Commissioning, startup, operation and shutdown.)

Where Flowserve has supplied only the bare shaft pump, the Ex rating applies only to the pump. The party responsible for assembling the pump set shall select the coupling, driver and any additional equipment, with the necessary CE Certificate/ Declaration of Conformity establishing it is suitable for the area in which it is to be installed.

START THE PUMP WITH OUTLET VALVE PARTLY OPENED (Unless otherwise instructed at a specific point in the User Instructions.) This is recommended to minimize the risk of overloading and damaging the pump motor at full or zero flow. Pumps may be started with the valve further open only on installations where this situation cannot occur. The pump outlet control valve may need to be adjusted to comply with the duty following the run-up process. (See section 5, Commissioning start-up, operation and shutdown.) NEVER RUN THE PUMP DRY INLET VALVES TO BE FULLY OPEN WHEN PUMP IS RUNNING Running the pump at zero flow or below the recommended minimum flow continuously will cause damage to the seal. DO NOT RUN THE PUMP AT ABNORMALLY HIGH OR LOW FLOW RATES Operating at a flow rate higher than normal or at a flow rate with no back pressure on the pump may overload the motor and cause cavitation. Low flow rates may cause a reduction in pump/bearing life, overheating of the pump, instability and cavitation/ vibration. 1.6.4 Products used in potentially explosive atmospheres • • • • •

The output from a variable frequency drive (VFD) can cause additional heating affects in the motor and so, for pumps sets with a VFD, the ATEX Certification for the motor must state that it is covers the situation where electrical supply is from the VFD. This particular requirement still applies even if the VFD is in a safe area. 1.6.4.2 Marking An example of ATEX equipment marking is shown below. The actual classification of the pump will be engraved on the nameplate. II 2 GD c 135 ºC (T4) Equipment Group I = Mining II = Non-mining Category 2 or M2 = High level protection 3 = normal level of protection Gas and/or Dust G = Gas; D= Dust c = Constructional safety (in accordance with prEn13463-5) Maximum surface temperature (Temperature Class) (See section 1.6.4.3.)

Measures are required to: Avoid excess temperature Prevent build up of explosive mixtures Prevent the generation of sparks Prevent leakages Maintain the pump to avoid hazard

1.6.4.3 Avoiding excessive surface temperatures ENSURE THE EQUIPMENT TEMPERATURE CLASS IS SUITABLE FOR THE HAZARD ZONE

The following instructions for pumps and pump units when installed in potentially explosive atmospheres must be followed to help ensure explosion protection. Both electrical and non-electrical equipment must meet the requirements of European Directive 94/9/EC.

Pumps have a temperature class as stated in the ATEX Ex rating on the nameplate. These are based on a maximum ambient of 40 °C (104 °F); refer to Flowserve for higher ambient temperatures. The surface temperature on the pump is influenced by the temperature of the liquid handled. The maximum permissible liquid temperature depends on the temperature class and must not exceed the values in the table that follows.

Page 5 of 36

LNN, LNNV and LNNC USER INSTRUCTIONS ENGLISH C953KH001 – 09/03 ®

The temperature rise at the seals and bearings and due to the minimum permitted flow rate is taken into account in the temperatures stated. Temperature class to prEN 13463-1 T6 T5 T4 T3 T2 T1

Maximum surface temperature permitted 85 °C (185 °F) 100 °C (212 °F) 135 °C (275 °F) 200 °C (392 °F) 300 °C (572 °F) 450 °C (842 °F)

Temperature limit of liquid handled (* depending on material and construction variant - check which is lower) Consult Flowserve Consult Flowserve 115 °C (239 °F) * 180 °C (356 °F) * 275 °C (527 °F) * 400 °C (752 °F) *

The responsibility for compliance with the specified maximum liquid temperature is with the plant operator. If an explosive atmosphere exists during the installation, do not attempt to check the direction of rotation by starting the pump unfilled. Even a short run time may give a high temperature resulting from contact between rotating and stationary components. Where there is any risk of the pump being run against a closed valve generating high liquid and casing external surface temperatures it is recommended that users fit an external surface temperature protection device. Avoid mechanical, hydraulic or electrical overload by using motor overload trips, temperature monitor or a power monitor and make routine vibration monitoring checks. In dirty or dusty environments, regular checks must be made and dirt removed from areas around close clearances, bearing housings and motors. 1.6.4.4 Preventing the build up of explosive mixtures ENSURE THE PUMP IS PROPERLY FILLED AND VENTED AND DOES NOT RUN DRY Ensure the pump and relevant suction and discharge pipeline system is totally filled with liquid at all times during the pump operation, so that an explosive atmosphere is prevented. In addition it is essential to make sure that seal chambers, auxiliary shaft seal systems and any heating and cooling systems are properly filled. If the operation of the system cannot avoid this condition the fitting of an appropriate dry run protection device is recommended (eg liquid detection or a power monitor). To avoid potential hazards from fugitive emissions of vapour or gas to atmosphere the surrounding area must be well ventilated.

1.6.4.5 Preventing sparks To prevent a potential hazard from mechanical contact, the coupling guard must be non-sparking and anti-static for Category 2. To avoid the potential hazard from random induced current generating a spark, the earth contact on the baseplate must be used. Avoid electrostatic charge: do not rub non-metallic surfaces with a dry cloth; ensure cloth is damp. The coupling must be selected to comply with 94/9/EC and correct alignment must be maintained. 1.6.4.6 Preventing leakage The pump must only be used to handle liquids for which it has been approved to have the correct corrosion resistance. Avoid entrapment of liquid in the pump and associated piping due to closing of suction and discharge valves, which could cause dangerous excessive pressures to occur if there is heat input to the liquid. This can occur if the pump is stationary or running. Bursting of liquid containing parts due to freezing must be avoided by draining or protecting the pump and ancillary systems. Where there is the potential hazard of a loss of a seal barrier fluid or external flush, the fluid must be monitored. If leakage of liquid to atmosphere can result in a hazard, the installation of a liquid detection device is recommended. 1.6.4.7 Maintenance to avoid the hazard CORRECT MAINTENANCE IS REQUIRED TO AVOID POTENTIAL HAZARDS WHICH GIVE A RISK OF EXPLOSION The responsibility for compliance with maintenance instructions is with the plant operator. To avoid potential explosion hazards during maintenance, the tools, cleaning and painting materials used must not give rise to sparking or adversely affect the ambient conditions. Where there is a risk from such tools or materials, maintenance must be conducted in a safe area. It is recommended that a maintenance plan and schedule is adopted. (See section 6, Maintenance.)

Page 6 of 36

LNN, LNNV and LNNC USER INSTRUCTIONS ENGLISH C953KH001 – 09/03 ®

1.7 Nameplate and warning labels 1.7.1 Nameplate For details of nameplate, see the Declaration of Conformity. 1.7.2 Warning labels

Oil lubricated units only:

Page 7 of 36

LNN, LNNV and LNNC USER INSTRUCTIONS ENGLISH C953KH001 – 09/03 ®

1.8 Specific machine performance

For performance parameters see section 1.5, Duty conditions. When the contract requirement specifies these to be incorporated into User Instructions these are included here. Where performance data has been supplied separately to the purchaser these should be obtained and retained with these User Instructions if required.

1.9 Noise level

When pump noise level exceeds 85 dBA attention must be given to prevailing Health and Safety Legislation, to limit the exposure of plant operating personnel to the noise. The usual approach is to control exposure time to the noise or to enclose the machine to reduce emitted sound. You may have already specified a limiting noise level when the equipment was ordered, however if no noise requirements were defined then machines above a certain power level will exceed 85 dBA. In such situations consideration must be given to the fitting of an acoustic enclosure to meet local regulations.

Pump noise level is dependent on a number of factors the type of motor fitted, the operating capacity, pipework design and acoustic characteristics of the building. Typical sound pressure levels measured in dB, and A-weighted are shown in the table below (LpfA). The figures are indicative only, they are subject to a +3 dB tolerance, and cannot be guaranteed. The values are based on the noisiest ungeared electric motors that are likely to be encountered. They represent sound pressure levels at 1 m (3.3 ft) from the directly driven pump, for "free field over a reflecting plane". If a pump unit only has been purchased, for fitting with your own driver, then the "pump only" noise levels from the table should be combined with the level for the driver obtained from the supplier. If the motor is driven by an inverter, it may show an increase in noise level at some speeds. Consult a Noise Specialist for the combined calculation.

For units driven by equipment other than electric motors or units contained within enclosures, see the accompanying information sheets and manuals.

Typical sound pressure level, LpfA – (dB, A-weighted) Motor size and speed

3550 r/min Pump and Pump motor only LpfA LpfA

2900 r/min Pump and Pump motor only LpfA LpfA

1750 r/min Pump and Pump motor only LpfA LpfA

kW

(hp)

<0.55 0.75 1.1 1.5 2.2 3 4 5.5 7.5 11 15 18.5 22 30 37 45 55 75 90 110 150 200

(<0.75) (1) (1.5) (2) (3) (4) (5) (7.5) (10) (15) (20) (25) (30) (40) (50) (60) (75) (100) (120) (150) (200) (270)

71 74 74 77 78 81 82 90 90 91 92 92 92 100 100 100 102 100 97 100 101

66 66 68 70 72 74 75 77 78 80 83 83 83 85 86 87 88 90 90 91 92

64 67 67 70 71 74 75 83 83 84 85 85 85 93 93 93 95 95 92 95 96

62 62 64 66 68 70 71 73 74 76 79 79 79 81 82 83 84 86 86 87 88

64 67 67 70 71 74 75 76 77 78 80 80 81 84 84 84 86 88 90 91 91

①

①

①

①

300

(400)

-

-

-

-

500

(670)

-

-

-

-

1000

(1300)

-

-

-

-

1500

(2000)

-

-

-

-

① ① ① ① ①

62 62 64 66 68 74 75 75 76 77 79 79 79 80 80 80 81 81 81 83 83 83 84 85 88 90

1450 r/min Pump and Pump motor only LpfA LpfA 63 63 65 66 68 70 71 72 73 74 76 76 77 80 80 80 82 83 85 86 86

① ① ① ① ①

① Motors in this range are generally job specific and noise levels should be calculated based on actual equipment installed. For 960 rpm reduce 1450 rpm values by 5 dBA.

Page 8 of 36

62 62 64 66 68 70 71 71 72 73 75 75 75 76 76 76 77 78 78 79 79 80 81 83 86 88

LNN, LNNV and LNNC USER INSTRUCTIONS ENGLISH C953KH001 – 09/03 ®

2 TRANSPORT AND STORAGE 2.1 Consignment receipt and unpacking

Immediately after receipt of the equipment it must be checked against the delivery and shipping documents for its completeness and that there has been no damage in transportation. Any shortage and or damage must be reported immediately to Flowserve Pump Division and received in writing within one month of receipt of the equipment. Later claims cannot be accepted. Check any crate, boxes and wrappings for any accessories or spare parts that may be packed separately with the equipment or attached to side walls of the box or equipment. Each product has a unique serial number. Check that this number corresponds with that advised and always quote this number in correspondence as well as when ordering spare parts or further accessories.

2.2 Handling

Boxes, crates, pallets or cartons may be unloaded using fork-lift vehicles or slings dependent on their size and construction.

2.3 Lifting To avoid distortion, the pump unit should be lifted as shown.

A crane must be used for all pump sets in excess of 25 kg (55 lb). Fully trained personnel must carry out lifting, in accordance with local regulations. The driver and pump weights are recorded on their respective nameplates or massplates.

2.4 Storage Store the pump in a clean, dry location away from vibration. Leave piping connection covers in place to keep dirt and other foreign material out of pump casing. Turn pump at intervals to prevent brinelling of the bearings and the seal faces, if fitted, from sticking. The pump may be stored as above for up to 6 months. Consult Flowserve for preservative actions when a longer storage period is needed.

2.5 Recycling and end of product life

At the end of the service life of the product or its parts, the relevant materials and parts should be recycled or disposed of using an environmentally acceptable method and with local regulations. If the product contains substances that are harmful to the environment, these should be removed and disposed of in accordance with current regulations. This also includes the liquids and or gases that may be used in the "seal system" or other utilities. Make sure that hazardous substances are disposed of safely and that the correct personal protective equipment is used. The safety specifications must be in accordance with the current regulations at all times.

Page 9 of 36

LNN, LNNV and LNNC USER INSTRUCTIONS ENGLISH C953KH001 – 09/03 ®

3 PUMP DESCRIPTION

3.3 Design of major parts

3.1 Configurations

3.3.1 Pump casing The pump has its main casing gasket axial to the shaft allowing maintenance to the rotating element by removing the top half casing. Suction and discharge branches are in the bottom half and therefore remain undisturbed.

The LNN can have the following configurations:

3.3.2 Impeller The impeller is fully shrouded and may be fitted with optional hub rings.

The LNN type pump is a single stage, double suction, horizontal split volute type centrifugal pump designed for water works, drainage, general service and circulating applications. It can be used with motor, steam turbine and gasoline or diesel engine drives.

3.3.3 Shaft The large diameter stiff shaft, mounted on bearings, has a keyed drive end. 3.3.4 Pump bearings and lubrication Ball bearings are fitted as standard and may be either oil or grease lubricated.

LNN horizontal suction and discharge nozzles (inline)

Oil lubrication is only available where the pump shaft is horizontal. The LNNV as standard has a liquid lubricated journal bearing fitted at the non-drive end. This bearing is lubricated by pumped product or from an external clean source. 3.3.5 Bearing housing Two grease nipples enable grease lubricated bearings to be replenished between major service intervals. For oil lubricated bearings, a constant level oiler is fitted.

LNNC bottom vertical suction, horizontal discharge

3.3.6 Stuffing box housing The stuffing box housing has a spigot (rabbet) fit between the pump casing and bearing housing for optimum concentricity. The design enables a number of sealing options to be fitted. LNNV horizontal suction/discharge, vertical shaft (inline)

3.2 Name nomenclature

The pump size will be engraved on the nameplate typically as below:

300-LNN-600

Nominal discharge branch size. Configuration – see 3.1 above. Nominal maximum impeller diameter. The typical nomenclature above is the general guide to the LNN configuration description. Identify the actual pump size and serial number from the pump nameplate. Check that this agrees with the applicable certification provided.

3.3.7 Shaft seal The mechanical seal(s), attached to the pump shaft, seals the pumped liquid from the environment. Gland packing may be fitted as an option. 3.3.8 Driver The driver is normally an electric motor. Different drive configurations may be fitted such as internal combustion engines, turbines, hydraulic motors etc driving via couplings, belts, gearboxes, drive shafts etc. 3.3.9 Accessories Accessories may be fitted when specified by the customer.

Page 10 of 36

LNN, LNNV and LNNC USER INSTRUCTIONS ENGLISH C953KH001 – 09/03 ®

3.4 Performance and operating limits

This product has been selected to meet the specifications of your purchase order. See section 1.5. The following data is included as additional information to help with your installation. It is typical, and factors such as temperature, materials, and seal type may influence this data. If required, a definitive statement for your particular application can be obtained from Flowserve. 3.4.1 Operating limits Pumped liquid temperature limits* Maximum ambient temperature*

up to +80 ºC (176 ºF) up to +40 ºC (104 ºF) up to 3 % by volume Maximum soft solids in suspension* (refer for size limits) Maximum pump speed refer to the nameplate *Subject to written agreement from Flowserve.

200LNN300 200LNN325 200LNN375 200LNN400 200LNN475 200LNN500 200LNN600 250LNN325 250LNN375 250LNN475 250LNN600 300LNN475 300LNN500 300LNN575 300LNN600 300LNN750 350LNN475 350LNN575 350LNN725 350LNN900 400LNN600 400LNN725 400LNN900 500LNN650 500LNN700 500LNN775 500LNN950 500LNN1150 600LNN950 600LNN975 600LNN1200 700LNN1225 300LNNC475 300LNNC500 300LNNC575 350LNNC475 350LNNC575 350LNNC725 350LNNC900 600LNNC950 600LNNC975 700LNNC1225

Impeller minimum passage size mm (in.) 22.4 (0.90) 24.3 (0.96) 25.5 (1.00) 29.6 (1.20) 24.0 (0.95) 17.5 (0.70) 16.0 (0.63) 30.1 (1.20) 27.8 (1.10) 32.5 (1.30) 22.0 (0.87) 36.3 (1.40) 36.8 (1.40) 42.9 (1.70) 30.0 (1.20) 27.9 (1.10) 45.4 (1.80) 41.2 (1.60) 48.0 (1.90) 33.0 (1.30) 45.3 (1.80) 54.3 (2.10) 45.1 (1.80) 60.4 (2.40) 48.7 (1.90) 54.6 (2.10) 64.0 (2.50) 41.9 (1.60) 67.6 (2.70) 72.2 (2.80) 60.2 (2.40) 89.3 (3.50) 36.3 (1.40) 36.8 (1.40) 42.9 (1.70) 45.4 (1.80) 41.2 (1.60) 48.0 (1.90) 33.0 (1.30) 67.6 (2.60) 72.2 (2.80) 89.3 (3.50)

Nominal wear ring diameter mm (in.) 215 (8.5) 240 (9.5) 215 (8.5) 240 (9.5) 240 (9.5) 215 (8.5) 240 (9.5) 240 (9.5) 264 (10.4) 264 (10.4) 264 (10.4) 330 (13.0) 300 (11.8) 350 (13.8) 300 (11.8) 330 (13.0) 380 (15.0) 380 (15.0) 380 (15.0) 380 (15.0) 420 (16.5) 420 (16.5) 440 (17.3) 470 (18.5) 440 (17.3) 500 (19.7) 500 (19.7) 500 (19.7) 620 (24.4) 564 (22.2) 580 22.8) 700 (27.6) 330 (13.0) 300 (11.8) 350 (13.8) 350 (13.8) 380 (15.0) 380 (15.0) 380 (15.0) 620 (24.4) 564 (22.2) 700 (27.6)

Equipment operated in hazardous locations must comply with the relevant explosion protection regulations. See section 1.6.4, Products used in potentially explosive atmospheres.

4.1 Location

The pump should be located to allow room for access, ventilation, maintenance and inspection with ample headroom for lifting and should be as close as practicable to the supply of liquid to be pumped. Refer to the general arrangement drawing for the pump set.

4.2 Part assemblies

3.4.2 Pump and impeller data Pump size

4 INSTALLATION

Mean radial wear ring clearance mm (in.) 0.3 (0.012) 0.3 (0.012) 0.3 (0.012) 0.3 (0.012) 0.3 (0.012) 0.3 (0.012) 0.3 (0.012) 0.3 (0.012) 0.3 (0.012) 0.3 (0.012) 0.3 (0.012) 0.3 (0.012) 0.3 (0.012) 0.3 (0.012) 0.3 (0.012) 0.3 (0.012) 0.3 (0.012) 0.3 (0.012) 0.3 (0.012) 0.3 (0.012) 0.3 (0.012) 0.3 (0.012) 0.3 (0.012) 0.3 (0.012) 0.3 (0.012) 0.3 (0.012) 0.3 (0.012) 0.3 (0.012) 0.4 (0.016) 0.4 (0.016) 0.4 (0.016) 0.4 (0.016) 0.3 (0.012) 0.3 (0.012) 0.3 (0.012) 0.3 (0.012) 0.3 (0.012) 0.3 (0.012) 0.3 (0.012) 0.4 (0.016) 0.4 (0.016) 0.4 (0.016)

Approx. oil capacity, both bearings litre (fl. oz.) 0.37 (12.5) 0.37 (12.5) 0.37 (12.5) 0.48 (16.2) 0.48 (16.2) 0.48 (16.2) 0.60 (20.3) 0.48 (16.2) 0.48 (16.2) 0.60 (20.3) 0.60 (20.3) 0.60 (20.3) 0.60 (20.3) 0.60 (20.3) 0.68 (23.0) 2.00 (67.6) 0.68 (23.0) 0.68 (23.0) 0.68 (23.0) 2.00 (67.6) 0.68 (23.0) 2.00 (67.6) 4.50 (152) 2.00 (67.6) 2.00 (67.6) 2.00 (67.6) 4.50 (152) 7.00 (237) 4.50 (152) 4.50 (152) 7.00 (237) 7.00 (237) 0.60 (20.3) 0.60 (20.3) 0.60 (20.3) 0.68 (23.0) 0.68 (23.0) 0.68 (23.0) 2.00 (67.6) 4.50 (152) 4.50 (152) 7.00 (237)

Motors may be supplied loose on LNNV pumps, typically on frame sizes 400 and above. It is the responsibility of the installer to ensure that the motor is assembled to the pump and lined up as detailed in section 4.5.2.

4.3 Foundation There are many methods of installing pump units to their foundations. The correct method depends on the size of the pump unit, its location and noise vibration limitations. Non-compliance with the provision of correct foundation and installation may lead to failure of the pump and, as such, would be outside the terms of the warranty. Ensure the following are met. a) The baseplate should be mounted onto a firm foundation, either an appropriate thickness of quality concrete or sturdy steel framework. (It should NOT be distorted or pulled down onto the surface of the foundation, but should be supported to maintain the original alignment.) b) Install the baseplate onto packing pieces evenly spaced and adjacent to foundation bolts.

c) Level with shims between baseplate and packing pieces. d) The pump and driver have been aligned before dispatch however the alignment of pump and motor half coupling must be checked. If this is incorrect, it indicates that the baseplate has become twisted and should be corrected by re-shimming.

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LNN, LNNV and LNNC USER INSTRUCTIONS ENGLISH C953KH001 – 09/03 ®

e) Vertical pumps should be mounted following the practices outlined for baseplate mounted pumps. (Larger sizes may need the motor fitting after installing the pump - refer to section 4.5.2.) f) If the pump is driven via a cardan shaft there may be a requirement to offset the pump shaft with respect to the driver to optimize the cardan shaft bearing life. This offset will typically be in the range 0 to 4 degrees depending on shaft design. Please consult the separate user instructions before installation. g) Any support for the cardan shaft plummer blocks must not exhibit resonant frequencies in the range 0.8 to 1.2 N where N = pump running speed. h) If not supplied, guarding shall be fitted as necessary to meet the requirements of EN292 and EN953.

4.4 Grouting

Where applicable, grout in the foundation bolts. After adding pipework connections and rechecking the coupling alignment, the baseplate should then be grouted in accordance with good engineering practice. Fabricated steel, cast iron and epoxy baseplates can be filled with grout. Folded steel baseplates should be grouted to locate their packing pieces. If in any doubt, please contact your nearest service centre for advice.

Although the pump will have been aligned at the factory it is most likely that this alignment will have been disturbed during transportation or handling. If necessary, align the motor to the pump, not the pump to the motor. Horizontal pumps – LNN and LNNC Alignment is achieved by adding or removing shims under the motor feet and also moving the motor horizontally as required. In some cases where the alignment cannot be achieved it will be necessary to move the pump before recommencing the above procedure. Vertical pumps – LNNV Adding or removing shims between the motor stool and the pump casing achieves alignment. The motor/motor stool assembly may also have to be moved horizontally at the interface with the pump casing, as required. It should be noted that the motor has a spigot (rabbet) fit into the motor stool and it is therefore not possible to achieve any horizontal movement at this interface. For couplings with narrow flanges use a dial indicator as shown below to check both parallel and angular alignment.

Parallel

Grouting provides solid contact between the pump unit and foundation, prevents lateral movement of running equipment and dampens resonant vibrations. Foundation bolts should only be fully tightened when the grout has cured.

Angular

4.5 Initial alignment 4.5.1 Thermal expansion The pump and motor will normally have to be aligned at ambient temperature and should be corrected to allow for thermal expansion at operating temperature. In pump installations involving high liquid temperatures, the unit should be run at the actual operating temperature, shut down and the alignment checked immediately. 4.5.2 Alignment methods Ensure pump and driver are isolated electrically and the half couplings are disconnected. The alignment MUST be checked.

Maximum permissible misalignment at working temperature: Parallel 0.2 mm (0.008 in.) TIR Angular 0.1 mm (0.004 in.) TIR Pumps with thick flanged non-spacer couplings can be aligned by using a straight-edge across the outside diameters of the coupling hubs and measuring the gap between the machined faces using feeler gauges, measuring wedge or calipers. When the electric motor has sleeve bearings it is necessary to ensure that the motor is aligned to run on its magnetic centreline.

Refer to the motor manual for details.

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LNN, LNNV and LNNC USER INSTRUCTIONS ENGLISH C953KH001 – 09/03 ®

A button (screwed into one of the shaft ends) is normally fitted between the motor and pump shaft ends to fix the axial position. If the motor does not run in its magnetic centre the resultant additional axial force may overload the pump thrust bearing. Complete piping as below and see sections 4.7, Final shaft alignment check up to and including section 5, Commissioning, startup, operation and shutdown before connecting driver and checking actual rotation.

4.6 Piping Protective covers are fitted to the pipe connections to prevent foreign bodies entering during transportation and installation. Ensure that these covers are removed from the pump before connecting any pipes. 4.6.1 Suction and discharge pipework In order to minimize friction losses and hydraulic noise in the pipework it is good practice to choose pipework that is one or two sizes larger than the pump suction and discharge. Typically main pipework velocities should not exceed 2 m/s (6 ft/sec) suction and 3 m/s (9 ft/sec) on the discharge. Take into account the available NPSH which must be higher than the required NPSH of the pump.

piping.

Never use the pump as a support for

4.6.2 Suction piping a) The inlet pipe should be one or two sizes larger than the pump inlet bore and pipe bends should be as large a radius as possible. b) Pipework reducers should have a maximum total angle of divergence of 15 degrees. c) On suction lift the piping should be inclined up towards the pump inlet with eccentric reducers incorporated to prevent air locks. d) On positive suction, the inlet piping must have a constant fall towards the pump. e) Flow should enter the pump suction with uniform flow, to minimize noise and wear. This is particularly important on large or high-speed pumps which should have a minimum of four diameters of straight pipe on the pump suction between the elbow and inlet flange. See section 10.3, Reference 1 for more detail. f) Inlet strainers, when used, should have a net ‘ free area’ of at least three times the inlet pipe area. g) Do not install elbows at an angle other than perpendicular to the shaft axis. Elbows parallel to the shaft axis will cause uneven flow. h) Except in unusual circumstances strainers are not recommended in inlet piping. If considerable foreign matter is expected a screen installed at the entrance to the wet well is preferable. i) Fitting an isolation valve will allow easier maintenance. j) Never throttle pump on suction side and never place a valve directly on the pump inlet nozzle. Typical design – flooded suction Schieber Discharge

valve

Maximum forces and moments allowed on the pump flanges vary with the pump size and type. To minimize these forces and moments that may, if excessive, cause misalignment, hot bearings, worn couplings, vibration and the possible failure of the pump casing, the following points should be strictly followed: • Prevent excessive external pipe load • Never draw piping into place by applying force to pump flange connections • Do not mount expansion joints so that their force, due to internal pressure, acts on the pump flange

Durchflußventil Suction valve

Typical design – suction lift Schieber Non Discharge valve Rückschlagventil return valve

The table in 4.6.3 summarizes the maximum forces and moments allowed on LNN pump casings. Refer to Flowserve for other configurations. before use.

Nonreturn valve

Rückschlagventil

Ensure piping and fittings are flushed

Ensure piping for hazardous liquids is arranged to allow pump flushing before removal of the pump.

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I I =Mindesteintauchtiefe: Submergence

I > 3D

LNN, LNNV and LNNC USER INSTRUCTIONS ENGLISH C953KH001 – 09/03 ®

4.6.3 Maximum forces and moments allowed on LNN pump flanges My

Fy -Fz

Fx

Shaft

Fy

-Mz

Mx

My Mx Fx Mz

SUCTION

Fz

DISCHARGE Maximum forces (F) in kN (lbf) and moments (M) in kNm (lbf•ft)

Type and size

200-LNN-300 to 200-LNN-600 250-LNN-325 to 250-LNN-600 300-LNN-500 300-LNN-575 300-LNN-600 300-LNN-750 350-LNN-475 to 350-LNN-900 400-LNN-600 to 400-LNN-900 500-LNN-650 to 500-LNN-1150 600-LNN-950 600-LNN-975 and 600-LNN-1200 700-LNN-1225

Suction

Discharge

Fx

Fy

Fz

Mx

My

Mz

Fx

Fy

Fz

Mx

My

Mz

5.34 (1200) 6.68 (1500) 7.12 (1600) 8.46 (1900) 7.12 (1600) 8.46 (1900) 10.70 (2400) 10.70 (2400) 12.90 (2900) 17.10 (3845) 14.90 (3350) 19.30 (4340)

4.45 (1000) 5.34 (1200) 5.79 (1300) 6.68 (1500) 5.79 (1300) 6.68 (1500) 8.58 (1930) 8.58 (1930) 10.50 (2360) 14.00 (3150) 12.10 (2720) 15.90 (3575)

6.68 (1500) 8.01 (1800) 8.90 (2000) 10.20 (2295) 8.90 (2000) 10.20 (2295) 12.90 (2900) 12.90 (2900) 15.60 (3500) 20.40 (4586) 17.80 (4000) 23.00 (5170)

5.02 (3700) 6.10 (4500) 6.37 (4700) 7.32 (5400) 6.37 (4700) 7.32 (5400) 9.12 (6725) 9.12 (6725) 10.90 (8040) 14.20 (10470) 12.40 (9145) 15.90 (11725)

3.80 (2800) 4.61 (3400) 4.75 (3500) 5.42 (4000) 4.75 (3500) 5.42 (4000) 6.74 (4970) 6.74 (4970) 8.05 (5935) 10.40 (7670) 9.14 (6740) 11.70 (8630)

2.44 (1800) 2.98 (2200) 3.12 (2300) 3.66 (2700) 3.12 (2300) 3.66 (2700) 4.90 (3615) 4.90 (3615) 6.14 (4530) 8.44 (6225) 7.22 (5325) 9.65 (7115)

3.78 (850) 5.34 (1200) 6.68 (1500) 6.68 (1500) 6.68 (1500) 6.68 (1500) 7.12 (1600) 8.46 (1600) 10.70 (2400) 12.90 (2900) 12.90 (2900) 14.90 (3350)

3.12 (700) 4.45 (1000) 5.34 (1200) 5.34 (1200) 5.34 (1200) 5.34 (1200) 5.79 (1300) 6.68 (1300) 8.58 (1930) 10.50 (2360) 10.50 (2360) 12.10 (2720)

4.90 (1100) 6.68 (1500) 8.01 (1800) 8.01 (1800) 8.01 (1800) 8.01 (1800) 8.90 (2000) 10.20 (2000) 12.90 (2900) 15.60 (3510) 15.60 (3510) 17.80 (4000)

3.53 (2600) 5.02 (3700) 6.10 (4500) 6.10 (4500) 6.10 (4500) 6.10 (4500) 6.37 (4700) 7.32 (5400) 9.12 (6725) 10.90 (8040) 10.90 (8040) 12.40 (9145)

2.58 (1900) 3.80 (2800) 4.61 (3400) 4.61 (3400) 4.61 (3400) 4.61 (3400) 4.75 (3500) 5.42 (4000) 6.74 (4970) 8.05 (5940) 8.05 (5940) 9.14 (6740)

1.76 (1300) 2.44 (1800) 2.98 (2200) 2.98 (2200) 2.98 (2200) 2.98 (2200) 3.12 (2300) 3.66 (2700) 4.90 (3615) 6.14 (4530) 6.14 (4530) 7.22 (5325)

Notes: 1) F = External force (tension or compression). M = External moment, clockwise or counter-clockwise. 2) Forces and moments may be applied simultaneously in any direction. 3) Values apply to all materials. 4) Higher loads may be applicable, if direction and magnitude of individual loads are known, but these need written approval from Flowserve Pump Division. 5) Pumps must be on rigid foundations and baseplates must be fully grouted. 6) Pump/baseplate should not used as pipe anchor. Expansion joints must be properly tied. 7) The pump mounting bolt torques specified must be used to prevent relative movement between the pump casing and baseplate. (See section 6.6, Fastener torques.) The bolt material must have a minimum yield strength of 600 N/mm2 (87 000 lb/in.2).

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LNN, LNNV and LNNC USER INSTRUCTIONS ENGLISH C953KH001 – 09/03 ®

4.6.4 Discharge piping A non-return valve should be located in the discharge pipework to protect the pump from excessive back pressure and hence reverse rotation when the unit is stopped. Pipework reducers should have a maximum total angle of divergence of 9 degrees.

Special seals may require different auxiliary piping to that described above. Consult separate User Instructions and/or Flowserve if unsure of correct method or arrangement. For pumping hot liquids, to avoid seal damage, it is recommended that any external flush/cooling supply be continued after stopping the pump. 4.6.6 Final checks Check the tightness of all bolts in the suction and discharge pipework. Check also the tightness of all foundation bolts.

Fitting an isolation valve will allow easier maintenance. Typical design Actual operating duty Kontrollmanometer should be verified

4.7 Final shaft alignment check

After connecting piping to the pump, rotate the shaft several times by hand to ensure there is no binding and all parts are free. Recheck the coupling alignment, as previously described, to ensure no pipe strain. If pipe strain exists, correct piping.

4.6.5 Auxiliary piping

4.8 Electrical connections

4.6.5.1 Drains Pipe pump casing drains and gland leakage to a convenient disposal point.

4.8.1 Electrical connections must be made by a qualified Electrician in accordance with relevant local national and international regulations.

4.6.5.2 Pumps fitted with gland packing When suction pressure is below ambient pressure it is necessary to feed the gland packing with liquid to provide lubrication and prevent the ingress of air. This is normally achieved with a supply from the pump discharge volute to the stuffing box. A control valve is fitted in the line to enable the pressure to the gland to be controlled.

4.8.2 It is important to be aware of the EUROPEAN DIRECTIVE on potentially explosive areas where compliance with IEC60079-14 is an additional requirement for making electrical connections.

If the pumped liquid is dirty and cannot be used for sealing, a separate clean compatible liquid supply to the gland at 1 bar (15 psi) above suction pressure is recommended. 4.6.5.3 Pumps fitted with mechanical seals Single seals requiring re-circulation will normally be provided with the auxiliary piping from pump casing already fitted. If the seal requires an auxiliary quench then a connection must be made to a suitable source of liquid flow, low pressure steam or static pressure from a header tank. Recommended pressure is 0.35 bar (5 psi) or less. Check General arrangement drawing.

It is important to be aware of the 4.8.3 EUROPEAN DIRECTIVE on electromagnetic compatibility when wiring up and installing equipment on site. Attention must be paid to ensure that the techniques used during wiring/installation do not increase electromagnetic emissions or decrease the electromagnetic immunity of the equipment, wiring or any connected devices. If in any doubt contact Flowserve for advice. 4.8.4 The motor must be wired up in accordance with the motor manufacturer’s instructions (normally supplied within the terminal box) including any temperature, earth leakage, current and other protective devices as appropriate. The identification nameplate should be checked to ensure the power supply is appropriate. 4.8.5 A device to provide emergency stopping must be fitted. 4.8.6 If not supplied pre-wired to the pump unit, the controller/starter electrical details will also be supplied within the controller/starter.

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LNN, LNNV and LNNC USER INSTRUCTIONS ENGLISH C953KH001 – 09/03 ®

4.8.7 For electrical details on pump sets with controllers see the separate wiring diagram. 4.8.8 See section 5.3, Direction of rotation before connecting the motor to the electrical supply.

When fitted with a constant level oiler, the bearing housing should be filled by unscrewing or hinging back the transparent bottle and filling the bottle with oil. Where an adjustable body Denco oiler is fitted this should be set to the height shown in the following diagram:

4.9 Protection systems The following protection systems are recommended particularly if the pump is installed in a potentially explosive area or is handling a hazardous liquid. If in doubt consult Flowserve. If there is any possibility of the system allowing the pump to run against a closed valve or below minimum continuous safe flow a protection device should be installed to ensure the temperature of the liquid does not rise to an unsafe level. If there are any circumstances in which the system can allow the pump to run dry, or start up empty, a power monitor should be fitted to stop the pump or prevent it from being started. This is particularly relevant if the pump is handling a flammable liquid. If leakage of product from the pump or its associated sealing system can cause a hazard it is recommended that an appropriate leakage detection system is installed. To prevent excessive surface temperatures at bearings it is recommended that temperature or vibration monitoring are carried out. See sections 5.7.4 and 5.7.5.

5 COMMISSIONING, START-UP, OPERATION AND SHUTDOWN These operations must be carried out by fully qualified personnel.

The oil filled bottle should then be refitted so as to return it to the upright position. Filling should be repeated until oil remains visible within the bottle. Approximate oil volumes are shown in section 3.4.2, Pump and impeller data. Grease lubricated pumps and electric motors are supplied pre-greased. Other drivers and gearboxes, if appropriate, should be lubricated in accordance with their manuals. In the case of product lubricated bearings the source of product supply should be checked against the order. There may be requirements for an external clean supply, particular supply pressure or the commencement of lubrication supply before pump start-up.

5.1 Pre-commissioning procedure 5.1.1 Lubrication Determine the mode of lubrication of the pump set, eg grease, oil, product lubrication etc. For oil lubricated pumps, fill the bearing housing with correct grade of oil to the correct level, ie sight glass or constant level oiler bottle.

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LNN, LNNV and LNNC USER INSTRUCTIONS ENGLISH C953KH001 – 09/03 ®

5.2 Pump lubricants

Centifugal pump lubrication

5.2.1 Recommended oil lubricants Oil

Splash lubrication

Viscosity mm ²/s 40 ºC Temp. maximum ºC (ºF) Designation according to DIN51502 ISO VG BP

Oil companies and lubricants

DEA Elf Esso Mobil Q8 Shell Texaco Wintershall (BASF Group)

32

68

46

65 (149)

80 (176)

-

HL/HLP 32

HL/HLP 68

HL/HLP 46

BP Energol HL32 BP Energol HLP32 Anstron HL32 Anstron HLP32 OLNA 32 HYDRELEF 32 TURBELF 32 ELFOLNA DS32 TERESSO 32 NUTO H32 Mobil DTE oil light Mobil DTE13 MobilDTE24 Q8 Verdi 32 Q8 Haydn 32 Shell Tellus 32 Shell Tellus 37 Rando Oil HD 32 Rando Oil HD-AZ-32 Wiolan HN32 Wiolan HS32

BP Energol HL68 BP Energol HLP68 Anstron HL68 Anstron HLP68

BP Energol HL46 BP Energol HLP46 Anstron HL46 Anstron HLP46

TURBELF SA68

TURBELF SA46

ELFOLNA DS68 TERESSO 68 NUTO H68 Mobil DTE oil heavy medium

ELFOLNA DS46 TERESSO 46 NUTO H46 Mobil DTE oil medium Mobil DTE15M Mobil DTE25 Q8 Verdi 46 Q8 Haydn 46 Shell Tellus 01 C 46 Shell Tellus 01 46 Rando Oil 46 Rando Oil HD B-46 Wiolan HN46 Wiolan HS46

Mobil DTE26 Q8 Verdi 68 Q8 Haydn 68 Shell Tellus 01 C 68 Shell Tellus 01 68 Rando Oil 68 Rando Oil HD C-68 Wiolan HN68 Wiolan HS68

5.2.2 Recommended grease lubricants

5.2.4

Grease nipples

Grease

Force feed lubrication

NLGI 2 *

NLGI 3 **

Temp. range ºC (ºF)

-20 to +100 (-4 to +212)

-20 to +100 (-4 to +212)

Designation according to DIN

K2K-20

K2K 30

BP

Energrease LS2

Energrease LS3

DEA

Glissando 20

Glissando 30

Elf

Elfmulti 2

Elfmulti 3

Esso

Beacon 2

Beacon 3

Mobil

Mobilux 2

Mobilux 3

Q8

Rembrandt 2

Rembrandt 3

Shell

Alvania Fett G2 Alvania Fett R2

Alvania R3

Texaco

Multilak 20 Multilak EP2

Multilak 30 Multilak EP3

Wintershall (BASF Group)

Wiolub LFK 2

-

SKF

LGMT 2

LGMT 3

Silkolene G55/T G56/T * NLGI 2 is an alternative grease and is not to be mixed with other grades ** Factory packed bearings for the temperature range with grease nipples

5.2.3 Recommended fill quantities Refer to section 3.4.2, Pump and impeller data.

Lubrication schedule

5.2.4.1 Oil lubricated bearings Normal oil change intervals are 4 000 operating hours or at least every 6 months. For pumps on hot service or in severely damp or corrosive atmosphere, the oil will require changing more frequently. Lubricant and bearing temperature analysis can be useful in optimizing lubricant change intervals. The lubricating oil should be a high quality mineral oil having foam inhibitors. Synthetic oils may also be used if checks show that the rubber oil seals will not be adversely affected. The bearing temperature may be allowed to rise to 50 ºC (122 ºF).above ambient, but should not exceed 82 ºC (180 ºF). A continuously rising temperature, or an abrupt rise, indicate a fault. 5.2.4.2 Grease lubricated bearings When grease nipples are fitted, one charge between grease changes is advisable for most operating conditions, ie 2 000 hours interval. Normal intervals between grease changes are 4 000 hours or at least every 6 months.

Page 17 of 36

LNN, LNNV and LNNC USER INSTRUCTIONS ENGLISH C953KH001 – 09/03 ®

Possible vent/priming Entlüftungsöffnung points

The characteristics of the installation and severity of service will determine the frequency of lubrication. Lubricant and bearing temperature analysis can be useful in optimising lubricant change intervals. The bearing temperature may be allowed to rise to 55 ºC (131 ºF) above ambient but should not exceed 95 °C (204 °F). For most operating conditions a quality grease having a lithium soap base and NLGI consistency of No 2 or No 3 is recommended. The drop point should exceed 175 ºC (350 ºF). Never mix greases containing different bases, thickeners or additives.

5.3 Direction of rotation

Entlüftungsöffnung

5.5.2 Suction lift with foot valve fitted Fill suction pipe and casing with liquid at a pressure of 1 to 2 bar from an external source. Vent as described in section 5.5.1. Possible priming Entlüftungsöffnung points

Ensure the pump is given the same rotation as the pump direction arrow cast on the pump casing.

Entlüftungsöffnung

Possible priming point Fülleitung

To avoid dry running the pump must either be filled with liquid or have the flexible coupling disconnected before driver is switched on. If maintenance work has been carried out to the site' s electricity supply, the direction of rotation should be re-checked as above in case the supply phasing has been altered.

5.4 Guarding Guarding is supplied fitted to the pump set. If this has been removed or disturbed ensure that all the protective guards around the pump coupling and exposed parts of the shaft are securely fixed.

5.5 Priming and auxiliary supplies Ensure all electrical, hydraulic, pneumatic, sealant and lubrication systems (as applicable) are connected and operational.

5.5.3 Suction lift without foot valve Pump casing vents on the suction volute must be connected to an external vacuum pump priming system. If in doubt please consult Flowserve.

5.6 Starting the pump Ensure flushing and/or cooling/ heating liquid supplies are turned ON before starting the pump. b) CLOSE the outlet valve. c) OPEN all inlet valves. d) Prime the pump.

a)

e)

Ensure the inlet pipe and pump casing are completely full of liquid before starting continuous duty operation. 5.5.1 Suction pressure above atmospheric pressure Horizontal pumps: open vent connection [102.11] on top of the pump upper casing to allow the trapped air to escape. Let liquid run out until free from air bubbles.

Ensure all vent connections are closed before starting. f) Start motor and check outlet pressure. g) If the pressure is satisfactory, slowly OPEN outlet control valve.

h) i)

Do not run the pump with the outlet valve closed for a period longer than 30 seconds. If NO pressure, or LOW pressure, STOP the pump. Refer to section 7, Faults; causes and remedies, for fault diagnosis.

Vertical pumps: open vent connection [102.11] at the front of the upper half casing and disconnect the seal flush line at the mechanical seal/stuffing box to allow the trapped air to escape. Let liquid run out until free from air bubbles.

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LNN, LNNV and LNNC USER INSTRUCTIONS ENGLISH C953KH001 – 09/03 ®

5.7 Running the pump

5.7.3 Pumps fitted with mechanical seal Mechanical seals require no adjustment. Any slight initial leakage will stop when the seal is run in.

5.7.1 Venting the pump Vent the pump to enable all trapped air to escape taking due care with hot or hazardous liquids. Under normal operating conditions, after the pump has been fully primed and vented, it should be unnecessary to re-vent the pump. 5.7.2 Pumps fitted with packed gland

Before pumping dirty liquids it is advisable, if possible, to run in the pump mechanical seal using clean liquid to safeguard the seal face. External flush or quench should be started before the pump is run and allowed to flow for a period after the pump has stopped. Never run a mechanical seal dry, even for a short time.

+P

5.7.4 Bearings If the pumps are working in a potentially explosive atmosphere, temperature or vibration monitoring at the bearings is recommended.

-P

If the pump has a packed gland there must be some leakage from the gland. Gland nuts should initially be finger-tight only. Leakage should take place soon after the stuffing box is pressurised. The gland must be adjusted evenly to give visible leakage and concentric alignment of the gland ring [409.00] to avoid excess temperature. If no leakage takes place the packing will begin to overheat. If overheating takes place the pump should be stopped and allowed to cool before being restarted. When the pump is re-started, check to ensure leakage is taking place at the packed gland. If hot liquids are being pumped it may be necessary to slacken the gland nuts to achieve leakage. The pump should be run for 30 minutes with steady leakage and the gland nuts tightened by 10 degrees at a time until leakage is reduced to an acceptable level, normally a minimum of 120 drops per minute is required. Bedding in of the packing may take another 30 minutes. Care must be taken when adjusting the gland on an operating pump. Safety gloves are essential. Loose clothing must not be worn to avoid being caught up by the pump shaft. Shaft guards must be replaced after the gland adjustment is complete. a short time.

Never run gland packing dry, even for

If bearing temperatures are to be monitored it is essential that a benchmark temperature is recorded at the commissioning stage and after the bearing temperature has stabilized. • Record the bearing temperature (t) and the ambient temperature (ta) • Estimate the likely maximum ambient temperature (tb) • Set the alarm at (t+tb-ta+5) °C [(t+tb-ta+10) °F] and the trip at 100 °C (212 °F) for oil lubrication and 105 °C (220 °F) for grease lubrication It is important, particularly with grease lubrication, to keep a check on bearing temperatures. After start up the temperature rise should be gradual, reaching a maximum after approximately 1.5 to 2 hours. This temperature rise should then remain constant or marginally reduce with time. (Refer to section 6.2.3.1 for further information.) 5.7.5 Normal vibration levels, alarm and trip For guidance, pumps generally fall under a classification for rigid support machines within the International rotating machinery standards and the recommended maximum levels below are based on those standards. Alarm and trip values for installed pumps should be based on the actual measurements (N) taken on the pump in the fully commissioned as new condition. Measuring vibration at regular intervals will then show any deterioration in pump or system operating conditions. Vibration velocity – unfiltered mm/s (in./s) r.m.s. Normal N Alarm

N x 1.25

Shutdown trip N x 2.0

Page 19 of 36

Horizontal pumps ≤ 5.6 (0.22)

Vertical pumps ≤ 7.1 (0.28)

≤ 7.1 (0.28)

≤ 9.0 (0.35)

≤ 11.2 (0.44)

≤ 14.2 (0.56)

LNN, LNNV and LNNC USER INSTRUCTIONS ENGLISH C953KH001 – 09/03 ®

5.7.6 Stop/start frequency Pump sets are normally suitable for the number of equally spaced stop/starts per hour shown in the table below. Check actual capability of the driver and control/starting system before commissioning. Motor rating kW (hp) Up to 15 (20) Between 15 (20) and 90 (120) Above 90 (120)

Maximum stop/starts per hour 15 10 6

Where duty and standby pumps are installed it is recommended that they are run alternately every week.

5.9.3 Pump speed Changing pump speed effects flow, total head, power absorbed, NPSHR, noise and vibration. Flow varies in direct proportion to pump speed, head varies as speed ratio squared and power varies as speed ratio cubed. The new duty, however, will also be dependent on the system curve. If increasing the speed, it is important therefore to ensure the maximum pump working pressure is not exceeded, the driver is not overloaded, NPSHA > NPSHR, and that noise and vibration are within local requirements and regulations.

5.8 Stopping and shutdown

5.9.4 Net positive suction head (NPSHA) NPSH available (NPSHA) is a measure of the head available in the pumped liquid, above its vapour pressure, at the pump suction branch.

Close the outlet valve, but ensure that the pump runs in this condition for no more than a few seconds. b) Stop the pump. c) Switch off flushing and/or cooling/heating liquid supplies at a time appropriate to the process.

NPSH required (NPSHR) is a measure of the head required in the pumped liquid, above its vapour pressure, to prevent the pump from cavitating. It is important that NPSHA > NPSHR. The margin between NPSHA > NPSHR should be as large as possible.

d)

If any change in NPSHA is proposed, ensure these margins are not significantly eroded. Refer to the pump performance curve to determine exact requirements particularly if flow has changed. If in doubt please consult your nearest Flowserve office for advice and details of the minimum allowable margin for your application.

a)

For prolonged shut-downs and especially when ambient temperatures are likely to drop below freezing point, the pump and any cooling and flushing arrangements must be drained or otherwise protected.

5.9 Hydraulic, mechanical and electrical duty

This product has been supplied to meet the performance specifications of your purchase order, however it is understood that during the life of the product these may change. The following notes may help the user decide how to evaluate the implications of any change. If in doubt contact your nearest Flowserve office. 5.9.1 Specific gravity (SG) Pump capacity and total head in metres (feet) do not change with SG, however pressure displayed on a pressure gauge is directly proportional to SG. Power absorbed is also directly proportional to SG. It is therefore important to check that any change in SG will not overload the pump driver or over-pressurize the pump. 5.9.2 Viscosity For a given flow rate the total head reduces with increased viscosity and increases with reduced viscosity. Also for a given flow rate the power absorbed increases with increased viscosity, and reduces with reduced viscosity. It is important that checks are made with your nearest Flowserve office if changes in viscosity are planned.

5.9.5 Pumped flow Flow must not fall outside the minimum and maximum continuous safe flow shown on the pump performance curve and or data sheet.

6 MAINTENANCE 6.1 General It is the plant operator’s responsibility to ensure that all maintenance, inspection and assembly work is carried out by authorized and qualified personnel who have adequately familiarized themselves with the subject matter by studying this manual in detail. (See also section 1.6.2.) Any work on the machine must be performed when it is at a standstill. It is imperative that the procedure for shutting down the machine is followed, as described in section 5.8. On completion of work all guards and safety devices must be re-installed and made operative again. Before restarting the machine, the relevant instructions listed in section 5, Commissioning, start up, operation and shut down must be observed.

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LNN, LNNV and LNNC USER INSTRUCTIONS ENGLISH C953KH001 – 09/03 ®

Oil and grease leaks may make the ground slippery. Machine maintenance must always begin and finish by cleaning the ground and the exterior of the machine. If platforms, stairs and guard rails are required for maintenance, they must be placed for easy access to areas where maintenance and inspection are to be carried out. The positioning of these accessories must not limit access or hinder the lifting of the part to be serviced. When air or compressed inert gas is used in the maintenance process, the operator and anyone in the vicinity must be careful and have the appropriate protection. Do not spray air or compressed inert gas on skin. Do not direct an air or gas jet towards other people. Never use air or compressed inert gas to clean clothes.

f)

Check vibration, noise level and surface temperature at the bearings to confirm satisfactory operation. g) Check dirt and dust is removed from areas around close clearances, bearing housings and motors. h) Check coupling alignment and re-align if necessary. Our specialist service personnel can help with preventative maintenance records and provide condition monitoring for temperature and vibration to identify the onset of potential problems. If any problems are found the following sequence of actions should take place: a) Refer to section 7, Faults; causes and remedies, for fault diagnosis. b) Ensure equipment complies with the recommendations in this manual. c) Contact Flowserve if the problem persists. 6.2.1 Routine inspection (daily/weekly)

Before working on the pump, take measures to prevent an uncontrolled start. Put a warning board on the starting device with the words: "Machine under repair: do not start". With electric drive equipment, lock the main switch open and withdraw any fuses. Put a warning board on the fuse box or main switch with the words: "Machine under repair: do not connect". Never clean equipment with inflammable solvents or carbon tetrachloride. Protect yourself against toxic fumes when using cleaning agents.

6.2 Maintenance schedule It is recommended that a maintenance plan and schedule is adopted, in line with these User Instructions. It should include the following: a) Any auxiliary systems installed must be monitored, if necessary, to ensure they function correctly. b) Gland packings must be adjusted correctly to give visible leakage and concentric alignment of the gland follower to prevent excessive temperature of the packing or follower. c) Check for any leaks from gaskets and seals. The correct functioning of the shaft seal must be checked regularly. d) Check bearing lubricant level, and if the hours run show a lubricant change is required. e) Check that the duty condition is in the safe operating range for the pump.

The following checks should be made and the appropriate action taken to remedy any deviations: a) Check operating behaviour. Ensure noise, vibration and bearing temperatures are normal. b) Check that there are no abnormal fluid or lubricant leaks (static and dynamic seals) and that any sealant systems (if fitted) are full and operating normally. c) Check that shaft seal leaks are within acceptable limits. d) Check the level and condition of oil lubricant. On grease lubricated pumps, check running hours since last recharge of grease or complete grease change. e) Check any auxiliary supplies eg heating/cooling, if fitted, are functioning correctly.

Refer to the manuals of any associated equipment for routine checks needed. 6.2.2 Periodic inspection (six monthly)

Check foundation bolts for security of attachment and corrosion. b) Check pump running records for hourly usage to determine if bearing lubricant requires changing. c) The coupling should be checked for correct alignment and worn driving elements.

a)

Refer to the manuals of any associated equipment for periodic checks needed.

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LNN, LNNV and LNNC USER INSTRUCTIONS ENGLISH C953KH001 – 09/03 ®

6.2 3 Re-lubrication Lubricant and bearing temperature analysis can be useful in optimizing lubricant change intervals. In general however, the following is recommended.

important.

Maintaining the correct oil level is very

If the pump is supplied with a constant level oiler the oil level will be automatically maintained and as long as oil is visible in the glass bottle there is no need to refill. If however a sight glass has been fitted then regular checks should be made to ensure the level is maintained at the centre of the glass window. Refer to section 5.1.1 for methods of oil fill, section 5.2.1 for oil grade recommendations and 5.2.4 for the schedule and temperature limits.

e) V-rings should be seated at the proper distance from the sealing surface to avoid overheating. f) The maximum allowable operating temperatures for anti friction bearings will vary from unit to unit, depending on ambient and fluid temperature. The rise above ambient should not normally exceed 55 °C (131 °F) or a combined maximum of 95 °C (204 °F). g) A continuously rising temperature or an abrupt temperature rise indicates a problem. If these symptoms occur, stop the pump immediately and investigate the cause. TEMPERATURE

6.2.3.1 Oil lubrication

6.2.3.2 Grease lubrication See section 5.2.2 for grease recommendations.

TIME

Regrease - via grease nipples every 2 000 hours or sooner depending on the severity of the application. a) It is important not to under or over grease the bearings as this will lead to over heating and premature failure. Grease lubricated bearing housings have grease nipples fitted in the bearing covers.

Grease change - every 4 000 hours or sooner depending on the severity of the application. a) Remove the bearing housing from the rotor assembly. b) Brush the bearing housing with hot kerosene (100 to 115 °C/212 to 240 °F) or other non-toxic solvent. c) Clean and flush out the housing with a light mineral oil.

b) Move the axial seal ring back so the gap between the pump shaft and bearing cover can be seen. d) Do not use waste oil to clean the housing. To clean the bearings: a) Wipe off as much grease as possible with a clean lint-free cloth. b) Brush bearings with hot kerosene (80 to 90 °C/ 175 to 195 °F) while gently spinning the outer bearing ring. c) Spin each ball to ensure that it is clean.

c) Connect grease gun to the nipple.

d) Press grease into the bearing housing until the first signs of it appear in the gap between the housing and shaft, then stop greasing.

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LNN, LNNV and LNNC USER INSTRUCTIONS ENGLISH C953KH001 – 09/03 ®

To remove badly oxidized grease which refuses to come off: a) Support the rotor in a vertical position and immerse the bearing in hot kerosene or a mixture of alcohol and light mineral solvent. b) Gently spin the bearing outer ring.

Any change to the original design specification (modification or use of a non-standard part) will invalidate the pump’s safety certification. 6.3.2 Storage of spares Spares should be stored in a clean dry area away from vibration. Inspection and re-treatment of metallic surfaces (if necessary) with preservative is recommended at 6 monthly intervals.

6.4 Recommended spares and consumable items c) Dry and reflush the bearing with clean light oil. d) It is important not to under or over grease the bearings as this will lead to over heating and premature failure. It is recommended that the bearings be filled with grease using a suitable spatula. In addition the housings should be no more than half filled. 6.2.4 Mechanical seals No adjustment is possible. When leakage reaches an unacceptable level the seal will need replacement. 6.2.5 Gland packing The stuffing box split gland can be completely removed for re-packing or to enable the addition of extra rings of packing. The stuffing box is normally supplied with a lantern ring to enable a clean or pressurised flush to the centre of the packing. If not required, this can be replaced by an extra 2 rings of packing. There must always be a small leakage, normally a minimum of 120 drops per minute to atmosphere to lubricate and cool the packing is required.

6.3 Spare parts

For start up purposes: 1 - complete set of gland packing 2 - shaft sleeves 1 - set of gaskets and seals (optional: 2 - mechanical seals) For 2 years operation: 1 - set of bearings (line and thrust) 2 - sets of gland packing 2 - shaft sleeves 2 - sets of gaskets and seals 2 - lantern rings 2 - casing wear rings (optional: 2 - mechanical seals 2 - impeller wear rings) For 4 years operation: 1 - set of bearings (line and thrust) 2 - sets of gland packing 2 - shaft sleeves 2 - sets of gaskets and seals 2 - lantern rings 2 - casing wear rings 1 - impeller (optional: 2 - mechanical seals 2 - impeller wear rings)

6.5 Tools required

6.3.1 Ordering of spares Flowserve keep records of all pumps that have been supplied. When ordering spares the following information should be quoted: 1) Pump serial number. 2) Pump size. 3) Part name – taken from section 8. 4) Part number – taken from section 8. 5) Number of parts required. The pump size and serial number are shown on the pump nameplate. To ensure continued satisfactory operation, replacement parts to the original design specification should be obtained from Flowserve.

A typical range of tools that will be required to maintain these pumps is listed below. Readily available in standard tool kits, and dependent on pump size: • Open ended spanners (wrenches) to suit up to M 48 screws/nuts • Socket spanners (wrenches), up to M 48 screws • Allen keys, up to 10 mm (A/F) • Range of screwdrivers • Soft mallet More specialized equipment: • Bearing pullers • Bearing induction heater • Dial test indicator • C-spanner (wrench) - for removing shaft nut. (If difficulties in sourcing are encountered, consult Flowserve.)

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LNN, LNNV and LNNC USER INSTRUCTIONS ENGLISH C953KH001 – 09/03 ®

6.6 Fastener torques Bolt size


 

M 16 ( M 20 (¾ in.)
  M 24 ( M 27 (1 in.)
  M 30 (1
  M 36 (1
  M 42 (1
  M 48 (1

Torque Nm (lb ft) Pump feet All other fasteners fasteners 170 (125) 84 (62) 340 (250) 165 (120) 590 (435) 285 (210) 770 (570) 375 (275) 1 100 (810) 540 (400) 1 840 (1 350) 900 (660) 2 000 (1 475) 1 410 (1 040) 2 240 (1 650) 2 060 (1 500)

6.7 Renewal clearances

As wear takes place between the impeller and casing ring the overall efficiency of the pump set will decrease. To maintain optimum efficiency it is recommended that rings are replaced and the impeller renovated when the radial clearance detailed in section 3.4.2 has doubled to 0.6 to 0.8 mm (0.024 to 0.032 in.), depending on pump size. On LNNV it is recommended that the product lubricated bearing is renewed at a diametrical clearance of 0.5 mm (0.02 in.).

6.8 Disassembly Refer to section 1.6, Safety, before dismantling the pump. Before dismantling the pump for overhaul, ensure genuine Flowserve replacement parts are available. Refer to sectional drawings for part numbers and identification. 6.8.1 Rotor unit 6.8.1.1 LNN and LNNC a) Isolate motor and lock off electrical supply in accordance with local regulations. b) Isolate suction and discharge valves. c) Remove coupling guards and disconnect the coupling halves. d) Drain pump casing. Remove any auxiliary piping if applicable. e) Unscrew and remove bearing housing setscrews [202.01]. f) Unscrew and remove nuts [101.05] above split flange on upper half casing. Drive out location pin [102.02] (if fitted) from casing flange halves. Remove upper half casing [102.00]. g) Take out complete rotor unit and place onto two support blocks.

6.8.1.2 LNNV This pump is best removed from the system to carry out complete strip down. It should be set down with the shaft horizontal to enable the rotor to be removed. a) Isolate motor and lock off electrical supply in accordance with local regulations. b) Isolate suction and discharge valves. c) Remove coupling guards and disconnect the coupling halves. d) Drain pump casing and, if applicable, remove any auxiliary piping. e) Remove motor complete with motor stool and set down carefully in a safe location. f) Retain any shimming between stool and pump casing. g) Remove bolts securing pump suction and discharge flanges. h) Sling pump as shown in section 2.3 and take the strain. Remove setscrews securing the pump baseplate to the pump casing. i) Remove the pump to a safe location and manoeuvre the pump shaft into a horizontal position. j) Unscrew and remove setscrews [202.01 and 251.01] securing the bearing housing and end cover [202.00 and 251.00]. Remove end cover. k) Unscrew and remove nuts [101.05] above split flange on upper half casing. If fitted to casing flange halves, drive out location pin [102.02]. l) Using jacking screws, remove upper half casing. m) Carefully remove non-drive end stuffing box housing [401.00] complete with lower bearing housing [221.00] and bearing bush [248.00]. The impeller now rests on the casing ring. n) Carefully take out complete rotor assembly. Protect the bearing surface on the outside diameter of the bearing sleeve [353.00] from damage and place rotor on two support blocks. 6.8.2 Bearing housing a) Remove bearing cover setscrews [252.01] and remove key [309.00] from shaft end. b) Remove V-ring seal [379.00] and pull off bearing housing [202.00] from the rotor. c) After unbolting the bearing housing, move it back using a wedge, as below:

Note: The bearing housings, ball bearings, and shaft seals can be removed without removal of the upper half casing, providing the pump is fitted with a spacer coupling.

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LNN, LNNV and LNNC USER INSTRUCTIONS ENGLISH C953KH001 – 09/03 ®

6.8.3 Line bearing 6.8.3.1 LNN and LNNC Remove bearing [241.00] from shaft using a bearing puller ensuring pulling force is applied to the inner race only. 6.8.3.2 LNNV If the bearing bush is showing signs of wear it must be removed by pressing out of the lower bearing housing [221.00]. 6.8.4 Thrust bearings a) Unscrew and remove the self locking bearing nut [366.00] and remove the bearing [231.00] using a puller as in 6.8.3.1 above. b)

On pumps with an unsleeved shaft, check if a bearing distance ring [382.00] is fitted between the bearing and the shaft shoulder. If so, this MUST be retained for refitting during assembly.

6.8.5 Shaft seal - gland packing a) Remove gland nuts [401.21] and gland [406.00]. b) Lever out gland ring [409.00] using its grip groove. c) Remove gland packing rings [404.00] and lantern rings [405.00] using a bent wire.

6.8.7.2 LNNV non-drive end a) Remove capscrew [371.01] and end cap [371.00] from shaft end. b) Remove lower bearing sleeve [353.00] and lower sleeve [352.00]. 6.8.8 Impeller and wear rings a) The impeller and wear rings can now be removed if required. b) When removing the rotor unit, the casing rings [161.00] will be attached to it as they are fixed by two diametrically opposite pins [161.01] inserted into the casing ring and located in grooves in the lower half casing. c) If impeller rings [321.00] are also fitted, they are shrunk onto the impeller and fixed with locking screws [321.01] between their diametral mating surfaces. d) To remove the impeller rings, remove the locking screws and heat up the ring until it slides off easily.

6.9 Examination of parts Used parts must be inspected before assembly to ensure the pump will subsequently run properly. In particular, fault diagnosis is essential to enhance pump and plant reliability.

6.8.6 Shaft seal - mechanical seal a) Remove seal cover screws [401.21] and pull off 6.9.1 Casing, seal housing and impeller seal cover [423.00] complete with the stationary a) Inspect for excessive wear, pitting, corrosion, seal ring which is held in place by the O-ring seal. erosion or damage and any sealing surface b) The mechanical seal cover can also be removed by irregularities. placing a wedge into the gland chamfer, as below: b) Replace as necessary. 6.9.2 Shaft and sleeve (if fitted) Replace if grooved, pitted or worn. 6.9.3 Gaskets and O-rings After dismantling, discard and replace.

Refer to any special instructions supplied with the mechanical seal. 6.8.7 Shaft sleeve 6.8.7.1 LNN, LNNC and LNNV drive end a) Loosen grub screw [363.01] and unscrew shaft nut [363.00]. Remove shaft sleeve [351.00] using its pulling groove. b) If after removing the seal cover, or cartridge seal, there is no shaft nut [363.00] visible, this means that an unsleeved shaft is fitted. See sectional drawing for details. c) The shaft nut and spacer are accessible after removing the stuffing box housing [401.00].

6.9.4 Bearings a) It is recommended that bearings are not re-used after any removal from the shaft. b) The plain liquid lubricated bearings may be reused if both the bearing bush and bearing sleeve show no sign of wear, grooving or corrosion attack. (It is recommended that both the bush and sleeve are replaced at the same time.) 6.9.5 Bearing isolators, labyrinths or lip seals (if fitted) a) The lubricant, bearings and bearing housing seals are to be inspected for contamination and damage. If oil bath lubrication is utilised, these provide useful information on operating conditions within the bearing housing.

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LNN, LNNV and LNNC USER INSTRUCTIONS ENGLISH C953KH001 – 09/03 ®

b) If bearing damage is not due to normal wear and the lubricant contains adverse contaminants, the cause should be corrected before the pump is returned to service. c) Labyrinth seals and bearing isolators should be inspected for damage but are normally nonwearing parts and can be re-used. d) Bearing seals are not totally leak free devices. Oil from these may cause staining adjacent to the bearings.

6.10 Assembly

To assemble the pump consult the sectional drawings, see section 8, Parts list and drawings. Ensure threads, gasket and O-ring mating faces are clean. Apply thread sealant to non-face sealing pipe thread fittings. 6.10.1 Wear rings a) Impeller rings (when fitted) should be heated up using a hotplate or hot oil bath and then slipped onto the impeller and pressed down to the shoulder. (Do NOT use a steel hammer to knock them into position.) b) Drill and tap 3 holes approximately 120 degrees apart into the diametral mating faces of the ring and impeller and insert grubscrews. (The existing half tapped holes from the removed impeller ring cannot be re-used).

c) Slip the casing wear rings over the impeller hubs before mounting the rotor unit into the lower half casing, ensuring the pins in the rings locate into the holes in the casing. d) Check the running clearance between impeller and casing ring against the appropriate pump size in section 3.4.2.

b) The rotor always rotates towards the expanding section of the volute. c) The two shaft sleeves and shaft nuts clamping the impeller define its position on the pump shaft and hence in the pump casing. d) The correct axial position of the impeller and mechanical seals can be checked with the grooved checking marks on the pump shaft.

6.10.3 Shaft seal - packed gland a) Fit impeller key and slide impeller onto shaft. b) Insert O-ring into shaft sleeves and slide sleeves along shaft and into the impeller hubs. Lightly lubricate the shaft and O-ring for easier assembly. No O-rings are fitted beneath LNNV product lubricated bearing sleeves (see sectional drawing). On unsleeved shaft versions an O-ring is fitted at each end of the spacer. c) Tighten and adjust the shaft nuts so that their distances to the grooved marks are equal at both ends. d) Lock the shaft nuts in place with grubscrews. Tighten capscrew (LNNV). e) Slide the stuffing box housings over the shaft and fit the O-ring [414.00].

f)

This O-ring must be replaced at each and every dismantling. Place the gland ring [409.00] over the sleeve.



6.10.4 Shaft seal - mechanical seal 6.10.2 Impeller setting a) When re-assembling the impeller on the shaft, it is important to mount the impeller so that the vane tips point away from the apparent flow direction.

a)

Refer to any special instructions supplied with the mechanical seal. b) Slide the rotating assembly of the mechanical seal along the shaft sleeve until the retaining ring has reached the correct setting distance along the sleeve. Tighten the grubscrews to lock it into position.

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LNN, LNNV and LNNC USER INSTRUCTIONS ENGLISH C953KH001 – 09/03 ®

c) Insert O-ring into shaft sleeves and slide sleeves along shaft and into the impeller hubs. Lightly lubricate the shaft and O-ring for easier assembly. d) Tighten and adjust the shaft nuts so that their distances to the grooved marks are equal at both ends. e) Lock the shaft nuts in place with grubscrews. Tighten capscrew (LNNV). f) Slide the stuffing box housings over the shaft and fit the O-ring [414.00]. This O-ring must be replaced at each and every dismantling. g) Slip the mechanical seal covers [423.00] together with their installed stationary seal parts and O-rings over the shaft. 6.10.5 Ball bearings Before mounting the bearings, proceed as follows: a) Mount the V-ring seals [378.00] onto the shaft and slide the bearing cover [252.00] over the shaft. b) Pumps with grease lubricated bearings have V-ring seals on the outside of the bearing cover only. Pumps with oil lubricated bearings have, in addition, V-rings on the inside of the bearing covers. The inner V-rings have two small perforations in the lip. Inner V-rings engage in the grooves in the shaft. c) Determine the thickness of the laminated shim on the thrust bearing side. Provisionally position the bearing into the bearing housing seated against the circlip and thrust washer. d) Measure distance ’Z’ to face of bearing housing. e) Measure distance ’Y’ on the bearing cover. f) Shim thickness to give correct clearance will be ’Z’ minus ’Y’. Place correct shims on the shaft.

g)

The shim is laminated material with an original thickness (T) of 1.0 mm (0.039 in.) and laminate thickness of 0.05 mm (0.002 in.). This allows the thickness to be varied in 0.05 mm (0.002 in.) increments by peeling off layers to achieve the required axial clearance.

h) For oil lubricated units only, place the inner perforated V-ring [379.00] in the shaft grooves for correct positioning. i) The bearing should be heated up to 100 °C (212 °F) using a hotplate, hot oil bath or induction heater and then slipped onto the shaft to the shoulder. j) On the thrust bearing side, mount the self-locking ring type nut.

k) Bearing arrangements using a pair of angular contact bearings should be mounted with the shoulders of the inner rings arranged face to face as shown:

6.10.6 Liquid lubricated bearing a) Press a new bearing bush [248.00] into the lower bearing housing [221.00], making sure the face of the bush is flush with the end of the housing. b) Secure the lower bearing housing into the stuffing box housing. c) Slide the stuffing box housing, complete with bearing bush, over the shaft and fit the O-ring [414.00]. 6.10.7 Rotor unit a) After completion of preceding steps, carefully place the rotor into the lower half pump casing. Make sure the fixing pins of the casing rings fit correctly in the casing grooves and ensure correct fit of locating pins at the stuffing box housing. b) Although both stuffing box housings are identical the locating pins in the lower half casing are different for drive and non-drive sides. The stuffing box housing should be rotated so that the correct slot engages with the pin. The long pin with small diameter must engage in the small deep slot whilst the short larger diameter pin engages in the shallower wider groove.

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LNN, LNNV and LNNC USER INSTRUCTIONS ENGLISH C953KH001 – 09/03 ®

6.10.8 Casing gasket a) The gasket must be cut from asbestos-free sheet 1 material of 0.8 + 0.1/-0.05 mm ( /32 in.) thickness, by following the actual inner casing contour of the lower half casing. b) Special care must be exercised at the bores and stuffing box housing. c) The gasket must be accurately cut and fit flush with these bores to prevent leakage at the O-ring. d) Position gasket carefully onto the cleaned surface of the lower half casing. e) Coat the casing bottom half flange surface of the wall between suction and discharge side with a contact adhesive, as below:

f)

To assist assembly, particularly on larger sizes it may be advisable to use further adhesive at key points around this flange. Push gasket flush against fit of stuffing box housing and secure gasket locally again using the above adhesive paste.

g) Place upper half casing onto pump, ensuring dowels or stuffing box and bearing housing make correct alignment. h) Tighten upper half casing flange nuts/screws to correct torque. 6.10.9 Bearing housing a) Insert the circlip [235.00] and thrust washer [218.00] at the thrust bearing end. The circlip and thrust washer must not be fitted at the line bearing end. LNN and LNNC - thrust bearing at non-drive end. LNNV - thrust bearing at drive end.

b) On oil lubricated units ensure that the inner perforated V-rings [379.00] are located in the grooves in the shaft. c) Slip the bearing housings over their respective bearings and insert them into the recesses of the pump casing. d) Fit bearing housing setscrews [202.01] and tighten. e) Apply liquid sealant to bearing cover flange. f) Ensure correct seating of shim. g) Turn bearing cover to correct position: Grease lubricated units - grease nipple towards top half casing. Oil lubricated units - plug towards bottom half casing. h) Tighten bearing cover at bearing housing and push the V-ring [378.00] against the bearing cover. The sealing surface of the V-ring must be covered with grease and pushed gently up to the bearing cover, otherwise it may run hot!

k) Refit plugs, vents, oiler etc as applicable. l) On oil lubricated units at the drive end, place the perforated V-ring [379.01] over the shaft and position in the groove to seal against the end cover. m) Fit the end cover [253] and the external V-ring seal [378.01], lubricate it with grease and push it up to the end cover. n) At the non-drive end, fit the end cover [251] and tighten the screws [251.01]. 6.10.10 Stuffing box assembly a) Gland packing: Insert inner two rings of packing, then lantern ring halves and finally 2 or 3 more rings of packing. Loosely fit the gland [406.00] and connect flush line. b) Mechanical seal: Fasten seal covers [423.00] complete with O-ring and connect flush line. Connect any auxiliary piping.

Page 28 of 36

LNN, LNNV and LNNC USER INSTRUCTIONS ENGLISH C953KH001 – 09/03 ®

7 FAULTS; CAUSES AND REMEDIES FAULT SYMPTOM P u m p o v e r h e a ts a n d s e i z e s



Bearings have short life



Pump vibrates or is noisy



Mechanical seal has short life



Mechanical seal leaks excessivel y



Pump requires excessive power



P u m p l o s e s p r i m e a ft e r s ta r t i n g



Insufficient pressure developed



Insuffi cient capaci t y deli vered



Pump does not deliver liquid



PROBABLE CAUSES



A. System troubles

 











 



Pump not primed.





 

 

Suction lift too high or level too low. Insufficient margin between suction pressure and vapour pressure.



 

 







 

Pump or suction pipe not completely filled with liquid.









 



  



















Check and purge pipes and system. Check suction line design for vapour pockets.

Air leaks into suction line.

Check suction pipe is airtight.

Air leaks into pump through mechanical seal, sleeve joints, casing joint or pipe plugs.

Check and replace faulty parts. CONSULT FLOWSERVE.

Foot valve too small.

Investigate replacing the foot valve.

Foot valve partially clogged.

Clean foot valve.

Inlet of suction pipe insufficiently submerged.

Check out system design.

Speed too low.

CONSULT FLOWSERVE.

Speed too high.

CONSULT FLOWSERVE.

Total head of system higher than differential head of pump.

Viscosity of liquid differs from that for which designed.



Check NPSHA>NPSHR, proper submergence, losses at strainers/fittings.

Air or vapour pocket in suction line.

Specific gravity of liquid different from design.



Check complete filling. Vent and/or prime.

Excessive amount of air or gas in liquid.

Total head of system lower than pump design head.





POSSIBLE REMEDIES

Check system losses. Remedy or CONSULT FLOWSERVE.

Check and CONSULT FLOWSERVE.

Operation at very low capacity.

Measure value and check minimum permitted. Remedy or CONSULT FLOWSERVE.



Operation at high capacity.

Measure value and check maximum permitted. Remedy or CONSULT FLOWSERVE.



Misalignment due to pipe strain.

Check the flange connections and eliminate strains using elastic couplings or a method permitted.

Improperly designed foundation.

Check setting of baseplate: tighten, adjust, grout base as required.

Shaft bent.

Check shaft runouts are within acceptable values. CONSULT FLOWSERVE.

Rotating part rubbing on stationary part internally.

Check and CONSULT FLOWSERVE, if necessary.

Bearings worn

Replace bearings.

Wearing ring surfaces worn.

Replace worn wear ring/surfaces.

B. Mechanical troubles







 



















 



 





Page 29 of 36

LNN, LNNV and LNNC USER INSTRUCTIONS ENGLISH C953KH001 – 09/03 ®

FAULT SYMPTOM P u m p o v e r h e a ts a n d s e i z e s



Bearings have short life



Pump vibrates or is noisy



Mechanical seal has short life



Mechanical seal leaks excessivel y



Pump requires excessive power





P u m p l o s e s p r i m e a ft e r s ta r t i n g Insufficient pressure developed



Insuffi cient capaci t y deli vered



 







Pump does not deliver liquid PROBABLE CAUSES

POSSIBLE REMEDIES

Impeller damaged or eroded.

Replace or CONSULT FLOWSERVE for improved material selection.

Leakage under sleeve due to joint failure.

Replace joint and check for damage.





Shaft sleeve worn or scored or running off centre.

Check and renew defective parts.







Mechanical seal improperly installed.

Check alignment of faces or damaged parts and assembly method used.







Incorrect type of mechanical seal for operating conditions.

CONSULT FLOWSERVE.

Shaft running off centre because of worn bearings or misalignment.

Check misalignment and correct if necessary. If alignment satisfactory check bearings for excessive wear.



















 

 



Internal misalignment of parts preventing seal ring and seat from mating properly.





Mechanical seal was run dry.

Check mechanical seal condition and source of dry running and repair.





Internal misalignment due to improper repairs causing impeller to rub.

Check method of assembly, possible damage or state of cleanliness during assembly. Remedy or CONSULT FLOWSERVE, if necessary.

Excessive thrust caused by a mechanical failure inside the pump.

Check wear condition of impeller, its clearances and liquid passages.



Impeller out of balance resulting in vibration.



Abrasive solids in liquid pumped.

Check and CONSULT FLOWSERVE.



 



Excessive grease in ball bearings.

Check method of regreasing.





Lack of lubrication for bearings.

Check hours run since last change of lubricant, the schedule and its basis.





Improper installation of bearings (damage during assembly, incorrect assembly, wrong type of bearing etc).

Check method of assembly, possible damage or state of cleanliness during assembly and type of bearing used. Remedy or CONSULT FLOWSERVE, if necessary.





Damaged bearings due to contamination.

Check contamination source and replace damaged bearings.







 







C. MOTOR ELECTRICAL PROBLEMS



Wrong direction of rotation.

Reverse 2 phases at motor terminal box.

Motor running on 2 phases only.

Check supply and fuses.

Motor running too slow.

Check motor terminal box connections and voltage.

Page 30 of 36

LNN, LNNV and LNNC USER INSTRUCTIONS ENGLISH C953KH001 – 09/03 ®

8 PARTS LIST AND DRAWINGS 8.1 Sectional drawings - LNN and LNNC

c Leak proofing material Casco 145 or Marston Hydrosil 100RTV silicone compound * Note:- Items 379.00 amd 379.01 – oil lubricated arrangement only (Drawing taken from C751/006)

Page 31 of 36

LNN, LNNV and LNNC USER INSTRUCTIONS ENGLISH C953KH001 – 09/03 ®

8.2 Parts list - LNN and LNNC Ref no 060.01 060.03 060.04 101.00 101.01 101.02 101.03 101.04 101.05 101.06 101.07 101.08 101.09 101.10 102.00 102.01 102.02 102.11 102.12 109.00 161.00 161.01 202.00 202.01 202.02 202.03 202.04 202.05 202.06 202.07 218.00 231.00 231.01 235.00 241.00 251.00 251.01 251.02

Description Nameplate Nameplate Drive screw Casing lower half Screw Screw Screw Stud Nut Plug Washer Plug Washer Stud Casing upper half Forcing screw Locating dowel Plug Washer Gasket Casing ring Pin Bearing housing Setscrew Plug Washer Plug Washer Plug Washer Thrust washer Thrust bearing - non drive end Shim set Circlip Bearing - drive end End cover - non drive end Capscrew O-ring

252.00 252.01 253.00 280.00 295.00 299.00 299.01 301.00 302.00 309.00 311.00 321.00 321.01 351.00 363.00 363.01 366.00 374.00 378.00 378.01 379.00 379.01 401.00 401.01 401.04 401.05 401.20 401.21 404.00 405.00 406.00 409.00 414.00 420.00 423.00 500.01

Bearing cover Setscrew End cover - drive end Grease nipple Constant level oiler Breather Reducing bush Shaft Key Key Impeller Wear ring Locking screw Sleeve Shaft nut Locking screw Locknut O-ring seal V-ring seal V-ring seal V-ring seal (perforated) V-ring seal (perforated) Stuffing box housing Plug Pin Pin Stud Nut Gland packing Lantern ring Gland Gland ring O-ring seal Mechanical seal assembly Seal cover(non-cartridge seal) Flush line assembly

Oil lubrication

Optional impeller ring Scrap section showing cartridge seal fitted directly to pump shaft

View A

Page 32 of 36

LNN, LNNV and LNNC USER INSTRUCTIONS ENGLISH C953KH001 – 09/03 ®

8.3 Sectional drawings - LNNV sleeve bearing type

c Leak proofing material Casco 145 or Marston Hydrosil 100RTV silicone compound d Screw secured with Casco ML type 119 or Loctite 270 (Drawing taken from C751/008)

Page 33 of 36

LNN, LNNV and LNNC USER INSTRUCTIONS ENGLISH C953KH001 – 09/03 ®

8.4 Parts list - LNNV Ref. no. 060.01 060.03 060.04 101.00 101.01 101.02 101.03 101.04 101.05 101.06 101.07 101.08 101.09 101.10 102.00 102.01 102.02 102.11 102.12 102.20 109.00 161.00 161.01 202.00 202.01 202.02 202.03 202.04 202.05 202.06

202.07 218.00 221.00 221.01 231.00 231.01 235.00 241.00 243.00 248.00 251.00 251.01 251.02 252.00 252.01 253.00 253.01 253.02 280.00 281.00 301.00 302.00 306.00 309.00 311.00 321.00 321.01 351.00 352.00 353.00 363.00

Description Nameplate Nameplate Drive screw Casing lower half Screw Screw Screw Stud Nut Plug Washer Plug Washer Stud Casing upper half Forcing screw Locating dowel Plug Washer Lifting eye bolt Gasket Casing ring Pin Bearing housing Setscrew Plug Washer Plug Washer Plug

Washer Thrust washer Lower bearing housing Setscrew Thrust bearing - drive end Shim set Circlip Bearing - non drive end Gland ring Bearing bush - non drive end End cover - non drive end Capscrew O-ring Bearing cover Setscrew End cover - drive end Capscrew O-ring Grease nipple O-ring Shaft Key Sleeve key Key Impeller Wear ring Locking screw Sleeve Lower sleeve Bearing sleeve Shaft nut

363.01 366.00 371.00 371.01 373.00 375.00 378.00 378.01 383.00 387.00 387.01 387.02 401.00 401.01 401.04 401.05 401.20 401.21 404.00 405.00 406.00 409.00 414.00 420.00 423.00 440.00 500.01

Locking screw Locknut (bearing) End cap Capscrew Circlip O-ring V-ring seal V-ring seal Back up ring Flinger O-ring Locking screw Stuffing box housing Plug Pin Pin Stud Nut Gland packing Lantern ring Gland Gland ring O-ring seal Mechanical seal assembly Seal cover (non-cartridge seal) Drain pan Flush line assembly

View A

Optional impeller ring

Optional grease lubricated bottom bearing (Drawing taken from C753/164)

Page 34 of 36

LNN, LNNV and LNNC USER INSTRUCTIONS ENGLISH C953KH001 – 09/03 ®

8.5 General arrangement drawing

The typical general arrangement drawing and any specific drawings required by the contract will be sent to the Purchaser separately unless the contract specifically calls for these to be included into the User Instructions. If required, copies of other drawings sent separately to the Purchaser should be obtained from the Purchaser and retained with these User Instructions.

9 CERTIFICATION

Certificates determined from the Contract requirements are provided with these Instructions where applicable. Examples are certificates for CE marking, ATEX marking etc. If required, copies of other certificates sent separately to the Purchaser should be obtained from the Purchaser for retention with these User Instructions.

10 OTHER RELEVANT DOCUMENTATION AND MANUALS 10.1 Supplementary User Instructions

10.3 Additional sources of information Reference 1: NPSH for Rotordynamic Pumps: a reference guide, Europump Guide No. 1, Europump & World Pumps, Elsevier Science, United Kingdom, 1999. Reference 2: th Pumping Manual, 9 edition, T.C. Dickenson, Elsevier Advanced Technology, United Kingdom, 1995. Reference 3: nd Pump Handbook, 2 edition, Igor J. Karassik et al, McGraw-Hill Inc., New York, 1993. Reference 4: ANSI/HI 1.1-1.5 Centrifugal Pumps - Nomenclature, Definitions, Application and Operation. Reference 5: ANSI B31.3 - Process Piping.

Supplementary instructions such as for a driver, instrumentation, controller, seals, sealant system etc are provided as separate documents in their original format. If further copies of these are required they should be obtained from the supplier for retention with these User Instructions.

10.2 Change notes

If any changes, agreed with Flowserve Pump Division, are made to the product after its supply, a record of the details should be maintained with these User Instructions.

Page 35 of 36

LNN, LNNV and LNNC USER INSTRUCTIONS ENGLISH C953KH001 – 09/03 ®

FLOWSERVE REGIONAL SALES OFFICES: Europe, Middle East & Africa Flowserve Limited (Pump Division) Harley House, 94 Hare Lane Claygate, Esher, Surrey KT10 0RB United Kingdom

Latin America Flowserve S.A. de C.V. Avenida Paseo de la Reforma 30 2nd Floor, Colonia Juarez Centro Mexico, D.F.Z.C. 06040

Tel +44 (0)1372 463 700 Fax +44 (0)1372 460 190

Tel +52 5705 5526 Fax +52 5705 1125

USA and Canada Flowserve Corporation (Pump Division) Millennium Center, 222 Las Colinas Blvd. 15th Floor, Irving, TX 75039-5421, USA

Asia Pacific Flowserve Pte Ltd (Pump Division) 200 Pandan Loop, 06-03/04 Pantech 21, Singapore 128388

Tel +1 972 443 6500 Toll free 800 728 PUMP (7867) Fax +1 972 443 6800

Tel +65 775 3003 Fax +65 779 4607

Visit our web site at: www.flowserve.com

Your Flowserve factory contact:

Your local Flowserve representative:

To find your local Flowserve representative, please use the Sales Support Locator System found at www.flowserve.com

Page 36 of 36

®

Flowserve Pump Division

I.O.M. Installation, Operation and Maintenance IDP® Pumps TYPES CPXS & CPXNS FRAME MOUNTED, MAGNETIC DRIVE CHEMICAL PROCESS PUMPS

Instruction Manual C957KH025 - 01/03

INSTRUCTION MANUAL CPXS & CPXNS ENGLISH

C957KH025 - 12/02

INTRODUCTION

Flowserve Flowserve's products and brands are the leading names in their fields: the CPXS range of process pumps specifically focus on demanding chemical process applications. The pumps are manufactured at modern facilities, utilising state of art equipment and sophisticated quality control techniques. Flowserve is proud of earning preferred supplier status to many of world's leading processing companies. Engineered, manufactured, sold and serviced to ISO 9001 quality certification, Flowserve pumps are truly world class products. With more than 120 years of experience in servicing the needs of world-wide process industries, Flowserve has become the unchallenged leader in hydraulic design engineering, materials expertise and application know-how. Committed to continuous quality improvement, Flowserve controls the complete product life cycle - from application engineering, design, melting and casting, to cellular manufacturing, to assembly and testing, to the supply of aftermarket products, repair and diagnostic services. Flowserve is on hand to provide technical support and special services specific to the needs of its customers. Copyright All rights reserved. No part of this manual may be reproduced, stored in a retrieval system or transmitted in any form or by any means without prior permission of Flowserve Pump Division. CE Mark System It is a legal requirement that machinery and equipment put into service within the European Union shall conform with the applicable European Union Directives covering Machinery and, where applicable, Low Voltage Equipment, Electromagnetic Compatibility (EMC), Pressure Equipment Directive (PED) and Equipment for Potentially Explosive Atmospheres (ATEX). Where applicable the European Union Directives cover important Safety aspects relating to machinery and equipment and the satisfactory provision of technical documents and safety instructions. This document incorporates information relevant to these Directives. The Manual should be read prior to installing, operating, using and maintaining the equipment. The equipment must not be put into service until all the conditions relating to safety noted in the Manual have been met. Disclaimer Flowserve manufactures products to exacting International Quality Management System Standards (ISO 9001). Genuine parts and accessories have been designed, tested and incorporated into the products to ensure their continued product quality and performance in use. As Flowserve cannot test parts and accessories sourced from other vendors the incorrect incorporation of such parts and accessories may adversely affect the performance and safety features of the products. The failure to properly select, install or use authorised Flowserve parts and accessories is considered to be misuse. Damage or failure caused by misuse is not covered by Flowserve's warranty. In addition, any modification of Flowserve products or removal of original components may impair the safety of these products in their use.

2 ®

INSTRUCTION MANUAL CPXS & CPXNS ENGLISH

C957KH025 - 12/02

1 NAMEPLATE & WARNING LABELS 1.1 Nameplate For details of nameplate, see the Declaration of conformity. 1.2 Warning labels

Oil lubricated units only.

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INSTRUCTION MANUAL CPXS & CPXNS ENGLISH

C957KH025 - 12/02

2 SAFETY 2.2.6 DO NOT RUN THE PUMP AT ABNORMALLY HIGH OR LOW FLOW RATES Operating at a flow rate higher than normal or at a flow rate with no back pressure on the pump may overload the motor and cause cavitation. Low flow rates may cause a reduction in pump/bearing life, overheating of the pump, instability and cavitation/ vibration.

2.1 Duty conditions 2.1.1 This pump has been selected to meet the duty and service conditions advised on your order. The acknowledgement of these conditions has been sent separately to the Purchaser. A copy should be kept with this manual. 2.1.2 If there is any doubt as to the suitability of the pump for the application intended, contact Flowserve for advice, quoting the pump serial number.

2.2.7 NEVER DO MAINTENANCE WORK WHILST THE UNIT IS CONNECTED TO POWER

2.2 Safety action Always co-ordinate repair activity with operations personnel, and follow all plant safety requirements and applicable safety and health laws/regulations.

2.2.8 NEVER APPLY HEAT TO REMOVE IMPELLER Trapped lubricant or vapour could cause an explosion.

THIS IS A SUMMARY OF CONDITIONS AND ACTIONS TO PREVENT INJURY TO PERSONNEL AND DAMAGE TO EQUIPMENT.

2.2.9 HANDLING COMPONENTS Many precision parts have sharp corners and the wearing of appropriate safety gloves and equipment is required when handling these components. To lift heavy pieces above 30kg (66lbs) use a crane corresponding to the mass and in accordance with current local regulations.

This sign indicates safety instructions where non-compliance would affect personal safety. This symbol indicates electrical safety instructions where non-compliance would affect personal safety.

2.2.10 DRAIN PUMP AND ISOLATE PIPEWORK BEFORE DISMANTLING THE PUMP The appropriate safety precautions should be taken where the pumped liquids are hazardous.

This symbol indicates safety instructions where non-compliance would affect the safe operation or protection of the pump or pump unit.

2.2.11 FLUORO-ELASTOMERS (When fitted to high temperature units.) When a pump has experienced temperatures over 250°C (482ºF), partial decomposition of fluoroelastomers (eg viton) will occur. In this condition these are extremely dangerous and skin contact must be avoided.

2.2.1 PREVENT EXCESSIVE EXTERNAL PIPE LOAD Do not use pump as a support for piping. Do not mount expansion joints so that their force, due to internal pressure, acts on the pump flange.

2.2.12 THERMAL SHOCK Rapid changes in the temperature of the liquid within the pump can cause thermal shock, which can result in damage or breakage of components. Thermal shock should be avoided, particularly so where the material of the pump is not resistant to such loading.

2.2.2 ONLY CHECK DIRECTION OF MOTOR ROTATION WITH COUPLING ELEMENT/PINS REMOVED Starting in reverse direction of rotation will damage the pump. 2.2.3 START THE PUMP WITH OUTLET VALVE CLOSED This is recommended to avoid the risk of overloading and damaging the pump motor at full flow. Pumps may be started with the valve open only on installations where this situation cannot occur. 2.2.4

2.2.13 HOT (and cold) PARTS If hot or freezing components or auxiliary heating supplies can present a danger to operators, they must be shielded to avoid accidental contact. If complete protection is not possible, machine access must be limited to maintenance staff only. (Note: drive motors and bearings may be hot.)

NEVER RUN THE PUMP DRY

IF THE TEMPERATURE IS GREATER THAN 80°C (175°F) OR BELOW 5°C (20°F), A VISUAL WARNING INDICATOR SUCH AS A WARNING PLATE MUST BE PLACED CLEARLY ON THE EQUIPMENT.

2.2.5 INLET VALVES TO BE FULLY OPEN WHEN PUMP IS RUNNING Running the pump at zero flow or below the recommended minimum flow continuously will cause damage to the magnets and/or bearings.

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INSTRUCTION MANUAL CPXS & CPXNS ENGLISH

C957KH025 - 12/02

2.2.14 HAZARDOUS LIQUIDS When the pump is handling hazardous liquids care must be taken to avoid liquid contact using the appropriate health and safety procedures. Pump location and personnel access/training should consider and address these site dangers. 2.2.15 HIGH MAGNETIC FIELDS Great care should be taken when assembling/ dismantling magnetic rotors, where fitted, because of the very high forces which can be created by the magnets. Persons with pacemakers and any instrumentation etc sensitive to magnetic fields should be kept well away from the magnetic drive unit during dismantling. 2.3 Potentially explosive atmospheres Always check that the driver, drive coupling assembly and pump equipment are suitably rated and/or certified for the classification of the specific atmosphere in which they are to be installed. See section 17, Certification.

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INSTRUCTION MANUAL CPXS & CPXNS ENGLISH

C957KH025 - 12/02

3 CONTENTS PAGE

PAGE

INTRODUCTION

2

11 STOPPING AND SHUTDOWN

13

1 NAMEPLATE & WARNING LABELS

3

2 SAFETY Duty conditions Safety action Potentially explosive atmospheres

4 4 5

12 PREVENTATIVE MAINTENANCE AND SERVICING Maintenance schedule Routine inspection (daily/weekly) Periodic inspection (6 monthly) Lubrication data

13 13 14 14

3 CONTENTS

6

13 DISMANTLING AND ASSEMBLY Dismantling Examination of parts Magnets Assembly

14 15 16 16

14 SPARE PARTS Ordering of spares Storage of spare parts Prices of spare parts Recommended spares

19 19 19 19

15 GENERAL ARRANGEMENT DRAWING

19

16 TOOLS REQUIRED

19

17 CERTIFICATION

19

18 SUPPLEMENTARY INSTRUCTION MANUALS

19

19 CHANGE NOTES

19

20 OPERATING DIFFICULTIES

20

4 PUMP TECHNICAL DATA Performance Noise level Pressure limits Recommended screw torques Flange loads Pump lubricants Temperature limits Wearing clearances

7 7 7 7 7 8 8 8

5 PRODUCT DESCRIPTION General Pump casing Impeller Shaft Bearing housing Pump bearings and lubrication Shaft seal Driver Accessories

9 9 9 9 9 9 9 9 9

6 STORAGE

9

7 INSTALLATION Unpacking and inspection Handling Location Foundation Grouting Alignment of couplings Electrical connections Pipework connections Final piping check Condition monitoring

9 9 9 9 10 10 10 11 11 11

8 MAKING READY FOR OPERATION Lubrication Direction of rotation Guarding Open impeller clearance Primary and auxiliary supplies Filling and priming

12 12 12 13 13 13

9 STARTING THE PUMP

13

10 RUNNING Normal vibration levels, alarm and trip levels Stop/start frequency

13 13

21 SECTIONAL ARRANGEMENT DRAWINGS AND PARTS LISTS

21-26

22 PARTS INTERCHANGEABILITY

27

6 ®

INSTRUCTION MANUAL CPXS & CPXNS ENGLISH 4 PUMP TECHNICAL DATA

For units driven by equipment other than electric motors or units contained within enclosures, see the accompanying information sheets and manuals.

4.1 Performance For performance parameters see the paragraph on Safety - Duty conditions. When specified by the contract, performance data has been supplied separately to the purchaser and should be obtained and retained with this manual if required.

4.3 Pressure limits The operating pressure has been selected to meet your specified requirements. See the paragraph on Safety - Duty conditions for details.

4.2 Noise level When pump noise level exceeds 85dBA attention must be given to prevailing Health and Safety Legislation, to limit the exposure of plant operating personnel to the noise. The usual approach is to control exposure time to the noise or to enclose the machine to reduce emitted sound.

The pressure and temperature operating limits for the flanges are in accordance with the relevant National or International standards unless advised otherwise. 4.4 Recommended screw torques Screw size M6 M8 M10 M12 M16 M20 Inner rotor Model 80 & 100 locknut Model 150

You may have already specified a limiting noise level when the equipment was ordered, however if no noise requirements were defined then machines above a certain power level will exceed 85dBA. Pump noise level is dependent on a number of factors - the type of motor fitted, the operating capacity, pipework design and acoustic characteristics of the building. The levels specified in the table below are estimated and not guaranteed.

When requested the permissible flange loading will have been supplied separately to the purchaser and should be obtained and retained with this manual.

If a pump unit only has been purchased, for fitting with your own driver, then the "pump only" noise levels from the table should be combined with the level for the driver obtained from the supplier. Consult a Noise Specialist for this calculation.

kW <0.55 0.75 1.1 1.5 2.2 3 4 5.5 7.5 11 15 18.5 22 30 37 45 55 75 90 110 150

(hp) (<0.75) (1) (1.5) (2) (3) (4) (5) (7.5) (10) (15) (20) (25) (30) (40) (50) (60) (75) (100) (120) (150) (200)

3550 rpm Pump & Pump motor only dBA dBA 71 66 74 66 74 68 77 70 78 72 81 74 82 75 90 77 90 78 91 80 92 83 92 83 92 83 100 85 100 86 100 87 102 88 100 90 97 90 100 91 101 92

Torque Nm (lbf ft) 11 (8) 16 (12) 25 (18) 35 (26) 80 (59) 130 (96) 70 (53) 120 (88)

4.5 Flange loads The permissible flange loading is dependent on a number of factors such as dimensions, flange rating, pressure, temperature, material, pump configuration, etc. The recommendations contained in the section on pipework connections should be followed to eliminate these loads.

The dBA values are based on the noisiest ungeared electric motors which are likely to be encountered. They are Sound Pressure levels at 1m (3.3ft) from the directly driven pump, for "free field over a reflecting plane".

Motor size

C957KH025 - 12/02

If in doubt contact Flowserve for information.

2900 rpm Pump & Pump motor only dBA dBA 64 62 67 62 67 64 70 66 71 68 74 70 75 71 83 73 83 74 84 76 85 79 85 79 85 79 93 81 93 82 93 83 95 84 95 86 92 86 95 87 96 88

1750 rpm Pump & Pump motor only dBA dBA 64 62 67 62 67 64 70 66 71 68 74 74 75 75 76 75 77 76 78 77 80 79 80 79 81 79 84 80 84 80 84 80 86 81 88 81 90 81 91 83 91 83

1450 rpm Pump & Pump motor only dBA dBA 63 62 63 62 65 64 66 66 68 68 70 70 71 71 72 71 73 72 74 73 76 75 76 75 77 75 80 76 80 76 80 76 82 77 83 78 85 78 86 79 86 79

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INSTRUCTION MANUAL CPXS & CPXNS ENGLISH

C957KH025 - 12/02

4.6 Pump lubricants

Centifugal pump lubrication

4.6.1 Recommended oil lubricants Oil Viscosity mm²/s 40ºC Temp. maximum ºC (ºF) Designation according to DIN51502 ISO VG BP

Oil companies and lubricants

DEA Elf

Esso Mobil

Q8 Shell Texaco Wintershall (BASF Group)

Splash lubrication 32

68

46

65 (149)

80 (176)

-

HL/HLP 32

HL/HLP 68

HL/HLP 46

BP Energol HL32 BP Energol HLP32 Anstron HL32 Anstron HLP32 OLNA 32 HYDRELEF 32 TURBELF 32 ELFOLNA DS32 TERESSO 32 NUTO H32 Mobil DTE oil light Mobil DTE13 MobilDTE24 Q8 Verdi 32 Q8 Haydn 32 Shell Tellus 32 Shell Tellus 37 Rando Oil HD 32 Rando Oil HD-AZ-32 Wiolan HN32 Wiolan HS32

BP Energol HL68 BP Energol HLP68 Anstron HL68 Anstron HLP68

BP Energol HL46 BP Energol HLP46 Anstron HL46 Anstron HLP46

TURBELF SA68

TURBELF SA46

ELFOLNA DS68 TERESSO 68 NUTO H68 Mobil DTE oil heavy medium

ELFOLNA DS46 TERESSO 46 NUTO H46 Mobil DTE oil medium Mobil DTE15M Mobil DTE25 Q8 Verdi 46 Q8 Haydn 46 Shell Tellus 01 C 46 Shell Tellus 01 46 Rando Oil 46 Rando Oil HD B-46 Wiolan HN46 Wiolan HS46

Mobil DTE26 Q8 Verdi 68 Q8 Haydn 68 Shell Tellus 01 C 68 Shell Tellus 01 68 Rando Oil 68 Rando Oil HD C-68 Wiolan HN68 Wiolan HS68

4.8 Wearing clearances

4.6.2 Grease lubricated bearing housings are fitted with "sealed for life" bearings.

4.8.1 Journal faces (as new)

4.6.3 Bearing sizes and oil capacities. Model 80 100

Oil lubricated bearings 6208C3 6208C3

Force feed lubrication

6208ZZ C3

80

Bush Sleeve/ Diametral inside dia. shaft OD clearance (mm) (mm) (mm) 35.03 -0.00/+0.025 34.95 -0.02/+0.00 0.08 to 0.125

100

38.03 -0.00/+0.025 37.95 -0.02/+0.00 0.08 to 0.125

Model

Grease lubricated Oil capacity bearings (approx) litres 6208ZZ C3 0.235 0.235

150 53.975 -0.00/+0.025 53.86 -0.02/+0.00 0.11 to 0.16 Replace when diametral clearance exceeds 0.2mm (models 80 and 100) or 0.25mm (model 150).

150 6210C3 6210ZZ C3 0.430 NB: The bearing sizes do not constitute a purchasing specification.

4.7 Temperature limits

4.8.2 Thrust faces (as new) 80

Thrust face thickness (mm) 6.00 -0.07/+0.00

Rear bush length (mm) 14.00 -0.00/+0.10

100

10.51 -0.07/+0.00

19.50 -0.00/+0.10

Model

4.7.1 The pump materials and construction have been selected for your application, however, the following fundamental limits should not be exceeded: Neodymium magnets -40 to +120°C Samarium cobalt magnets -40 to +250°C Peek shell (depending on pressure) -40 to +120°C

150 12.00 -0.05/+0.05 33.20 -0.00/+0.10 Replace when total wear on both faces exceeds 0.15mm.

4.7.2 Ambient temperature These pumps are generally fitted with TEFC motors with an ambient temperature limit of 40°C. Specific pumps may be fitted with motors to suit client's requirements with other ambient temperature limits - see motor nameplate for details.

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INSTRUCTION MANUAL CPXS & CPXNS ENGLISH 5 PRODUCT DESCRIPTION

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achieved with the prior agreement of Flowserve and would necessitate inspection prior to putting the pump into service.

5.1 General The pump is a modular designed centrifugal pump that can be built to achieve almost all chemical liquid pumping requirements. For ultimate safety the pump has been fitted with a magnetic drive.

7 INSTALLATION 7.1 Unpacking and inspection The pump should be checked against the delivery advice note and any damage or shortage reported immediately to Flowserve. Any crate, carton or wrapping should be checked for any spare parts or accessories that may be packed with the pump.

5.2 Pump casing The pump casing is designed with a horizontal centreline end inlet and a vertical centreline top outlet which makes it self venting.

7.2 Handling Boxes, crates, pallets or cartons may be unloaded using fork lift vehicles or slings dependent on their size and construction. The pump set should be handled as shown in the appropriate drawing:

For ease of maintenance, the pump is constructed so that pipe connectors do not have to be disturbed when internal maintenance is required. 5.3 Impeller An open impeller is fitted. 5.4 Shaft The large diameter stiff drive shaft, mounted on bearings, has a keyed drive end. The pump shaft is fitted with magnetic rotor and product lubricated bearings. 5.5 Bearing housing For oil lubricated bearings, a sight glass enables the oil level to be viewed. Additional lubrication and cooling options may be fitted. 5.6 Pump bearings and lubrication The ball bearings fitted in the bearing housing may be oil or grease lubricated. The magnetic drive journal bearings may be lubricated by product or from an external source.

7.3 Location The pump should be located to allow room for access, ventilation, maintenance and inspection with ample headroom for lifting and should be as close as practicable to the supply of liquid to be pumped.

5.7 Shaft seal The magnetic drive design utilises the shell between the magnets to prevent leakage of the pumped fluid. 5.8 Driver The DRIVER is normally an electric motor. Different drive configurations may be fitted such as an internal combustion engine, turbines, hydraulic motors etc driving via couplings, belts, gearboxes, drive shafts etc.

7.4 Foundation There are many methods of installing pump units to their foundations. The correct method depends on the size of the pump unit, its location and noise vibration limitations. Non-compliance with the provision of correct foundation and installation may lead to failure of the pump and, as such, would be outside the terms of the warranty.

5.9 Accessories Accessories may be fitted when specified by the customer.

The baseplate should be mounted onto a firm foundation, either an appropriate thickness of quality concrete or sturdy steel framework. It should NOT be distorted or pulled down onto the surface of the foundation, but should be supported to maintain the original alignment.

6 STORAGE Store the pump in a clean, dry location away from vibration. Leave piping connection covers in place to keep dirt and other foreign material out of pump casing. Turn pump at intervals to prevent brinelling of the bearings and the seal faces, if fitted, from sticking.

Install the baseplate onto packing pieces evenly spaced and adjacent to foundation bolts. Level with shims between baseplate and packing pieces. The pump and driver have been aligned before dispatch. Check alignment of pump and motor half coupling. If this is incorrect, it indicates that the baseplate has become twisted and should be corrected by reshimming.

The pump may be stored as above for up to 6 months. Consult Flowserve for preservative actions when a longer storage period is needed. Warranty for the pumps will normally be for 12 months. Extension of this period can only be 9

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INSTRUCTION MANUAL CPXS & CPXNS ENGLISH 7.5 Grouting Where applicable, grout in the foundation bolts.

C957KH025 - 12/02

Alignment is achieved by adding or removing shims from under the motor feet and also moving the motor horizontally as required. In some cases, where the alignment cannot be achieved, it will be necessary to move the pump before recommencing the above procedure.

After adding pipework connections and rechecking the coupling alignment, the baseplate should then be grouted in accordance with good engineering practice. Fabricated steel, cast iron and epoxy baseplates can be filled with grout. Folded steel baseplates should be grouted to locate their packing pieces. If in any doubt, please contact your nearest service centre for advice.

7.6.2.3 For couplings with narrow flanges, use a dial indicator gauge as shown. The alignment values are maximums for continuous service.

Grouting provides solid contact between the pump unit and foundation, prevents lateral movement of vibrating equipment and dampens resonant vibrations. 7.6 Alignment of couplings 7.6.1 Thermal expansion The pump and motor will normally have to be aligned at ambient temperature and should be corrected to allow for thermal expansion at operating temperature. In pump installations involving high liquid temperatures, the unit should be run at the actual operating temperature, shut down and the alignment checked immediately.

Permissible misalignment limits at working temperature: • Parallel alignment - 0.25mm (0.010in) TIR maximum • Angular alignment - 0.3mm (0.012in) TIR maximum for couplings not exceeding 100mm (4in) flange diameter - 0.5mm (0.020in) TIR maximum for couplings over 100mm (4in) diameter

Motor and pump centre line height adjustment:

7.6.2.4 When checking parallel alignment, the total indicator read-out (TIR) shown is twice the value of the actual shaft displacement. 7.6.2.5 When the electric motor has sleeve bearings it is necessary to ensure that the motor is aligned to run on its magnetic centreline. Refer to the motor manual for details. A button (screwed into one of the shaft ends) is normally fitted between the motor and pump shaft ends to fix the axial position. 7.7 Electrical connections 7.7.1 Electrical connections should be made by a qualified Electrician in accordance with the relevant local national and international regulations. 7.7.2 It is important to be aware of the EUROPEAN DIRECTIVE on electromagnetic compatibility when wiring up and installing equipment on site. Attention must be paid to ensure that the techniques used during wiring/installation do not increase electromagnetic emissions or decrease the electromagnetic immunity of the equipment, wiring or any connected devices. If in any doubt contact Flowserve for advice.

Graph based on the assumptions that: a) Operating temperature rise of the motor frame is 50°C b) Packing piece/motor stool is not affected Operation 1 Enter graph at base to shaft centre line height 2 Read line for frame material 3 Set motor shaft and coupling LOW by figure on left-hand side

7.6.2 Alignment methods

7.7.3 The motor must be wired up in accordance with the motor manufacturer's instructions (normally supplied within the terminal box) including any temperature, earth leakage, current and other protective devices as appropriate. The identification nameplate should be checked to ensure the power supply is appropriate.

7.6.2.1 Ensure the pump and motor half couplings are disconnected. 7.6.2.2 The alignment MUST be checked. Although the pump will have been aligned at the factory it is most likely that this alignment will have been disturbed during transportation or handling. 10

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INSTRUCTION MANUAL CPXS & CPXNS ENGLISH

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7.8.6 Fitting an isolator and non-return valves can allow easier maintenance. Never throttle pump on the suction side and never place a valve directly on the pump inlet nozzle.

7.7.4 A device to provide emergency stopping shall be fitted. 7.7.5 If not supplied pre-wired to the pump unit the controller/starter electrical details will also be supplied within the controller/starter.

7.8.7 A non-return valve should be located in the discharge pipework to protect the pump from excessive back pressure and hence reverse rotation when the unit is stopped.

7.7.6 For electrical details on pump sets with controllers see the wiring diagram.

7.8.8 Piping and fittings should be flushed before use to ensure any debris from installation work etc is not passed through the pump.

7.7.7 See paragraphs on 'direction of rotation' before connecting the motor to the electrical supply.

7.8.9 Piping for corrosive liquids should be arranged to allow pump flushing before removal of a unit.

7.8 Pipework connections

7.8.10 By-pass lines should be fed back to the suction source, NOT to the pump entry.

7.8.1 Protective covers are fitted to the pipe connections to prevent foreign bodies entering during transportation and installation. Ensure that these covers are removed from the pump before connecting any pipes.

7.9 Final piping check After connecting piping to the pump, rotate the shaft several times by hand to ensure there is no binding and all parts are free.

7.8.2 Maximum forces and moments allowed on the pump flanges vary with the pump size and type. To minimise these forces and moments that may cause misalignment, hot bearings, worn couplings, vibration and the possible failure of the pump casing, the following points should be strictly followed: • Prevent excessive external pipe load. • Never draw piping into place by applying force to pump flange connections. • Do not mount expansion joints so that their force, due to internal pressure, acts on the pump flange.

Recheck the coupling alignment, as previously described, to ensure no pipe strain. If pipe strain exists, correct piping. 7.10 Condition monitoring Magnetic drive pumps are inherently safe and are ideal for toxic, corrosive and highly volatile liquids. However if abused and allowed to run dry, for example, the consequences can be expensive to repair. A few minutes dry running will cause severe damage to the magnetic drive.

7.8.3 The inlet pipe should be one or two sizes larger than the pump inlet bore and pipe bends should be as large a radius as possible.

The main potential risks of failure are: 1. Dry running due to blocking of lubrication ports with solids in pumped liquid. 2. Dry running due to loss of liquid to pump suction. 3. Dry running due to impeller seizing, caused by debris in the pump casing. 4. Dry running due to solidification of liquid in the shell, eg due to poor control of temperature.

On suction lift the piping should be inclined up towards the pump inlet with eccentric reducers incorporated to prevent air locks. On positive suction, the inlet piping must have a constant fall towards the pump. 7.8.4 Allow a minimum of two pipe diameters of straight section between the elbow and inlet flange. Inlet strainers, when used, should have a net `free area' of at least three times the inlet pipe area.

If any of these conditions occur, the system must be switched off within one minute and the most universal way of achieving this, for all the above conditions, is by using a power or current monitor fitted into the starter.

7.8.5 The pump is fitted with silicon carbide bearings therefore small non-abrasive solids less than 0.3mm in diameter can be handled providing they constitute no more than 2.5% by volume of liquid handled.

One other potential problem that can be monitored when pumping hazardous fluids is leakage from the shell.

Solids must be non-magnetic, must not have a tendency to coagulate and must not be fibrous. They should also be non-abrasive and must not scale wetted surfaces.

In this instance the drive should be fitted with dual containment and monitoring of the space between the two shells can be carried out using a pressure switch connected to either motor starter or alarm.

For services other than above you are recommended to contact Flowserve for advice.

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INSTRUCTION MANUAL CPXS & CPXNS ENGLISH

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If required, temperature of the liquid in the drive and the metal shell (when single containment) can also be monitored from the tapping points shown.

The oil filled bottle should then be refitted so as to return it to the upright position. Filling of the bottle should be repeated until oil remains visible within the bottle.

A G¼ B G½ C Rp ? F Rp ? H Rp ¼ J Rp ¼ K G½ L Rp ¼ V Rp ? W Rp ? Y1 Rp ½ Y2 Rp ¼ Z Rp ½

8.1.2 Oil lubricated units are supplied without oil and must be filled to the marked level before starting the pump.

Auxiliary connection (Rp) Bearing housing drain (plugged when provided) Casing drain (plugged when provided) Leakage (plugged when provided) External flush connection (plugged when provided) Discharge gauge connection (plugged when provided) Suction gauge connection (plugged when provided) Sight glass (when fitted) Constant level oiler (when fitted) Fluid temperature connection (plugged when provided) Bearing housing vent (when provided) Shell temperature connection (plugged as standard) Dual containment pressure connection (plugged when provided) Assembly access (plugged as standard)

8.1.3 To fill the bearing housing with oil, unscrew the oil filler/breather and fill through the orifice. 8.1.4 Grease lubricated pumps and electric motors are supplied pre-greased. 8.1.5 Other drivers and gearboxes, if appropriate, should be lubricated in accordance with their manuals. 8.1.6 Pump bearings are product lubricated. 8.1.7 In the case of product lubricated bearings the source of product supply should be checked against the order. There may be requirements for an external clean supply, particular supply pressure or the commencement of lubrication supply before pump start-up.

As each system has its unique requirements it is recommended that Flowserve is consulted when advice is required.

8.2 Direction of rotation

8 MAKING READY FOR OPERATION

8.2.1 Serious damage can result if the pump is started or run in the wrong direction of rotation.

8.1 Lubrication 8.1.1 For oil lubricated pumps, fill the bearing housing with correct grade of oil to the correct level, ie sight glass or constant level oiler bottle.

8.2.2 The pump is shipped with the coupling element removed. Ensure the direction of rotation of the motor is correct before fitting the coupling element. Direction of rotation must correspond to the direction arrow. 8.2.3 Ensure the pump is given the same rotation as the pump direction arrow. 8.2.4 If maintenance work has been carried out to the site's electricity supply, the direction of rotation should be re-checked, as above, in case the supply phasing has been altered.

When fitted with a constant level oiler, the bearing housing should be filled by unscrewing or hinging back the transparent bottle and filling the bottle with oil. Where an adjustable body Denco oiler is fitted this should be set to the height shown in the following diagram.

8.3 Guarding Guarding is supplied fitted to the pump set. If this has been removed or disturbed ensure that all the protective guards are securely refitted. 12 ®

INSTRUCTION MANUAL CPXS & CPXNS ENGLISH 8.4 Open impeller clearance The impeller clearance is set in the factory. This may require adjustment because of piping attachment or increase in temperatures. For setting instructions refer to the Operation and maintenance section of this manual.

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Vibration velocity mm/s rms (unfiltered) Normal (N) Alarm (N x 1.25) Shutdown trip (N x 2.0)

≤ 3.0

> 15kW & vertical ≤ 4.5

≤ 3.8 ≤ 6.0

≤ 5.6 ≤ 9.0

≤ 15kW

10.2 Stop/start frequency Generally 6 stop/starts per hour may be satisfactory. Refer frequent stop/starting to the motor manufacturer.

8.5 Primary and auxiliary supplies Ensure all electrical, hydraulic, pneumatic, sealant and lubrication systems (as applicable) are connected and operational.

STANDBY PUMPS SHOULD BE RUN ALTERNATELY.

8.6 Filling and priming

11 STOPPING AND SHUTDOWN

8.6.1 Ensure inlet pipe and pump casing are completely full of liquid before starting continuous duty operation.

11.1 Close the outlet valve, but ensure that the pump runs in this condition for no more than a few seconds. 11.2 Stop the pump.

8.6.2 Priming may be carried out with an ejector, vacuum pump interceptor or other equipment, or by flooding from the inlet source.

11.3 Switch off flushing and/or cooling/heating liquid supplies at a time appropriate to the process.

8.6.3 When in service, pumps using inlet pipes with foot valves may be primed by passing liquid back from the outlet pipe through the pump.

11.4 For prolonged shut-downs and especially when ambient temperatures are likely to drop below freezing point, the pump and any cooling and flushing arrangements must be drained or otherwise protected.

9 STARTING THE PUMP 9.1 Ensure flushing and/or cooling/heating liquid supplies are turned ON, before starting pump.

12 PREVENTATIVE MAINTENANCE AND SERVICING 12.1 Maintenance schedule Our specialist service personnel can help with preventative maintenance records and provide condition monitoring for temperature and vibration to identify the onset of potential problems.

9.2 CLOSE the outlet valve. 9.3 OPEN all inlet valves. 9.4 Prime the pump.

12.2 Routine inspection (daily/weekly) The following checks should be made and the appropriate action taken to remedy any deviations: • Check operating behaviour; ensure noise, vibration and bearing temperatures are normal. • Check the level and condition of oil lubricant. On grease lubricated pumps, check running hours since last recharge of grease or complete grease change. When "sealed for life" bearings are fitted it is recommended they are renewed every 12,000 hours running life or every 2 years, whichever is the sooner. • Check any auxiliary supplies eg heating/cooling (if fitted) are functioning correctly. • Refer to the manuals of any associated equipment for routine checks needed. • Pumps having ferrous wetted components may rust internally if stood for periods longer than say 2 weeks. In such cases it is recommended that the pump shaft be turned a few revolutions at least once a week to break any rust or algae which may have built up in the clearances between rotating parts. On units where the shaft is accessible it may be turned by hand. In other cases a flick of the starter is permissible after ensuring that the pump casing is full of liquid to prevent seals, bearings etc running dry.

9.5 Start motor and check outlet pressure. 9.6 If pressure is satisfactory, slowly OPEN outlet valve. 9.7 Do not run the pump with the outlet valve closed for a period longer than 30 seconds. 9.8 If NO pressure, or LOW pressure, STOP the pump. Refer to fault finding chart for fault diagnosis. 10 RUNNING 10.1 Normal vibration levels, alarm and trip levels For guidance, pumps generally fall under a classification for rigid support machines within the International rotating machinery standards and the recommended guideline levels below are based on those standards. Alarm and trip values for installed pumps should be based on the actual measurements taken on the pump in the fully commissioned condition. Measuring vibration at regular intervals will then show any deterioration in pump or system operating conditions. 13

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INSTRUCTION MANUAL CPXS & CPXNS ENGLISH 12.3 Periodic inspection (6 monthly) • Check foundation bolts for security of attachment and corrosion. • Check pump running records for hourly usage to determine if bearing lubricant requires changing. • The coupling should be checked for correct alignment and worn driving elements. • Refer to the manuals of any associated equipment for periodic checks needed.

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13.1.9 The dismantled magnet assemblies have a very strong attraction. They should be handled separately at a safe distance from each other and stored in a clean area. 13.1.10 The units are designed so that the outer assembly or the inner assembly can be removed without disturbing the pump end. This allows the outer rotor and the ball bearings of the bearing housing to be examined.

12.4 Lubrication data

13.1.11 Dismantling frame mounted outer assembly • Loosen the bearing housing footbolts. • Remove the 4 bolts holding the bearing housing to the casing cover. • Insert 2 bolts into the bearing housing flange threaded holes. M14 x 60mm for Model 80 M14 x 75mm for Model 100 M12 x 140mm for Model 150 • Alternately jack the 14mm bolts into the flange, 32mm (1.25in) for frame 80 and 53mm (2.1in) for frame 100. This will release the outer rotor from the flux of the inner rotor.

12.4.1 Oil lubricated bearings • Normal oil change intervals are 4000 operating hours. For pumps on hot service or in severely damp or corrosive atmosphere, the oil will require changing more frequently. Lubricant and bearing temperature analysis can be useful in optimising lubricant change intervals. • The lubricating oil should be a high quality oil having oxidisation and foam inhibitors, or synthetic oil. • The bearing temperature may be allowed to rise to 50°C above ambient, but should not exceed 82°C (API 610 limit). A continuously rising temperature, or an abrupt rise, indicate a fault. 12.4.2 Grease lubricated bearings • The bearings are sealed for life. 13 DISMANTLING AND ASSEMBLY 13.1 Dismantling 13.1.1 Refer to Safety section before dismantling the pump.

• • •

13.1.2 Before dismantling the pump for overhaul, ensure genuine Flowserve replacement parts are available. 13.1.3 Refer to Sectional arrangement drawings for identification of parts.

•

13.1.4 The "back pullout" design of these units enables the pump casing to be left in line. The outer assembly or inner assembly can be removed without disturbing the pump end thus allowing examination of the outer rotor and bearing housing ball bearings.

Slide the outer assembly out past the shell assembly. Secure the outer assembly in a horizontal position. Using a dial indicator, determine the play of the outer magnet carrier within the bore. If there is contact between the outer carrier and the skid ring, (if fitted), then the ball bearings need to be replaced. The nominal diametral clearance of the outer rotor to the skid ring is 1mm.

13.1.2 Removal of outer rotor • The outer rotor is removed by first removing the pipe plug from the side of the housing.

13.1.5 Refer to Tools required, section 16 for instruments needed during dismantling and assembly procedures. These tools are not supplied with the pump but can be ordered from Flowserve if required. 13.1.6 Lock and tag power source. 13.1.7 Ensure the work area is clean of grease, oil and metallic chips or dust. Ferritic dust will attract to magnetic assemblies.

•

13.1.8 Drilling, grinding or machining should NOT be attempted near the work area.

Rotate the outer rotor so that its 14mm diam. hole aligns with the plug hole. Insert a bar or screw into the hole provided in the bearing housing and outer rotor to lock the rotor.

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INSTRUCTION MANUAL CPXS & CPXNS ENGLISH • • • •

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Place a coupling hub and key onto the shaft coupling end and loosen the outer magnet assembly. (NB: outer rotor is right hand thread.) Continue to unscrew and remove the outer rotor from the large bore of the bearing housing. Scuff marks on the skid ring can be removed with a light file. If the skid ring has excessive scuffing, it can be removed by first sawing a small cut in the ring, horizontally. Then place the end of a chisel under the outer diameter and tap one side of the cut up and over the other side. Continue hammering inwardly until the ring comes loose. Clean the groove into which the skid ring fits. Remove the four screws fastening the bearing end cover to the bearing housing face. Remove cover and gasket. Slide the bearing shaft assembly out of the bearing housing.

• • • • • •

•

Inspect both inboard and outboard ball bearings. Replace as necessary.

•

13.1.3 Dismantling of casing cover assembly • Loosen and remove the screws holding the casing and casing cover assembly together. • Pull the casing cover assembly out of the casing and secure horizontally in a vice. On some of the heavier assemblies it is advisable to fit two studs into the top two holes in the casing to temporarily support the casing cover, whilst getting a firm grip. • Remove the 6mm hexagon socket head containment shell capscrews and washers. • Remove the containment shell(s) and containment shell gasket. Discard the gasket (and O-ring if dual containment).

• • • • • •

C957KH025 - 12/02

section 16 for proper sizes. (WARNING: When removing the shaft nut, residual fluid may be in the undercut.) Rotate the inner rotor so that the shaft key is at 12 o'clock position. Slide the assembly off the pump shaft and place in a clean, non-ferrous area. Remove the inner magnet assembly, key and shims from the shaft. Slide the shaft, complete with impeller, out of the casing cover. Inspect the coated surface of the shaft, or sleeves, if fitted. If damaged, it should be replaced. The impeller/shaft assembly may now be disassembled after first removing the sleeves and spacer, if fitted. Secure the pump shaft in a vice in the vertical position using soft jaws with a pin through the bypass holes. Care must be exercised so that the coating is not damaged.

Loosen impeller using a strap wrench or similar tool. (NB: right hand thread.) Inspect BOTH sleeve bearing-bushings in the casing cover. Polishing in the bores and thrust faces is normal. No removal is required. If either bushing appears to be cracked, chipped or severely worn, remove by laying the casing cover flat face uppermost. Remove two setscrews, item 53K, if fitted (see drawings in section 21.5 and 21.6). Using an arbor, press out the bushings. Remove tolerance rings and discard. Remove the thrust collar from the inner magnet rotor. The collar is a loose fit in the carrier bore. Remove the thrust collar gasket, discard and replace.

13.2 Examination of parts Used parts must be inspected before assembly to ensure the pump will subsequently run properly. In particular, fault diagnosis is essential to enhance pump and plant reliability. THE MAGNETS MUST BE KEPT AT A SAFE DISTANCE FROM OTHER PARTS AND TOOLS. • • Frame 80

•

Frame 100

•

Loosen and remove the locknut using the appropriate socket and spanner wrench pair. (NB: locknut has left hand thread.) A handle extension may be required. See Tools required,

•

Clean the internal pump parts thoroughly. Inspect for excessive wear, pitting, corrosion, erosion or damage and any sealing surface irregularities. Replace as necessary. For units equipped with a wash flow strainer, be sure to clean the filter removing any debris which may be blocking the strainer holes. Clean lubrication holes in the casing cover, inner magnet carrier and shaft.

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INSTRUCTION MANUAL CPXS & CPXNS ENGLISH • • • • • •

13.4 Assembly

On the casing cover, inspect the injection, vent, drain and return holes. Clean if necessary. The balance holes of the inner rotor should also be inspected and cleared of any debris. Replace all gaskets and O-rings. Check the driver manufacturer's instructions. As a minimum, check the bearings and shaft for straightness. The lubricant, bearings and bearing seals in the motor should be inspected for contamination and damage. Ensure all lubrication passageways in the bearing housing are clean and free from damage.

13.4.1 To assemble the pump consult the Sectional drawings. 13.4.2 Ensure threads, gasket and O-ring mating faces are clean. Apply thread sealant to non-face sealing pipe thread fittings. 13.4.3 Outer assembly - bearing housing assembly • If removed, replace the skid ring. • If the radial ball bearings are found to be damaged, press 2 new bearings onto the shaft. (NB: be sure to press on only the inner race of the bearing whilst pressing it onto the shaft.) Press bearings up to the shaft shoulders. • Install the bearing/shaft assembly into the bore of the bearing housing. • Seat the end cover gasket. • Bolt the bearing end cover to the bearing housing face. (NB: be sure that the oil return grooves on the gasket and end cover line up.) • Tighten screws to 13Nm torque. • Turn the shaft coupling end to ensure freedom of rotation. • Install the flinger over the shaft. (NB: be sure that the flinger is not pressed down hard against the bearing end cover.) • Position the bearing housing horizontally. • Coat the outer magnet rotor threads with antiseize compound. • Insert the outer magnet rotor into the large bore of the bearing housing and screw it onto the frame shaft. (NB: right hand thread.) • Insert a bar or screw into the holes provided in the bearing housing and outer rotor to lock the rotor. • Place a coupling hub and key onto the shaft coupling end and torque the outer magnet assembly as follows: Frame 80 54Nm Frame 100 54Nm • Remove the bar or screw and check shaft for freedom of rotation. • Re-install the pipe plug.

13.3 Magnets 13.3.1 Demagnetization of the magnet material can be the result of either high operating temperatures around the magnet assemblies or decoupled magnets operating around a metallic containment shell. 13.3.2 High ambient temperatures are detrimental to the attraction properties of the magnets. 13.3.3 The inner magnet assembly is most susceptible to high operating temperatures and cannot tolerate operation above its upper critical temperature limit. 13.3.4 NB: If decoupling has occurred or if a system upset has caused the temperature limits to be exceeded, the original strength of the magnets may have decreased. The following torque test procedure should be followed in such a situation. 13.3.5 Magnet torque test procedure • Remove the casing from the pump. • Lock the outer rotor assembly in position. Insert bolt in assembly hole. • Remove the impeller by using a strap wrench around the periphery of the impeller. Turn counterclockwise. • Install a shaft adapter on the shaft threaded connection. Models 80 and 100 M22-1.5 pitch Model 150 M30-1.5 pitch • Secure the bearing housing on a stable worktable. • Use a torque wrench on the nut and turn clockwise to measure the force required to break the magnetic coupling. Adjust the wrench setting such that the torque value is determined prior to breaking the magnetic couple. This is the torque capability of the magnetic coupling.

Model 80 80 100 100 150 150 150

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13.4.4 Inner assembly - casing cover assembly • If the sleeve bearing bushings were not removed during dismantling, proceed to the Rotor assembly section of this manual. • The inboard tolerance ring must be trimmed to length using tin snips. This is due to the difference in diameter of the 2 bushings. Trim off 3 corrugations prior to placement in the casing cover. • There must be a minimum gap of 1.5mm between the cut ends of the tolerance ring before assembly of bushes. • Install the cut tolerance ring into the inner bore. (NB: rotate the ring to guarantee that it is secure.) • Position the casing cover with "top" designation at 12 o'clock position.

Factory torque specification Minimum torque (Nm) Series Neodymium Samarium Cobalt 8 13 11 15 27 24 25 45 40 50 90 80 50 80 70 100 160 140 150 240 210

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INSTRUCTION MANUAL CPXS & CPXNS ENGLISH •

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•

Insert the front bearing bushing so that the lubrication groove is at a 9 o'clock position. The face with the white spot should be visible. The opposite face is lapped and should locate against the shoulder in the casing cover. Press in the front bearing bushing up to the shoulder in the casing cover. Fit bearing bush spacer (if appropriate). Insert the second tolerance ring, rotating it for a secure fit. Insert the rear bearing bushing until resistance is felt. Position the bushing so that one lubrication groove is at 6 o'clock position and press in up to the shoulder in the casing cover. The face with the white spot should mate up to the shoulder in the casing cover. If bearing bush spacer has been fitted, secure with two radial socket head setscrews.

• • • •

Install 0.8mm shim on the pump shaft, between inner rotor and shaft shoulder. Install the inner rotor key. Slide the inner rotor onto the shaft -pump end. Thread the locknut onto the shaft-pump end. (NB: left hand threads.) Tighten the locknut and torque up to the values given in section 4.4.

13.4.6 Impeller setting procedure • Position the pump casing with the suction flange facing down on the bench. • Install the casing gasket followed by the inner rotor/casing cover assembly. • Tighten the casing bolts. • Locate a dial indicator on top of the pump shaft to enable its vertical movement to be recorded.

13.4.5 Rotor assembly • Thoroughly clean and degrease impeller and shaft threads. • Install the impeller on the pump shaft, after first applying anti-seize compound (which does not contain copper) at the impeller to shaft thread to assist in subsequent removal. • Secure the pump shaft in a vice in the vertical position using soft jaws with a pin through the bypass holes. Care must be exercised so that the coating is not damaged. • Tighten the impeller using a strap wrench or similar tool. • If fitted, slide sleeves and spacer onto shaft. • Position the casing cover upright with the through bore in the horizontal plane and "TOP" designation at 12 o'clock. • Slide the shaft through the casing cover. • Diametral journal bearing clearances are 0.08 to 0.13mm. Extreme care must be exercised.

•

•

Loosen inner rotor locknut and record shaft drop when the impeller touches the pump casing. This is a measure of the current front clearance. 0 Ideally this should be done 3 times at 120 intervals and the smallest value recorded. Record this reading on line 1 below. Subtract gasket compression factor (0.15mm) and record on line 3 below: Line

•

• • •

•

C957KH025 - 12/02

Place the thrust collar gasket into the inner rotor. Make sure it is fitted over the drive pin, which should be replaced if damaged. If a new pin is fitted it must be ensured it does not protrude more than 3mm from inner rotor face. The top of the pin should not foul the silicon carbide thrust face. Shorten if necessary. Install the thrust collar into the inner rotor. Be sure that the slot in the thrust collar is in alignment with the drive pin. Small spots of grease may be used to hold the gasket and collar in place, if necessary.

Detail

1

Shaft drop

2

Gasket compression factor

3

Line (1 - 2)

4

Design front clearance

5

Shim adjustment required (4 - 3)

mm 0.15

Subtract the design front clearance from the total shaft drop recorded in line 3 to arrive at the shim adjustment. IF POSITIVE - ADD SHIMS IF NEGATIVE - SUBTRACT SHIMS Design front clearances are:

• •

Impeller diameters up to 210mm inclusive

0.3mm

Impeller diameters 211mm to 254mm

0.4mm

Re-install the inner rotor and tighten up the locknut in accordance with torques shown in section 4.4. Check axial float is between 0.5 and 1.5mm.

17 ®

INSTRUCTION MANUAL CPXS & CPXNS ENGLISH 13.4.7 Containment shell assembly - single containment metallic shell • Install the containment shell gasket in the groove in the casing cover. • Install the metallic shell. • Install the backing ring. • Install and tighten the 6mm capscrews to 11Nm.

C957KH025 - 12/02

13.4.10 Final bearing housing assembly • Install the 2 jack bolts, (previously used for disassembly), into the bearing housing and thread in until they extend as follows: Model 80 26 mm Model 100 45 mm Model 150 95 mm • It is recommended that a liquid sealant, Hylomar Universal Blue or equivalent, is applied between the bearing housing and casing cover. (73C on Sectional arrangement refers.) • Position the outer assembly over the shell until the jacking screws rest against the casing cover. • Back out the jack screws ALTERNATELY, ensuring spigot between casing cover and bearing housing is fully located and square. • Remove the jacking bolts. • Install the 4 bearing housing to cover bolts and tighten to 54Nm. • Tighten the foot bolt(s). • Re-check the frame shaft for freedom of rotation. • Install the coupling. • Align the coupling (as detailed in section 7.6). • Ensure that all other items have been re-attached and all fasteners tightened, then follow the instructions in the Installation sections of this manual.

13.4.8 Containment shell assembly - single containment peek shell • Install the containment shell gasket in the groove in the casing cover. • Install shell with slotted temperature tap as shown, overleaf. • Install and tighten the 6mm capscrews and washers to 11Nm. It is important that washers are not omitted otherwise excessive damage to PEEK shell flange will occur.

slotted tap for temperature monitoring

13.4.9 Containment shell assembly - dual containment • Install the containment shell gasket in the groove in the casing cover. • Install the metallic shell. • Install dual containment O-ring. • Install PEEK shell with pressure tap as shown. • Install and tighten the 6mm capscrews to 11Nm. It is important that washers are not omitted otherwise excessive damage to PEEK shell flange will occur.

threaded tap for dual containment pressure monitoring

18 ®

INSTRUCTION MANUAL CPXS & CPXNS ENGLISH 14 SPARE PARTS

C957KH025 - 12/02

15 GENERAL ARRANGEMENT DRAWING The typical general arrangement drawing and any specific drawings required by the Contract will be sent to the Purchaser separately. If required these should be obtained from the Purchaser and retained with this manual.

14.1 Ordering of spares Flowserve keep records of all pumps that have been supplied. When ordering spares the following information should be quoted: 1. Pump serial number. 2. Pump size. 3. Part name. 4. Part number. 5. Number of parts required.

16 TOOLS REQUIRED The only special tools required are pin wrenches ('C' spanners) which are available for purchase from Flowserve. Order as follows: Part No. N.919GZ172 for Model 80 units. Part No. N.919GZ173 for Model 100 and 150 units.

14.2 The pump size and serial number are shown on the pump nameplate. 14.3 To ensure continued satisfactory operation, replacement parts to the original design specification should be obtained from Flowserve. Any change to the original design specification (modification or use of a non-standard part) will invalidate the pump's safety certification.

17 CERTIFICATION Any certificates eg materials, hydraulic tests, conformities, Ex protection for an explosive atmosphere, performance test curves etc as determined by the contract requirements, will be sent to the Purchaser separately. If required, copies of these should be obtained from the Purchaser for retention with this manual.

14.4 Storage of spare parts Spares should be stored in a clean dry area away from vibration. Inspection and retreatment of metallic surfaces (if necessary) with preservative is recommended at 6 monthly intervals.

18 SUPPLEMENTARY INSTRUCTION MANUALS See also the supplementary instruction manuals supplied with this manual eg for electric motors, controllers, engines, gearboxes, sealant systems etc.

14.5 Prices of spare parts For spares prices refer to Flowserve at Newark.

19 CHANGE NOTES If changes are made to the pump after supply, this manual will require updating.

14.6 Recommended spares for two years operation Part no.

Number of pumps (including stand-by)

Designation 2

3

4

5

10(+)

Impeller

3

30%

6

Pump shaft

1

2

3

30%

235

Bearing bush - front

1

2

3

30%

237

Bearing bush - rear

1

2

3

30%

37A

Tolerance ring

4

8

12

120%

306

Sleeve (if fitted)

2

4

6

60%

72

Thrust collar

2

4

6

60%

73E

Gasket - thrust collar

4

8

12

120%

52

Drive pin - thrust collar

4

8

12

120%

67

Shim pack

1

2

3

30%

230

Inner rotor

1

2

3

30%

232

Outer rotor

1

2

3

30%

3

30%

16

Shell (see note 1)

2

8/9

2

217A or 217B

1

6/7

1

2

Ball bearing

2

Casing gasket

4

6

8

9

10

100%

73B & 119

Shell O-ring set (see note 2)

2

3

4

5

6

60%

73D & 73C

Remaining gasket set

1

73A

252 Skid ring (if fitted) Note 1: 217A PEEK (polymer). 217B metallic. Note 2: 73B primary O-ring. 119 secondary O-ring if fitted.

4

6

2 1

60%

3 2

30% 3

30%

19 ®

INSTRUCTION MANUAL CPXS & CPXNS ENGLISH

C957KH025 - 12/02

20 OPERATING DIFFICULTIES SYMPTOMS PUMP OVERHEATS AND SEIZES BALL BEARINGS HAVE SHORT LIFE ⇓ PUMP VIBRATES OR IS NOISY ⇓ LIQUID LUBRICATED BEARINGS HAVE SHORT LIFE ⇓ LEAKAGE FROM PUMP ⇓ PUMP REQUIRES EXCESSIVE POWER ⇓ PUMP LOSES PRIME AFTER STARTING ⇓ INSUFFICIENT PRESSURE DEVELOPED ⇓ INSUFFICIENT CAPACITY DELIVERED ⇓ PUMP DOES NOT DELIVER LIQUID ⇓ ⇓

SUCTION TROUBLES Pump not primed. Pump or suction pipe not completely filled with liquid. Suction lift too high. Insufficient margin between suction pressure and vapour pressure. Excessive amount of air or gas in liquid. Air or vapour pocket in suction line. Air leaks into suction line. Air leaks into pump through casing and pipework gaskets. Foot valve too small. Foot valve partially clogged. Inlet of suction pipe insufficiently submerged.

l l l l l l

l l l l l l l l l l l

SYSTEM TROUBLES Speed too low. Speed too high. Total head of system higher than head of pump. Total head of system lower than pump design head. Specific gravity of liquid different from design. Viscosity of liquid differs from that for which designed. Operation at very low capacity or pump run dry. Operation at high capacity.

l l l

l

l l

l

l

l

l l l

l

l

l l l

l l l

l

MECHANICAL TROUBLES Misalignment due to pipe strain. Improperly designed foundation. Shaft bent. Rotating part rubbing on stationary part internally. Bearings worn. Wearing ring surfaces worn. Impeller damaged or eroded. Shell O-ring/casing gasket failure. Magnetic coupling de-coupled. Shell corroded/eroded through. Shaft running off centre because of worn bearings or misalignment. Impeller out of balance resulting in vibration. Pump was run dry. Internal misalignment due to improper repairs causing impeller to rub. Excessive thrust caused by a mechanical failure inside the pump. Excessive grease in ball bearings. Lack of lubrication for bearings. Improper installation of bearings (damage during assembly, incorrect assembly, wrong type of bearing etc). Damaged bearings due to contamination. MOTOR ELECTRICAL PROBLEMS Wrong direction of rotation. Motor running on 2 phases only. Motor running too slow, check terminal box.

l l l l l

l l

l

l l l l

l

l l

l l

l l l

l

l l l l l

l

l

l l l

l

l l l

l l

l l

l l

l l

l l l l

l

l

l l l l

l

l

l

l

l l

l l

l

l l l l l l l

l l l

l

l l

l l

l

20 ®

INSTRUCTION MANUAL CPXS & CPXNS ENGLISH

C957KH025 - 12/02

21 SECTIONAL ARRANGEMENT DRAWINGS AND PARTS LISTS 21.1 CPXS - frame mounted - coated shaft - peek shell

Item 1 2 6 12 16 19 19A 19B* 19C* 19D* 19E* 19F* 30A

Description Casing Impeller Pump shaft Drive shaft Ball bearing Bearing housing Support foot Oil filler plug Constant level oiler Sight glass Drain plug (magnetic) Drain plug - bearing housing Plug - casing drain

37A 46 52 53B 53C 53D** 53F 53J 62A 66 67 72 73A 73B

Tolerance ring Key - coupling Thrust bearing drive pin Capscrew & washer - shell Bolt - casing Bolt - casing cover/bearing housing Bolt - end cover Bolt - bearing housing foot Flinger Inner rotor lock nut Shim pack Thrust collar Gasket - casing Gasket - shell

73C 73D 73E 76 123 217A 230 232 235 237 239 252*

Gasket - housing Gasket - end cover Gasket - thrust collar Key - inner rotor assembly End cover Shell - PEEK Inner-magnet rotor Outer-magnet rotor Bearing bush - front Bearing bush - rear Casing cover Skid ring

Notes: * When supplied. ** Not shown.

CPXNS CASING HAS CENTRE-LINE MOUNTING FEET.

21 ®

INSTRUCTION MANUAL CPXS & CPXNS ENGLISH

C957KH025 - 12/02

21.2 CPXS - frame mounted - sleeved shaft - peek shell

Item 1 2 6 12 16 19 19A 19B* 19C* 19D* 19E* 19F* 30A

Description Casing Impeller Pump shaft Drive shaft Ball bearing Bearing housing Support foot Oil filler plug Constant level oiler Sight glass Drain plug (magnetic) Drain plug - bearing housing Plug - casing drain

37A 46 52 53B 53C 53D** 53F 53J 62A 66 67 72 73A 73B

Tolerance ring Key - coupling Thrust bearing drive pin Capscrew & washer - shell Bolt - casing Bolt - casing cover/bearing housing Bolt - end cover Bolt - bearing housing foot Flinger Inner rotor lock nut Shim pack Thrust collar Gasket - casing Gasket - shell

73C 73D 73E 76 123 217A 230 232 235 237 239 252* 306

Gasket - housing Gasket - end cover Gasket - thrust collar Key - inner rotor assembly End cover Shell - PEEK Inner-magnet rotor Outer-magnet rotor Bearing bush - front Bearing bush - rear Casing cover Skid ring Sleeve bearing

Notes: * When supplied. ** Not shown.

CPXNS CASING HAS CENTRE-LINE MOUNTING FEET.

22 ®

INSTRUCTION MANUAL CPXS & CPXNS ENGLISH

C957KH025 - 12/02

21.3 CPXS - frame mounted - coated shaft - dual containment

Item 1 2 6 12 16 19 19A 19B* 19C* 19D* 19E* 19F* 30A 37A

Description Casing Impeller Pump shaft Drive shaft Ball bearing Bearing housing Support foot Oil filler plug Constant level oiler Sight glass Drain plug (magnetic) Drain plug - bearing housing Plug - casing drain Tolerance ring

46 52 53B 53C 53D** 53F 53J 62A 66 67 72 73A 73B 73C 73D

Key - coupling Thrust bearing drive pin Capscrew & washer - shell Bolt - casing Bolt - casing cover/bearing housing Bolt - end cover Bolt - bearing housing foot Flinger Inner rotor lock nut Shim pack Thrust collar Gasket - casing Gasket - shell Gasket - housing Gasket - end cover

73E 76 119 123 217A 217B 230 232 235 237 239 252*

Gasket - thrust collar Key - inner rotor assembly O-ring dual containment End cover Shell - PEEK Shell - metallic Inner-magnet rotor Outer-magnet rotor Bearing bush - front Bearing bush - rear Casing cover Skid ring

Notes: * When supplied. ** Not shown.

CPXNS CASING HAS CENTRE-LINE MOUNTING FEET.

23 ®

INSTRUCTION MANUAL CPXS & CPXNS ENGLISH

C957KH025 - 12/02

21.4 CPXS - frame mounted - sleeved shaft - dual containment

Item 1 2 6 12 16 19 19A 19B* 19C* 19D* 19E* 19F* 30A 37A

Description Casing Impeller Pump shaft Drive shaft Ball bearing Bearing housing Support foot Oil filler plug Constant level oiler Sight glass Drain plug (magnetic) Drain plug - bearing housing Plug - casing drain Tolerance ring

46 52 53B 53C 53D** 53F 53J 62A 66 67 72 73A 73B 73C 73D

Key - coupling Thrust bearing drive pin Capscrew & washer - shell Bolt - casing Bolt - casing cover/bearing housing Bolt - end cover Bolt - bearing housing foot Flinger Inner rotor lock nut Shim pack Thrust collar Gasket - casing Gasket - shell Gasket - housing Gasket - end cover

73E 76 78 119 123 217A 217B 230 232 235 237 239 252* 306

Gasket - thrust collar Key - inner rotor assembly Spacer sleeve O-ring dual containment End cover Shell - PEEK Shell - metallic Inner-magnet rotor Outer-magnet rotor Bearing bush - front Bearing bush - rear Casing cover Skid ring Sleeve bearing

Notes: * When supplied. ** Not shown.

CPXNS CASING HAS CENTRE-LINE MOUNTING FEET.

24 ®

INSTRUCTION MANUAL CPXS & CPXNS ENGLISH

C957KH025 - 12/02

21.5 CPXS - frame mounted - coated shaft - metal shell

Item 1 2 6 7 12 16 19 19A 19B* 19C* 19D* 19E* 19F* 30A

Description Casing Impeller Pump shaft Back-up ring Drive shaft Ball bearing Bearing housing Support foot Oil filler plug Constant level oiler Sight glass Drain plug (magnetic) Drain plug - bearing housing Plug - casing drain

37A 46 52 53B 53C 53D** 53F 53J 53K* 62A 66 67 72 73A 73B

Tolerance ring Key - coupling Thrust bearing drive pin Capscrew & washer - shell Bolt - casing Bolt - casing cover/bearing housing Bolt - end cover Bolt - bearing housing foot Setscrew - bearing spacer Flinger Inner rotor lock nut Shim pack Thrust collar Gasket - casing Gasket - shell

73C 73D 73E 76 78B* 123 217B 230 232 235 237 239 252*

Gasket - housing Gasket - end cover Gasket - thrust collar Key - inner rotor assembly Bearing spacer End cover Shell - metallic Inner-magnet rotor Outer-magnet rotor Bearing bush - front Bearing bush - rear Casing cover Skid ring

Notes: * When supplied. ** Not shown.

CPXNS CASING HAS CENTRE-LINE MOUNTING FEET.

25 ®

INSTRUCTION MANUAL CPXS & CPXNS ENGLISH

C957KH025 - 12/02

21.6 CPXS - frame mounted - sleeved shaft - metal shell

Item 1 2 6 7 12 16 19 19A 19B* 19C* 19D* 19E* 19F* 30A 37A

Description Casing Impeller Pump shaft Back-up ring Drive shaft Ball bearing Bearing housing Support foot Oil filler plug Constant level oiler Sight glass Drain plug (magnetic) Drain plug - bearing housing Plug - casing drain Tolerance ring

46 52 53B 53C 53D** 53F 53J 53K* 62A 66 67 72 73A 73B 73C 73D

Key - coupling Thrust bearing drive pin Capscrew & washer - shell Bolt - casing Bolt - casing cover/bearing housing Bolt - end cover Bolt - bearing housing foot Setscrew - bearing spacer Flinger Inner rotor lock nut Shim pack Thrust collar Gasket - casing Gasket - shell Gasket - housing Gasket - end cover

73E 76 78 78B* 123 217B 230 232 235 237 239 252* 306

Gasket - thrust collar Key - inner rotor assembly Spacer sleeve Bearing spacer End cover Shell - metallic Inner-magnet rotor Outer-magnet rotor Bearing bush - front Bearing bush - rear Casing cover Skid ring Sleeve bearing

Notes: * When supplied. ** Not shown.

CPXNS CASING HAS CENTRE-LINE MOUNTING FEET.

26 ®

INSTRUCTION MANUAL CPXS & CPXNS ENGLISH

C957KH025 - 12/02

22 CPXS - PARTS INTERCHANGEABILITY Drive shaft

Bearing housing cover & bearings

Shell single/dual containent

Magnetic drive

Pump shaft & journals

Casing cover & bushes

Pump casing & impeller

Notes: 1. * ball bearings and bearing cover common on models 80 and 100. 2. All Ø125 and Ø160 impellers have deeper back vanes than standard CPX and are therefore not interchangeable.

27 ®

INSTRUCTION MANUAL CPXS & CPXNS ENGLISH

C957KH025 - 12/02

Europe, Middle East & Africa Flowserve Limited (Pump Division) Harley House, 94 Hare Lane Claygate, Esher, Surrey KT10 0RB United Kingdom

Latin America Flowserve S.A. de C.V. Avenida Paseo de la Reforma #30 2nd Floor, Colonia Juarez Centro Mexico, D.F.Z.C. 06040

Tel +44 (0)1372 463 700 Fax +44 (0)1372 460 190

Tel +52 5705 5526 Fax +52 5705 1125

USA and Canada Flowserve Corporation (Pump Division) Millennium Center, 222 Las Colinas Blvd. 15th Floor, Irving, TX 75039-5421, USA

Asia Pacific Flowserve Pte Ltd (Pump Division) 200 Pandan Loop, #06-03/04 Pantech 21, Singapore 128388

Tel +1 972 443 6500 Toll free 800 728 PUMP (7867) Fax +1 972 443 6800

Tel +65 775 3003 Fax +65 779 4607

Visit our web site at: www.flowserve.com

Your Flowserve factory contact:

Your local Flowserve representative:

Flowserve Pumps Limited PO Box 17, Newark Notts NG24 3EN United Kingdom Telephone (24 hours) +44 (0)1636 494 600 Sales & Admin Fax +44 (0)1636 705 991 Repair & Service Fax +44 (0)1636 494 833 E.mail [email protected]

To find your local Flowserve representative, please use the Sales Support Locator System found at www.flowserve.com

28 ®

®

Pump Division

Types:

LR, LRV, LLR and LR-S

CENTRIFUGAL PUMPS

USER INSTRUCTIONS: INSTALLATION, OPERATION, MAINTENANCE User Instructions C953KH013 - 09/03 (E)

LR, LRV, LLR and LR-S USER INSTRUCTIONS ENGLISH C953KH013 - 09/03 ®

CONTENTS

PAGE

PAGE

1 INTRODUCTION AND SAFETY............................4 1.1 General............................................................4 1.2 CE marking and approvals ..............................4 1.3 Disclaimer........................................................4 1.4 Copyright .........................................................4 1.5 Duty conditions ................................................4 1.6 Safety ..............................................................5 1.7 Warning labels summary.................................8 1.8 Specific machine performance ........................5 1.9 Noise level .......................................................9

7 FAULTS; CAUSES AND REMEDIES................... 32

2 TRANSPORT AND STORAGE............................10 2.1 Consignment receipt and unpacking .............10 2.2 Handling ........................................................10 2.3 Lifting .............................................................10 2.4 Storage ..........................................................10 2.5 Recycling and end of product life ..................10 3 PUMP DESCRIPTION .........................................11 3.1 Configurations ...............................................11 3.2 Name nomenclature ......................................11 3.3 Design of major parts ....................................11 3.4 Performance and operating limits..................12 4 INSTALLATION ....................................................12 4.1 Location .........................................................12 4.2 Part assemblies .............................................12 4.3 Foundation.....................................................12 4.4 Grouting.........................................................13 4.5 Initial alignment..............................................13 4.6 Piping.............................................................14 4.7 Final shaft alignment check ...........................16 4.8 Electrical connections....................................16 4.9 Protection systems ........................................16

8 PARTS LISTS AND DRAWINGS......................... 34 8.1 Sectional drawings – LR single entry impeller, grease lubricated, gland packed (pump sizes 2.5LR-10 and 2.5LR-13 only) ... 34 8.2 Sectional drawings – LR double entry impeller, grease lubricated, gland packed..... 36 8.3 Sectional drawings – LLR grease lubricated, gland packed ............................... 38 8.4 Sectional drawings – LR-S double entry impeller, grease lubricated, gland packed..... 40 8.5 Sectional drawings – LRV double entry impeller, grease lubricated, mechanical seal, SiC bearing ........................................... 41 8.6 Interchangeability chart for LR, LLR and LR-S ....................................................... 42 8.7 Interchangeability chart for LRV.................... 43 8.8 General arrangement drawing ...................... 43 9 CERTIFICATION ................................................. 43 10 OTHER RELEVANT DOCUMENTATION AND MANUALS ............................................... 43 10.1 Supplementary User Instructions................ 43 10.2 Change notes.............................................. 43 10.3 Additional sources of information ................ 43

5 COMMISSIONING, START-UP, OPERATION AND SHUTDOWN............................................17 5.1 Pre-commissioning procedure.......................17 5.2 Pump lubricants.............................................18 5.3 Direction of rotation .......................................19 5.4 Guarding........................................................19 5.5 Priming and auxiliary supplies.......................19 5.6 Starting the pump ..........................................19 5.7 Running the pump .........................................20 5.8 Stopping and shutdown .................................21 5.9 Hydraulic, mechanical and electrical duty .....21 6 MAINTENANCE ...................................................21 6.1 General..........................................................21 6.2 Maintenance schedule ..................................22 6.3 Spare parts ....................................................24 6.4 Recommended spares and consumable items ..........................................24 6.5 Tools required ................................................24 6.6 Fastener torques ...........................................24 6.7 Renewal clearances ......................................24 6.8 Disassembly ..................................................24 6.9 Examination of parts......................................27 6.10 Assembly .....................................................27

Page 2 of 44

LR, LRV, LLR and LR-S USER INSTRUCTIONS ENGLISH C953KH013 - 09/03 ®

INDEX PAGE

PAGE

Additional sources (10.3) .........................................43 Alignment of shafting (4.3, 4.5 and 4.7) Assembly (6.10) .......................................................27 ATEX marking (1.6.4.2)..............................................6 CE marking and approvals (1.2) ................................4 Certification (9) .........................................................43 Change notes (10.2) ................................................43 Clearances (6.7, Renewal clearances) ....................24 Commissioning and operation (5) ............................17 Compliance, ATEX (1.6.4.1) ......................................6 Configurations (3.1)..................................................11 Consumable items (6.4) ...........................................24 Copyright (1.4) ...........................................................4 Design of major parts (3.3) ......................................11 Direction of rotation (5.3)..........................................19 Disassembly (6.8) ....................................................24 Disclaimer (1.3) ..........................................................4 Dismantling (6.8, Disassembly) ...............................24 Drawings (8) .............................................................34 Duty conditions (1.5) ..................................................4 Electrical connections (4.8) ......................................16 End of product life (2.5)............................................10 Examination of parts (6.9) ........................................27 Fastener torques (6.6)..............................................24 Faults; causes and remedies (7)..............................32 Forces and moments (4.6.3) ....................................15 Foundation (4.3).......................................................12 General arrangement drawing (8.8) .........................43 General assembly drawings (8) ...............................34 Grouting (4.4) ...........................................................13 Guarding (5.4) ..........................................................19 Handling (2.2)...........................................................10 Hydraulic, mechanical and electrical duty (5.9) .......21 Inspection (6.2.1 and 6.2.2) .....................................22 Installation (4)...........................................................12 Interchangeability charts (8.6 and 8.7) Lifting (2.3) ...............................................................10 Location (4.1) ...........................................................12 Lubrication (5.1.1, 5.2 and 6.2.3) Lubrication schedule (5.2.4).....................................18 Maintenance (6) .......................................................21 Maintenance schedule (6.2).....................................22 Name nomenclature (3.2) ........................................11 Nameplate (1.7.1) ......................................................8 Operating limits (3.4.1).............................................12 Options (8) ...............................................................34 Ordering spare parts (6.3.1).....................................24 Part assemblies (4.2) ...............................................12 Parts lists (8) ............................................................34 Performance (3.4) ....................................................12 Piping (4.6) ...............................................................14 Pre-commissioning (5.1) ..........................................17 Priming and auxiliary supplies (5.5) .........................19 Protection systems (4.9) ..........................................16

Reassembly (6.10, Assembly) ................................. 27 Receipt and unpacking (2.1).................................... 10 Recommended fill quantities (see 3.4.2) ................. 12 Recommended grease lubricants (5.2.2) ................ 18 Recommended oil lubricants (5.2.1)........................ 18 Recommended spares (6.4) .................................... 24 Recycling (2.5)......................................................... 10 Renewal clearances (6.7) ........................................ 24 Replacement parts (6.3 and 6.4) ............................. 24 Running the pump (5.7) ........................................... 20 Safety action (1.6.3)................................................... 5 Safety markings (1.6.1) ............................................. 5 Safety, protection systems (1.6 and 4.9) Sectional drawings (8) ............................................. 34 Sound pressure level (1.9, Noise level)..................... 9 Sources, additional information (10.3)..................... 43 Spare parts (6.3)...................................................... 24 Specific machine performance (1.8).......................... 9 Starting the pump (5.6) ............................................ 19 Stop/start frequency (5.7.6) ..................................... 20 Stopping and shutdown (5.8)................................... 21 Storage, pump (2.4)................................................. 10 Storage, spare parts (6.3.2)..................................... 24 Supplementary manuals or information sources ..... 43 Supplementary User Instructions (10.1) .................. 43 Tools required (6.5) ................................................. 24 Torques for fasteners (6.6) ...................................... 24 Trouble-shooting (see 7) ......................................... 32 Vibration (5.7.5) ....................................................... 20 Warning labels (1.7.2)................................................ 8

Page 3 of 44

LR, LRV, LLR and LR-S USER INSTRUCTIONS ENGLISH C953KH013 - 09/03 ®

1 INTRODUCTION AND SAFETY

1.3 Disclaimer

1.1 General

Information in these User Instructions is believed to be reliable. In spite of all the efforts of Flowserve Pump Division to provide sound and all necessary information the content of this manual may appear insufficient and is not guaranteed by Flowserve as to its completeness or accuracy.

These Instructions must always be kept close to the product’s operating location or directly with the product. Flowserve’s products are designed, developed and manufactured with state-of-the-art technologies in modern facilities. The unit is produced with great care and commitment to continuous quality control, utilising sophisticated quality techniques, and safety requirements. Flowserve is committed to continuous quality improvement and being at service for any further information about the product in its installation and operation or about its support products, repair and diagnostic services. These instructions are intended to facilitate familiarization with the product and its permitted use. Operating the product in compliance with these instructions is important to help ensure reliability in service and avoid risks. The instructions may not take into account local regulations; ensure such regulations are observed by all, including those installing the product. Always coordinate repair activity with operations personnel, and follow all plant safety requirements and applicable safety and health laws and regulations. These instructions should be read prior to installing, operating, using and maintaining the equipment in any region worldwide. The equipment must not be put into service until all the conditions relating to safety, noted in the instructions, have been met.

1.2 CE marking and approvals

It is a legal requirement that machinery and equipment put into service within certain regions of the world shall conform with the applicable CE Marking Directives covering Machinery and, where applicable, Low Voltage Equipment, Electromagnetic Compatibility (EMC), Pressure Equipment Directive (PED) and Equipment for Potentially Explosive Atmospheres (ATEX). Where applicable, the Directives and any additional Approvals, cover important safety aspects relating to machinery and equipment and the satisfactory provision of technical documents and safety instructions. Where applicable this document incorporates information relevant to these Directives and Approvals. To confirm the Approvals applying and if the product is CE marked, check the serial number plate markings and the Certification. (See section 9, Certification.)

Flowserve manufactures products to exacting International Quality Management System Standards as certified and audited by external Quality Assurance organisations. Genuine parts and accessories have been designed, tested and incorporated into the products to help ensure their continued product quality and performance in use. As Flowserve cannot test parts and accessories sourced from other vendors the incorrect incorporation of such parts and accessories may adversely affect the performance and safety features of the products. The failure to properly select, install or use authorised Flowserve parts and accessories is considered to be misuse. Damage or failure caused by misuse is not covered by Flowserve’s warranty. In addition, any modification of Flowserve products or removal of original components may impair the safety of these products in their use.

1.4 Copyright

All rights reserved. No part of these instructions may be reproduced, stored in a retrieval system or transmitted in any form or by any means without prior permission of Flowserve Pump Division.

1.5 Duty conditions

This product has been selected to meet the specifications of your purchaser order. The acknowledgement of these conditions has been sent separately to the Purchaser. A copy should be kept with these instructions. The product must not be operated beyond the parameters specified for the application. If there is any doubt as to the suitability of the product for the application intended, contact Flowserve for advice, quoting the serial number. If the conditions of service on your purchase order are going to be changed (for example liquid pumped, temperature or duty) it is requested that the user seeks Flowserve’s written agreement before start up.

Page 4 of 44

LR, LRV, LLR and LR-S USER INSTRUCTIONS ENGLISH C953KH013 - 09/03 ®

1.6 Safety 1.6.1 Summary of safety markings These User Instructions contain specific safety markings where non-observance of an instruction would cause hazards. The specific safety markings are: This symbol indicates electrical safety instructions where non-compliance will involve a high risk to personal safety or the loss of life. This symbol indicates safety instructions where non-compliance would affect personal safety and could result in loss of life. This symbol indicates “hazardous and toxic fluid” safety instructions where non-compliance would affect personal safety and could result in loss of life. This symbol indicates safety instructions where non-compliance will involve some risk to safe operation and personal safety and would damage the equipment or property. This symbol indicates explosive atmosphere zone marking according to ATEX. It is used in safety instructions where non-compliance in the hazardous area would cause the risk of an explosion. This sign is not a safety symbol but indicates an important instruction in the assembly process. 1.6.2 Personnel qualification and training All personnel involved in the operation, installation, inspection and maintenance of the unit must be qualified to carry out the work involved. If the personnel in question do not already possess the necessary knowledge and skill, appropriate training and instruction must be provided. If required the operator may commission the manufacturer/supplier to provide applicable training. Always coordinate repair activity with operations and health and safety personnel, and follow all plant safety requirements and applicable safety and health laws and regulations. 1.6.3 Safety action This is a summary of conditions and actions to help prevent injury to personnel and damage to the environment and to equipment. For products used in potentially explosive atmospheres section 1.6.4 also applies. NEVER DO MAINTENANCE WORK WHEN THE UNIT IS CONNECTED TO POWER GUARDS MUST NOT BE REMOVED WHILE THE PUMP IS OPERATIONAL

DRAIN THE PUMP AND ISOLATE PIPEWORK BEFORE DISMANTLING THE PUMP The appropriate safety precautions should be taken where the pumped liquids are hazardous. FLUORO-ELASTOMERS (When fitted.) When a pump has experienced temperatures over 250 ºC (482 ºF), partial decomposition of fluoroelastomers (example: Viton) will occur. In this condition these are extremely dangerous and skin contact must be avoided. HANDLING COMPONENTS Many precision parts have sharp corners and the wearing of appropriate safety gloves and equipment is required when handling these components. To lift heavy pieces above 25 kg (55 lb) use a crane appropriate for the mass and in accordance with current local regulations. THERMAL SHOCK Rapid changes in the temperature of the liquid within the pump can cause thermal shock, which can result in damage or breakage of components and should be avoided. APPLYING HEAT TO REMOVE IMPELLER There may be occasions when the impeller has either been shrunk fit on to the pump shaft or has become difficult to remove due to products of corrosion. If you elect to use heat to remove the impeller, it must be applied quickly to the impeller boss. TAKE GREAT CARE! Before applying heat ensure any residual hazardous liquid trapped between the impeller and pump shaft is thoroughly drained out through the impeller keyway to prevent an explosion or emission of toxic vapour. This must be carried out with the shaft in the vertical position. On some pump sizes a cavity exists in the impeller bore so on occasions a significant volume of liquid may drain out. HOT (and cold) PARTS If hot or freezing components or auxiliary heating supplies can present a danger to operators and persons entering the immediate area action must be taken to avoid accidental contact. If complete protection is not possible, the machine access must be limited to maintenance staff only, with clear visual warnings and indicators to those entering the immediate area. Note: bearing housings must not be insulated and drive motors and bearings may be hot. If the temperature is greater than 68 ºC (175 ºF) or below 5 ºC (20 ºF) in a restricted zone, or exceeds local regulations, action as above shall be taken.

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LR, LRV, LLR and LR-S USER INSTRUCTIONS ENGLISH C953KH013 - 09/03 ®

HAZARDOUS LIQUIDS When the pump is handling hazardous liquids care must be taken to avoid exposure to the liquid by appropriate siting of the pump, limiting personnel access and by operator training. If the liquid is flammable and/or explosive, strict safety procedures must be applied. Gland packing must not be used when pumping hazardous liquids. PREVENT EXCESSIVE EXTERNAL PIPE LOAD Do not use pump as a support for piping. Do not mount expansion joints, unless allowed by Flowserve in writing, so that their force, due to internal pressure, acts on the pump flange. ENSURE CORRECT LUBRICATION (See section 5, Commissioning, startup, operation and shutdown.) START THE PUMP WITH OUTLET VALVE PARTLY OPENED (Unless otherwise instructed at a specific point in the User Instructions.) This is recommended to minimize the risk of overloading and damaging the pump motor at full or zero flow. Pumps may be started with the valve further open only on installations where this situation cannot occur. The pump outlet control valve may need to be adjusted to comply with the duty following the run-up process. (See section 5, Commissioning start-up, operation and shutdown.) NEVER RUN THE PUMP DRY INLET VALVES TO BE FULLY OPEN WHEN PUMP IS RUNNING Running the pump at zero flow or below the recommended minimum flow continuously will cause damage to the pump and mechanical seal. DO NOT RUN THE PUMP AT ABNORMALLY HIGH OR LOW FLOW RATES Operating at a flow rate higher than normal or at a flow rate with no back pressure on the pump may overload the motor and cause cavitation. Low flow rates may cause a reduction in pump/bearing life, overheating of the pump, instability and cavitation/ vibration. 1.6.4 Products used in potentially explosive atmospheres • • • • •

Measures are required to: Avoid excess temperature Prevent build up of explosive mixtures Prevent the generation of sparks Prevent leakages Maintain the pump to avoid hazard

The following instructions for pumps and pump units when installed in potentially explosive atmospheres must be followed to help ensure explosion protection. Both electrical and non-electrical equipment must meet the requirements of European Directive 94/9/EC. 1.6.4.1 Scope of compliance Use equipment only in the zone for which it is appropriate. Always check that the driver, drive coupling assembly, seal and pump equipment are suitably rated and/or certified for the classification of the specific atmosphere in which they are to be installed. Where Flowserve has supplied only the bare shaft pump, the Ex rating applies only to the pump. The party responsible for assembling the pump set shall select the coupling, driver and any additional equipment, with the necessary CE Certificate/ Declaration of Conformity establishing it is suitable for the area in which it is to be installed. The output from a variable frequency drive (VFD) can cause additional heating affects in the motor and so, for pumps sets with a VFD, the ATEX Certification for the motor must state that it is covers the situation where electrical supply is from the VFD. This particular requirement still applies even if the VFD is in a safe area. 1.6.4.2 Marking An example of ATEX equipment marking is shown below. The actual classification of the pump will be engraved on the nameplate. II 2 GD c 135 ºC (T4) Equipment Group I = Mining II = Non-mining Category 2 or M2 = High level protection 3 = normal level of protection Gas and/or Dust G = Gas; D= Dust c = Constructional safety (in accordance with prEn13463-5) Maximum surface temperature (Temperature Class). (See section 1.6.4.3.)

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1.6.4.3 Avoiding excessive surface temperatures ENSURE THE EQUIPMENT TEMPERATURE CLASS IS SUITABLE FOR THE HAZARD ZONE Pumps have a temperature class as stated in the ATEX Ex rating on the nameplate. These are based on a maximum ambient of 40 ºC (104 ºF); refer to Flowserve for higher ambient temperatures. The surface temperature on the pump is influenced by the temperature of the liquid handled. The maximum permissible liquid temperature depends on the temperature class and must not exceed the values in the table that follows. The temperature rise at the seals and bearings and due to the minimum permitted flow rate is taken into account in the temperatures stated. Temperature class to prEN 13463-1 T6 T5 T4 T3 T2 T1

Maximum surface temperature permitted 85 °C (185 °F) 100 °C (212 °F) 135 °C (275 °F) 200 °C (392 °F) 300 °C (572 °F) 450 °C (842 °F)

Temperature limit of liquid handled (* depending on material and construction variant - check which is lower) Consult Flowserve Consult Flowserve 115 °C (239 °F) * 180 °C (356 °F) * 275 °C (527 °F) * 400 °C (752 °F) *

The responsibility for compliance with the specified maximum liquid temperature is with the plant operator. If an explosive atmosphere exists during the installation, do not attempt to check the direction of rotation by starting the pump unfilled. Even a short run time may give a high temperature resulting from contact between rotating and stationary components. Where there is any risk of the pump being run against a closed valve generating high liquid and casing external surface temperatures it is recommended that users fit an external surface temperature protection device. Avoid mechanical, hydraulic or electrical overload by using motor overload trips, temperature monitor or a power monitor and make routine vibration monitoring checks. In dirty or dusty environments, regular checks must be made and dirt removed from areas around close clearances, bearing housings and motors.

1.6.4.4 Preventing the build up of explosive mixtures ENSURE THE PUMP IS PROPERLY FILLED AND VENTED AND DOES NOT RUN DRY Ensure the pump and relevant suction and discharge pipeline system is totally filled with liquid at all times during the pump operation, so that an explosive atmosphere is prevented. In addition it is essential to make sure that seal chambers, auxiliary shaft seal systems and any heating and cooling systems are properly filled. If the operation of the system cannot avoid this condition the fitting of an appropriate dry run protection device is recommended (for example liquid detection or a power monitor). To avoid potential hazards from fugitive emissions of vapour or gas to atmosphere the surrounding area must be well ventilated. 1.6.4.5 Preventing sparks To prevent a potential hazard from mechanical contact, the coupling guard must be non-sparking and anti-static for Category 2. To avoid the potential hazard from random induced current generating a spark, the earth contact on the baseplate must be used. Avoid electrostatic charge: do not rub non-metallic surfaces with a dry cloth; ensure cloth is damp. The coupling must be selected to comply with 94/9/EC and correct alignment must be maintained. 1.6.4.6 Preventing leakage The pump must only be used to handle liquids for which it has been approved to have the correct corrosion resistance. Avoid entrapment of liquid in the pump and associated piping due to closing of suction and discharge valves, which could cause dangerous excessive pressures to occur if there is heat input to the liquid. This can occur if the pump is stationary or running. Bursting of liquid containing parts due to freezing must be avoided by draining or protecting the pump and ancillary systems. Where there is the potential hazard of a loss of a seal barrier fluid or external flush, the fluid must be monitored. If leakage of liquid to atmosphere can result in a hazard, the installation of a liquid detection device is recommended.

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1.6.4.7 Maintenance to avoid the hazard CORRECT MAINTENANCE IS REQUIRED TO AVOID POTENTIAL HAZARDS WHICH GIVE A RISK OF EXPLOSION The responsibility for compliance with maintenance instructions is with the plant operator.

To avoid potential explosion hazards during maintenance, the tools, cleaning and painting materials used must not give rise to sparking or adversely affect the ambient conditions. Where there is a risk from such tools or materials, maintenance must be conducted in a safe area. It is recommended that a maintenance plan and schedule is adopted. (See section 6, Maintenance.)

1.7 Warning labels summary 1.7.1 Nameplate For details of nameplate, see the Declaration of Conformity. 1.7.2 Warning labels

Oil lubricated units only:

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1.8 Specific machine performance

Pump noise level is dependent on a number of factors - the type of motor fitted, the operating capacity, pipework design and acoustic characteristics of the building. Typical sound pressure levels measured in dB, and A-weighted are shown in the table below (LpfA). The figures are indicative only, they are subject to a +3 dB tolerance, and cannot be guaranteed.

1.9 Noise level

The values are based on the noisiest ungeared electric motors that are likely to be encountered. They represent sound pressure levels at 1 m (3.3 ft) from the directly driven pump, for "free field over a reflecting plane".

For performance parameters see section 1.5, Duty conditions. When the contract requirement specifies these to be incorporated into User Instructions these are included here. Where performance data has been supplied separately to the purchaser these should be obtained and retained with these User Instructions if required. When pump noise level exceeds 85 dB(A) attention must be given to prevailing Health and Safety Legislation, to limit the exposure of plant operating personnel to the noise. The usual approach is to control exposure time to the noise or to enclose the machine to reduce emitted sound. You may have already specified a limiting noise level when the equipment was ordered, however if no noise requirements were defined then machines above a certain power level will exceed 85 dB(A). In such situations consideration must be given to the fitting of an acoustic enclosure to meet local regulations.

If a pump unit only has been purchased, for fitting with your own driver, then the "pump only" noise levels from the table should be combined with the level for the driver obtained from the supplier. If the motor is driven by an inverter it may show an increase in noise level at some speeds. Consult a Noise Specialist for the combined calculation.



For units driven by equipment other than electric motors or units contained within enclosures, see the accompanying information sheets and manuals. Typical sound pressure level, LpfA – (dB, A-weighted) Motor size and speed

3550 r/min Pump Pump and motor only LpfA LpfA

2900 r/min Pump and Pump motor only LpfA LpfA

1750 r/min Pump and Pump motor only LpfA LpfA

kW

(hp)

<0.55 0.75 1.1 1.5 2.2 3 4 5.5 7.5 11 15 18.5 22 30 37 45 55 75 90 110 150 200

(<0.75) (1) (1.5) (2) (3) (4) (5) (7.5) (10) (15) (20) (25) (30) (40) (50) (60) (75) (100) (120) (150) (200) (270)

71 74 74 77 78 81 82 90 90 91 92 92 92 100 100 100 102 100 97 100 101

66 66 68 70 72 74 75 77 78 80 83 83 83 85 86 87 88 90 90 91 92

64 67 67 70 71 74 75 83 83 84 85 85 85 93 93 93 95 95 92 95 96

62 62 64 66 68 70 71 73 74 76 79 79 79 81 82 83 84 86 86 87 88

64 67 67 70 71 74 75 76 77 78 80 80 81 84 84 84 86 88 90 91 91

①

①

①

①

300

(400)

-

-

-

-

500

(670)

-

-

-

-

① ① ①

62 62 64 66 68 74 75 75 76 77 79 79 79 80 80 80 81 81 81 83 83 83 84 85

1450 r/min Pump and Pump motor only LpfA LpfA 63 63 65 66 68 70 71 72 73 74 76 76 77 80 80 80 82 83 85 86 86

① ① ①

① Motors in this range are generally job specific and noise levels should be calculated based on actual equipment installed. For 960 rpm reduce 1450 rpm values by 5 dBA.

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62 62 64 66 68 70 71 71 72 73 75 75 75 76 76 76 77 78 78 79 79 80 81 83

LR, LRV, LLR and LR-S USER INSTRUCTIONS ENGLISH C953KH013 - 09/03 ®

2 TRANSPORT AND STORAGE 2.1 Consignment receipt and unpacking

Immediately after receipt of the equipment it must be checked against the delivery and shipping documents for its completeness and that there has been no damage in transportation. Any shortage and or damage must be reported immediately to Flowserve Pump Division and received in writing within one month of receipt of the equipment. Later claims cannot be accepted. Check any crates, boxes and wrappings for any accessories or spare parts that may be packed separately with the equipment or attached to side walls of the box or equipment. Each product has a unique serial number. Check that this number corresponds with that advised and always quote this number in correspondence as well as when ordering spare parts or further accessories.

2.2 Handling

Boxes, crates, pallets or cartons may be unloaded using fork lift vehicles or slings dependent on their size and construction.

2.3 Lifting To avoid distortion, the pump unit should be lifted as shown.

When there are no specific lifting points on the baseplate

A crane must be used for all pump sets in excess of 25 kg (55 lb.). Fully trained personnel must carry out lifting, in accordance with local regulations. The driver and pump weights are recorded on their respective nameplates or massplates.

2.4 Storage Store the pump in a clean, dry location away from vibration. Leave piping connection covers in place to keep dirt and other foreign material out of pump casing. Turn pump at intervals to prevent brinelling of the bearings and the seal faces, if fitted, from sticking. The pump may be stored as above for up to 6 months. Consult Flowserve for preservative actions when a longer storage period is needed.

2.5 Recycling and end of product life

At the end of the service life of the product or its parts, the relevant materials and parts should be recycled or disposed of using an environmentally acceptable method and local regulations. If the product contains substances that are harmful to the environment, these should be removed and disposed of in accordance with current regulations. This also includes the liquids and or gases that may be used in the "seal system" or other utilities. Make sure that hazardous substances are disposed of safely and that the correct personal protective equipment is used. The safety specifications must be in accordance with the current regulations at all times.

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LR, LRV, LLR and LR-S USER INSTRUCTIONS ENGLISH C953KH013 - 09/03 ®

3 PUMP DESCRIPTION

3.3 Design of major parts

3.1 Configurations

3.3.1 Pump casing The pump has its main casing gasket axial to the shaft allowing maintenance to the rotating element by removing the top half casing. Suction and discharge branches are in the bottom half and therefore remain undisturbed.

The range can have the following configurations:

3.3.2 Impeller The impeller is fully shrouded and may be fitted with optional hub rings.

The LR range of pumps are horizontal split casing volute type centrifugal pumps designed for water works, drainage, general service and circulating applications. They can be used with motor, steam turbine and gasoline or diesel engine drives.

3.3.3 Shaft The large diameter stiff shaft, mounted on bearings, has a keyed drive end. 3.3.4 Pump bearings and lubrication Ball bearings are fitted as standard and may be either oil or grease lubricated. LR single-stage horizontal suction and discharge nozzles. LLR two-stage horizontal suction and discharge nozzles. LR-S single stage horizontal suction and discharge nozzles.

Oil lubrication is only available where the pump shaft is horizontal. The LRV as standard has a liquid lubricated journal bearing fitted at the non-drive end. This bearing is lubricated by pumped product or from an external clean source. 3.3.5 Bearing housing Two grease nipples enable grease lubricated bearings to be replenished between major service intervals. LR-S pumps have sealed for life bearings and cannot be re-greased.

LRV single-stage LR horizontal suction/discharge nozzles, with vertical pump shaft.

3.2 Name nomenclature

The pump size will be engraved on the nameplate typically as below:

6LR-18S

Nominal discharge branch size Configuration – see 3.1 above

For oil lubricated bearings, a constant level oiler is fitted. 3.3.6 Seal housing The design enables one of a number of sealing options to be fitted. 3.3.7 Shaft seal The mechanical seal(s), attached to the pump shaft, seals the pumped liquid from the environment. Gland packing may be fitted as an option on the LR, LR-S, and LLR. 3.3.8 Driver The driver is normally an electric motor. Different drive configurations may be fitted such as internal combustion engines, turbines, hydraulic motors etc driving via couplings, belts, gearboxes, drive shafts etc.

Nominal maximum impeller diameter Tongue and groove casing rings fitted The typical nomenclature above is the general guide to the LR configuration description. Identify the actual pump size and serial number from the pump nameplate. Check that this agrees with the applicable certification provided.

3.3.9 Accessories Accessories may be fitted when specified by the customer.

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3.4 Performance and operating limits

This product has been selected to meet the specifications of your purchase order, see section 1.5.

The following data is included as additional information to help with your installation. It is typical, and factors such as temperature, materials, and seal type may influence this data. If required, a definitive statement for your particular application can be obtained from Flowserve. 3.4.1 Operating limits - 20 to + 150 ºC (- 4 to + 302 ºF) - 20 to + 40 ºC Maximum ambient temperature* (- 4 to +104 ºF) up to 3 % by volume Maximum soft solids in suspension* (refer for size limits) Maximum pump speed refer to the nameplate *Subject to written agreement from Flowserve. Pumped liquid temperature limits*

3.4.2 Pump and impeller data Impeller Nominal Mean Approx. oil minimum radial capacity, wear passage wear ring both ring clearance bearings diameter size mm (in.) mm (in.) * litres (fl. oz.) mm (in.) 1.5LLR-7 6 (0.24) 95.25 (3.75) 0.19 (0.007) 0.16 (5.4) 1.5LLR-10 9 (0.35) 95.25 (3.75) 0.19 (0.007) 0.16 (5.4) 2LLR-9 7.5 (0.29) 103.2 (4.06) 0.22 (0.009) 0.16 (5.4) 2LLR-11 7.5 (0.29) 124 (4.88) 0.22 (0.009) 0.19 (6.4) 3LLR-11 10.5 (0.41) 139.9 (5.51) 0.22 (0.009) 0.19 (6.4) 4LLR-11 17 (0.67) 157.3 (6.19) 0.22 (0.009) 0.19 (6.4) 2.5LR10 9 (0.35) 95.25 (3.75) 0.19 (0.007) 0.16 (5.4) 2.5LR-13 13 (0.51) 123.8 (4.88) 0.22 (0.009) 0.17 (5.8) 3LR-9 8 (0.31) 103.2 (4.06) 0.22 (0.009) 0.16 (5.4) 3LR-12 14.5 (0.57) 123.8 (4.88) 0.22 (0.009) 0.16 (5.4) 4LR-10 16.5 (0.65) 123.8 (4.88) 0.22 (0.009) 0.16 (5.4) 4LR-11 18 (0.71) 123.8 (4.88) 0.22 (0.009) 0.17 (5.8) 4LR-12 12 (0.47) 123.8 (4.88) 0.22 (0.009) 0.17 (5.8) 4LR-14 16 (0.63) 123.8 (4.88) 0.22 (0.009) 0.16 (5.4) 5LR-10 16.5 (0.65) 139.7 (5.5) 0.22 (0.009) 0.16 (5.4) 5LR-13 15 (0.59) 139.7 (5.5) 0.22 (0.009) 0.17 (5.8) 5LR-15 17 (0.67) 139.7 (5.5) 0.22 (0.009) 0.19 (6.4) 5LR-19 17 (0.67) 168.4 (6.63) 0.22 (0.009) 0.19 (6.4) 6LR-10 21 (0.83) 157.2 (6.19) 0.22 (0.009) 0.16 (5.4) 6LR-13 17.5 (0.69) 157.2 (6.19) 0.22 (0.009) 0.19 (6.4) 6LR-16 17.5 (0.69) 157.2 (6.19) 0.22 (0.009) 0.19 (6.4) 6LR-18 23.5 (0.93) 190.5 (7.5) 0.22 (0.009) 0.21 (7.1) 6LR-18S 26.5 (1.04) 215.9 (8.5) 0.13 (0.005) 0.47 (15.9) 8LR-12 22 (0.87) 190.5 (7.5) 0.22 (0.009) 0.19 (6.4) 8LR-14 22 (0.87) 190.5 (7.5) 0.22 (0.009) 0.19 (6.4) 8LR-18S 38 (1.50) 247.7 (9.75) 0.13 (0.005) 0.47 (15.9) 8LR-20 27 (1.06) 228.6 (9.0) 0.22 (0.009) 0.21 (7.1) 8LR-23S 23 (0.91) 235 (9.25) 0.13 (0.005) 0.47 (15.9) 10LR-14 44.5 (1.75) 228.6 (9.0) 0.22 (0.009) 0.21 (7.1) 10LR-14S 42.5 (1.67) 247.7 (9.75) 0.13 (0.005) 0.47 (15.9) 10LR-16 39 (1.54) 228.6 (9.0) 0.22 (0.009) 0.21 (7.1) 10LR-17 41 (1.61) 278 (10.95) 0.22 (0.009) 0.28 (9.46) 10LR-18 22 (0.87) 278 (10.95) 0.22 (0.009) 0.28 (9.46) 10LR-18S 57.5 (2.26) 273.1 (10.75) 0.13 (0.005) 0.47 (15.9) 12LR-14S 58.5 (2.30) 273.1 (10.75) 0.13 (0.005) 0.47 (15.9) * May be up to 0.13 mm (0.005 in.) larger if casing ring and impeller have a tendency to gaul. Pump size

4 INSTALLATION Equipment operated in hazardous locations must comply with the relevant explosion protection regulations. See section 1.6.4, Products used in potentially explosive atmospheres.

4.1 Location

The pump should be located to allow room for access, ventilation, maintenance and inspection with ample headroom for lifting and should be as close as practicable to the supply of liquid to be pumped. Refer to the general arrangement drawing for the pump set.

4.2 Part assemblies

Motors may be supplied loose on LRV pumps, typically on frame sizes 250 and above. It is the responsibility of the installer to ensure that the motor is assembled to the pump and lined up as detailed in section 4.5.2.

4.3 Foundation There are many methods of installing pump units to their foundations. The correct method depends on the size of the pump unit, its location and noise vibration limitations. Non-compliance with the provision of correct foundation and installation may lead to failure of the pump and, as such, would be outside the terms of the warranty. Ensure the following are met. a) The baseplate should be mounted onto a firm foundation, either an appropriate thickness of quality concrete or sturdy steel framework. (It should NOT be distorted or pulled down onto the surface of the foundation, but should be supported to maintain the original alignment.) b) Install the baseplate onto packing pieces evenly spaced and adjacent to foundation bolts.

c) Level with shims between baseplate and packing pieces. d) The pump and driver have been aligned before dispatch however the alignment of pump and motor half coupling must be checked. If this is incorrect, it indicates that the baseplate has become twisted and should be corrected by re-shimming.

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LR, LRV, LLR and LR-S USER INSTRUCTIONS ENGLISH C953KH013 - 09/03 ®

e) Vertical pumps should be mounted following the practices outlined for baseplate mounted pumps. (Larger sizes may need the motor fitting after installing the pump - refer to section 4.5.2.) f) If the pump is driven via a cardan shaft there may be a requirement to offset the pump shaft with respect to the driver to optimize the cardan shaft bearing life. This offset will typically be in the range 0 to 4 degrees depending on shaft design. Please consult the separate user instructions before installation. g) Any support for the cardan shaft plummer blocks must not exhibit resonant frequencies in the range 0.8 to 1.2 N where N = pump running speed. h) If not supplied, guarding shall be fitted as necessary to meet the requirements of EN292 and EN953.

4.4 Grouting

Where applicable, grout in the foundation bolts. After adding pipework connections and rechecking the coupling alignment, the baseplate should then be grouted in accordance with good engineering practice. Fabricated steel, cast iron and epoxy baseplates can be filled with grout. Folded steel baseplates should be grouted to locate their packing pieces. If in any doubt, please contact your nearest service centre for advice.

Although the pump will have been aligned at the factory it is most likely that this alignment will have been disturbed during transportation or handling. If necessary, align the motor to the pump, not the pump to the motor. Horizontal pumps – LR, LLR and LR-S Alignment is achieved by adding or removing shims under the motor feet and also moving the motor horizontally as required. In some cases where the alignment cannot be achieved it will be necessary to move the pump before recommencing the above procedure. Vertical pumps – LRV Adding or removing shims between the motor stool and the pump casing achieves alignment. The motor/motor stool assembly may also have to be moved horizontally at the interface with the pump casing, as required. It should be noted that the motor has a spigot (rabbet) fit into the motor stool and it is therefore not possible to achieve any horizontal movement at this interface. For couplings with narrow flanges use a dial indicator as shown below to check both parallel and angular alignment.

Parallel

Grouting provides solid contact between the pump unit and foundation, prevents lateral movement of running equipment and dampens resonant vibrations. Foundation bolts should only be fully tightened when the grout has cured.

Angular

4.5 Initial alignment 4.5.1 Thermal expansion The pump and motor will normally have to be aligned at ambient temperature and should be corrected to allow for thermal expansion at operating temperature. In pump installations involving high liquid temperatures, the unit should be run at the actual operating temperature, shut down and the alignment checked immediately. 4.5.2 Alignment methods Ensure pump and driver are isolated electrically and the half couplings are disconnected. The alignment MUST be checked.

Maximum permissible misalignment at working temperature: Parallel 0.2 mm (0.008 in.) TIR Angular 0.1 mm (0.004 in.) TIR Pumps with thick flanged non-spacer couplings can be aligned by using a straight-edge across the outside diameters of the coupling hubs and measuring the gap between the machined faces using feeler gauges, measuring wedge or callipers. When the electric motor has sleeve bearings it is necessary to ensure that the motor is aligned to run on its magnetic centreline.

Refer to the motor manual for details.

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A button (screwed into one of the shaft ends) is normally fitted between the motor and pump shaft ends to fix the axial position. If the motor does not run in its magnetic centre the resultant additional axial force may overload the pump thrust bearing. Complete piping as below and see sections 4.7, Final shaft alignment check up to and including section 5, Commissioning, startup, operation and shutdown before connecting driver and checking actual rotation.

4.6 Piping Protective covers are fitted to the pipe connections to prevent foreign bodies entering during transportation and installation. Ensure that these covers are removed from the pump before connecting any pipes. 4.6.1 Suction and discharge pipework In order to minimize friction losses and hydraulic noise in the pipework it is good practice to choose pipework that is one or two sizes larger than the pump suction and discharge. Typically main pipework velocities should not exceed 2 m/s (6 ft/sec) suction and 3 m/s (9 ft/sec) on the discharge. Take into account the available NPSH which must be higher than the required NPSH of the pump. piping.

Never use the pump as a support for

Maximum forces and moments allowed on the pump flanges vary with the pump size and type. To minimize these forces and moments that may, if excessive, cause misalignment, hot bearings, worn couplings, vibration and the possible failure of the pump casing, the following points should be strictly followed: • Prevent excessive external pipe load • Never draw piping into place by applying force to pump flange connections • Do not mount expansion joints so that their force, due to internal pressure, acts on the pump flange The table in 4.6.3 summarizes the maximum forces and moments allowed on horizontal shaft pump casings. Refer to Flowserve when the pump shaft is vertical. before use.

4.6.2 Suction piping a) The inlet pipe should be one or two sizes larger than the pump inlet bore and pipe bends should be as large a radius as possible. b) Pipework reducers should be conical and have a maximum total angle of divergence of 15 degrees. c) On suction lift the piping should be inclined up towards the pump inlet with eccentric reducers incorporated to prevent air locks. d) On positive suction, the inlet piping must have a constant fall towards the pump. e) Flow should enter the pump suction with uniform flow, to minimize noise and wear. This is particularly important on large or high-speed pumps which should have a minimum of five diameters of straight pipe on the pump suction between the elbow and inlet flange. See section 10.3, Reference 1, for more detail. f) Inlet strainers, when used, should have a net ‘ free area’ of at least three times the inlet pipe area. g) Do not install elbows at an angle other than perpendicular to the shaft axis. Elbows parallel to the shaft axis will cause uneven flow. h) Except in unusual circumstances strainers are not recommended in inlet piping. If considerable foreign matter is expected a screen installed at the entrance to the wet well is preferable. i) Fitting an isolation valve will allow easier maintenance. j) Never throttle pump on suction side and never place a valve directly on the pump inlet nozzle. Typical design – flooded suction Discharge isolating valve

Non return valve

Concentric conical reducer

Eccentric conical reducer

Suction isolating valve

>5D

Slope up from pump suction

Note: Ideally reducers should be limited to one pipe diameter change, ie 150 mm (6 in.) to 200 mm (8 in.). Must have a maximum total angle of divergence of 15 degrees.

Typical design – suction lift Discharge isolating valve

Non return valve

Concentric conical reducer

Eccentric conical reducer

>5D

Long radius bend

Ensure piping and fittings are flushed

Ensure piping for hazardous liquids is arranged to allow pump flushing before removal of the pump.

Slope down from pump suction

Notes: 1. S = Minimum submergence >3E. 2. Ideally reducers to be limited to one pipe diameter change, ie 150 mm (6 in.) to 200 mm (8 in.). Must have a maximum total angle of divergence of 15 degrees.

Page 14 of 44

LR, LRV, LLR and LR-S USER INSTRUCTIONS ENGLISH C953KH013 - 09/03 ®

4.6.3 Maximum forces and moments allowed on the pump suction and discharge flanges of horizontal shaft pumps

Type and size 1.5LLR-7 1.5LLR-10 2LLR-9 to 2LLR-11 3LLR-11 4LLR-11 2.5LR-10 2.5LR-13 3LR-9 3LR-12 4LR-10 4LR-11 to 4LR-14 5LR-10 to 5LR-15 5LR-19 6LR-10 to 6LR-16 6LR-18 6LR-18S 8LR12/14 8LR18S 8LR-20 8LR-23S 10LR-14/16 10LR-14S 10LR-17/18 10LR-18S 12LR-14S

Mx 0.85 (627) 1.36 (1003) 1.36 (1003) 1.70 (1254) 2.55 (1880) 1.36 (1003) 1.70 (1254) 1.70 (1254) 2.12 (1563) 2.12 (1563) 2.55 (1880) 2.55 (1880) 3.40 (2507) 3.40 (2507) 4.25 (3134) 4.25 (3134) 4.25 (3134) 5.10 (3761) 5.10 (3761) 4.25 (3134) 5.10 (3761) 5.10 (3761) 5.95 (4388) 5.95 (4388) 5.95 (4388)

My 0.45 (332) 0.72 (531) 0.72 (531) 0.90 (664) 1.35 (996) 0.72 (531) 0.90 (664) 0.90 (664) 1.12 (826) 1.12 (826) 1.35 (996) 1.35 (996) 1.80 (1327) 1.80 (1327) 2.25 (1659) 2.25 (1659) 2.25 (1659) 2.70 (1991) 2.70 (1991) 2.25 (1659) 2.70 (1991) 2.70 (1991) 3.15 (2323) 3.15 (2323) 3.15 (2323)

Notes: 1) F = External force (tension or compression). M = External moment, clockwise or counter-clockwise. 2) Forces and moments may be applied simultaneously in any direction. 3) Values apply to all materials. 4) Higher loads may be applicable, if direction and magnitude of individual loads are known, but these need written approval from Flowserve Pumps. 5) Pumps must be on rigid foundations and baseplates must be fully grouted 6) Pump/baseplate should not be used as pipe anchor. Suction and discharge piping should be anchored as close as possible to the pump flanges to reduce vibration and prevent strain on the pump casing. Expansion joints are recommended. They must be properly tied and located on the side of the pipe anchor away from the pump. 7) The pump mounting bolt torques specified must be used to prevent relative movement between the pump casing and baseplate. (See section 6.6, Fastener torques.) The bolt material must have a minimum yield strength of 600 N/mm2 (87 000 lb/in.2). Maximum moments (M) in kNm (lbf·ft) and maximum forces (F) in kN (lbf) Suction Discharge Mz Fx Fy Fz Mx My Mz Fx Fy Fz 0.6 1.47 1.15 1.34 0.44 0.24 0.32 0.77 0.88 0.64 (442) (330) (259) (189) (324) (177) (236) (173) (198) (144) 0.96 2.35 1.85 1.34 0.44 0.24 0.32 0.77 0.88 0.64 (708) (529) (415) (302) (324) (177) (236) (173) (198) (144) 0.96 2.35 1.85 1.34 0.59 0.32 0.43 1.03 1.17 0.85 (708) (529) (415) (302) (431) (235) (314) (230) (263) (192) 1.20 2.94 2.31 1.68 0.88 0.48 0.64 1.54 1.76 1.28 (885) (661) (519) (378) (649) (354) (472) (346) (396) (288) 1.80 4.41 3.47 2.52 1.10 0.60 0.80 1.92 2.20 1.60 (1327) (991) (779) (566) (811) (442) (590) (432) (495) (360) 0.96 2.35 1.85 1.34 0.72 0.39 0.52 1.25 1.43 1.00 (708) (529) (416) (302) (531) (288) (383) (281) (321) (234) 1.20 2.94 2.31 1.68 0.72 0.39 0.52 1.25 1.43 1.00 (885) (661) (519) (378) (531) (288) (383) (281) (321) (234) 1.20 2.94 2.31 1.68 0.88 0.48 0.64 1.54 1.76 1.28 (885) (661) (519) (378) (649) (354) (472) (346) (396) (288) 1.50 3.68 2.88 2.10 0.88 0.48 0.64 1.54 1.76 1.28 (1106) (826) (648) (472) (649) (354) (472) (346) (396) (288) 1.50 3.68 2.88 2.10 1.10 0.60 0.80 1.92 2.20 1.60 (1106) (826) (648) (472) (811) (442) (590) (432) (495) (360) 1.80 4.41 3.47 2.52 1.10 0.60 0.80 1.92 2.20 1.60 (1327) (991) (779) (566) (811) (442) (590) (432) (495) (360) 1.80 4.41 3.47 2.52 1.37 0.75 1.00 2.40 2.75 2.00 (1327) (991) (779) (566) (1010) (553) (737) (540) (618) (450) 2.40 5.88 4.62 3.36 1.37 0.75 1.00 2.40 2.75 2.00 (1770) (1322) (1039) (755) (1010) (553) (737) (540) (618) (450) 2.40 5.88 4.62 3.36 1.65 0.90 1.20 2.88 3.30 2.40 (1770) (1322) (1039) (755) (1217) (664) (885) (648) (742) (540) 3.00 7.35 5.78 4.20 1.65 0.90 1.20 2.88 3.30 2.40 (2212) (1653) (1299) (944) (1217) (664) (885) (648) (742) (540) 3.00 7.35 5.78 4.20 1.65 0.90 1.20 2.88 3.30 2.40 (2212) (1653) (1299) (944) (1217) (664) (885) (648) (742) (540) 3.00 7.35 5.78 4.20 2.20 1.20 1.60 3.84 4.40 3.20 (2212) (1653) (1299) (944) (1622) (885) (1180) (863) (989) (719) 3.60 8.82 6.93 5.04 2.20 1.20 1.60 3.84 4.40 3.20 (2655) (1983) (1558) (1133) (1622) (885) (1180) (863) (989) (719) 3.60 8.82 6.93 5.04 2.20 1.20 1.60 3.84 4.40 3.20 (2655) (1983) (1558) (1133) (1622) (885) (1180) (863) (989) (719) 3.00 7.35 5.78 4.20 2.20 1.20 1.60 3.84 4.40 3.20 (2212) (1653) (1299) (944) (1622) (885) (1180) (863) (989) (719) 3.60 8.82 6.93 5.04 2.75 1.50 2.00 4.80 5.50 4.00 (2655) (1983) (1558) (1133) (2028) (1106) (1475) (1079) (1237) (899) 3.60 8.82 6.93 5.04 2.75 1.50 2.00 4.80 5.50 4.00 (2655) (1983) (1558) (1133) (2028) (1106) (1475) (1079) (1237) (899) 4.20 10.29 8.09 5.88 2.75 1.50 2.00 4.80 5.50 4.00 (3097) (2314) (1818) (1322) (2028) (1106) (1475) (1079) (1237) (899) 4.20 10.29 8.09 5.88 2.75 1.50 2.00 4.80 5.50 4.00 (3097) (2314) (1818) (1322) (2028) (1106) (1475) (1079) (1237) (899) 4.20 10.29 8.09 5.88 3.30 1.80 2.40 5.76 6.60 4.80 (3097) (2314) (1818) (1322) (2434) (1327) (1770) (1295) (1484) (1079)

Page 15 of 44

LR, LRV, LLR and LR-S USER INSTRUCTIONS ENGLISH C953KH013 - 09/03 ®

4.6.4 Discharge piping See section 4.6.2 for typical pipework design. a) A non-return valve should be located in the discharge pipework to protect the pump from excessive back pressure and hence reverse rotation when the unit is stopped. b) Pipework reducers should have a maximum total angle of divergence of 15 degrees. c) Fitting an isolation valve will allow easier maintenance. 4.6.5 Auxiliary piping 4.6.5.1 Drains Pipe pump casing drains and gland leakage to a convenient disposal point. 4.6.5.2 Pumps fitted with packed gland When suction pressure is below ambient pressure it is necessary to feed the gland packing with liquid to provide lubrication and prevent the ingress of air. This is normally achieved with a supply from the pump discharge volute to the stuffing box. If the pumped liquid is dirty and cannot be used for sealing, a separate clean compatible liquid supply to the gland at 1 bar (15 psi) above suction pressure is recommended. 4.6.5.3 Pumps fitted with mechanical seals Single seals requiring re-circulation will normally be provided with the auxiliary piping from pump casing already fitted. If the seal requires an auxiliary quench then a connection must be made to a suitable source of liquid flow, low pressure steam or static pressure from a header tank. Recommended pressure is 0.35 bar (5 psi) or less. Check General arrangement drawing. Special seals may require different auxiliary piping to that described above. Consult separate User Instructions and or Flowserve if unsure of correct method or arrangement. For pumping hot liquids, to avoid seal damage, it is recommended that any external flush/cooling supply be continued after stopping the pump. 4.6.6 Final checks Check the tightness of all bolts in the suction and discharge pipework. Check also the tightness of all foundation bolts.

4.7 Final shaft alignment check

After connecting piping to the pump, rotate the shaft several times by hand to ensure there is no binding and all parts are free.

Recheck the coupling alignment, as previously described, to ensure no pipe strain. If pipe strain exists, correct piping.

4.8 Electrical connections 4.8.1 Electrical connections must be made by a qualified Electrician in accordance with relevant local, national and international regulations. It is important to be aware of the 4.8.2 EUROPEAN DIRECTIVE on potentially explosive areas where compliance with IEC60079-14 is an additional requirement for making electrical connections. 4.8.3 It is important to be aware of the EUROPEAN DIRECTIVE on electromagnetic compatibility when wiring up and installing equipment on site. Attention must be paid to ensure that the techniques used during wiring/installation do not increase electromagnetic emissions or decrease the electromagnetic immunity of the equipment, wiring or any connected devices. If in any doubt contact Flowserve for advice. 4.8.4 The motor must be wired up in accordance with the motor manufacturer’s instructions (normally supplied within the terminal box) including any temperature, earth leakage, current and other protective devices as appropriate. The identification nameplate should be checked to ensure the power supply is appropriate. 4.8.5 A device to provide emergency stopping must be fitted. 4.8.6 If not supplied pre-wired to the pump unit, the controller/starter electrical details will also be supplied within the controller/starter. 4.8.7 For electrical details on pump sets with controllers see the separate wiring diagram. 4.8.8 See section 5.3, Direction of rotation before connecting the motor to the electrical supply.

4.9 Protection systems The following protection systems are recommended particularly if the pump is installed in a potentially explosive area or is handling a hazardous liquid. If in doubt consult Flowserve. If there is any possibility of the system allowing the pump to run against a closed valve or below minimum continuous safe flow a protection device should be installed to ensure the temperature of the liquid does not rise to an unsafe level.

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LR, LRV, LLR and LR-S USER INSTRUCTIONS ENGLISH C953KH013 - 09/03 ®

If there are any circumstances in which the system can allow the pump to run dry, or start up empty, a power monitor should be fitted to stop the pump or prevent it from being started. This is particularly relevant if the pump is handling a flammable liquid. If leakage of product from the pump or its associated sealing system can cause a hazard it is recommended that an appropriate leakage detection system is installed.

The oil filled bottle should then be refitted so as to return it to the upright position. Filling should be repeated until oil remains visible within the bottle. The LR-S pumps are fitted with a different oiler - set oil level D as below: 7

6LR-18S, 10LR-14S and 12LR-14S = 48 mm (1 /8in.) 1 8LR-18S, 8LR-23S and 10LR-18S = 53 mm (2 /16in.) 33

To prevent excessive surface temperatures at bearings it is recommended that temperature or vibration monitoring are carried out. See sections 5.7.4 and 5.7.5.

5 COMMISSIONING, START-UP, OPERATION AND SHUTDOWN

D

These operations must be carried out by fully qualified personnel. Approximate oil volumes are shown in section 3.4.2, Pump and impeller data.

5.1 Pre-commissioning procedure 5.1.1 Lubrication Determine the mode of lubrication of the pump set, eg grease, oil, product lubrication etc. For oil lubricated pumps, fill the bearing housing with correct grade of oil to the correct level, ie sight glass or constant level oiler bottle.

Grease lubricated pumps and electric motors are supplied pre-greased. Other drivers and gearboxes, if appropriate, should be lubricated in accordance with their manuals. In the case of product lubricated bearings the source of product supply should be checked against the order. There may be requirements for an external clean supply, particular supply pressure or the commencement of lubrication supply before pump start-up.

When fitted with a constant level oiler, the bearing housing should be filled by unscrewing or hinging back the transparent bottle and filling the bottle with oil. Where an adjustable body Denco oiler is fitted this should be set to the height shown in the following diagram:

Page 17 of 44

LR, LRV, LLR and LR-S USER INSTRUCTIONS ENGLISH C953KH013 - 09/03 ®

5.2 Pump lubricants

Centifugal pump lubrication

5.2.1 Recommended oil lubricants Oil

Splash lubrication

Viscosity mm²/s 40 ºC Temp. maximum ºC (ºF) Designation according to DIN51502 ISO VG BP

Oil companies and lubricants

DEA Elf Esso Mobil Q8 Shell Texaco Wintershall (BASF Group)

32

68

46

65 (149)

80 (176)

-

HL/HLP 32

HL/HLP 68

HL/HLP 46

BP Energol HL32 BP Energol HLP32 Anstron HL32 Anstron HLP32 OLNA 32 HYDRELEF 32 TURBELF 32 ELFOLNA DS32 TERESSO 32 NUTO H32 Mobil DTE oil light Mobil DTE13 MobilDTE24 Q8 Verdi 32 Q8 Haydn 32 Shell Tellus 32 Shell Tellus 37 Rando Oil HD 32 Rando Oil HD-AZ-32 Wiolan HN32 Wiolan HS32

BP Energol HL68 BP Energol HLP68 Anstron HL68 Anstron HLP68

BP Energol HL46 BP Energol HLP46 Anstron HL46 Anstron HLP46

TURBELF SA68

TURBELF SA46

ELFOLNA DS68 TERESSO 68 NUTO H68 Mobil DTE oil heavy medium

ELFOLNA DS46 TERESSO 46 NUTO H46 Mobil DTE oil medium Mobil DTE15M Mobil DTE25 Q8 Verdi 46 Q8 Haydn 46 Shell Tellus 01 C 46 Shell Tellus 01 46 Rando Oil 46 Rando Oil HD B-46 Wiolan HN46 Wiolan HS46

Mobil DTE26 Q8 Verdi 68 Q8 Haydn 68 Shell Tellus 01 C 68 Shell Tellus 01 68 Rando Oil 68 Rando Oil HD C-68 Wiolan HN68 Wiolan HS68

5.2.2 Recommended grease lubricants

5.2.4

Grease nipples

Grease

Force feed lubrication

NLGI 2 *

NLGI 3 **

Temp. range ºC (ºF)

-20 to +100 (-4 to +212)

-20 to +100 (-4 to +212)

Designation according to DIN

K2K-20

K2K 30

BP

Energrease LS2

Energrease LS3

DEA

Glissando 20

Glissando 30

Elf

Elfmulti 2

Elfmulti 3

Esso

Beacon 2

Beacon 3

Mobil

Mobilux 2

Mobilux 3

Q8

Rembrandt 2

Rembrandt 3

Shell

Alvania Fett G2 Alvania Fett R2

Alvania R3

Texaco

Multilak 20 Multilak EP2

Multilak 30 Multilak EP3

Wintershall (BASF Group)

Wiolub LFK 2

-

SKF

LGMT 2

LGMT 3

Silkolene G55/T G56/T * NLGI 2 is an alternative grease and is not to be mixed with other grades ** Factory packed bearings for the temperature range with grease nipples

5.2.3 Recommended fill quantities Refer to section 3.4.2, Pump and impeller data.

Lubrication schedule

5.2.4.1 Oil lubricated bearings Normal oil change intervals are 4 000 operating hours or at least every 6 months. For pumps on hot service or in severely damp or corrosive atmosphere, the oil will require changing more frequently. Lubricant and bearing temperature analysis can be useful in optimizing lubricant change intervals. The lubricating oil should be a high quality mineral oil having foam inhibitors. Synthetic oils may also be used if checks show that the rubber oil seals will not be adversely affected. The bearing temperature may be allowed to rise to 50 °C (122 °F).above ambient, but should not exceed 82 °C (180 °F). A continuously rising temperature, or an abrupt rise, indicate a fault. 5.2.4.2 Grease lubricated bearings When grease nipples are fitted, one charge between grease changes is advisable for most operating conditions, ie 2 000 hours interval. Normal intervals between grease changes are 4 000 hours or at least every 6 months.

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LR, LRV, LLR and LR-S USER INSTRUCTIONS ENGLISH C953KH013 - 09/03 ®

The characteristics of the installation and severity of service will determine the frequency of lubrication. Lubricant and bearing temperature analysis can be useful in optimising lubricant change intervals. The bearing temperature may be allowed to rise to 55 °C (131 °F) above ambient but should not exceed 95 °C (204 °F). For most operating conditions a quality grease having a lithium soap base and NLGI consistency of No 2 or No 3 is recommended. The drop point should exceed 175 °C (350 °F). Never mix greases containing different bases, thickeners or additives.

5.3 Direction of rotation Ensure the pump is given the same rotation as the pump direction arrow cast on the pump casing.

(1) Possible priming points

5.5.2 Suction lift with foot valve fitted Fill suction pipe and casing with liquid at a pressure of 1 to 2 bar from an external source. Vent as described in section 5.5.1. 5.5.3 Suction lift without foot valve Pump casing vents on the suction volute must be connected to an external vacuum pump priming system. If in doubt please consult Flowserve. Possible priming points

To avoid dry running the pump must either be filled with liquid or have the flexible coupling disconnected before driver is switched on.

To vacuum pump

If maintenance work has been carried out to the site’s electricity supply, the direction of rotation should be re-checked as above in case the supply phasing has been altered.

5.4 Guarding Guarding is supplied fitted to the pump set. If this has been removed or disturbed ensure that all the protective guards around the pump coupling and exposed parts of the shaft are securely fixed.

5.5 Priming and auxiliary supplies Ensure all electrical, hydraulic, pneumatic, sealant and lubrication systems (as applicable) are connected and operational. Ensure the inlet pipe and pump casing are completely full of liquid before starting continuous duty operation. 5.5.1 Suction pressure above atmospheric pressure Horizontal pumps: open vent connection (1) on top of the pump upper casing to allow the trapped air to escape. Let liquid run out until free from air bubbles. Vertical pumps: open vent connection (1) at the front of the upper half casing and disconnect the seal flush line at the mechanical seal/stuffing box to allow the trapped air to escape. Let liquid run out until free from air bubbles.

5.6 Starting the pump a)

Ensure flushing and/or cooling/ heating liquid supplies are turned ON before starting the pump. b) CLOSE the outlet valve. c) OPEN all inlet valves. d) Prime the pump. e)

Ensure all vent connections are closed before starting. f) Start motor and check outlet pressure. g) If the pressure is satisfactory, slowly OPEN outlet control valve.

h) i)

Do not run the pump with the outlet valve closed for a period longer than 30 seconds. If NO pressure, or LOW pressure, STOP the pump. Refer to section 7, Faults; causes and remedies, for fault diagnosis.

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LR, LRV, LLR and LR-S USER INSTRUCTIONS ENGLISH C953KH013 - 09/03 ®

5.7 Running the pump

5.7.4 Bearings

5.7.1 Venting the pump

If the pumps are working in a potentially explosive atmosphere, temperature or vibration monitoring at the pump bearings is recommended

Vent the pump to enable all trapped air to escape taking due care with hot or hazardous liquids. Under normal operating conditions, after the pump has been fully primed and vented, it should be unnecessary to re-vent the pump. 5.7.2 Pumps fitted with packed gland If the pump has a packed gland there must be some leakage from the gland. Gland nuts should initially be finger-tight only. Leakage should take place soon after the stuffing box is pressurised. The gland must be adjusted evenly to give visible leakage and concentric alignment of the gland to avoid excess temperature. If no leakage takes place the packing will begin to overheat. If overheating takes place the pump should be stopped and allowed to cool before being re-started. When the pump is re-started, check to ensure leakage is taking place at the packed gland. If hot liquids are being pumped it may be necessary to slacken the gland nuts to achieve leakage. The pump should be run for 30 minutes with steady leakage and the gland nuts tightened by 10 degrees at a time until leakage is reduced to an acceptable level, normally a minimum of 120 drops per minute is required. Bedding in of the packing may take another 30 minutes. Care must be taken when adjusting the gland on an operating pump. Safety gloves are essential. Loose clothing must not be worn to avoid being caught up by the pump shaft. Shaft guards must be replaced after the gland adjustment is complete. a short time.

Never run gland packing dry, even for

It is important, particularly with grease lubrication, to keep a check on bearing temperatures. After start up the temperature rise should be gradual, reaching a maximum after approximately 1.5 to 2 hours. This temperature rise should then remain constant or marginally reduce with time. (Refer to section 6.2.3.1 for further information.) 5.7.5 Normal vibration levels, alarm and trip For guidance, pumps generally fall under a classification for rigid support machines within the International rotating machinery standards and the recommended maximum levels below are based on those standards. Alarm and trip values for installed pumps should be based on the actual measurements (N) taken on the pump in the fully commissioned as new condition. Measuring vibration adjacent to the pump bearings at regular intervals will then show any deterioration in pump or system operating conditions. Vibration velocity – unfiltered mm/s (in./s) r.m.s. Normal N Alarm

N x 1.25

Shutdown trip N x 2.0

5.7.3 Pumps fitted with mechanical seals Mechanical seals require no adjustment. Any slight initial leakage will stop when the seal is run in. Before pumping dirty liquids it is advisable, if possible, to run in the pump mechanical seal using clean liquid to safeguard the seal face. External flush or quench should be started before the pump is run and allowed to flow for a period after the pump has stopped. Never run a mechanical seal dry, even for a short time.

If bearing temperatures are to be monitored it is essential that a benchmark temperature is recorded at the commissioning stage and after the bearing temperature has stabilized. • Record the bearing temperature (t) adjacent to the bearing and the ambient temperature (ta) • Estimate the likely maximum ambient temperature (tb) • Set the alarm at (t+tb-ta+5) °C [(t+tb-ta+10) °F] and the trip at 100 °C (212 °F) for oil lubrication and 105 °C (220 °F) for grease lubrication

Horizontal pumps ≤ 5.6 (0.22)

Vertical pumps ≤ 7.1 (0.28)

≤ 7.1 (0.28)

≤ 9.0 (0.35)

≤ 11.2 (0.44)

≤ 14.2 (0.56)

5.7.6 Stop/start frequency Pump sets are normally suitable for the number of equally spaced stop/starts per hour shown in the table below. Check actual capability of the driver and control/starting system before commissioning. Motor rating kW (hp) Up to 15 (20) Between 15 (20) and 90 (120) Above 90 (120)

Maximum stop/starts per hour 15 10 6

Where duty and standby pumps are installed it is recommended that they are run alternately every week.

Page 20 of 44

LR, LRV, LLR and LR-S USER INSTRUCTIONS ENGLISH C953KH013 - 09/03 ®

5.8 Stopping and shutdown a)

Close the outlet valve, but ensure that the pump runs in this condition for no more than a few seconds. b) Stop the pump. c) Switch off flushing and/or cooling/heating liquid supplies at a time appropriate to the process.

d)

For prolonged shut-downs and especially when ambient temperatures are likely to drop below freezing point, the pump and any cooling and flushing arrangements must be drained or otherwise protected.

5.9 Hydraulic, mechanical and electrical duty

This product has been supplied to meet the performance specifications of your purchase order, however it is understood that during the life of the product these may change. The following notes may help the user decide how to evaluate the implications of any change. If in doubt contact your nearest Flowserve office. 5.9.1 Specific gravity (SG) Pump capacity and total head in metres (feet) do not change with SG, however pressure displayed on a pressure gauge is directly proportional to SG. Power absorbed is also directly proportional to SG. It is therefore important to check that any change in SG will not overload the pump driver or over-pressurize the pump. 5.9.2 Viscosity For a given flow rate the total head reduces with increased viscosity and increases with reduced viscosity. Also for a given flow rate the power absorbed increases with increased viscosity, and reduces with reduced viscosity. It is important that checks are made with your nearest Flowserve office if changes in viscosity are planned. 5.9.3 Pump speed Changing pump speed effects flow, total head, power absorbed, NPSHR, noise and vibration. Flow varies in direct proportion to pump speed, head varies as speed ratio squared and power varies as speed ratio cubed. The new duty, however, will also be dependent on the system curve. If increasing the speed, it is important therefore to ensure the maximum pump working pressure is not exceeded, the driver is not overloaded, NPSHA > NPSHR, and that noise and vibration are within local requirements and regulations. 5.9.4 Net positive suction head (NPSHA) NPSH available (NPSHA) is a measure of the head available in the pumped liquid, above its vapour pressure, at the pump suction branch.

NPSH required (NPSHR) is a measure of the head required in the pumped liquid, above its vapour pressure, to prevent the pump from cavitating. It is important that NPSHA > NPSHR. The margin between NPSHA > NPSHR should be as large as possible. If any change in NPSHA is proposed, ensure these margins are not significantly eroded. Refer to the pump performance curve to determine exact requirements particularly if flow has changed. If in doubt please consult your nearest Flowserve office for advice and details of the minimum allowable margin for your application. 5.9.5 Pumped flow Flow must not fall outside the minimum and maximum continuous safe flow shown on the pump performance curve and or data sheet.

6 MAINTENANCE 6.1 General It is the plant operator’s responsibility to ensure that all maintenance, inspection and assembly work is carried out by authorized and qualified personnel who have adequately familiarized themselves with the subject matter by studying this manual in detail. (See also section 1.6.2.) Any work on the machine must be performed when it is at a standstill. It is imperative that the procedure for shutting down the machine is followed, as described in section 5.8. On completion of work all guards and safety devices must be re-installed and made operative again. Before restarting the machine, the relevant instructions listed in section 5, Commissioning, start up, operation and shut down must be observed. Oil and grease leaks may make the ground slippery. Machine maintenance must always begin and finish by cleaning the ground and the exterior of the machine. If platforms, stairs and guard rails are required for maintenance, they must be placed for easy access to areas where maintenance and inspection are to be carried out. The positioning of these accessories must not limit access or hinder the lifting of the part to be serviced. When air or compressed inert gas is used in the maintenance process, the operator and anyone in the vicinity must be careful and have the appropriate protection.

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LR, LRV, LLR and LR-S USER INSTRUCTIONS ENGLISH C953KH013 - 09/03 ®

Do not spray air or compressed inert gas on skin.

6.2.1 Routine inspection (daily/weekly)

Do not direct an air or gas jet towards other people.

The following checks should be made and the appropriate action taken to remedy any deviations: a) Check operating behaviour. Ensure noise, vibration and bearing temperatures are normal. b) Check that there are no abnormal fluid or lubricant leaks (static and dynamic seals) and that any sealant systems (if fitted) are full and operating normally. c) Check that shaft seal leaks are within acceptable limits. d) Check the level and condition of oil lubricant. On grease lubricated pumps, check running hours since last recharge of grease or complete grease change. e) Check any auxiliary supplies eg. heating/cooling, if fitted, are functioning correctly.

Never use air or compressed inert gas to clean clothes. Before working on the pump, take measures to prevent an uncontrolled start. Put a warning board on the starting device with the words: "Machine under repair: do not start". With electric drive equipment, lock the main switch open and withdraw any fuses. Put a warning board on the fuse box or main switch with the words: "Machine under repair: do not connect". Never clean equipment with inflammable solvents or carbon tetrachloride. Protect yourself against toxic fumes when using cleaning agents.

Refer to the manuals of any associated

6.2 Maintenance schedule

equipment for routine checks needed.

It is recommended that a maintenance plan and schedule is adopted, in line with these User Instructions. It should include the following: a) Any auxiliary systems installed must be monitored, if necessary, to ensure they function correctly. b) Gland packings must be adjusted correctly to give visible leakage and concentric alignment of the gland follower to prevent excessive temperature of the packing or follower. c) Check for any leaks from gaskets and seals. The correct functioning of the shaft seal must be checked regularly. d) Check bearing lubricant level, and if the hours run show a lubricant change is required. e) Check that the duty condition is in the safe operating range for the pump. f) Check vibration, noise level and surface temperature at the bearings to confirm satisfactory operation. g) Check dirt and dust is removed from areas around close clearances, bearing housings and motors. h) Check coupling alignment and re-align if necessary.

6.2.2 Periodic inspection (six monthly)

Our specialist service personnel can help with preventative maintenance records and provide condition monitoring for temperature and vibration to identify the onset of potential problems. If any problems are found the following sequence of actions should take place: a) Refer to section 7, Faults; causes and remedies, for fault diagnosis. b) Ensure equipment complies with the recommendations in this manual. c) Contact Flowserve if the problem persists.

a)

Check foundation bolts for security of attachment and corrosion. b) Check pump running records for hourly usage to determine if bearing lubricant requires changing. c) The coupling should be checked for correct alignment and worn driving elements.

Refer to the manuals of any associated equipment for periodic checks needed.

6.2.3 Re-lubrication Lubricant and bearing temperature analysis can be useful in optimizing lubricant change intervals. In general however, the following is recommended. 6.2.3.1 Grease lubrication See section 5.2.2 for grease recommendations. Regrease - via grease nipples every 2 000 hours or sooner depending on the severity of the application. a) It is important not to under or over grease the bearings as this will lead to over heating and premature failure. Grease lubricated bearing housings have grease nipples fitted in the bearing brackets. b) The maximum allowable operating temperatures for anti friction bearings will vary from unit to unit, depending on ambient and fluid temperature. The rise above ambient should not normally exceed 55 °C (131 °F) or a combined maximum of 95 °C (204 °F).

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LR, LRV, LLR and LR-S USER INSTRUCTIONS ENGLISH C953KH013 - 09/03 ®

TEMPERATURE

c) A continuously rising temperature or an abrupt temperature rise indicates a problem. If these symptoms occur, stop the pump immediately and investigate the cause.

c) Dry and reflush the bearing with clean light oil. d) It is important not to under or over grease the bearings as this will lead to over heating and premature failure. It is recommended that the bearings be filled with grease using a suitable spatula. In addition the housings should be no more than half filled.

TIME

Grease change - every 4 000 hours or sooner depending on the severity of the application. a) Remove the bearing bracket from the rotor assembly. b) Brush the bearing bracket with hot kerosene (100 to 115 °C/212 to 240 °F) or other non-toxic solvent. c) Clean and flush out the housing with a light mineral oil.

6.2.3.2 Oil lubrication important.

Maintaining the correct oil level is very

If the pump is supplied with a constant level oiler the oil level will be automatically maintained and as long as oil is visible in the glass bottle there is no need to refill. If however a sight glass has been fitted then regular checks should be made to ensure the level is maintained at the centre of the glass window. Refer to section 5.1.1 for methods of oil fill, section 5.2.1 for oil grade recommendations and 5.2.4 for the schedule and temperature limits. 6.2.4 Mechanical seals No adjustment is possible. When leakage reaches an unacceptable level the seal will need replacement.

d) Do not use waste oil to clean the housing. To clean the bearings: a) Wipe off as much grease as possible with a clean lint-free cloth. b) Brush bearings with hot kerosene (80 to 90 °C/ 175 to 195 °F) while gently spinning the outer bearing ring. c) Spin each ball to ensure that it is clean.

6.2.5 Gland packing The stuffing box gland can be backed off for re-packing or to enable the addition of extra rings of packing. The stuffing box is normally supplied with a lantern ring to enable a clean or pressurised flush to the centre of the packing. If not required, this can be replaced by an extra 2 rings of packing. There must always be a small leakage, normally a minimum of 120 drops per minute to atmosphere to lubricate and cool the packing is required.

To remove badly oxidized grease that refuses to come off: a) Support the rotor in a vertical position and immerse the bearing in hot kerosene or a mixture of alcohol and light mineral solvent. b) Gently spin the bearing outer ring.

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LR, LRV, LLR and LR-S USER INSTRUCTIONS ENGLISH C953KH013 - 09/03 ®

6.5 Tools required

6.3 Spare parts 6.3.1 Ordering of spares Flowserve keeps records of all pumps that have been supplied. When ordering spares the following information should be quoted. 1) Pump serial number 2) Pump size 3) Part name – taken from section 8 4) Part number – taken from section 8 5) Number of parts required The pump size and serial number are shown on the pump nameplate. To ensure continued satisfactory operation, replacement parts to the original design specification should be obtained from Flowserve. Any change to the original design specification (modification or use of a non-standard part) will invalidate the pump’s safety certification. 6.3.2 Storage of spares Spares should be stored in a clean dry area away from vibration. Inspection and re-treatment of metallic surfaces (if necessary) with preservative is recommended at 6 monthly intervals.

6.4 Recommended spares and consumable items For start up purposes: 1 - complete set of gland packing 2 - shaft sleeves 1 - set of gaskets and seals (optional: 2 - mechanical seals) For 2 years operation: 1 - set of bearings (line and thrust) 2 - sets of gland packing 2 - shaft sleeves 2 - sets of gaskets and seals 2 - lantern rings 2 - casing wear rings (optional: 2 - mechanical seals 2 - impeller wear rings) For 4 years operation: 1 - set of bearings (line and thrust) 2 - sets of gland packing 2 - shaft sleeves 2 - sets of gaskets and seals 2 - lantern rings 2 - casing wear rings 1 - impeller (optional: 2 - mechanical seals 2 - impeller wear rings)

A typical range of tools that will be required to maintain these pumps is listed below. Readily available in standard tool kits, and dependent on pump size: • Open ended spanners (wrenches) to suit up to 7 M24 ( /8in.) screws/nuts 7 • Socket spanners (wrenches), up to M24 ( /8in.) screws • Allen keys, up to 6 mm (¼in.) A/F • Range of screwdrivers • Soft mallet More specialized equipment: • Bearing pullers • Bearing induction heater • Dial test indicator • C-spanner (wrench) - for removing shaft nut. (If difficulties in sourcing are encountered, consult Flowserve.) See also section 6.8.1.k.

6.6 Fastener torques Bolt size M8 (5/16 in.)
  M10 ( M12 (½ in.)
  M16 ( M20 (¾ in.)
  M24 (

Torque Nm (lbf·ft) Pump feet All other fasteners fasteners 10 (7) 20 (15) 63 (46) 34 (25) 170 (125) 84 (62) 340 (250) 165 (120) 590 (435) 285 (210)

6.7 Renewal clearances

As wear takes place between the impeller and casing ring the overall efficiency of the pump set will decrease. To maintain optimum efficiency it is recommended that rings are replaced and the impeller renovated when the radial clearance detailed in section 3.4.2 has doubled. On the LRV it is recommended that the product lubricated bearing is renewed at a diametrical clearance of 0.5 mm (0.02 in.).

6.8 Disassembly Refer to section 1.6, Safety, before dismantling the pump. Before dismantling the pump for overhaul, ensure genuine Flowserve replacement parts are available. To dismantle the pump consult the sectional drawings. See section 8, Parts lists and drawings. LR single entry (2.5LR sizes only) – section 8.1 LR double entry (all other sizes) – section 8.2 LLR two-stage – section 8.3 LR-S double entry – section 8.4 LRV double entry – section 8.5

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6.8.1 LR and LLR a) Isolate motor and lock off electrical supply in accordance with local regulations. b) Isolate suction and discharge valves. c) Remove coupling guards and disconnect the coupling halves. d) Drain pump casing. Remove any auxiliary piping if applicable. e) If bearings are oil lubricated drain oil from both bearing brackets. f) Remove the pump half coupling. g) Unbolt the glands/seal covers from the casing. If glands are split type, remove completely. h) With a suitable punch, drive out the two straight roll pins which are used on the horizontal split flange to align the upper and lower half casings. i) Remove the bolts, which hold the upper and lower half of the casing together, and remove the upper half. Tapped holes are provided in the joint flange to enable the use of forcing bolts to loosen the joint. j) Lift the casing upper half using the cast on lifting lugs where provided. Where there are no integral lifting lugs, remove the pipe plug or fittings, if used, from the volute vent connection located on top of the casing upper half and install a special steel lifting eye with a threaded shank to match the pipe tap opening in the casing. Do NOT use these methods to lift the bottom half or complete pump casing. k) Remove the bearing bracket to casing bolts and remove the 2 dowel bushings on each side. A tool for removing the bushings can be easily and economically made as shown in the following diagrams.

Material: 25mm (1 in.) standard weight steel pipe

l)

Lift out rotor assembly. Use care in slinging, handling and supporting of the rotor for subsequent dismantling. Place rotor securely on two support blocks.

m) When removing the rotor assembly, the casing rings will be attached to it. They are fixed by two diametrically opposite pins inserted into the casing ring and located in grooves in the lower half casing. On the LLR design the interstage bushing between the two impellers will also be attached to the shaft. n) Remove bearing covers and slide bearing brackets off bearings. Some pump sizes have a shim fitted at the non-drive end – retain for future use. o) Release the bearing lockwasher at the non-drive end and remove the bearing locknut. Pull off both ball bearings using a suitable puller, ensuring force is applied to inner race only. Retain the non-drive end bearing spacer fitted to the shaft on some pump sizes for future use. Remove the bearing covers. p) Depending on configuration remove glands/seal covers, packing and lantern ring/mechanical seal.

Refer to any special instructions supplied

with the mechanical seal. q) Remove the two socket head screws securing each shaft nut. Using C-spanner remove shaft nuts. Slide off shaft sleeves. r) Remove impeller(s), casing rings, impeller key, and interstage bushing if fitted. The 2.5LR10 and 2.5LR13 are not fitted with shaft sleeves and removal of the impeller nut will allow the impeller to be withdrawn. If impellers prove difficult to remove, the use of heat is permissible. Refer to Section 1.6, Safety, APPLYING HEAT TO REMOVE IMPELLER, for more details. s) If impeller rings are also fitted, they are shrunk onto the impeller and fixed with locking screws between their diametral mating surfaces. t) To remove the impeller rings, remove the locking screws and heat up the ring until it slides off easily. 6.8.2 LR-S a) Isolate motor and lock off electrical supply in accordance with local regulations. b) Isolate suction and discharge valves. c) Remove coupling guards and disconnect the coupling halves. d) Drain pump casing. Remove any auxiliary piping if applicable. e) If bearings are oil lubricated drain oil from both bearing brackets. f) Remove the pump half coupling. g) Unbolt the glands/seal covers from the casing. If glands are split type, remove completely. h) With a suitable punch, drive out the two straight roll pins which are used on the horizontal split flange to align the upper and lower half casings. i) Remove the bolts, which hold the upper and lower half of the casing together, and remove the upper half. Tapped holes are provided in the joint flange to enable the use of forcing bolts to loosen the joint.

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LR, LRV, LLR and LR-S USER INSTRUCTIONS ENGLISH C953KH013 - 09/03 ®

j)

k) l) m) n)

o)

p)

Lift the casing upper half using the cast on lifting lugs where provided. Where there are no integral lifting lugs, remove the pipe plug or fittings, if used, from the volute vent connection located on top of the casing upper half and install a special steel lifting eye with a threaded shank to match the pipe tap opening in the casing. Do NOT use these methods to lift the bottom half or complete pump casing. Remove the bearing bracket to casing bolts and dowels on each side. Lift out rotor assembly. Use care in slinging, handling, and supporting of the rotor for subsequent dismantling. Place rotor securely on two support blocks. When removing the rotor assembly, the casing rings will be attached to it. They are both secured against rotation by one roll pin inserted into the casing ring and located in a hole at 6 o’clock in the lower half casing. Remove both bearing covers and circlip from nondrive end bearing. Using bearing pullers remove bearing bracket complete with bearings and lip seals from the pump shaft. If double row bearings are fitted the non-drive end bearing will be secured by bearing locknut and lockwasher. Depending on configuration remove glands/seal covers, packing and lantern rings/mechanical seal.

Refer to any special instructions supplied

with the mechanical seal. q) Remove the socket head screws securing each shaft nut. Using C-spanner remove shaft nuts. Slide off shaft sleeves. r) Remove impeller, casing rings, and impeller key. The impeller is a shrink fit on the shaft and the boss of the impeller must be heated in order to remove it. Refer to Section 1.6, Safety, APPLYING HEAT TO REMOVE IMPELLER, for more details. s) This work must only be undertaken when both shaft nuts and sleeves have been fully removed. t) If impeller rings are also fitted, they are shrunk onto the impeller and fixed with locking screws between their diametral mating surfaces. u) To remove the impeller rings, remove the locking screws and heat up the ring until it slides off easily. 6.8.3 LRV The pump is best removed from the system to carry out a complete strip down. It should be set down with the shaft horizontal to enable the pump to be dismantled in a similar fashion to the LR and LLR. a) Isolate motor and lock off electrical supply in accordance with local regulations. b) Isolate suction and discharge valves. c) Remove coupling guards and disconnect the coupling halves.

d) Drain pump casing and, if applicable, remove any auxiliary piping. e) Remove motor complete with motor stool and set down carefully in a safe location. f) Retain any shimming between stool and pump casing. g) Remove bolts securing pump suction and discharge flanges. h) Sling pump as shown in section 2.3 and allow lifting gear to just take the pump weight. Remove setscrews securing the pump casing to the baseplate. i) Remove the pump to a safe location and manoeuvre the pump shaft into a horizontal position. j) Remove the pump half coupling. k) Unbolt the seal cover from the casing at the drive end. l) Remove the bottom bearing housing (non-drive end) complete with bearing bush, taking care not to damage the bearing surfaces. m) With a suitable punch, drive out the two straight roll pins which are used on the horizontal split flange to align the upper and lower half casings. n) Carry on as for LR and LRR section 6.8.1, j to k. o) Lift out rotor assembly. Use care in slinging, handling, and supporting of the rotor for subsequent dismantling. Place rotor securely on two support blocks. Protect the bearing surface on the outside diameter of the bottom sleeve from damage. p) Remove drive end bearing cover, outboard V-ring and slide bearing bracket off bearing. Some pump sizes have a shim fitted – retain for future use. q) Release the bearing lockwasher and remove the bearing locknut. Pull off drive end thrust ball bearing using a suitable puller, ensuring force is applied to inner race only. Remove the bearing cover and inboard V-ring. r) Depending on configuration remove the gland/seal cover, packing and lantern ring/mechanical seal from the drive end.

Refer to any special instructions supplied

with the mechanical seal. s) Remove the two socket head screws securing the shaft sleeve nut at the drive end. Using a C-spanner remove shaft sleeve nut. Slide off top shaft sleeve. t) Remove cap screw, sleeve end cap, and bottom sleeve. Take care not to damage the bearing surface on the sleeve. u) Remove impeller, casing rings and impeller key.

v)

If impeller proves difficult to remove, the use of heat is permissible. Refer to Section 1.6, Safety, APPLYING HEAT TO REMOVE IMPELLER, for more details. If impeller rings are fitted, they are shrunk onto the impeller and fixed with locking screws as for LR/LLR. To remove refer to 6.8.1, paragraph t.

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6.9 Examination of parts

6.10.1 LR/LLR

Used parts must be inspected before assembly to ensure the pump will subsequently run properly.

6.10.1.1 Impeller wear rings a) Impeller rings (when fitted) should be heated up to approximately 100 ºC (212 ºF) using a hotplate or hot oil bath and then slipped onto the impeller and pressed down to the shoulder. (Do NOT use a steel hammer to knock them into position.) b) Drill and tap 3 holes approximately 120 degrees apart into the diametral mating faces of the ring and impeller and insert socket head screws. (The existing half tapped holes from the removed impeller ring cannot be re-used.)

In particular, fault diagnosis is essential to enhance pump and plant reliability. 6.9.1 Casing, seal housing and impeller a) Inspect for excessive wear, pitting, corrosion, erosion or damage and any sealing surface irregularities. b) Replace as necessary.



6.9.2 Shaft and sleeve (if fitted) Replace if grooved, pitted or worn.



6.9.3 Gaskets and O-rings After dismantling, discard and replace. 6.9.4 Bearings a) It is recommended that bearings are not re-used after any removal from the shaft. b) The plain liquid lubricated bearings may be reused if both the bearing bush and bearing sleeve show no sign of wear, grooving or corrosion attack. (It is recommended that both the bush and sleeve are replaced at the same time.) 6.9.5 Bearing isolators, labyrinths or lip seals (if fitted) a) The lubricant, bearings and bearing housing seals are to be inspected for contamination and damage. If oil bath lubrication is utilised, these provide useful information on operating conditions within the bearing housing. b) If bearing damage is not due to normal wear and the lubricant contains adverse contaminants, the cause should be corrected before the pump is returned to service. c) Labyrinth seals and bearing isolators should be inspected for damage but are normally nonwearing parts and can be re-used. d) Bearing seals are not totally leak free devices. Oil from these may cause staining adjacent to the bearings.

6.10.1.2 Pre-assembly of casing gasket a) Fit casing gasket to the bottom half horizontal flange using a small amount of contact adhesive to prevent movement when the top half is fitted. Do not apply adhesive to the top surface of the gasket. b) It is important that the external corner of the casing gasket face and the stuffing box face is as sharp as possible. Do not chamfer with a file. c) If necessary trim gasket to match volute profile. Do not trim to stuffing box face at this stage. 6.10.1.3 Rotating element and bearing bracket a) Ensure all gaskets and O-rings are renewed and replaced in the correct position during assembly. b) Assemble the impeller on the shaft. It is important to mount the impeller so that the vane tips point away from the apparent flow direction.

6.10 Assembly

To assemble the pump consult the sectional drawings, see section 8, Parts list and drawings. Ensure threads, gasket and O-ring mating faces are clean. Apply thread sealant to non-face sealing pipe thread fittings. Coat the outside diameter of the dowel bushings with pipe compound prior to installation.

The rotor always rotates towards the expanding section of the volute

c) If working on a two-stage LLR pump, the spacer sleeve and interstage bushing, complete with anti-rotation screw, must be fitted on to the shaft between the two impellers.

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d) Fit the two shaft sleeves, O-rings and shaft nuts and lightly secure the impeller(s) on the shaft. Take care to protect the sleeve O-rings from damage on the shaft threads. The sleeves and nuts define the impeller position on the pump shaft and hence in the pump casing. Initially position the impeller(s) centrally on its keyway. This position may be adjusted slightly later on in the assembly process. 2.5LR pumps do not have sleeves fitted and the impeller is positively held against the shaft shoulder by the impeller nut and cannot subsequently be adjusted. e) It is recommended that gasket sealing compound Loctite 574 or equivalent is used between sleeve and impeller mating faces to protect the shaft from the liquid pumped. f) When mechanical seals are fitted the rotating parts can be slid onto the sleeves before the sleeves are fitted onto the shaft. The seal locking collars should be left loose. On some sizes of nd LLR a 2 . stage stuffing box bush is fitted, this must be slid onto the shaft before the seal.

Refer to any special instructions supplied

with the mechanical seal. g) If gland packing is used fit stuffing box bush and glands. h) Fit seal covers complete with seal seat, water throwers and bearing covers complete with gaskets. i) Fit the bearings onto the shaft. The main thrust bearing is at the non-drive end. Where double row bearings are fitted these must be assembled ’back to back’ as below

j) k) l) m) n)

o) p)

If grease lubricated, fill both sides of bearing with grease. Fit the bearing lockwasher and tighten the bearing shaft nut. Peen over a tab of lockwasher into locknut slot. Slip casing rings, complete with anti-rotation screws, loosely over the impeller hubs. Slide the bearing brackets over the bearings. All pumps except the 10LR17 and 10LR18 have a shim fitted between the outside diameter of the non-drive end bearing and the bearing bracket. Ensure shim is seated against the shoulder in the bearing bracket before sliding bracket over the bearing. Ensure bearings are located square in the bracket bore. If grease lubricated, one third fill the space between bearing cover and bearing with grease. Secure bearing cover, complete with gasket. Fit the coupling hub.

6.10.1.4 Casing lower half a) Coat the faces of the bearing housing brackets with liquid sealant to protect against corrosion. b) Place the complete rotating assembly into the casing ensuring that impeller rings are in the correct position and the anti-rotation screws are located in the slots on the horizontal flange. If working on an LLR pump the anti-rotation screw in the second stage stuffing box bush, when fitted, and the interstage bushing must also locate in the slot on the horizontal flange. c) Locate the dowel bushings within the holes in the lower half casing and bolt the bearing brackets to the casing. The dowel bushes must be sprayed with anti-seize compound Molyslip or equivalent before assembly in to the bracket/casing. d) Torque up the fixing bolts. e) Check rotor for free rotation. f) Centralize the impeller(s) within the casing waterway by adjusting the shaft nut, if necessary. Using a C-spanner fully tighten the shaft nuts and lock with the two radial socket head screws. g) Set the seals, if fitted, to the correct working length and tighten the seal collar screws.

Refer to any special instructions supplied

with the mechanical seal. h) Check for free rotation.

The 2.5LR13, 4LR11, 4LR14 and 5LR13 pumps have bearing spacers fitted to the shaft at the non-drive end. Ensure this is fitted before the bearing is assembled to the shaft. The bearings must be heated up to 100 ºC (212 ºF) using a hot plate, oil bath or induction heater and slid onto the shaft. Ensure bearing is fully seated against the shaft shoulder and bearing spacer, where fitted.

6.10.1.5 Casing upper half a) Lower the casing upper half over the lower half. Take care to ensure the renewable rings are correctly located in the upper half bores. b) Drive home the two casing roll pins to accurately position the casing and torque up all horizontal flange bolts. c) Check for free rotation. d) Using a sharp flexible-bladed knife, cut off the exposed casing gasket in the stuffing box area flush with the stuffing box face.

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e) If mechanical seals are fitted apply a small amount of silicon rubber sealant along the horizontal joint line on the stuffing box face and fit seal cover complete with gasket or O-ring. Take care not to damage O-ring, if fitted, and locate cover squarely on stuffing box face. f) Torque up seal cover bolts and check shaft/sleeve does not rub on seal cover or stationary seal seat bore. Ensure any spare holes in mechanical seal cover, particularly on cartridge seals, have sealing plugs fitted. g) If gland packing is fitted pack the gland, ensuring that the cut ends in each ring are staggered by 120 degrees. The glands are packed as shown on the relevant drawings in Sections 8.1 to 8.3. Finger tighten the gland nuts. h) Check coupling alignment, fit coupling drive element(s) and fit guards. i) Pipe up any external auxiliary connections. j) Check for free rotation. 6.10.2 LR-S 6.10.2.1 Impeller wear rings a) Impeller rings (when fitted) should be heated up to approximately 100 ºC (212 ºF) using a hotplate or hot oil bath and then slipped onto the impeller and pressed down to the shoulder. (Do NOT use a steel hammer to knock them into position.) b) Drill and tap 3 holes approximately 120 degrees apart into the diametral mating faces of the ring and impeller and insert socket head screws. (The existing half tapped holes from the removed impeller ring cannot be re-used.)

 

6.10.2.2 Pre-assembly of casing gasket a) Fit casing gasket to the bottom half horizontal flange using a small amount of contact adhesive to prevent movement when the top half is fitted. Do not apply adhesive to the top surface of the gasket. b) It is important that the external corner of the casing gasket face and the stuffing box face is as sharp as possible. Do not chamfer with a file. c) If necessary trim gasket to match volute profile. Do not trim to stuffing box face at this stage. 6.10.2.3 Rotating element and bearing bracket a) Ensure all gaskets and O-rings are renewed and replaced in the correct position during assembly.

b) Assemble the impeller on the shaft. It is important to mount the impeller so that the vane tips point away from the apparent flow direction.

The rotor always rotates towards the expanding section of the volute.

c)

The impeller is an interference fit on the shaft and the impeller boss needs quickly heating up to allow it to be fitted to the shaft. Take extreme care when handling hot components. Position impeller centrally on its keyway. d) Fit the two shaft sleeves, O-rings and shaft nuts. Take care to protect the sleeve O-rings from damage on the shaft threads. e) It is recommended that gasket sealing compound Loctite 574 or equivalent is used between sleeve and impeller mating faces to protect the shaft from the liquid pumped. f) When mechanical seals are fitted the rotating parts can be slid onto the sleeves before the sleeves are fitted onto the shaft. The seal locking collars should be left loose.

Refer to any special instructions supplied

with the mechanical seal. g) If gland packing is used fit stuffing box bush and glands. h) Fit seal covers complete with seal seat, water throwers and bearing brackets complete with lip seals. i) Fit the bearings on to the shaft. Main thrust bearing is at the non-drive end. The bearings must be heated up to 100 ºC (212 ºF) using a hot plate, oil bath or induction heater and slid onto the shaft. Ensure bearing is fully seated against the shaft shoulder. j) If bearings are grease lubricated they will be sealed for life and do not require any extra grease. k) Fit the bearing circlip or locknut/lockwasher at the non-drive end. Peen over a tab of lockwasher into locknut slot, if fitted. l) Slide bearing brackets over bearings ensuring bearings are located square in the bores. m) Fit bearing cover complete with gasket and lip seal. n) Slip casing rings complete with anti-rotation roll pins loosely over the impeller hubs. o) Fit the coupling hub.

Page 29 of 44

LR, LRV, LLR and LR-S USER INSTRUCTIONS ENGLISH C953KH013 - 09/03 ®

6.10.2.4 Casing lower half a) Coat the faces of the bearing housing brackets with liquid sealant to protect against corrosion. b) Place the complete rotating assembly into the casing ensuring that impeller rings are fitted in the grooves machined into the casing and the antirotation pins are located in the holes in the bottom of the casing bore. c) Locate the dowels in the bearing bracket into the holes in the lower half casing and bolt the bearing brackets to the casing. The dowels must be sprayed with anti-seize compound Molyslip or equivalent before assembly into the bracket/ casing. d) Torque up the fixing bolts. e) Check rotor for free rotation. f) Check end float of rotor is between 0.2 mm to 1.5 mm (0.008 in. to 0.060 in.). g) Using a C-spanner fully tighten the shaft nuts and lock with the two radial socket head screws. h) Set the seals, if fitted, to the correct working length and tighten the seal collar screws.

Refer to any special instructions supplied

i)

with the mechanical seal. Check for free rotation.

6.10.2.5 Casing upper half a) Lower the casing upper half over the lower half. Take care to ensure that the renewable rings are correctly located in the upper half bores. b) Drive home the two casing roll pins to accurately position the casing and torque up all horizontal flange bolts. c) Check for free rotation. d) Using a sharp flexible bladed knife, cut off the exposed casing gasket in the stuffing box area flush with the stuffing box face. e) If mechanical seals are fitted apply a small amount of silicon rubber sealant along the horizontal joint line on the stuffing box face and fit the seal cover complete with gasket or O-ring. Take care not to damage the O-ring, if fitted, and locate the cover squarely on the stuffing box face. f) Torque up seal cover bolts and check shaft/ sleeve does not rub in seal cover bore. Ensure any spare holes in seal cover, particularly on cartridge seals, have sealing plugs fitted. g) If gland packing is fitted, pack the gland, ensuring that the cut ends in each ring are staggered by 120 degrees. h) The glands are packed as shown in section 8.4. Finger tighten the gland nuts. i) Check coupling alignment, fit coupling drive element(s) and fit guards. j) Pipe up any external auxiliary connections.

6.10.3 LRV 6.10.3.1 Impeller wear rings As for LR/LLR. (See section 6.10.1.1.) 6.10.3.2 Pre-assembly of casing gasket As for LR/LLR. (See section 6.10.1.2.) 6.10.3.3 Pre-assembly of bottom bearing housing - SiC bearing a) Insert tolerance ring in bottom bearing housing and press in sleeve bearing until it is square against the shoulder at the bottom of the housing. b) Fit bearing retaining ring and secure with radial locking screw. 6.10.3.4 Pre-assembly of bottom bearing housing - cutless rubber bearing Press cutless rubber bearing into bottom bearing housing. 6.10.3.5 Rotating element and bearing bracket a) Ensure all gaskets and O-rings are renewed and replaced in the correct position during assembly. b) Assemble the impeller on the shaft. It is important to mount the impeller so that the vane tips point away from the apparent flow direction.

The rotor always rotates towards the expanding section of the volute

c) Fit the two shaft sleeves, top O-ring, drive end shaft nut, bottom sleeve end cap and socket headed cap screw. Take care to protect the sleeve O-ring from damage on the shaft thread. It is recommended Loctite 243 or equivalent is used to lock the socket headed cap screw in the shaft. d) Lightly secure the impeller on the shaft. Take care to protect the bearing surface on the bottom sleeve. The sleeves and shaft nuts define the impeller position on the pump shaft and hence in the pump casing. Initially position the impeller centrally on its keyway. This position may be adjusted slightly, later on in the assembly process. e) It is recommended that gasket sealing compound Loctite 574 or equivalent is used between sleeve and impeller mating faces to protect the shaft from the liquid pumped.

Page 30 of 44

LR, LRV, LLR and LR-S USER INSTRUCTIONS ENGLISH C953KH013 - 09/03 ®

f)

The rotating parts of the mechanical seal can be slid onto the top sleeve before it is fitted on to the shaft. The seal locking collar should be left loose.

Refer to any special instructions supplied

with the mechanical seal. g) If gland packing is used fit stuffing box bush and gland. h) Fit seal cover complete with seal seat, inboard V-ring seal and drive end bearing cover complete with gasket. i) Fit the ball thrust bearing at the drive end onto the shaft. The 5LRV15, 6LRV13, 6LRV16 and 6LRV18 pumps have bearing spacers fitted to the shaft at the drive end - ensure this is fitted before the bearing is assembled to the shaft. The bearing must be heated up to 100 ºC (212 ºF) using a hot plate, oil bath or induction heater and slid onto the shaft. Ensure bearing is fully seated against the shaft shoulder and bearing spacer, where fitted. j) Fill both sides of bearing with grease. k) Fit the bearing lockwasher and tighten the bearing shaft nut. Peen over a tab of lockwasher into locknut slot. l) Slip casing rings complete with anti-rotation screws loosely over impeller hubs. m) Slide the bearing bracket over the drive end bearing. All pumps except the 10LR17 and 10LR18 have a shim fitted between the outside diameter of the drive end bearing and the bearing bracket. Ensure shim is seated against the shoulder in the bearing bracket before sliding bracket over the bearing. Ensure bearing is located square in the bracket bore. n) One third fill the space between bearing cover and bearing with grease. o) Secure bearing cover, complete with gasket. p) Fit outboard V-ring seal. q) Fit the coupling hub. 6.10.3.6 Casing lower half a) Coat the faces of the bearing housing bracket with liquid sealant to protect against corrosion. b) Place the complete rotating assembly into the casing ensuring that impeller rings are in the correct position and the anti-rotation screws are located in the slots on the horizontal flange. c) Locate the dowel bushings within the holes in the lower half casing and bolt the bearing bracket to the casing. The dowel bushes must be sprayed with anti-seize compound Molyslip or equivalent before assembly into the bracket/casing. d) Lightly torque up the fixing bolts. e) Centralize the impeller within the casing waterway by adjusting the shaft nut and capscrew, if necessary. Apply locking compound Loctite 222 or equivalent to the threads of the capscrew.

f)

Fully tighten the shaft nut and capscrew and lock the shaft nut with the two radial socket head screws. g) Set the mechanical seal to the correct working length and tighten the seal collar screws.

Refer to any special instructions supplied

with the mechanical seal. 6.10.3.7 Casing upper half a) Lower the casing upper half over the lower half. Take care to ensure the renewable rings are correctly located in the upper half bores. b) Drive home the two casing roll pins to accurately position the casing and torque up all horizontal flange bolts. c) Using a sharp flexible bladed knife, cut off the exposed casing gasket in the stuffing box area flush with the stuffing box face. d) Secure bottom bearing housing complete with sleeve bearing and O-ring seal into stuffing box bore. e) Fully torque up bearing bracket fixing screws. f) Check for free rotation. g) Apply silicon rubber sealant along the horizontal joint line on the stuffing box face at drive end and fit seal cover complete with gasket or O-ring. Take care not to damage O-ring, if fitted, and locate cover squarely on stuffing box face. h) Torque up seal cover bolts and check shaft/sleeve does not rub in seal cover or stationary seal seat bore. Ensure any spare holes in seal cover, particularly on cartridge seals, have sealing plugs fitted. i) If gland packing is fitted, pack the gland, ensuring that the cut ends in each ring are staggered by 120 degrees. j) The glands are packed as shown in section 8.2. Finger tighten gland nuts. k) Add the same grease as used on the bearings under the lips of the inboard and outboard V-ring seal and slide up to bearing cover and bearing bracket faces to give light contact. l) Pipe up any external auxiliary connections. m) Check for free rotation.

Page 31 of 44

LR, LRV, LLR and LR-S USER INSTRUCTIONS ENGLISH C953KH013 - 09/03 ®

7 FAULTS; CAUSES AND REMEDIES FAULT SYMPTOM Pump overheats and seizes



Bearings have short life



Pump vibrates or is noisy



Mechanical seal has short life



Mechanical seal leaks excessively



Pump requires excessive power



Pump loses prime after starting



Insufficient pressure developed



Insufficient capacity delivered



Pump does not deliver liquid



PROBABLE CAUSES



A. SYSTEM TROUBLES

 

























 

 

 







 





 



    





 









POSSIBLE REMEDIES

Pump not primed.

Check complete filling

Pump or suction pipe not completely filled with liquid.

Check and complete filling

Suction lift too high or level too low.

Check NPSHa>NPSHr, proper submergence, losses at strainers and fittings

Excessive amount of air or gas in liquid.

Check and purge from pipes

Air or vapour pocket in suction line.

Check suction line design for pockets

Air leaks into suction line.

Check airtight pipe then joints and gaskets

Air leaks into pump through mechanical seal, sleeve joints, casing joint or pipe lugs.

Check airtight assembly then joints and gaskets

Foot valve too small.

Investigate replacing the foot valve

Foot valve partially clogged.

Clean foot valve

Inlet of suction pipe insufficiently submerged.

Check cut out system design

Total head of system higher than differential head of pump.

Check discharge head and head losses in discharge pipe at the valve settings. Check back pressure is not too high

Total head of system lower than pump design head.

Throttle at discharge valve or ask Flowserve if the impeller can be trimmed

Specific gravity of liquid different from design.

Consult Flowserve

Viscosity of liquid differs from design.

Consult Flowserve

Operation at very low capacity.

Measure value and check minimum permitted

Operation at high capacity.

Measure value and check maximum permitted B. MECHANICAL TROUBLES









 



















 





 

 

 





 



Misalignment due to pipe strain.

Check the flange connections and eliminate strains using elastic couplings or a method permitted

Improperly designed foundation.

Check setting of baseplate: tighten, adjust, grout base as required

Shaft bent.

Check shaft runouts within acceptable values

Rotating part rubbing on stationary part internally.

Check for signs of this and consult Flowserve if necessary

Bearings worn

Replace bearings

Wearing ring surfaces worn.

Replace worn wear ring/ surfaces

Impeller damaged or eroded.

Replace impeller and check reason

Leakage under sleeve due to joint failure.

Replace joint and check for damage

Mechanical seal improperly installed.

Check alignment of faces or damaged parts and assembly method used

Page 32 of 44

LR, LRV, LLR and LR-S USER INSTRUCTIONS ENGLISH C953KH013 - 09/03 ®

FAULT SYMPTOM Pump overheats and seizes



Bearings have short life



Pump vibrates or is noisy



Mechanical seal has short life



Mechanical seal leaks excessively



Pump requires excessive power



Pump loses prime after starting



Insufficient pressure developed



Insufficient capacity delivered



Pump does not deliver liquid









 















 

 

 













 

 

 

 





 





PROBABLE CAUSES

POSSIBLE REMEDIES

Incorrect type of mechanical seal for operating conditions.

Consult Flowserve

Shaft running off centre because of worn bearings or misalignment.

Check misalignment and correct if necessary. If alignment satisfactotry check bearings for excessive wear

Impeller out of balance resulting in vibration.

Check and consult Flowserve

Abrasive solids in liquid pumped.

Check and consult Flowserve

Mechanical seal was run dry.

Check mechanical seal condition and source of dry running and repair

Internal misalignment due to improper repairs causing impeller to rub.

Check method of assembly, possible damage or state of cleanliness during assembly

Excessive thrust caused by a mechanical failure inside the pump.

Check wear condition of impeller, its clearances and liquid passages

Excessive grease in ball bearings.

Check method of regreasing

Lack of lubrication for bearings.

Check hours run since last change of lubricant, the schedule and its basis

Improper installation of bearings

Check method of assembly, possible damage or state of cleanliness during assembly and type of bearing used

Damaged bearings due to contamination.

Check contamination source and replace damaged bearings

 



C. ELECTRICAL TROUBLES



Wrong direction of rotation.

Reverse 2 phases on motor terminal box

Motor running too slow,

Check motor terminal box connections

Page 33 of 44

LR, LRV, LLR and LR-S USER INSTRUCTIONS ENGLISH C953KH013 - 09/03 ®

8 PARTS LISTS AND DRAWINGS 8.1 Sectional drawings – LR single entry impeller, grease lubricated, gland packed (pump sizes 2.5LR-10 and 2.5LR-13 only) 2

4

5

7

8

9

10

11

14

16

17

18

21

22

29

28

30

23

24

25

26

27

’Z’

1

3

6

12

13

15

Taken from A336/036

8.1.1 Parts list – LR single entry impeller Ref. no. 1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16

Description Dust cover Bearing locknut Lockwasher – bearing locknut Shim (c) Ball bearing - thrust Thrust bearing spacer (c) Gland Lantern ring Gland packing Stuffing box bush Casing – upper half Casing – lower half Gasket – horizontal split Casing wear ring Anti-rotation screw (b) Impeller

17 Key - impeller 18 Seal pipe assembly 21 Socket head screw for 22 22 Impeller nut 23 Dowel bushing 24 Bearing bracket 25 Water shield 26 Bearing cover 27 Gasket – bearing cover 28 Ball bearing - line 29 Grease nipple 30 Shaft 31 Key - coupling a) When fitted, depends on type of mechanical seal. b) True position is on casing bottom half gasket face. c) Fitted on 21/2LR13 only.

Page 34 of 44

31

LR, LRV, LLR and LR-S USER INSTRUCTIONS ENGLISH C953KH013 - 09/03 ®

8.1.2 Options - LR single entry impeller



 

 View on ‘Z’ impeller ring

  

Oil lubrication





  Component mechanical seal

Cartridge mechanical seal

8.1.3 Options parts list - LR single entry impeller Ref. no.

Description

32 Breather 33 Constant level oiler 34 Mechanical seal 35 Mechanical seal cover 36 O-ring – mech. seal cover 37 Socket head screw for 38 (a) 38 Seal abutment ring (a) 39 Cartridge seal assembly 45 Impeller wear ring 46 Socket head screw for 45 a) When fitted, depends on type of mechanical seal.

Page 35 of 44

LR, LRV, LLR and LR-S USER INSTRUCTIONS ENGLISH C953KH013 - 09/03 ®

8.2 Sectional drawings – LR double entry impeller, grease lubricated, gland packed 

 )

( *

%'&

+,+

+.-

+./

+10

5

243

2

2 5

6 3









"!

"#

$

6

7 897













Taken from A336/035

8.2.1 Parts list – LR double entry impeller Ref. no. 1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17 18 19

Description Dust cover Bearing locknut Lockwasher – bearing locknut Shim (d) Ball bearing - thrust Thrust bearing spacer (c) Gland Lantern ring Gland packing Stuffing box bush Casing – upper half Casing – lower half Gasket – horizontal split Casing wear ring Anti-rotation screw (b) Impeller Key - impeller Seal pipe assembly O-ring – shaft sleeve

a) b) c) d)

20 Shaft sleeve 21 Socket head screw for 22 & 40 22 Shaft nut 23 Dowel bushing 24 Bearing bracket 25 Water shield 26 Bearing cover 27 Gasket – bearing cover 28 Ball bearing - line 29 Grease nipple 30 Shaft 31 Key - coupling When fitted, depends on type of mechanical seal. True position is on casing bottom half gasket face. Fitted on 4LR11, 4LR14 and 5LR13 only. Not fitted on 10LR 17 and 10LR18.

Page 36 of 44

LR, LRV, LLR and LR-S USER INSTRUCTIONS ENGLISH C953KH013 - 09/03 ®

8.2.2 Options - LR double entry impeller







 View on ‘Z’ impeller ring

:@A :@? Oil lubrication

:> :<;

DE FHG BC

:"= Component mechanical seal

Cartridge mechanical seal

8.2.3 Options parts list - LR double entry impeller Ref. no.

Description

32 Breather 33 Constant level oiler 34 Mechanical seal 35 Mechanical seal cover 36 O-ring – mech. seal cover 37 Socket head screw for 38 (a) 38 Seal abutment ring (a) 39 Cartridge seal assembly 40 Impeller nut 45 Impeller wear ring 46 Socket head screw for 45 a) When fitted, depends on type of mechanical seal.

Page 37 of 44

LR, LRV, LLR and LR-S USER INSTRUCTIONS ENGLISH C953KH013 - 09/03 ®

8.3 Sectional drawings – LLR grease lubricated, gland packed 8.3.1 All but 2LLR11, 3LLR11 and 4LLR11 pumps 



















$























=























Taken from A336/037

8.3.2 2LLR11, 3LLR11 and 4LLR11 pumps From first stage suction

From first stage suction

B A

First stage packing arrangement

Second stage packing arrangement

Seal piping arrangement A = Seal pipe - plan 11 first stage. B = Balance pipe first stage suction to second stage suction.

Page 38 of 44

LR, LRV, LLR and LR-S USER INSTRUCTIONS ENGLISH C953KH013 - 09/03 ®

8.3.3 Parts list – LLR Ref. no. 1 2 3 5 7 8 9 10 11 12 13 14 15 16 16A 17 18 19 20

21 Socket head screw for 22 and 40 22 Shaft nut 23 Dowel bushing 24 Bearing bracket 25 Water shield 26 Bearing cover 27 Gasket – bearing cover 28 Ball bearing - line 29 Grease nipple 30 Shaft 31 Key - coupling 48 Interstage bushing 49 Anti-rotation screw for 48 (b) 50 Spacer sleeve 51 Stuffing box bushing (c) 52 Socket head screw for 51(c) 53 Balance pipe assembly (c) a) When fitted, depends on type of mechanical seal. b) True position is on casing bottom half gasket face. c) When fitted.

Description Dust cover Bearing locknut Lockwasher – bearing locknut Ball bearing - thrust Gland Lantern ring Gland packing Stuffing box bush Casing – upper half Casing – lower half Gasket – horizontal split Casing wear ring Anti-rotation screw (b) Impeller – first stage Impeller – second stage Key - impeller Seal pipe assembly O-ring – shaft sleeve Shaft sleeve



8.3.4 Options - LLR





 View on ‘Z’ impeller ring

TVU

RS

From first stage suction (when fitted)

Oil lubrication

IKQ ]^

IMP ION

_`

From first stage suction (when fitted)

[\ YOZ

IKJ WX

IML Component mechanical seal

Cartridge mechanical seal

8.3.5 Options parts list - LLR Ref. no. 32 33 34 35 36

Description Breather Constant level oiler Mechanical seal Mechanical seal cover O-ring – mechanical seal cover

37 Socket head screw for 38 (a) 38 Seal abutment ring (a) 39 Cartridge seal assembly 40 Impeller nut 45 Impeller wear ring 46 Socket head screw for 45 a) When fitted, depends on type of mechanical seal.

Page 39 of 44

LR, LRV, LLR and LR-S USER INSTRUCTIONS ENGLISH C953KH013 - 09/03 ®

8.4 Sectional drawings – LR-S double entry impeller, grease lubricated, gland packed hi

j

g

i

m

bKd

bf

~'

non

npqnrsnt

yx

bc

aKa

be

u

z

z

v

wx

p

t

{x

{

r Taken from A336/038

8.4.1 Parts list – LR-S Ref. no. 5 7 8 9 10 11 12 13 14 15 16 17 18 19

k@l

Description Ball bearing - thrust Gland Lantern ring Gland packing Stuffing box bush Casing – upper half Casing – lower half Gasket – horizontal split Casing wear ring Anti-rotation roll pin Impeller Key - impeller Seal pipe assembly O-ring – shaft sleeve

20 21 22 23 24 25 26 27 28 30 31 48 49 50

Shaft sleeve Socket head screw for 22 & 40 Shaft nut Dowel Bearing bracket Water shield Bearing cover Gasket – bearing cover Ball bearing - line Shaft Key - coupling Circlip – thrust bearing Lip seal – inboard Lip seal - outboard

8.4.2 Options parts list – LR-S

2 Bearing locknut 3 Lockwasher – bearing locknut 33 Constant level oiler 34 Mechanical seal 35 Mechanical seal cover 36 O-ring – mech. seal cover 37 Socket head screw for 38 (a) 38 Seal abutment ring (a) 51 Double row thrust bearing 52 Shaft for 51 a) When fitted, depends on type of mechanical seal.

‚O…

€K

‚O‡

|

ˆ<ˆ

‚O†

}

‚„ ‚Kƒ Double row thrust bearing

Component mechanical seal

Page 40 of 44

Oil lubrication

LR, LRV, LLR and LR-S USER INSTRUCTIONS ENGLISH C953KH013 - 09/03 ®

8.5 Sectional drawings – LRV double entry impeller, grease lubricated thrust bearing, component mechanical seal, SiC bearing £ ¥ ¦

¥"¢ ¥

©4¤ ¡£

©¨

¡¤

©§ ©"¡

¡¥

¡¡

©¢

¡¢

¡ ¦

£

¡¨ ¡§ ¤

‰Š

•o˜ •o—



‘’ •–• “”

š ›œš  

Œ  Ÿ

  Ž ž

 ™ ‹

8.5.1 Parts list – LRV Ref. no. 1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16

Description Socket head capscrew Sleeve end cap Bottom bearing housing O-ring, bottom bearing housing Tolerance ring Bearing bush - SiC Retaining ring, bearing bush Socket head screw Bearing flush pipe assembly Shaft sleeve, bottom Impeller Key, impeller Casing - upper half Casing - lower half Gasket - horizontal split Casing wear ring

Taken from A336/033 17 18 19 20 21 22 23 24 25 26 27 28 29 30 31 32 33

Anti-rotation screw (b) Seal pipe assembly O-ring - top sleeve Shaft sleeve, top Abutment ring - mech. seal (a) Socket head screw for 21 (a) Mechanical seal Mechanical seal cover O-ring, mech. seal cover Dowel bushing Shaft nut Socket head screw for 27 & 43 Shaft V-ring, inboard Bearing cover, drive end Gasket, bearing cover Grease nipple

Page 41 of 44

34 Thrust bearing spacer (c) 35 Ball bearing, thrust 36 Shim (d) 37 Bearing locknut 38 Lockwasher, bearing locknut 39 Bearing bracket 40 V-ring, outboard 41 Key, coupling a) When fitted, depends on type of mechanical seal. b) True position is on casing bottom half gasket face. c) When fitted, depends on pump size. d) Not fitted on 10LR17 and 10LR18. The shim is inboard of the bearing on the 6LR18, 8LR20, 10LR14 and 10LR16.

LR, LRV, LLR and LR-S USER INSTRUCTIONS ENGLISH C953KH013 - 09/03 ®

8.6 Interchangeability chart for LR, LLR and LR-S 1

8.5.2 Options - LRV



Pump size

 View on ‘Z’ impeller ring



Cutlass rubber bearing bush

ª«

«<¬ ª ­

Cartridge mechanical seal

8.5.3 Options parts list - LRV Ref. no. 42 43 44 45 46

Description Cartridge seal assembly Impeller nut Bearing bush, cutless rubber Impeller ring Socket head screw

Impeller Shaft 2

Sleeve Bearing Casing assembly 3 assembly 4 wear ring

1.5LLR-7 A A A A A 1.5LLR-10 B 2LLR-9 C B 2LLR-11 D C B B B 3LLR-11 E D 4LLR-11 F E 2.5LR-10 cw G C C A 2.5LR-10 ccw H D None fitted 2.5LR-13 cw I E D F 2.5LR-13 ccw J F 3LR-9 K B G C 3LR-12 L F 4LR-10 M E 4LR-12 N H D 5LR-10 O G 6LR-10 P I H J5 4LR-11 Q F E D 4LR-14 R J 5LR-13 S G 5LR-15 T 5LR-19 U I 6LR-13 V H K F F 6LR-16 W 8LR-12 X J 8LR-14 Y 6LR-18 Z 8LR-20 AA L G G K 10LR-14 BB 10LR-16 CC 10LR-17 DD M H H L 10LR-18 EE 6LR-18S FF N M I I 10LR-14S GG N O 12LR-14S HH O 8LR-18S II N P J J 10LR-18S JJ O 8LR-23S KK P Notes: 1) All the above pump casings can be supplied for clockwise or counter clockwise rotation. The casings are interchangeable with each other although pump suction and discharge positions change – refer to the relevant Sectional arrangement drawing. The LLR has a range of clockwise and counter clockwise impellers, which are not interchangeable with each other. 2) The same shaft is used for gland packing and component mechanical seals. Cartridge mechanical seals use a different series of shafts fitted with impeller nuts – see relevant Sectional arrangement drawings. 3) Gland packing and inch-type component mechanical seals use the same shaft sleeve. Includes shaft sleeve, shaft nut, gland, packing, lantern ring, stuffing box bush and mechanical seal. When metric seals are fitted a different diameter shaft sleeve and nut is used. 4) Includes bearing bracket, bearing cover, ball bearing set, bearing locknut and washer (except 2.5LR-10 which uses the bearing locknut and washer from the 1.5LLR7 group). 5) Dimensionally the same as other J, but may have had a material upgrade to handle some of the higher power applications.

Page 42 of 44

LR, LRV, LLR and LR-S USER INSTRUCTIONS ENGLISH C953KH013 - 09/03 ®

8.7 Interchangeability chart for LRV 1 Pump size Impeller Shaft

2

Top sleeve Top bearing assembly 3 assembly 4

SiC bearing Bearing Bottom bush shaft assembly 5 sleeve

Cutlass rubber bearing Bottom bearing housing

Bearing bush

Bottom shaft sleeve

Bottom bearing housing

Casing wear ring

3LRV-9 K B A Q A 3LRV-12 L C E A A F 4LRV-10 M 4LRV-12 N R B B A A 5LRV-10 O G 4LRV-11 Q F S D F C B C B 4LRV-14 R 5LRV-13 S G 6LRV-10 P T C E B C B C H 5LRV-15 T G 6LRV-13 V H U E G D D B D D 6LRV-16 W 8LRV-12 X J 8LRV-14 Y B 6LRV-18 Z 8LRV-20 AA E V F H E E E K 10LRV-14 BB C 10LRV-16 CC 10LRV-17 DD W G I F F F F L 10LRV-18 EE 1) The LRV impeller, top sleeve assembly, and casing wear ring are also interchangeable with the equivalent LR pump. The LRV pump casing is not interchangeable with the LR. The LRV clockwise and counter clockwise casings are not interchangeable with each other. 2) For component mechanical seals. Cartridge mechanical seals use different shafts fitted with impeller nuts – see relevant Sectional arrangement drawings. 3) For inch-type component mechanical seals, includes sleeve, shaft nut and mechanical seal. When metric seals are fitted a different diameter shaft sleeve and nut is used. 4) Includes bearing bracket, bearing cover, ball bearing set, bearing locknut and washer. 5) Includes tolerance ring and bearing retaining ring.

8.8 General arrangement drawing

The typical general arrangement drawing and any specific drawings required by the contract will be sent to the Purchaser separately unless the contract specifically calls for these to be included into the User Instructions. If required, copies of other drawings sent separately to the Purchaser should be obtained from the Purchaser and retained with these User Instructions.

9 CERTIFICATION

Certificates determined from the Contract requirements are provided with these Instructions where applicable. Examples are certificates for CE marking, ATEX marking etc. If required, copies of other certificates sent separately to the Purchaser should be obtained from the Purchaser for retention with these User Instructions.

10 OTHER RELEVANT DOCUMENTATION AND MANUALS 10.1 Supplementary User Instructions

Supplementary instructions such as for a driver, instrumentation, controller, seals, sealant system etc are provided as separate documents in their original format. If further copies of these are required they should be obtained from the supplier for retention with these User Instructions.

10.2 Change notes

If any changes, agreed with Flowserve Pump Division, are made to the product after its supply, a record of the details should be maintained with these User Instructions.

10.3 Additional sources of information Reference 1: NPSH for Rotordynamic Pumps: a reference guide, Europump Guide No. 1, Europump & World Pumps, Elsevier Science, United Kingdom, 1999. Reference 2: th Pumping Manual, 9 edition, T.C. Dickenson, Elsevier Advanced Technology, United Kingdom, 1995. Reference 3: nd Pump Handbook, 2 edition, Igor J. Karassik et al, McGraw-Hill Inc., New York, 1993. Reference 4: ANSI/HI 1.1-1.5 Centrifugal Pumps - Nomenclature, Definitions, Application and Operation. Reference 5: ANSI B31.3 - Process Piping.

Page 43 of 44

LR, LRV, LLR and LR-S USER INSTRUCTIONS ENGLISH C953KH013 - 09/03 ®

FLOWSERVE REGIONAL SALES OFFICES: Europe, Middle East & Africa Flowserve Limited (Pump Division) Harley House, 94 Hare Lane Claygate, Esher, Surrey KT10 0RB United Kingdom

Latin America Flowserve S.A. de C.V. Avenida Paseo de la Reforma 30 2nd Floor, Colonia Juarez Centro Mexico, D.F.Z.C. 06040

Tel +44 (0)1372 463 700 Fax +44 (0)1372 460 190

Tel +52 5705 5526 Fax +52 5705 1125

USA and Canada Flowserve Corporation (Pump Division) Millennium Center, 222 Las Colinas Blvd. 15th Floor, Irving, TX 75039-5421, USA

Asia Pacific Flowserve Pte Ltd (Pump Division) 200 Pandan Loop, 06-03/04 Pantech 21, Singapore 128388

Tel +1 972 443 6500 Toll free 800 728 PUMP (7867) Fax +1 972 443 6800

Tel +65 775 3003 Fax +65 779 4607

Visit our web site at: www.flowserve.com

Your Flowserve factory contact:

Your local Flowserve representative:

To find your local Flowserve representative, please use the Sales Support Locator System found at www.flowserve.com

Page 44 of 44

PETRONAS PENAPISAN (T) SDN.BHD KERTEH REFINERY

HAND TOOLS WRENCHES 1)

Objective To teach how wrenches work and the techniques of using them

2)

Tools and Equipment Wrenches of various types as described below.

3)

General Information As in all technical professions and trades within the refinery, the beginner and the apprentice must first educate himself with the various tools he or she will meet, and must also learn their different shapes and uses. (4)

Safety

Obtain a cold work permit from process operator. Check with the process controller to see that no pressure exists on equipment that is to be work upon. Use the proper protective clothing and equipment where necessary. (A) OPEN JAW TYPE (Open ended type) Let us first of all consider an everyday tool commonly called the “spanner” or wrenches. This is, of course, used for the screwing up or the unscrewing of nuts and bolts etc. They are usually made of hard, tough steel in order to stand up to the duties which they are called upon to perform. The jaws are made so as to fit comfortably over the flat faces of the nut without any extreme slackness There are three chief types of open-jaw wrenches: (a) The single-ended (b) The double-ended (c) The spud or Prong-ended wrench These are shown in fig. 1 below:

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Fig.1 To use a wrench, it should be securely placed on the nut or bolt-head so that the jaws extend completely over the nut. Otherwise, the wrench will probably slip off when an attempt is made to handle it. The smaller jaw-arm should be placed in the direction of the hand-pull. The larger jaw-arm, having more metal in it, is thus better enabled to take care of the tension force which is exerted or to bring to bear steadily or forcefully while the short jaw-arm tends to be compressed. When a wrench is correctly used in this manner the jaws are less liable to become strained or opened out, and this considerably reduces the possibility of the wrench slipping off, thus avoiding the risks of bruising the operators’ hand (fig.2).

Fig.2

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The ‘Single-ended’ wrench is used when only one size of nut is to be released or tightened, and the ‘Double-ended’ type is used in cases where two different sizes of nuts or bolt-head are to be dealt with. The ‘Spud’ or ‘prong-ended’ or ‘podger’ wrench has an open-jaw at one end to fit a specific size of nut and a fairly long round (circular) stem or handle. The handle gradually tapers i.e. diminishes in diameter towards the opposite end from the jaw. This tapered stem in addition to being used as a handle may also be used for prising any two plates (fig.3). It is thus chiefly used for steel-plate work, tank work, aligning two flanges in piping work, in cases where any two bolt holes or rivet holes of plate do not particularly get together. This means that if any two plates which have been previously drilled with rows of holes are to be bolted or riveted together (one on top of the other) it is possible that on placing them into position one or more holes of one plate may be found slightly out of mark with the corresponding hole of the other plate. In such cases the handle end of the spud wrench may often be used as a drift to pries the plates slightly in order to make the two holes coincide. This procedure, however, is not allowed in boiler or high pressure vessel work, where to two rivets or bolt holes must be accurately drilled so as to coincide with each other.

Fig.3

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CAUTION: It is extremely dangerous for this single-ended and spud wrenches to be exploited with an extension bar or pipe to loosen seized bolts and nuts. They should only be used as they are and without any amplification of any kind. (B) BOX WRENCH In maintenance work, several other kinds of wrenches are used, the type selected depending chiefly on the accessibility of the bolt-head or nut. In the construction of various machines, it is sometimes essential to place bolts in positions which are difficult to access by ordinary single or double-ended wrenches. For such cases a special type commonly known as a box wrench is used. This type, instead of fitting the nut on two of its faces or sides only, fits on all its sides, and this permits of greater force being applied without the danger of the wrench opening out. It therefore has an advantage over the jaw type wrench. (C) RING SPANNER This is a special design of fixed-end wrench. Each end takes the form of a ‘ring’, and from this it derives its name. Although of comparatively light construction, it is made of a special type of very tough, hard steel alloy. It is therefore very strong. The wrench usually takes the form of a round or oval-shape bar having at each end an ‘annulus’ or ring. Combination of both the ring and open-ended of the same size is also available. The inner face of each ring has teeth on it which fit exactly over the corners of the nut. One or both ends of the wrench are frequently cranked or arched in order to cater for nuts or bolts which are in shallow recesses. Although the rings have teeth formation on them, they are engineered or arranged to fit snugly over the nut, so it will not damage the nut faces or sides. This type of wrench has therefore an advantage over the open-jaw type. In as much as it fits over the whole of a nut in a similar way to a ‘box’ or ‘socket’ type wrench. It is a modern type and is usually supplied in sets which cater for a wide range of nut of both the imperial and MYG/2001

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metric sizes. In addition to its use in general maintenance engineering trade, this wrench is also extensively used in the automotive industry.

Fig.5

(D) TUBULAR TYPE BOX WRENCH This is perhaps one of the commonest forms of box wrench also called the ‘spark-plug’ wrench. As. its name implies, it is of steel of Tubular section. Both ends of the spanner are, however, hexagonal shaped to fit The intended sizes of nuts. Near each end, immediately above the hexagon formation, a hole is drilled through the tube for accommodating the steel handle or 'Tommy bar’ as it is called. The hole at one end of the tube is formed at right angles to that at the other end, i.e. the holes are at ninety degrees to each other. To use this type of wrench, one end is placed over the nut and the tommy bar is inserted through the hole at the other end. It is sometimes preferable, however, to place the tommy bar through the tube at the end nearer to the nut on which it rests, if this can be done conveniently. The advantage is that none of the pressure applied to bar is then lost through misalignment of the wrench. It will, perhaps, be realised by the user that if the bar is inserted through the tube at its top end a certain amount of pressure may be lost by the ‘torque’ or tendency of the tube to twist slightly throughout its length. There may be instances, however, where it is found impossible to insert The bar through the lower end of the tube, in which case the hole at the top end must be used.

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Fig.6 Now, let it be assumed that the particular nut is situated in fairly deep rectangular recess of a machine. Let it also be assumed that the length of the recess will easily accommodate the tommy bar, also, that it will leave room for reasonable hand movement during the turning operation. The width of the recess may be only half of its length. This being the case, and the tubular box wrench having been placed with its lower end over the nut in such a manner that the lower holes are lengthwise or parallel with the ‘length’ of the recess, it is possible for a turn on the bar to be effected. It may be found that the turn obtained was only 25.4 mm or 50.8 mm (an inch or two), say through an arc of about 20o. In that case the hand on the bar will ‘foul’ to side of the recess, thus preventing a greater turning movement. The tommy bar can than be withdrawn and reinserted, but this time in the holes through the top end of the tube. As these holes are at right angle to those at the lower end, a further turn can thus be made. This procedure can be repeated by applying the bar first to one, and then the other end of the tube until the nut is tightened or secured. In certain cases, and in order to overcome this constant changing of the bar positions, a specially long tubular wrench may be used, by means of which the complete turns can be achieved from the top end position without withdrawing the bar. The tube should be of suitable strength or rigidity, however, to reduce the torque to a minimum or a certain pressure may be lost.

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(E) SOCKET WRENCH In several respects, this type is similar to the box wrench. It fits completely over the whole of the nut and it also has a separate handle. The socket head is made of one piece of tough steel and is recessed on its lower face to accommodate the nut. On the top of the socket a head or ‘lug’ projects. This is either square or hexagonal according to the form adopted for the handle or ‘key’. Socket wrench are supplied in sizes which cater for the whole range of conventional imperial and ISO metric bolts and nuts and all the socket heads, however, are made to fit one key.

Fig.7 This type of wrench is more rigid in its application than the box type but it can be used only in places where the nut is either above the face of the work or in a shallow recess which does not exceed the depth of the socket head. It does not damage the corners of the nut, so it has an advantage over the open-jaw type of wrenches in this respect. Some modern types of socket wrenches have ratchet mechanism attached to their handle through an arc of approximately 30o to 40o and then return it to its MYG/2001

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starting point. This is frequently more convenient than having to turn the handle through a complete circle or half-circle for each operation. Further, these modern type are so devised that it is possible to operate the ratchet mechanism in the opposite direction. This is accomplished by reversing the ratchet ‘catch’ which is operated by a small push knob causing the ratchet to work in the other direction. It is clearly a great advantage to be able to either tighten up or loosen up a nut without having to remove to remove the socket or handle. The principle of the ratchet operation is illustrated below. Other types of socket wrenches may have a ‘female’ recess instead of a ‘male’ lug at their top ends. In these cases the key or handle than has a lug for fitting into the socket recess. How does ratchet work Ratchet is a mechanical device that transmits intermittent rotary motion or permits a shaft to rotate in one direction but not in the opposite one. In the Figure below, the arm and the ratchet wheel are both swiveled or pivoted. The stem of the pawl can slide in the arm and is kept in its lowest position by a spring. If the arm oscillates through the specified angle, the pawl rotates the wheel intermittently in a counter-clockwise direction; if the arm rotates clockwise, the sloping side of the pawl rides over the teeth and has no turning effect on the wheel. If the pawl is rotated half a turn so that its sloping side is on the left, oscillation of the arm rotates the wheel in a clockwise direction mechanisms only. Reversing ratchets of the type described are used on socket wrench handles and are convenient for tightening or loosening bolts in positions where a complete revolution of a wrench handle is impossible. They are also used to obtain an intermittent feeding motion (workpiece movement) on machine-tool worktables; the ratchet wheel is attached to the screw that moves the table, and the arm is driven by a crank, the throw of which can be varied to change the required angle of attack. Although ratchets with pawls and toothed wheels are the most common, other types are used. In one such type, an oscillating member works through a one-way clutch to rotate a wheel intermittently.

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(F) SPECIAL DOUBLE-ENDED (DUMB-BELL) This is perhaps a special type of double-ended hexagonal box wrench that may be similar in concept to that of a 'Swiss-army knife’ and takes its name from its shape which resembles that of an athlete’s ‘dumb-bell’. The ends are somewhat spherical, but have their faces flattened. Each face is pierced to form a hexagonal recess of suitable depth to fit effectively over the nut. Each end-piece has four flattened faces with hexagonal gaps of various sizes. This wrench therefore caters for eight different sizes of nuts. Frequently, the two extreme ends of the wrench are also drilled thus providing for two additional nut sizes.

Fig.9 (G) ADJUSTABLE WRENCH This wrench has jaws, one of which is made intact with the handle, so is therefore ‘fixed’. The other jaw is attached to a movable slide and is adjustable. By this means several sizes of nuts and bolts can be attended. About midway along the handle or shaft at the end of which is the fix jaw, teeth are formed on the inner edge, attached to the movable jaw-slide is a small ‘worm-wheel’. This rotates on a small, steel spindle which is screwed at the tail end, and it is thus fixed horizontally into the slide-jaw frame. By turning the worm-wheel, the sliding jaw can be moved along the teeth of the shaft, as these teeth are ‘meshed’ or engaged with those of the worm-wheel. This system gives the adjustment which accommodates many sizes of nuts. The chief advantage claimed for this type of wrench is that it is capable of dealing with many sizes of bolts and nuts, thus eliminating the necessity of using several different sizes of ‘fixed-jaw’ wrenches. This is suitable, especially for maintenance or repair work of a machine where time is a vital factor. However, as the jaws are not of the rigid or fixed design, they are ‘most of the time’ liable to be strained or opened-out when in use. This may cause them to slip off a nut and also to damage the nut corners. It is therefore important to ensure that the sliding jaw is screwed up to the nut face as tightly as possible and that it is in compression before applying pressure or torque. MYG/2001

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When new, these wrenches are quite effective, but after prolonged use they are liable to give all sort of trouble. This may occur when wear has taken place on the worm-wheel teeth or the teeth of the ‘rack’ with which the former teeth engage. The best manufacturers pays special attention to this point and make these components fit as well as possible so as to avoid any undue slackness or ‘play’ in this machanism. However, in spite of the maker’s precautions, after prolonged use ‘wear and tear’ will result. These parts should be kept clean and occasionally oiled, for if small particles of grits are allowed to enter the worm-wheel mechanism they will quickly cause undue wear and is unsafe to use. The other form of adjustable wrench equivalents to that describe above where the lower jaws forms part of the handle is called a ‘King Dick' adjustable wrench.

Fig.10 Several other types of adjustable wrench are made. All operate on the principle of a movable or adjustable upper or lower jaw and the difference in design is usually in the mechanism causing the jaw to move. A very large type of adjustable wrench is sometimes referred to as a ‘monkey wrench’. Another with ‘serrated’ gripping jaws designed for tubular duty is called a ‘pipe wrench’.

General Use of Wrenches From these brief descriptions it will be appreciated that each of the types described above has its specific use or application. Although all wrenches are primarily intended for screwing or unscrewing nuts and bolts, they may occasionally have to be used for other purposes. It is fully realised that in the case of emergencies, such as a temporary breakdown of a machine, which might involve loss of production and valuable MYG/2001

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time, a certain amount of improvisation or modification often has to be done, and in the absence of the specific tool, others have to be used. It may happen that a particular size of nut has to be removed and that the correct size of wrench is not available. Furthermore, there may not be an adjustable wrench to hand, but there might be an open-jaw type of slightly larger size than the nut concerned. In such cases an ‘old trick’ of the trade is to use a small ‘packing’ between one side of the nut and one jaw of the wrench. If this is done the ‘packing-piece’ should be placed adjacent or touching to the small jaw-face. Small broken pieces of hacksaw blades can often be used to advantage for this purpose. Further, a small coin has occasionally been used as packing, especially in the case of a automotive breakdown. Sometimes, when using a wrench, especially the single-ended or double-ended open-jaw types, for securely tightening nuts, it is the practice to place a length of steel pipe over one end of the spanner to obtain a greater leverage for the hand-pull. This method is sometimes unavoidably used, especially for the tightening up of heavy foundation bolt-nuts to ensure that they are sufficiently tight to prevent the vibration of the particular machine loosening them. Great caution should be used, however, if such practice has to be responded to, as if undue pressure is exerted, the bolt may be broken by “tearing” or “twisting” across the threaded part. This method should never be approved with the smaller sized bolts, for the bolts may sometimes break when undue pressure is used on an ordinary wrench. However, the length of a wrench or spanner has been designed to provide sufficient “leverage” for a normal hand-pull or a fair average pressure is to be applied. This is why the length of double-ended wrenches varies in proportion to the size of their end-jaws. This point should be borne in mind when about to apply a length of pipe to the wrench with a view to increasing the leverage for the hand-pull. For this purpose, a special hydraulically operated ‘torque wrench’ is the correct tool. However, in these cases, it is strictly prohibited to use any types of adjustable wrenches for the procedures decsribed above.

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BARRIER AND BUFFER FLUID SELECTION By current terminology (API 682), a fluid between the two seals in a dual seal at higher pressure than process pressure is a barrier fluid. (A dual seal pressurized this way was formerly called a double seal.) A barrier fluid completely isolates the pumped process fluid from the environment. A buffer fluid is instead at lower pressure than pump process pressure. (Dual seals with buffer fluids were previously called tandem seals.) Most modern seals are double-balanced, so it can be pressurized either way. Neither barrier nor buffer fluids should be confused with a flush, which process medium is injected directly into the pumpage through the seal gland or the pump's seal chamber. As recommended by seal manufacturer, a barrier or buffer fluid should be: i)--compatible with the process (tolerable to contamination), ii)--compatible with the seal materials (does not attack or erodes seal faces), iii)--a good lubricant and heat transfer medium for the seal faces, iv)--kind and safe to the environment and the workers in the plant. Some good choices for barrier and buffer fluids: Water/Ethylene Glycol mixture: Almost as good as water for heat transfer; doesn't freeze in outdoor applications. 50/50 mix by volume is easiest to mix, and gives good freeze protection. Use corrosion-inhibited industrial grade. Water: Cheap, safe, available, excellent heat transfer characteristics (high specific heat, low viscosity, high thermal conductivity). Don’t use in freezing conditions. Water/Propylene Glycol mixture: Like water/ethylene glycol, but usable in food applications. Kerosene or Diesel fuel: Good where an oil is required. Low enough viscosity to flow well and transfer heat, and a good lubricant. Vapor pressure low enough that emissions aren't a problem. Light mineral and synthetic oils: Generally good. Within synthetics, polyalphaolefin (PAO) based fluids usually better than ester-based. Synthetic oils specifically formulated for use as mechanical seal barrier fluids are available, including grades accepted by Food and Drug Administrations. Some common, but bad, choices: Automotive antifreeze: Based on ethylene glycol, but contains additives to prevent corrosion of automotive engine components and to stop small radiator leaks. These additives cause excessive seal face wear. Uninhibited ethylene glycol: Without corrosion inhibitors, can attack seal parts, notably the nickel binder in tungsten carbide.

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Automatic transmission fluid (ATF): Contains additives to increase friction in the bands and clutches in automatic transmissions. These also increase wear and friction in seal faces. Silicone oils: Inert, but often form glassy particles that abrade and clog seals.

BARRIER AND BUFFER FLUID CONVECTION SYSTEMS Dual mechanical seals require that a barrier or buffer liquid be introduced between the seals for cooling and lubrication. The most economical and commonly used method to provide this liquid to the seal is a reservoir piped to the seal. The reservoir can be used in remote locations without elaborate piping systems. A dual seal with barrier fluid pressure greater than process pressure is often called a double seal. This mode of pressurization prevents leakage of product into the reservoir or to the environment. Reservoir pressure is usually maintained at 15-30 psi (1-2 Kg/Cm2) greater than process pressure. Bottled nitrogen or compressed air available in the plant are common sources of pressure. If process pressure exceeds buffer pressure, the seal is said to be operated as a tandem seal. The reservoir may be pressurized to split the pressure loading between the inboard and outboard seals or to force buffer liquid between the outboard seal faces for lubrication. Any product leakage is contained by the reservoir. A process fluid that is pumped as a liquid but exists as a gas at atmospheric conditions, such as a light hydrocarbon, may be sealed using a dual seal with an unpressurized buffer fluid. An immiscible buffer fluid such as diesel fuel is used. Any product that leaks through the inboard seal bubbles up through the buffer liquid in the reservoir, where it can be vented to a flare. Since most modern seals has a double-balanced inboard seal, it may be pressurized as either a double or tandem seal.

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Typical installation of dual seal with tangent ports for clockwise shaft rotation (viewed from coupling)

Connections for counterclockwise shaft rotation

Connections for seal with radial ports for any shaft rotation

INSTALLATION: Locate the reservoir not more than four feet (1.2 m) from the seal, with the bottom of the reservoir 12-18 inches (30-45 cm) above the centerline of the pump shaft. Mount the reservoir to a rigid support where the sightglass is easily visible for inspection and where the fill ports are accessible. Fill the reservoir to the center of the sightglass. This provides enough barrier liquid to allow for losses, while leaving headspace in the reservoir to allow for thermal expansion. Eliminate any air trapped in the seal or piping by loosening the fittings on the seal gland temporarily. MONITORING: Reservoir pressure variation caused by thermal expansion is normal. Rising liquid level in the reservoir indicates leakage of product past the inboard seal of a tandem seal. If the level drops slowly, the barrier liquid should be replenished. Rapidly dropping liquid level without visible leakage of the outboard seal or piping indicates inboard seal leakage into the product.

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BARRIER AND BUFFER FLUID CONVECTION SYSTEMS

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BARRIER AND BUFFER FLUID CONVECTION SYSTEMS Dual mechanical seals require that a barrier or buffer liquid be introduced between the seals for cooling and lubrication. The most economical and commonly used method to provide this liquid to the seal is a reservoir piped to the seal. The reservoir can be used in remote locations without elaborate piping systems. A dual seal with barrier fluid pressure greater than process pressure is often called a double seal. This mode of pressurization prevents leakage of product into the reservoir or to the environment. Reservoir pressure is usually maintained at 15-30 psi (1-2 bar) greater than process pressure. Bottled nitrogen or compressed air available in the plant are common sources of pressure. If process pressure exceeds buffer pressure, the seal is said to be operated as a tandem seal. The reservoir may be pressurized to split the pressure loading between the inboard and outboard seals or to force buffer liquid between the outboard seal faces for lubrication. Any product leakage is contained by the reservoir. A process fluid that is pumped as a liquid but exists as a gas at atmospheric conditions, such as a light hydrocarbon, may be sealed using a dual seal with an unpressurized buffer fluid. An immiscible buffer fluid such as diesel fuel is used. Any product that leaks through the inboard seal bubbles up through the buffer liquid in the reservoir, where it can be vented to a flare. Since the AST 80 has a double-balanced inboard seal, it may be pressurized as either a double or tandem seal.

Typical installation of AST 80 dual seal with tangent ports for clockwise shaft rotation (viewed from coupling)

Connections for counterclockwise shaft rotation

Connections for seal with radial ports for any shaft rotation

BARRIER AND BUFFER FLUID CONVECTION SYSTEMS

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INSTALLATION: Locate the reservoir not more than four feet (1.2 m) from the seal, with the bottom of the reservoir 1218 inches (30-45 cm) above the centerline of the pump shaft. Mount the reservoir to a rigid support where the sightglass is easily visible for inspection and where the fill ports are accessible. Fill the reservoir to the center of the sightglass. This provides enough barrier liquid to allow for losses, while leaving headspace in the reservoir to allow for thermal expansion. Eliminate any air trapped in the seal or piping by loosening the fittings on the seal gland temporarily. MONITORING: Reservoir pressure variation caused by thermal expansion is normal. Rising liquid level in the reservoir indicates leakage of product past the inboard seal of a tandem seal. If the level drops slowly, the barrier liquid should be replenished. Rapidly dropping liquid level without visible leakage of the outboard seal or piping indicates inboard seal leakage into the product. Copyright © 1998 Advanced Sealing Technology Join Our Affiliate Program | Terms & Conditions Home | Services | Online Store | EPM In The News | FAQs | Contact Us | E-mail

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